Monitoring of a heat pump system using deep borehole heat exchangers

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Master of Science Thesis

KTH School of Industrial Engineering and Management Energy Technology TRITA-ITM-EX 2018-753 Division of Applied Thermodynamics and Refrigeration

SE-100 44 STOCKHOLM

Monitoring of a heat pump system using deep borehole heat exchangers

Baptiste Delattre

Supervisor at KTH: Willem Mazzotti Supervisor at INSA: Marc Clausse

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Master of Science Thesis TRITA-ITM-EX 2018-753

Monitoring of a heat pump system using deep borehole heat exchangers

Baptiste Delattre

Approved Examiner

Pr. Björn Palm

Supervisor

Willem Mazzotti

Commissioner Contact person

Abstract

Sweden has been one of the first countries in the world to use ground-source heat pumps (GSHP) to supply heating and cooling to its buildings. Today, it is the leading country in Europe and new installations tend to have larger capacities. The use of deeper borehole heat exchangers (BHE) is an opportunity to extract larger amounts of heat over small land areas. However, there are only few case studies on BHE deeper than 300 m, hence such systems may be optimized.

This study focuses on a deep GSHP system recently installed in central Stockholm, composed of four 510 m deep BHE. The objectives were to get the system ready for monitoring and to analyze the first sets of data recorded. First, a review of all the sensors already installed, of the data needed, and of the different ways to extract it, has been led. Practically, an acquisition system has been set up and connected to new and existing sensors such as thermometers and flow-meters. Theoretically, a method to derive the thermodynamic cycles of the different heat pumps has been determined. It led to the determination of COPs for several days during late spring 2017. The system globally showed reasonable efficiency, with an overall performance factor (equivalent to SPF2) of 3.42 including the circulation pumps of the ground loop.

However, it could certainly be improved in several ways, for example by avoiding short cycles or by finding an optimum flow in the secondary ground loop. Furthermore, these results should be juxtaposed with those that will be obtained during winter, when the heating demand will be the highest.

Sammanfattning

Sverige var ett av de första länderna i världen som använt bergvärmepumpar (GSHP) för att täcka värme- och kylbehoven i byggnader. Sverige är, idag, det ledande landet i Europa och nya bergvärmeanläggningar tenderar att vara större, åtminstone kapacitetsmässigt. Användningen av djupa borrhålsvärmeväxlare (BHEs) ger möjligheten att extrahera en större mängd värme i areabegränsade egendomar. Det finns dock bara få studiefall om anläggningar med borrhål djupare än 300 m och de anläggningarna skulle därför kunna optimeras.

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Den här studien fokuserar på ett GSHP system med djupa borrhål som nyligen installerades i centrala Stockholm och som består av fyra 510 m djupa borrhål, bland annat. Målen var att förberedda prestandauppföljningssystemet och analysera de första insamlade mätningarna. Första steget var att samla information om de sensorerna som redan var installerade, bestämma vilka mätvärde var nödvändiga och hur skulle de kunna mättas. Praktiskt har en datainsamlingsenhet iordningställts och anslutits till befintliga och nya sensorer såsom temperaturgivare och flödesmätare. Teoretiskt har en metod frambringats för att bestämma den termodynamiska cykeln av varje värmepump. Det möjliggjorde beräkningen av COP:n under vissa dagar under våren 2017. Globalt visade systemet rimlig prestanda med en prestandafaktor (likvärdig SPF2) på 3.42, inklusive cirkulationspumpar i bergvärmeskretsen. Det skulle dock kunna förbättras på olika sätt, t.ex. genom att undvika kort-cykling av kompressorer eller gnom att hitta ett optimalt köldbärarflöde i bergvärmeskretsen. Resultaten som det här arbetet kom fram till borde dessutom jämföras med en liknande analys under vintertid, då värmebehovet är högst.

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Acknowledgements

This internship has been a significant and stimulating experience of my university program. First, I would like to give a special thanks to my supervisor at KTH, Willem Mazzotti, for his support, good mood and guidance during this year.

