Experimental characterization of a transcritical CO2 direct expansion ground source heat pump for heating applications

Journal Pre-proof

Experimental Characterization of a Transcritical CO2 Direct Expansion Ground Source Heat Pump for Heating Applications

Arash Bastani Conceptualization;Methodology;Formal analysis;Investigation;Resource;Writing – Original Draft;Writ Parham Eslami-Nejad Conceptualization;Methodology;Formal analysis;Investigation;Resource;Writing – Original D Messaoud Badache Conceptualization;Methodology;Investigation;Resource;Writing – Original Draft , Alain Tuan Anh Nguyen Conceptualization;Writing – Original Draft PII: DOI: Reference:

S0378-7788(19)32630-1 https://doi.org/10.1016/j.enbuild.2020.109828 ENB 109828

To appear in:

Energy & Buildings

Received date: Revised date: Accepted date:

28 August 2019 23 January 2020 30 January 2020

Please cite this article as: Arash Bastani Conceptualization;Methodology;Formal analysis;Investigation;Resource;W Parham Eslami-Nejad Conceptualization;Methodology;Formal analysis;Investigation;Resource;Writing – Original D Messaoud Badache Conceptualization;Methodology;Investigation;Resource;Writing – Original Draft , Alain Tuan Anh Nguyen Conceptualization;Writing – Original Draft , Experimental Characterization of a Transcritical CO2 Direct Expansion Ground Source Heat Pump for Heating Applications, Energy & Buildings (2020), doi: https://doi.org/10.1016/j.enbuild.2020.109828

This is a PDF file of an article that has undergone enhancements after acceptance, such as the addition of a cover page and metadata, and formatting for readability, but it is not yet the definitive version of record. This version will undergo additional copyediting, typesetting and review before it is published in its final form, but we are providing this version to give early visibility of the article. Please note that, during the production process, errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain. © 2020 Published by Elsevier B.V.

Highlight   

DX-CO2 GSHP has a better performance in applications with lower temperature demand There would be an optimum number of borehole providing the highest heating capacity GHE heat extraction rate increases in lower temperature demand

1

Experimental Characterization of a Transcritical CO2 Direct Expansion Ground Source Heat Pump for Heating Applications Authors name Arash Bastani*1, Parham Eslami-Nejad1, Messaoud Badache1, Alain Tuan Anh Nguyen1 1

CanmetENERGY-Varennes, Natural Resources Canada (NRCan), Canada

*[email protected]

Corresponding Author: Arash Bastani

Abstract: Space heating and domestic hot water account for a large portion of energy consumption in buildings. Direct expansion ground source heat pump (DX-GSHP) is a promising renewable energy technology to provide heat efficiently in a wide range of applications in buildings. Moreover, there is a renewed interest on CO2, which is a viable option to replace synthetic refrigerants with an environmentally friendly alternative, yet limited studies have examined the performance of a CO2 DX-GSHP for different heating applications in buildings. A transcritical DX-CO2 GSHP had been built, and a number of tests were carried out to evaluate the performance of such a system and the associated ground heat exchanger (GHE) for three heating applications in buildings. Those applications represent low to high water temperature demand. DX-CO2 GSHP demonstrated a better performance and a higher heating capacity in applications with lower temperature demand. The highest COP was around 4 when DX_GSHP provides heat to radiant floor heating systems. Moreover, the performance of the system was examined with different number of active boreholes. The results implied that there could be an optimum number of boreholes to provide the highest heating capacity.

Keywords Direct Expansion Ground Source Heat Pump; CO2; Heating Applications; Renewable Energy Resources

2

Nomenclature Symbol A Cp DP k P Qbor Qh q T U W ρ Subscript bor c dis in out w

Area [m2] Specific Heat Capacity at Constant Pressure [kJ/kg/K] Pressure Drop [kPa] Conductivity [W/m/K] Pressure [bar] Borehole Heat Extraction Rate [kW] Gas Cooler Heat Capacity [kW] Flowrate [l/s] Temperature [°C] Overall Heat Transfer Coefficient [kW/K] Work [kW] Density [kg/m3]

Abbreviation COP DHW DX-GSHP LPEV GHE GC GWP HCSH HP HPCV IHE RFSH RV

Coefficient of Performance Domestic Hot Water Direct Expansion Ground Source Heat Pump Low-Pressure Expansion Valve Ground Heat Exchanger Gas Cooler Global Warming Potential Heating Coil Space Heating Heat Pump High-Pressure Control Valve Internal Heat Exchanger Radiant Floor Space Heating Regulating Valve

Borehole Compressor Compressor discharge Inlet Outlet Water

1. Introduction: The building sector in cold climates is one of the largest energy end-use sectors, accounting for a large proportion of the total energy consumption in many developed countries. In Canada, energy consumption for space heating represents 64% and 56% of total energy used in residential and commercial buildings, respectively [1]. This high-energy consumption has created the need to integrate energy efficient technologies. Amongst today’s available renewable energy technologies, vaporcompression heat pumps are considered one of the most efficient technologies for space heating and can be used in a wide range of applications. Generally, the efficiency of heat pumps (HPs) depends on the temperature lift between load and source. The load temperature of a heating system varies depending on the heat exchange equipment and the application. For instance, compact panel or cast iron radiators usually require temperatures between 55°C to 75°C. In systems with a large surface heat emitter, such as in hydronic radiant-floor, the load temperature is often less than 35°C. Lowering the load temperature can significantly increase the performance of HPs. Ploskic [2] showed that a HP efficiency could increase by 25 % when the load temperature reduces from 55°C to 40°C. On the other hand, the source temperature of a HP depends on the environment in which heat is extracted. Ground source heat pumps (GSHPs) perform better than conventional air-source HPs due to higher and more stable temperatures. GSHP systems extract heat from the ground via a ground heat exchanger (GHE), through which a heat carrier fluid circulates. Among the various GSHP systems, the conventional vertical secondary-loop GSHP 3