I enjoyed working in the good atmosphere of the lab thanks to, among others, the helpfulness and the kindness of José, Alberto, Mohammad, Monika and Patricia.

I am also grateful towards the sponsors which made this project possible, in particular the condominium Ingemar 5 and the heat pump installer Wermer, for their help and the precious information they provided me.

Moreover, I would like to thanks my supervisor at INSA, Marc Clausse.

Finally, I thank all the other students with whom I worked in the lab for their warm welcome and their friendship.

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Nomenclature

Acronyms:

GSHP Ground-source heat pump BHE Borehole heat exchanger COP Coefficient of performance

SPF Seasonal performance factor DHW Domestic hot water

HP Heat pump (1 and 2)

M Heat pump’s module (1 to 4)

CPb Heat pump’s circulation pump for cold side secondary loop (1 and 2) CPb3 Cold side secondary loop additional circulation pump

CPh Module’s circulation pump for hot side secondary loop (1 to 4)

Variables:

𝑄𝑄₁̇ Heating power delivered by the heat pump at the condenser (W) 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡̇ Power consumed by the compressor of the heat pump (W)

𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑘𝑘̇ Power transferred to the refrigerant by the compressor of the heat pump (W) 𝑇𝑇_𝑎𝑎^′′ Temperature of the vaporous refrigerant exiting the evaporator (°C)

𝑇𝑇_𝑏𝑏 Compressor inlet temperature (°C) 𝑇𝑇𝑐𝑐 Compressor outlet temperature (°C) 𝑇𝑇_𝑐𝑐^′′ Evaporator inlet temperature (°C)

𝑇𝑇𝑑𝑑 Condenser outlet temperature (°C)

𝑃𝑃_{ℎ𝑖𝑖𝑖𝑖ℎ} Condensing pressure (bar)

𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ,𝑐𝑐𝑖𝑖𝑚𝑚 Condensing pressure when the refrigerant is not subcooled (bar)

𝑃𝑃𝑙𝑙𝑐𝑐𝑙𝑙 Evaporating pressure (bar)

ℎ_𝑎𝑎 Expansion valve outlet enthalpy (kJ/kg) ℎ_𝑏𝑏 Compressor inlet enthalpy (kJ/kg) ℎ_𝑐𝑐 Compressor outlet enthalpy (kJ/kg) ℎ𝑑𝑑 Condenser outlet enthalpy (kJ/kg) 𝑠𝑠_𝑏𝑏 Compressor inlet entropy (kJ/kg·K)

𝑇𝑇_{𝑏𝑏,𝑖𝑖𝑚𝑚} Evaporator inlet, cold side secondary fluid temperature (°C) 𝑇𝑇_{𝑏𝑏,𝑐𝑐𝑜𝑜𝑡𝑡} Evaporator outlet, cold side secondary fluid temperature (°C)

𝑇𝑇ℎ,𝑖𝑖𝑚𝑚 Condenser inlet, hot side secondary fluid temperature (°C) 𝑇𝑇_{ℎ,𝑐𝑐𝑜𝑜𝑡𝑡} Condenser outlet, hot side secondary fluid temperature (°C)

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𝐸𝐸_ℎ𝑠𝑠̇ Heating power delivered to the space heating system (W) 𝑉𝑉ℎ𝑠𝑠̇ Secondary loop flow in the space heating system (m³/h) 𝐸𝐸_{𝐷𝐷𝐷𝐷𝐷𝐷}̇ Heating power delivered to the DHW (W)

𝑉𝑉_{𝐷𝐷𝐷𝐷𝐷𝐷}̇ Secondary loop flow in DHW system (m³/h) 𝐸𝐸𝐵𝐵𝐷𝐷𝐵𝐵̇ Power extracted from the BHEs (W)

𝑉𝑉_𝑏𝑏̇ Cold side secondary loop flow (m³/h)

𝑇𝑇_𝑥𝑥 BHEs outlet temperatures (𝑥𝑥 varying from 1 to 4) (°C)

𝐸𝐸_{𝐶𝐶𝐶𝐶}̇ Power consumed by the circulation pumps (CPb1 to CPb3) (W) 𝑉𝑉𝐶𝐶𝐶𝐶̇ Flow through the circulation pump (CPb1 and CPb2) (m³/h)

𝑉𝑉_{𝑏𝑏,𝑐𝑐}̇ Secondary loop flow through the evaporator of the module (𝑚𝑚 varying from 1 to 4) (m³/h) 𝑃𝑃𝑎𝑎𝑎𝑎 Available pressure in the circulation pumps (kPa)

𝜌𝜌_𝑏𝑏 Ground loop secondary fluid density (kg/m³)

𝐶𝐶𝐶𝐶_𝑏𝑏 Ground loop secondary fluid volumetric heat capacity (kJ/kgK) 𝜂𝜂𝑐𝑐 Mechanical and electrical efficiency of the compressors (-)

𝑇𝑇_{𝑐𝑐𝑜𝑜𝑡𝑡} Outdoor temperature (°C)

𝑇𝑇₀ Undisturbed ground temperature (°C) 𝑧𝑧 Ground depth (m)

𝑞𝑞 Geothermal heat flux (W/m²) 𝑘𝑘 Ground conductivity (W/m·K)

∆𝐶𝐶 Borehole pressure drop (bar)

∆𝑉𝑉 Corresponding voltage of the borehole pressure drop provided by the transducer (V) 𝑢𝑢∆𝑐𝑐 Uncertainty over the borehole pressure drop (bar)

𝑢𝑢_{∆𝑐𝑐𝑝𝑝} Uncertainty over the corresponding voltage of the borehole pressure drop (V)

𝑢𝑢_𝑥𝑥 Uncertainty over the 1^st coefficient of the linear regression (V^-1) 𝑢𝑢_𝑦𝑦 Uncertainty over the 2^nd coefficient of the linear regression (bar)

𝑈𝑈 Expanded uncertainty over the borehole pressure drop (bar)

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Table of Figures

Figure 1. Ground-source heat pump installations in the world ranked on a per-capita basis (Rees, 2016).... 9

Figure 2. Ground-source heat pump cycle (after Lacombe, 2016)...10

Figure 3. Log(p)-h diagram of a simple refrigeration cycle (after Granryd et al., 2009) ...11

Figure 4. T-s diagram of a simple refrigeration cycle (after Granryd et al., 2009) ...11

Figure 5-a and Figure 5-b. Single U-tube (left) and coaxial tube (right) in a vertical BHE (Rees, 2016) ...12

Figure 6. Seasonal performance factor scope according to SEPEMO definitions (Rees, 2016) ...13

Figure 7. Measured ground temperature in a 270 m deep borehole at Norra Frescati, Stockholm (Gehlin et al., 2016) ...14

Figure 8. Location of the condominium “Ingemar 5” (after SGUs Kartvisare) ...15

Figure 10. Simplified hydraulic flowchart ...16

Figure 11. Undisturbed temperature profile derived from directional measurements (Mazzotti, 2016) ...17

Figure 12. Detailed hydraulic flowchart with data measurement points ...19

Figure 13. Data acquisition system overview (after Agilent 34970A User’s Guide) ...20

Figure 14. Thermometers connected to the 4 boreholes exit ...20

Figure 15-a and Figure 15-b. Thermal paste (left) and silicone (right) on a borehole thermometer ...21

Figure 16-a and Figure 16-b. Differential pressure transducer (left) and connection to a pipe (right) ...21

Figure 17. CPb3 energy-meter ...22

Figure 18. Acquisition card module connected to a 4-20 mA signal (yellow cables) ...23

Figure 19. Part of the dataflow program realized ...23

Figure 20. Example of system monitoring ...24

Figure 21. Temperatures in the condenser for refrigerant (blue) and water (black) (from Lundqvist, 2009) ...25