system has attracted the most interest in both research and engineering, owing to its simplicity and reliability. Although a considerable number of studies have been carried out to investigate the development and applications of the conventional GSHP systems during the past few decades, these systems still struggle to penetrate the market due to their high capital cost and long pay back periods. Recently, systems with a direct connection between the HP and the soil have gained more attention. These systems, known as direct-expansion ground source heat pumps (DX-GSHPs), consist of a GHE in which the refrigerant evaporates (fully or partially) and extracts heat from the ground. A typical DXGSHP connected to single U-tube GHEs is illustrated in Figure 1-a. In contrast with SL-GSHP, circulating two-phase refrigerant and using the latent heat energy requires lower heat transfer surface area in GHEs. Therefore, shorter boreholes and eventually lower initial cost are expected to provide the same capacity as SL-GSHP. Earlier studies on DX-GSHP were conducted between 1940 and 1990 ([3], [4], [5], [6] and [7]). In 1991, Mei and Baxter [8] derived a guideline for the design of the DX systems. According to Ndiaye [9], since early 2000s, some researchers have carried out in-depth investigation on the performance of DX-GSHP for heating applications in buildings. They highlight the significant energy savings of such a system in comparison with other competent heating systems. For instance, Johnson [10] tested two DX-GSHPs, operated with R-22 refrigerant, installed in residential buildings in Knoxville, Tennessee. The research indicated 30% to 50% energy saving, comparing with air-to-air HPs. Minea [11] experimentally studied the performance of an R-410A-DX-GSHP, connected with a radiant floor and a forced air heating system in a cold climate region (Shawinigan, Canada). With an average monthly COP of 3.07, they could achieve 30% energy savings compared to a natural gas heating system. Some researchers investigated different types of the GHE as an integral component of a DX-GSHP. Goulburn and Fearon [5] studied the performance of a DX-GSHP with both horizontal and vertical GHEs in heating mode. The installation operated with condensation temperatures ranging between 40°C and 78°C. While the heat extraction rate per meter of the horizontal GHE was higher, the vertical installation presented a better performance in terms of the calculated COP. An inclined 30 m GHE was investigated by Beauchamps [12] in heating and cooling mode. The COP of the HP was around 4.7 for 10h continuous operation in heating mode. Lenarduzzi and Bennett [13] studied a 2.8 ton system with a spiral copper GHE operating at 7°C evaporating temperature and 54°C condensing temperature in a Canadian climate. The system’s COP varied between 2.5 and 2.8. Various refrigerants have been investigated in the literature as a working fluid in DX-GSHP systems. Wang et al. [14] evaluated the application of R134a as a refrigerant in a DX-GSHP test bench with three, 30 m long, vertical boreholes. Their system provided heat around 50°C with an average COP of 2.28. They noted mal-distribution of the refrigerant in the boreholes as a drawback. Wang et al. [15] employed R-22 as the refrigerant in their DX-GSHP test bench, which provided heating energy with a COP of 2.88. The system was equipped with four 20 m vertical boreholes and a copper coil system to facilitate oil return. Yang [16] reported a detailed energy analysis of a DX-GSHP system using R-22 as the working fluid for both heating and cooling applications. Fannou et al. [17] experimented the application of R-22 in a DX-GSHP test bench. Their test bench continuously operated for 30 days and provided heat with average COP of 2.87. Considering a GHE as an integral component of a DX-GSHP, the selected refrigerant affects the size of a GHE and the overall performance of a HP. Selection of the operating fluid comes with some considerations in addition to the primary performance requirements such as: safety issues, chemical 4

compatibility with the components’ material, environmental protection, cost and availability. Meeting all these requirements is a challenge for DX-GSHP technology. A promising refrigerant to address these criteria is CO2.It is a natural refrigerant, non-toxic, and non-flammable with a very low global warming potential (GWP=1). In addition to its environmental benefits, CO2 favorable thermophysical properties help to operate a system with significantly lower pressure ratio, and relatively lower pressure drop in pipes and heat exchangers. Moreover, having low critical temperature (31.1°C), a transcritical CO2 system rejects heat in a wider range of temperature suitable for different heating applications in buildings. Badache et al. [18], comprehensively summarized the advantages of employing CO2 in DXGSHP. Only few researches have investigated the technical aspects of using CO2 as the refrigerant in DX-GSHP. Eslami-Nejad et al. [19] were among the first who constructed and utilized a transcritical CO2 DX-GSHP test bench. They employed the measurements to validate their numerical model. Furthermore, they used the model to characterize the influential parameters on the performance of a CO2 DX-GSHP [20]. Another research work in their group [21] experimentally investigated CO2 evaporation in the GHE and the overall performance of the system for different opening of the expansion valves and different number of boreholes. Nguyen and Eslami-Nejad [22] numerically studied variable speed compression of a CO2 DX-GSHP for space heating and cooling of a residential building in Canadian climate. The DX-GSHP was simulated in a ventilation system environment, producing 36°C in heating and 16°C in cooling. Badache et al. [23] conducted an experimental study to characterize the effect of CO2 mass flowrate on the performance of a DX GHE. They also carried out a theoretical analysis to investigate the performance of a GHE with different inlet and outlet vapor qualities. Their analysis presented that with an identical heat extraction rate, both increasing the inlet vapour quality and decreasing the outlet vapour quality affect the mass flowrate inside a GHE. Recently, Nguyen et al. [24] investigated the effect of an internal heat exchanger (IHE) on the heating performance of a CO2 DX-GSHP. Results highlighted that adding an IHE to the system improves its efficiency by mitigating thermal short-circuiting between downward and upward legs of boreholes and marginalizing the effect of ground surface temperature variations. According to this literature survey, one notices that there is a lack of study to characterize the operation of a CO2 DX-GSHP for different heating applications in buildings. The performance of a CO2 DX-GSHP is influenced based on the application and the demand temperature. Therefore, as the main contribution of this research, a number of experiments were designed to emulate three types of heating applications in buildings and the performance of a CO2 DX-GSHP was characterized for these applications: Domestic Hot Water (DHW), Heating Coil Space Heating (HCSH), and Radiant Floor Space Heating (RFSH). These applications require water in relatively high to low range of temperature. The experiments were carried out using the test apparatus presented in [18] and [21] at CanmetENERGY research laboratory located in Varennes, Canada.

2. System description: Figure 1-a schematically presents the single-stage transcritical CO2 DX-GSHP test bench (shown in Figure 1-b), working only in heating mode, with the hot gas bypass. The main components of this test apparatus include a semi hermetic compressor (3.52 kW (one ton) nominal refrigeration capacity) (1-2), a gas cooler (GC) (2-3), a receiver (5-6) and boreholes (7-8) as the GHE (evaporator). While the ground works as the heat source, a water loop is connected to the cold side of the GC to emulate the load side. 5

As shown in Figure 1, the test apparatus is fully equipped with different measuring devices including pressure sensors, temperature sensors and flow meters. The system operates in three pressure levels: high, intermediate, and low. The thermodynamic cycle of the test bench is presented in Figure 2. CO2 as the refrigerant is compressed to the high-pressure level (supercritical pressure) at point 2 with a corresponding temperature rise. Then, the high pressure/high temperature gas enters the gas cooler to heat the water. After the internal heat exchanger (IHE) (point 4), CO2 is expanded through the highpressure control valve (HPCV) to the intermediate-pressure level of the cycle (point 5). Two-phase CO2 enters the separator; the vapor portion is bypassed around the boreholes, while the liquid portion flows toward the boreholes. Then, liquid CO2 is expanded through the low-pressure expansion valves (LPEV) and enters the boreholes with low vapor quality (point 7). The exhaust of the boreholes mixes with the bypassed vapor, gains superheat in IHE and enters the compressor to complete the cycle.

a) b) Figure 1: a) Test setup schematic, b) Test bench unit The heat pump is also equipped with an oil management system consisting of a separator and a receiver, with a pressure regulating valve to ensure adequate lubrication of the compressor. Sufficient insulation was installed all around the piping and the components to minimize the heat loss.

6

Figure 2: Thermodynamic cycle with hot gas bypass Two counter-flow plate heat exchangers were used for the GC (AGC=0.74 m2 and UGC=0.3 kW/K at test conditions) and for the IHE (AIHE=0.09 m2 and UIHE=0.1 kW/K at test conditions). The cold side of the GC is connected to a water loop. It supplies water at different temperatures and mass flowrates. The GHE consists of four 30-meter vertical boreholes with single copper U-tube arranged in a square pattern with, uniform spacing of 6.25 m. Due to CO2 evaporation, the upward pipe is filled more with gas while the downward pipe mainly contains liquid. As a result of a difference between liquid and gas density, the pipe sectional area in the liquid side needs to be reduced to maintain fluid’s velocity. Maintaining CO2 velocity avoids any possible chance of oil trapping in the bottom of the U-pipe. One of the boreholes was equipped with temperature sensors in different levels to determine the temperature profile of CO2 in both two legs of the U-pipe. Table 1 presents borehole dimensions, compressor characteristics, and in-situ thermal properties of the soil obtained from a thermal response test. The test bench has the flexibility to work with all or some of those boreholes based on the test requirements. The system can be operated in two different modes; automatic and manual. In the manual mode, the opening ratio of the expansion valves located before the boreholes are regulated manually as presented in [21]. On the other hand, in automatic mode, the operation of the system is automatically controlled using the following four commands: 1. The valve that discharges the vapor from the receiver maintains the intermediate pressure (Preceiver) at a certain value set by the operator. It should always be higher than the evaporating pressure in the GHE. 2. The HPCV installed after the IHE controls the discharge pressure (high-pressure level). The valve operation is controlled following an equation giving the optimum pressure as a function of GC outlet temperature. The optimum pressure is expected to provide the maximum COP at the operating conditions. 3. LPEVs’ opening changes automatically to be able to provide superheat temperature setpoint either at the exit of the boreholes or at the suction of the compressor depending on the assigned thermocouple. 4. Two regulating valves (RV) installed at the low-pressure level before the IHE, which can be set to 7

a certain opening to recover some heat from the high-pressure line. In case the LPEVs are controlled to deliver a given superheat at the exit of the boreholes, the RVs are controlled automatically to satisfy the second superheat setpoint at the compressor suction by modulating the flow of CO2 passing through the IHE. In this study, to avoid the drastic fluctuation of the LPEVs’ opening and deliver a reliable measurement, the RVs’ opening were set manually at a constant percentage and the LPEVs were controlled using the temperature and pressure readings at the suction of the compressor. Table 1: System specification

3. Test cases: 3.1 Designed tests: This study characterizes the performance of CO2 DX-GSHP water heater for three different heating systems in buildings: RFSH, HCSH, and DHW. Each system requires a certain range of water outlet temperature provided by the GC described later in “Application” section. To produce experimentally the desired temperature range, 15 tests were designed using different configurations of water inlet flowrate ( ̇ ) and water inlet temperature ( ). The assigned magnitude to the mentioned parameters are as follows:  

Water flowrate ( ̇ ) (0.05, 0.15, 0.25 l/s) Water temperatures (25, 30, 35, 40, 45°C)

Fifteen different experiments were conducted, using all the configuration of the aforementioned parameters as the water inlet conditions to the GC. In those experiments, all four boreholes were operational, the superheat at the suction of the compressor was set at 20°C, and the Preceiver (intermediate pressure) was regulated at 3930 kPa (corresponding to 4.61°C saturated temperature). 8

The HPCV regulates the discharge pressure using a pre-established correlation as a function of CO2 GC outlet temperature ( ). This correlation is presented in “Discharge Pressure Control” section, later in this manuscript. The correlation is intended to provide the optimum discharge pressure rewarding the highest COP in any operating condition. To evaluate how optimal is this correlation performs, two extra tests were conducted at and ̇ with three active boreholes, in which the discharge pressure was manually set. The COP of the HP in those manually adjusted discharge pressure was compared with the one provided by the pre-established correlation. Moreover, three extra tests with three, two and one operating boreholes were conducted to quantify the effect on the performance of GHE and the HP. Overall, twenty tests were conducted with the conditions tabulated in Table 2. Table 2: Tests conditions Test Scenario 1 2 3

Test number 1-15 16-18 19-20

[°C] 25-45 35 35

̇ [l/s] 0.05-0.25 0.15 0.15

Number of boreholes 4 3,2,1 3

Discharge Pressure [kPa] Pre-established correlation Pre-established correlation 7800, 8900