Figure 22. Positions of a p-h diagram where point c must be located ...26

Figure 23-a and Figure 23-b. Pressure gain against flow for CPb1 (left) and CPb2 (right) (from NIBE F1345 manual) ...27

Figure 24. Resulting available pressure against flow for CPb1 and CPb2 in parallel (100 %) ...28

Figure 25-a and Figure 25-b. Power against flow for CPb1 (left) and CPb2 (right) (from NIBE F1345 Installer manual) ...29

Figure 26. Modules percentage of operation time for three periods of several days (2017) ...30

Figure 27. BHEs outlet temperature during summer operation ...32

Figure 28. M1 (red) and M2 (green) typical summer cycle in R410A p-h diagram ...33

Figure 29. M1 (red) and M2 (green) typical early spring cycle in R410A p-h diagram ...33

Figure 30. M3 (red) and M4 (green) typical summer cycle in R407C p-h diagram ...34

Figure 31. M3 (red) and M4 (green) typical early spring cycle in R407C p-h diagram ...35

Figure 32. System’s COPs recorded in June (low space heating demand) ...36

Figure 33. System’s COPs recorded in June (DHW demand only) ...36

Figure 34. Energy consumption repartition (%) between pumps and compressors during summer operation. Remind that CPbs are the heat pumps circulation pumps for the cold side secondary loop, and that CPhs are the modules circulation pumps for the hot side secondary loop. ...37

Figure 35. Energy production repartition (%) during summer operation ...37

Table 1. Average COPH1 of each module for early spring and summer operations ...32

1 Introduction

In the current climate of global warming and need for more efficient energy systems, the use of ground- source heat pumps (GSHP) is one of the solution to reduce energy consumption. For both residential and commercial buildings, these systems are commonly used in Europe for heating, but also sometimes for cooling (Lund and Boyd, 2016). To do so, stored energy is extracted from the ground using only electricity.

Sweden, with its cold climate, was one of the first country to use geothermal energy in the 1970s and is now a leading country in Europe with some 500 000 GSHP systems installed (Juhlin and Gehlin, 2017). When looking at figures per capita, Sweden ranks first among all other countries in the world with roughly five GSHP for 100 inhabitants, as it is shown on Figure 1.

Figure 1. Ground-source heat pump installations in the world ranked on a per-capita basis (Rees, 2016)

Two different configurations of GSHP are possible: horizontal and vertical (Granryd, 2005). This report will focus on the second kind of systems, when used for heating. Indeed, the vast majority of the installations in Sweden are single vertical boreholes for single-family houses used only for heating (Juhlin and Gehlin, 2017).

The average depth of vertical borehole heat exchanger (BHE) has increased through the years with the development of new drilling equipment and techniques. From 100 m in 1995, it reached 171 m in 2013 and today, boreholes of about 250 m are not uncommon (Gehlin et al., 2016). The use of deeper boreholes allows to extract a larger amount of energy within a small surface area. This makes particularly sense in cities such as Stockholm where the concentration of population is high and the demand for geothermal energy is important.

Therefore, geothermal wells for GSHP can nowadays reach very high depths, such as the four boreholes of about 510 m drilled in 2016 in central Stockholm for the condominium “Ingemar 5”. KTH’s division of Applied Thermodynamics and Refrigeration has the chance to be involved in this project, and through this internship we have the opportunity to monitor such an uncommonly deep GSHP system.

The aims of the project are:

- Determine the data needed for the system monitoring.

- Set up new sensors and acquisition card.

- Analyse the performances of the system from the data recorded.

This report will give an overview of the GSHP technology, present the system studied and summarize the work done during the project. Some results derived from the measurements will be presented and analysed.

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2 Theory about ground-source heat pumps 2.1 Introduction to the technology

Heat pumps extract energy from a low-temperature source to a high-temperature sink using the mechanical work of a compressor. To do so, they rely on a vapor-compression cycle such as the one represented on Figure 2 (refrigerant loop), in which a refrigerant is evaporated, compressed, condensed and expanded in order to transfer the heat. Many configurations are possible (using ambient or exhaust air, ground soil, ground water or lake water as a source or sink) and a heat pump can be used either for heating or air- conditioning purpose (Granryd, 2005). In this report, we will consider the first case only since it is the most common one in Sweden.