3.2 Measurement: As presented in Figure 1, the test apparatus has been equipped to measure the temperature, pressure, and mass flowrate of CO2 and water along the process circuit, in addition to the electrical power of the compressor. Moreover, one of the boreholes has been instrumented with 24 T-type thermocouples installed on the U-pipe surface measuring pipe surface temperature representing CO2 temperature profile along both legs of the GHE. The uncertainty of the employed instruments is presented in Table 3. Data were recorded at 30 s intervals using data acquisition system. Table 3: Uncertainty of the instrumentation Instrument T-type Thermocouples Pressure Transmitter CO2 Mass Flow Meter Water Mass Flow Meter Wattmeter

Uncertainty ±0.5°C ±0.6% ±0.2% ±0.5% ±0.2%

To characterize the performance of the HP, the following parameters are calculated using the measured temperature, pressure and flowrate: heating coefficient of performance (COP), GC heating capacity ( ), GHE heat extraction rate ( ), and GHE pressure drop (DP). These parameters are calculated as follow: (1) ̇

(

)

(2) (3) (4)

where is the measured work of the compressor [kW], is the water specific heat capacity [kJ/kg/K], is the water density [kg/m3], is the water outlet temperature from the gas cooler

9

[°C], is CO2 pressure at the inlet of the boreholes [bar], and the boreholes.

is CO2 pressure at the exit of

Having single-phase flow (water) in the cold side of the GC, is determined using the measured data from the water-side. The average difference between the calculated heat transfer rate on the CO2-side and Water-side was 1.83% which conveys negligible heat loss in the GC. The heating capacity of the GC has an uncertainty of ±3.26%, and the uncertainty of the determined COP was estimated to be ±3.27%.

3.3 Data collection procedures and test duration: Each test was started after undisturbed ground temperature had been observed as the initial condition of the GHE. In fact, fully recovered ground was confirmed at the beginning of each test by premonitoring the temperature profile of the borehole. The water temperature and flowrate, number of active boreholes, and the controlled discharge pressure were set to a certain desired magnitude according to the test conditions presented in Table 2. Moreover, the opening of the RVs were adjusted manually to bypass 15% of the low-pressure flow to the suction of the compressor. As mentioned earlier, with one single superheat setpoint and no modulating RVs, the superheat was only controlled with the LPEVs upstream of the boreholes. All the tests were carried out in one-month time span with a comparable ground temperature profile. Operation of the test bench continued until the measured superheat and discharge pressure presented stable reading for 30 minutes. The average of the readings during the final 30 minutes was selected to represent the state conditions of the system. The results of this study project the short-term performance of CO2 DX-GSHP for heating applications. With the relatively high thermal mass of the ground, almost constant ground temperature and consequently evaporation temperature are expected in short-term operation of the system. In long-term operation, however, the performance of heat pumps is impaired due to the ground temperature-drop. It is worth highlighting that a DX-GSHP functioning in automatic mode imposes the following limitations in the operation of experiments:  It is impossible to execute a test with a certain adjusted evaporation temperature.  There is no direct procedure to set a certain CO2 flowrate in the boreholes.

3.4 Discharge pressure control: In a GC, the pressure and temperature are decoupled due to the transcritical operating region. The slope of the CO2 isotherms is quite modest for specific pressure ranges close to the critical pressure, where it is quite steep at higher pressure. This unique behavioral pattern of CO2 results in the existence of an optimal discharge pressure ( ) of a transcritical CO2 refrigeration system where the system reaches the highest COP under certain operating conditions. Many investigations were devoted to identify the operating parameters that significantly influence the optimal . Moreover, the effort was made to derive reliable correlation to determine the optimal pressure. Some of those correlations, which were noticeably cited in literature, are presented in Figure 3 and described here. Sarkar et al. [25] identified optimal as a function of inlet temperature of the fluid on the cold side of a GC (Pdis_Sarkar in Figure 3). Chen and Gu [26] concluded that an IHE with high effectiveness is very crucial for a transcritical system to achieve high COP and that evaporating temperature has little influence on the optimal . They developed the correlation of optimal , either as a function of the ambient temperature (Pdis_Chen_air in Figure 3) or the CO2 gas cooler outlet temperature, TCO2,out (Pdis_Chen_gc in Figure 3). 10

On the contrary, for a constant inlet water temperature in an air-source transcritical CO2 heat pump, Wang et al. [27] experimentally showed that the optimal depends on both evaporating temperature and . Meanwhile, at a fixed water inlet temperature, the optimal increases in higher ambient temperature and higher outlet water temperature. Qi et al. [28] found that at different ambient temperatures, the optimal varies with in a transcritical HP water heater (Pdis_Qi in Figure 3). Their experimental studies show that the optimal COP reduces substantially with the increase of in the range of 25°C to 45°C.

Figure 3: Discharge pressure control correlations

Figure 4: Pre-established discharge pressure control

Figure 3 presents the correlations mentioned earlier in this section. In addition, the optimal of the test apparatus in the current study is illustrated in the same figure (square symbols) at different . Employing the results of the experiments, the optimal pressure correlation was regressed verses and displayed in Figure 4. According to Figure 3, the optimal in the present tested system is different, in similar operating conditions, from the specified optimal pressure provided by the correlations in the literature. To the knowledge of the authors, the HPCV is not small enough for the relatively low CO 2 flowrate of this system. Thus, at higher GC inlet water temperature, the expected high discharge pressure cannot be achieved due to employing an inappropriate size of the valve. One practical challenge in building of a CO2 transcritical HP system is to find an appropriate size of equipment available in the market due to the relatively lower mass flowrate of CO2 in comparison with the other refrigerant for a similar capacity. While, there are no studies reporting the optimal pressure for CO2 DX-GSHP; one objective of this research is to investigate the deviation of the current implemented correlation to provide the optimal COP.

4. Applications:

11

Temperature demand in heating applications depends on particular aspects such as the type of heating system, the efficiency of the system, and the regional regulations. In building heating applications, water temperature demand may vary between 25°C to 75°C [29]. In this study, three applications were considered to be addressed using CO2 DX-GSHP: RFSH, HCSH, and DHW.  RFSH: A heating system that maintains the indoor condition by circulating water under the floor. The system has a relatively low demanded temperature ranging between 25°C to 37°C.  HCSH: A common heating system providing heated air from the supply outlet located on the inner envelope of buildings. In heating mode, the system requires a heating coil containing water at a temperature ranging between 37°C to 50°C.  DHW: DHW accounts for one of the largest end-use energy consuming system in buildings. The setpoint of the tank may range between 50°C to 65°C. The aim of this study is to characterize the performance of a CO2 DX-GSHP in the abovementioned applications in order to provide water in those temperature range.

5. Results: 5.1 Characterizing heating performance of CO2 DX-GSHP: Experimental results characterize the operation of the system providing hot water for three applications in buildings. Based on the demand, CO2 DX-GSHP produced hot water for DHW, HCSH, and RFSH in the following temperature range; over 50°C, 37°C to 50°C, and 25°C to 37°C. In order to provide water with the required temperature for these applications, both inlet water temperature and flowrate to the GC were changed. Figure 5 presents the COP of the system associated to different water flowrates and inlet water temperatures. The size of the circles presents the magnitude of COP at that certain conditions. Moreover, the figure visualizes three heating applications based on the required temperature and categorizes the test cases appropriate for those applications. As expected, lower flowrate results in a higher outlet temperature due to the lower water mass that needs to be heated. Globally, the results express that the production of hot water for higher temperature demand reduces the performance of the system. In fact, either increasing water inlet temperature or reducing water flowrate in the GC, negatively influences the performance of CO2 DX-GSHP. The conducted experiments clearly characterize the response of the system to the increase of inlet water temperature in the GC as follow: 

Lower CO2 GC outlet temperature is inevitable with lower water inlet temperature; hence, the HPCV imposes lower discharge pressure in accordance with the optimal discharge pressure control in Figure 4. The change in specific heat capacity of CO2 is drastic around 30°C and close to its critical pressure, thus higher heating capacity of the GC is inevitable (displayed in Figure 6a)



Increasing inlet water temperature is associated with more work of the compressor due to the higher optimal discharge pressure. On the other hand, reducing water inlet temperature to the 12

GC prompts higher liquid CO2 mass flowrate into the boreholes, which consequently enhances the total borehole heat-extraction rate. Meanwhile, higher liquid CO2 in the GHE requires lower evaporation pressure to satisfy the superheat setpoint. However, the measured work of the compressor is lower in comparison with the high inlet water temperature as presented in Figure 6-b (the size of the circles in this figure presents the magnitude of the work). The abovementioned phenomena and observations confirm a higher performance of transcritical CO2 DX-GSHP for the applications with lower demand temperature. It is worth mentioning that these heating demand are addressed with a significantly higher effectiveness than the conventional systems such as electrical or gas boilers, which have a COP lower than one.

Figure 5: System coefficient of performance (COP) Water flowrate has its own influence on the heating capacity of the GC and the discharge pressure, particularly in lower inlet temperature. Reducing water flowrate to the GC causes a drop in the overall heat transfer coefficient, and the heating capacity decreases. As a result, increases; hence, the compressor’s work increases to discharge CO2 at a higher pressure (Figure 6-b). Taking these phenomena into account, the performance of transcritical CO2 DX-GSHP degrades in applications with lower water inlet flowrates.

a) b) Figure 6: a) System heating capacity [kW] b) Work of compressor [kW]

13

5.2 Characterizing CO2 evaporation in the GHE a. Effect of gas cooler water inlet temperature: As mentioned earlier, one of the boreholes was equipped with 24 T-type thermocouples (12 at each leg). Figure 7 illustrates the CO2 temperature profile for the first test scenario with 0.15 l/s water inlet flowrate and different inlet temperatures. Each column depicts the temperature profile of both legs at different inlet water temperatures. It also describes heat extraction rate and CO2 mass flowrate per borehole. This measurement delivers the following outcomes: 











Two regions can be distinguished in each profile: two-phase CO2 and single-phase CO2. The former region presents the evaporation of CO2 in the GHE at almost constant temperature, indicating low pressure-drop in the flow. The latter one, with escalating temperature, represents the sensible heat extraction of CO2. The location on the GHE separating those two regions is an indicator to evaluate the effectiveness of the GHE. The highest efficiency of GHE happens when the two-phase region occurs in the entire borehole. As mentioned earlier, CO2 liquid flowrate reduces in boreholes when water inlet temperature increases. In the test carried out with 25°C water temperature, CO2 flowrate increased by 85% compared to the scenario with 45°C water temperature. In a fixed receiver pressure, at lower water temperature, the outlet of the GC is expanded to a lower vapor quality; hence, more liquid flows to boreholes. Higher mass flowrate to the borehole enhances the extraction rate in GHE. The case with 25°C water temperature presented a 141% higher heat extraction rate than the case with 45°C water temperature. This case also presents the longest two-phase region in the borehole. In DX-GSHP, evaporation temperature is under the influence of the inlet water temperature. Draining higher CO2 liquid from the receiver demands lower evaporation temperature to satisfy superheat setpoint. Limited control on the evaporation temperature and the mass flowrate in a GHE is a barrier in DX-GSHP. This shortcoming could diminish the effectiveness of the GHE in scenarios similar to the test with 40°C inlet water temperature. Despite satisfying the superheat set point at the suction of the compressor, complete evaporation occurred only in one leg of the borehole. Theoretically, higher mass flowrate to the boreholes could be a plausible option for this scenario with an enhanced heat extraction rate. In addition, appropriate sizing of the LPEVs provides a closer evaporation temperature to the soil temperature. Setting an appropriate superheat setpoint regarding the application and the required temperature would remedy such a drawback. CO2 temperature reduces at the exit of the upward leg due to the lower shallow ground temperature and the interaction with the downward leg.