Figure 2. Ground-source heat pump cycle (after Lacombe, 2016)

The refrigerant cycle in the primary loop can be represented in a “pressure-enthalpy” (p-h) or in a

“temperature-entropy” (T-s) diagram, depicted respectively on Figure 3 and Figure 4. In simple cycles such as these ones, the expansion is assumed to be isenthalpic (step d-a). Furthermore, we suppose there is no pressure drop in the heat exchangers (evaporator, condenser). An ideal compressor would result in an isentropic transformation (step b-cis), but in reality additional work is required to compensate for friction

a b

c

Borehole heat exchanger (BHE)

Circulation pump

Refrigerant loop Secondary loop Expansion

valve Compressor

𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡̇ Condenser

Evaporator Supply heat to the building

d

𝑄𝑄1̇

𝑄𝑄2̇

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and other losses in the compressor (step b-c). The refrigerant is superheated before reaching the compressor (step a’’-b), which means that it becomes pure vapor (step a’’-b). Similarly, it is subcooled when it becomes pure liquid before reaching the expansion valve (step c’-d).

Figure 3. Log(p)-h diagram of a simple refrigeration cycle (after Granryd et al., 2009)

Figure 4. T-s diagram of a simple refrigeration cycle (after Granryd et al., 2009)

An isotherm is drawn on Figure 3. In most of the cases, the temperature can be assumed to be constant with the pressure for a given enthalpy in the liquid zone. Note that the isotherm, just like the cycle on the T-s diagram Figure 4, are valid for a pure refrigerant (that is, constituted of one compound). Indeed, in the case of a mixture, the temperature in the “vapor + liquid” region may not be constant with the pressure (Makhnatch, 2017). This is the case of zeotropic mixtures for which boiling point changes as more vapor is formed. On the contrary, azeotropic mixtures behave like a pure fluid.

More specifically, a GSHP uses the ground as a source for heating. A secondary loop containing an anti- freeze secondary fluid circulates between the ground and the evaporator of the heat pump as represented

a b

c

Enthalpy (h) Evaporator

Compressor Condenser

Pressure (p)

Expansion valve

d

a b

d

a’’

c

c’’

Temperature (T)

Entropy (s)

c’

Vapor + Liquid Liquid

Vapor Vapor + Liquid

Vapor Liquid

Isotherm

c^is

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on Figure 2 (secondary loop). Heat is pumped from the ground and transferred to the refrigerant of the primary loop via the evaporator.

The ground temperature is not uniform, and increases with depth. It is quantified by the geothermal gradient, which is about 1-3 K per 100 m in Scandinavia (Gehlin et al., 2016). Without BHEs, the temperature remains constant over the years under the first 15 meters approximately (Derouet, 2014).

However, due to the GSHP extraction of heat, the ground temperature may decrease from one year to another if nothing is done to counteract this heat extraction. For example, heat could be reloaded during summer by doing free-cooling, that is, heat injection from the building to the ground, which is cold enough to directly provide cooling trough a fluid circulation.

Figure 5-a and Figure 5-b. Single U-tube (left) and coaxial tube (right) in a vertical BHE (Rees, 2016)

Sometimes, the secondary fluid is in direct contact with the surrounding soil (without any pipe) in an open system (Acuña, 2013). But the most common type of secondary loops for vertical BHE are the U-tube and the coaxial tube as represented on Figure 5.

2.2 Performance metrics

The most common way to measure the efficiency of a heat pump system is to determine the coefficient of performance (COP), which is defined as the ratio between the heating power delivered and the power consumed. It is important to notice that this is an instantaneous value that can be calculated at any time.