14

Figure 7: CO2 temperature profile for tests with different gas cooler water inlet temperature

b. Effect of the number of active boreholes in a GHE: Following the discussion in the previous section, one concludes that the mass flowrate in a GHE cannot be adjusted independently in DX-GSHP. This uncontrolled characteristic is a concomitant of having GHE as an integral part of the GSHP. Consequently, there would be an optimum number of boreholes or an optimum borehole configuration associated with the design criteria of CO2 DX-GSHP. Accordingly, additional experiments were conducted with different number of active boreholes with similar water inlet temperature and flowrate ( ̇ ). Figure 8 presents the temperature profile in the borehole, the heat extraction rate, and the CO2 mass flowrate per borehole for four experiments conducted with different number of boreholes. Limited by the size of compressor, the mass flowrate per borehole increases for the scenarios with fewer number of boreholes. To satisfy a certain superheat, the system is expected to operate at lower evaporation temperatures when CO2 flowrate increases. Meanwhile, heat transfer surface area reduces with lower number of active boreholes. According to Figure 8, similar phenomenon was observed in this study, except for the scenario with three active boreholes. The scenarios with three and four boreholes operated almost in an equal evaporation temperature. Despite the same evaporation temperature, CO2 evaporation was completed almost in the middle of the second leg of the U-pipe in the case with four boreholes, unlike the scenario with three boreholes where the CO2 flowrate per borehole is higher. As a result of lower mass flowrate, increasing the number of boreholes reduces the extraction rate per borehole. The case with one active borehole presented the highest extraction rate, with 300% more than the lowest case with four boreholes. It is worth mentioning that the extracted heat from the singlephase portion of the GHE in the test with four active boreholes accounts for 2% of the total extracted heat, although it covers almost one-quarter of the total GHE length. This quantification highlights the 15

importance of avoiding the sensible heat extraction in boreholes and a lower superheat in the outlet of the DX-GHE. Proper system control (receiver pressure and the superheat at the exit of a GHE) and correct size of the expansion valves would reduce the length of the sensible heat extraction portion in boreholes and improve the effectiveness of a GHE. Temperature (C)

0

0

3

-3

6

0

DOWN UP

20

4 boreholes 0.59 kW heat extraction per borehole 2.8 g/s mass flow rate per borehole

15

3 boreholes 0.81 kw heat extraction per borehole 4.1 g/s mass flow rate per borehole

1 borehole 1.75 kW heat extraction per borehole 10 g/s mass flow rate per borehole

10

Depth (m)

2 boreholes 1.04 kW kW heat extraction per borehole 4.9 g/s mass flow rate per borehole

5

3

5

10

15

20

25

25

30

6 0

Depth (m)

-3

2% of Heat exctraction

Temperature (C)

-3

0

3

6

-3

0

3

6

30

Temperature (C) Temperature (C) Figure 8: CO2 temperature profile for tests with different number of boreholes

Figure 9 presents the CO2 pressure drop and the total heat extraction rate of the GHE with different number of boreholes. As expected, the pressure drop reduces in scenarios with more active boreholes due to the lower CO2 mass flowrate. In this study, a trivial pressure drop of 10 kPa was observed in the GHE with four boreholes. Eslami-Nejad et al. [20] concluded that even some pressure gain would be observed in cases with low flow in boreholes. According to Figure 9, heat extraction rate of the GHE ( ) was enhanced by increasing the number of boreholes from one to three; yet, it was reduced from three to four boreholes. The higher is attributed to the higher heat transfer surface area. On the other hand, the case with four boreholes presented lower than the case with three boreholes, despite the largest heat transfer’s surface area. Considering the comparable superheat setpoint for both scenarios, latent heat extraction all along the borehole at an almost constant temperature (Figure 8) resulted in a more effective GHE and higher extraction rate in the case with three active boreholes. In addition, the COP and the GC heating capacity were calculated for different number of active boreholes and they are presented in Figure 10. Increasing the number of boreholes enhances heating capacity and COP, until the efficiency of a GHE is not impaired due to the reduced CO2 mass flowrate. Sufficient numbers of borehole in a CO2 DX-GSHP not only provides the highest heating capacity but also avoids the cost of building unnecessary wells.

16

120

2.2

90

2

60 1.8 30 1.6 0

0

1

2

3

4

5

Number of active boreholes

Figure 9: Borehole heat transfer rate and pressure drop for tests with different number of boreholes

Qh COP

3.8 3.6

2.8 3.4 2.7

3.2

2.6

3 2.8

2.5

Gas cooler heat transfer rate (kW)

2.4

4

2.9

Coefficient of performance

Borehole pressure drop (kPa)

150

Borehole total heat transfer rate (kW)

3

2.6

Q_borehole DP

180

2.6 2.4

0

1

2

3

4

5

Number of active boreholes

Figure 10: System COP and heating capacity for tests with different number of boreholes

5.3 Effect of discharge pressure: As described earlier, transcritical CO2 DX-GSHP would operate with an optimal discharge pressure wherein the system achieves the highest COP in certain operating conditions. Appropriate discharge pressure would significantly improve heating capacity of a HP and the performance of the system. In order to quantify the impact of this parameter, the test bench was operated with three different discharge pressures with similar inlet water temperature and flowrate to the GC. The water inlet was set at 35°C and 0.15 l/s, which could emulate all the heating applications mentioned in this study according to Figure 5. The change of the discharge pressure theoretically alters both the heating capacity and the performance of a CO2 DX-GSHP. Manually, two different discharge pressures (7.8 and 8.9 MPa) were set on the system working with three boreholes, as alternatives to the already pre-implemented pressure (8.3 MPa), pre-set on the test apparatus. The automatically set pressure is considered as the reference to make the comparison. Figure 11 illustrates the measured heating capacity and COP of the system with those discharge pressures. The reduced pressure impaired COP and the heating capacity of the system by 26% and 32%, respectively. Alternatively, the heating capacity of the GC increased by 15% in the higher discharge pressure in comparison with the reference case. Despite the higher work of the compressor, COP of the system improved by 8% due to the higher delivered heat in the GC.

17

4.5 4.2

%15 %8

3

3.9 3.6

2.8

%32

2.6

3.3 3

%26

Coefficient of performance

COP 3.2

2.7

2.4

2.4

2.2

Gas cooler heat transfer rate (kW)

Qh

3.4

2.1 2 7500

7800

8100

8400

8700

9000

Discharge pressure (kPa)

Figure 11: System COP and heating capacity for tests with different discharge pressure The superior heating capacity can be described by consulting the specific heat and enthalpy profiles of CO2 at the operated discharge pressure, presented in Figure 12. The peaks in the profiles of the specific heat illustrate the transcritical area in CO2 phase diagram where the isotherms are plateau. In the reference case, CO2 at the outlet of the GC had almost its maximum specific heat at that certain pressure, while the specific heat did not even reach the peak of the profile in the case with the lowest discharge pressure. Therefore, the lowest change in the enthalpy was observed and the lowest heating capacity was calculated in that certain discharge pressure. On the other hand, the largest change of the enthalpy in the GC was observed in the test conducted at 8.9 MPa discharge pressure, wherein its maximum specific heat occurred inside the GC. Hence, the higher portion of CO2 heating capacity is transferred to the water in that certain discharge pressure unlike the other tested pressures. CO2 gas cooler outlet temperature (C) 40

60

80

100

120

80 500 70 Cp (left-bottom axes) @ 7800kPa h (right-top axis) @ 7800kPa Cp (left-bottom axes) @ 8300kPa h (right-top axes) @ 8300kPa Cp (left-bottom axes) @ 8900kPa h (right-top axes) @ 8900kPa Tgc_out Gas cooler inlet

60 50 40

450

400

3 boreholes water flow rate=0.15L/s

350

30

Enthalpy (kJ/kg)

CO2 spesific heat capacity (kJ/kg/K)

20 90

20 300 10 0 20

25

30

35

40

45

250 50

CO2 gas cooler outlet temperature (C)

Figure 12: Transcritical CO2 specific heat and enthalpy profiles

18

6. Conclusion Space heating and DHW account for a large portion of end-use energy consumption in buildings and compel the employment of energy efficient technologies. Recently, DX-GSHP has gained the attention as a relatively cost-efficient alternative to the conventional secondary-loop GSHP. Build with a GHE as an integral part of the HP, selecting an appropriate refrigerant is a challenge in design and construction of a DX-GSHP. CO2 is a viable option to replace synthetic refrigerants with an environmentally friendly alternative, yet limited studies have examined the performance of a CO2 DX-GSHP for different heating applications in buildings. This lack of study encouraged the authors to characterize the operation of such a system for three heating applications: DHW, HCSH, and RFSH which represent high, medium and low temperature demand in buildings, respectively. Fifteen experiments were conducted with different inlet water conditions to the GC to produce hot water in temperature ranges required for the mentioned heating applications. The study characterizes CO2 DX-GSHPs as follow:  The system has a better performance and larger heating capacity in applications with lower temperature demand.  Reducing water inlet temperature to the GC induces higher CO2 mass flowrate to the GHE and increases heat extraction rate from the ground.  Reducing water inlet temperature causes lower evaporation pressure in the GHE. On the other hand, increasing water inlet temperature requires higher discharge pressure. The latter case resulted in a higher work of the compressor, in addition to the reduced GC heating capacity.  The system in this study presented COP ranging between 2 to 4, which depends on its application. In comparison with other DX-GSHPs using different refrigerant ([3], [14], [15] and [17]), CO2 presented higher COP in lower demand temperature and comparable COP in higher demand temperature.  Limited control on the evaporation temperature and the mass flowrate in a GHE is a barrier in DX-GSHP. Having an appropriate system control is crucial for such a system. Moreover, the performance of the system was examined with different number of active boreholes. The results imply that:  A lower number of active boreholes associates with higher mass flowrate per borehole and lower evaporation pressure.  Increasing the number of boreholes enhances heating capacity and COP of a system, until the efficiency of a GHE is not impaired due to the reduced CO2 mass flowrate. There would be an optimum number of borehole which provides the highest heating capacity.  Although the test bench has not been equipped with a finely tuned controlling system and correctly sized expansion valves, the boreholes heat extraction rate was around 50 W/m in some tests which is higher than a SL-GSHP in a comparable operating conditions. One specification, which singles out in CO2 transcritical HP, is the optimum discharge pressure, where the maximum COP is achieved. In this study, one operating condition was repeated with three different discharge pressure to quantify the impact of this parameter on the performance of the system. By examining the specific heat profile at each discharge pressure, the case in which the peak of the profile occurred inside the GC presented the highest heating capacity and COP of the system. Regarding the test bench, the current implemented control function does not provide the optimum discharge pressure. Nevertheless, the HPCV is not small enough for the relatively low CO2 flowrate of this system, 19

which needs to be replaced with an appropriately sized valve to achieve the maximum discharge pressure. From the practical design point-of-view, the outcomes of this study emphasize on having appropriate control system to provide optimum discharge pressure and correct sizing of a GC to provide the highest capacity based on the application and temperature demand. Moreover, the cost of the system can be reduced using the optimum number of boreholes based on heating demands and heating applications.

CRediT Author Statement Arash Bastani: Conceptualization, Methodology, Formal analysis, Investigation, Resource, Writing – Original Draft, Writing – Review & Editing, Visualization. Parham Eslami-Nejad: Conceptualization, Methodology, Formal analysis, Investigation, Resource, Writing – Original Draft, Supervision, Project administration, Visualization. Messaoud Badache: Conceptualization, Methodology, Investigation, Resource, Writing – Original Draft. Alain Tuan Anh Nguyen: Conceptualization, Writing – Original Draft.

Author Declaration We wish to confirm that there are no known conflicts of interest associated with this publication and there as been no significant financial support for this work that could have influenced its outcome.

Acknowledgment The authors would like to acknowledge the funding received by the Office of Energy Research and Development (OERD) of Canada through the ecoENERGY Innovation Initiative and Energy Innovation Program for test bench construction and supporting the research.

20

References [1] [2] [3] [4]

[5] [6] [7] [8] [9]

[10] [11]

[12] [13] [14]

[15]

[16] [17]

[18] [19]

NRCan, Paving the Road to 2030 and Beyond: Market transformation road map for energy efficient equipment in the building sector, (2018). A. Ploskid, Technical solutions for low-temperature heat emission in buildings, KTH, Stockholm, (2013). C.H. Coogan, Heat Transfer Rates Experimental Determination for Heating Absorbing Coils Buried in the Earth, AS ME Transactions. (1949). E.A. Freund, G.S. Whitlow, Earth source heat pumps: Characteristics, design, and operation, Transactions of the American Institute of Electrical Engineers, Part II: Applications and Industry. 77 (1959) 540–551. doi:10.1109/TAI.1959.6371497. J.R. Goulburn and J. Fearon, Deep ground coil evaporators for heat pumps, Applied Energy. (1978). J. Fearon, R.J. Goulburn, Domestic Heat Pump with Deep Hole Ground Source Evaporator, Applied Energy. 14 (1983) 99–113. G.S. Smith, Intermittent ground grids for heat pumps, University of Washington, (1956). V.C. Mei, V.D. Baxter, Experimental analysis of direct-expansion ground-coupled heat pump systems, (1991). D. Ndiaye, Reliability and performance of direct-expansion ground-coupled heat pump systems : Issues and possible solutions, Renewable and Sustainable Energy Reviews. 66 (2016) 802–814. doi:10.1016/j.rser.2016.08.049. W.S. Johnson, Field tests of two residential direct exchange geothermal heat pumps/Discussion, ASHRAE Transactions. 108 (2002) 99. V. Minea, Combined radiant floor and forced air heating with direct expansion GSHP in a northern greenhouse, in: Proceedings of the 21st International Congress of Refrigeration, Washington, DC, USA, (2003): pp. 17–22. B. Beauchamps, Modélisation et validation expérimentale d’une pompe à chaleur, École de technologie superieur, (2011). http://espace.etsmtl.ca/932/. F.J. Lenarduzzi, A direct-expansion ground-source heat pump with spiral ground coil-heating mode, in: ASHRAE Transactions: Symposia, (1991): pp. 902–908. X. Wang, C. Ma, Y. Lu, An experimental study of a direct expansion ground-coupled heat pump system in heating mode, International Journal of Energy Research. 33 (2009) 1367–1383. doi:10.1002/er.1541. H. Wang, Q. Zhao, J. Wu, B. Yang, Z. Chen, Experimental investigation on the operation performance of a direct expansion ground source heat pump system for space heating, Energy & Buildings. 61 (2013) 349–355. doi:10.1016/j.enbuild.2013.02.042. W. Yang, Experimental performance analysis of a direct-expansion ground source heat pump in Xiangtan, China, Energy. 59 (2013) 334–339. doi:10.1016/j.energy.2013.07.036. J.L. Fannou, C. Rousseau, L. Lamarche, K. Stanislaw, Experimental analysis of a direct expansion geothermal heat pump in heating mode, Energy & Buildings. 75 (2014) 290–300. doi:10.1016/j.enbuild.2014.02.026. M. Badache, P. Eslami-nejad, A. Bastani, Z. Aidoun, M. Ouzzane, Carbon dioxide evaporation process in direct expansion geothermal boreholes, in: IGSHPA Research Track, Stockholm, (2018). P. Eslami-nejad, M. Badache, M. Ouzzane, Z. Aidoun, Direct expansion ground source heat pump using carbon dioxide as refrigerant : Test facility and and theoretical model presentation, in: IGSHPA Technical/Research Conference and Expo, Denver, (2017). doi:10.22488/okstate.17.000526. 21

[20]

[21]

[22]

[23]

[24]

[25]

[26] [27]

[28]

[29]

P. Eslami-nejad, M. Badache, A. Bastani, Z. Aidoun, Detailed Theoretical Characterization of a transcritical CO2 direct expansion ground source heat pump water heater, Energies. 11, 387 (2018). doi:10.3390/en11020387. M. Badache, M. Ouzzane, P. Eslami-nejad, Z. Aidoun, Experimental study of a carbon dioxide direct-expansion ground source heat pump ( CO2 -DX-GSHP ), Applied Thermal Engineering. 130 (2018) 1480–1488. doi:10.1016/j.applthermaleng.2017.10.159. A. NGuyen, P. Elsami-Nejad, A transient coupled model of a variable speed transcritical CO2 direct expansion ground source heat pump for space heating and cooling, Renewable Energy. 140 (2019) 1012–1021. doi:10.1016/j.renene.2019.03.110. M. Badache, P. Eslami-nejad, A. Bastani, Z. Aidoun, A.Nguyen, Theoretical and experimental analysis of a vertical direct expansion geothermal evaporator using CO2 as refrigerant Theoretical and experimental analysis of a vertical direct expansion geothermal evaporator using CO2 as refrigerant, Science and Technology for the Built Environment. 25 (2019) 1081-1094. doi:10.1080/23744731.2019.1648131. A. Nguyen, P. Eslami-nejad, M. Badache, A. Bastani, Energy & Buildings Influence of an internal heat exchanger on the operation of a CO2 direct expansion ground source heat pump, Energy & Buildings. 202 (2019) 109343. doi:10.1016/j.enbuild.2019.109343. J. Sarkar, S. Bhattacharyya, M.R. Gopal, Simulation of a transcritical CO2 heat pump cycle for simultaneous cooling and heating applications, 29 (2006) 735–743. doi:10.1016/j.ijrefrig.2005.12.006. Y. Chen, J. Gu, The optimum high pressure for CO2 transcritical refrigeration systems with internal heat exchangers, 28 (2005) 1238–1249. doi:10.1016/j.ijrefrig.2005.08.009. S. Wang, H. Tuo, F. Cao, Z. Xing, Experimental investigation on air-source transcritical CO2 heat pump water heater system at a fixed water inlet temperature, International Journal of Refrigeration. 36 (2012) 701–716. doi:10.1016/j.ijrefrig.2012.10.011. P. Qi, Y. He, X. Wang, X. Meng, Experimental investigation of the optimal heat rejection pressure for a transcritical CO2 heat pump water heater, Applied Thermal Engineering. 56 (2013) 120–125. doi:10.1016/j.applthermaleng.2013.03.045. HVAC Systems and Equipment, American Society Of Heating, Refrigeration, And Air Conditioning Engineers (ASHRAE Handbook), Atlanta, GA, 2008.

22