𝐶𝐶𝐶𝐶𝑃𝑃𝐷𝐷1= 𝑄𝑄₁̇

𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡̇ (1)

However, the concept of seasonal performance factor (SPF) has to be introduced to determine the heat pump performances over a season (Granryd, 2005). Instead of measuring the efficiency at a specific time, it is considering a full range of operating conditions that are varying throughout the year, but also throughout a single operation cycle (for example, the start-up may not be as efficient as the steady state operation).

Hence, the SPF becomes more useful when making comparison with other technologies or estimating the running costs. In this case, the heating power delivered and the power consumed are integrated for a full year of operation:

𝑆𝑆𝑃𝑃𝐹𝐹_𝐷𝐷1= ∫ 𝑄𝑄1̇

∫ 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡̇ (2)

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In order to account for the whole system, other sources of power demand such as the circulation pumps need to be considered. A European project called “SEPEMO” introduced several definitions for the SPF that are regrouped on Figure 6 depending on the system boundaries (Nordman, 2012):

- SPF^H1 includes only the heat pump unit (compressor).

- SPF^H2 additionally includes the ground loop circulation pump.

- SPFH3 additionally includes the supplementary heating system, used when the heat pump capacity is overcome.

- SPF^H4 additionally includes all circulating pumps for the heat distribution.

Figure 6. Seasonal performance factor scope according to SEPEMO definitions (Rees, 2016)

In the same way, we can define COPH1 to COPH4 depending on the scope of the system. It is important to use such normalized values for performance indicators, especially when making comparisons with other systems. It prevents making unfair comparisons, for instance by using COPH1 for a system and COPH4 for another.

2.3 Deep boreholes considerations

Because of the geothermal gradient, deeper boreholes will reach higher temperatures and provide higher heating power. As an example, one coaxial borehole heat exchanger of 800 m deep could provide the equivalent heat load of several conventional U-tube boreholes of 300 m (Holmberg et al., 2016). However, as the temperature increases, the potential of using GSHP for free-cooling is reduced. Consequently, deep boreholes are suitable for buildings with large heating and small cooling demand, in areas where the surface is limited and when other boreholes have already been drilled on neighboring properties.

Nevertheless, deep vertical GSHP systems remain poorly known (Le Lous et al., 2015), and some phenomena, which would not have been an issue for shallow boreholes, need to be taken into consideration.

In a BHE, the secondary fluid temperature varies along the depth. Hence, there is heat exchange between the “hot” upward-flowing and the “cold” downward-flowing pipe (in the heat extraction case). This phenomenon is called “thermal shunt” (Javed and Spitler, 2016). While it can be neglected for typical small boreholes, it becomes stronger as depth increases, resulting in a rise of the heat losses from one leg to another inside the BHE, reducing the efficiency of the GSHP system. It is possible to reduce this effect by increasing the resistance between the legs of the borehole, and by increasing the mass-flow rate of the circulating fluid, which is done to reduce the temperature difference between the two legs, i.e. the “shunt potential”.

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Besides, extra care should be taken with the pressure drop in the secondary loop. Indeed, the regular head losses are higher due to longer pipes and to the possibly increased mass-flow rate (to avoid thermal shunt).

Consequently, the total head losses and the energy required by the circulation pump are much more important than the case of a shallow BHE. To avoid this, larger pipes and coaxial or double U-tubes may be needed.

Furthermore, while the ground temperature increases with depth and is generally not disturbed by the weather under 15 m deep, it may be influenced by heat leakages in old urban areas, which come principally from buildings. In this case, the temperature of the top ground layers is more important, and the temperature profile is modified up to 100 m deep or more. Instead of increasing from the beginning, the ground temperature will reach a minimum along the depth. Consequently, deeper borehole does not necessarily mean higher heat power, especially in cities where the ground temperature should be correctly investigated in order to design a new system. For example, in Figure 7, the average temperature of a borehole of 350 m and 100 m deep are equivalent (in this case, it is assumed that the average ground surface temperature for an equivalent undisturbed area is 8°C).

Figure 7. Measured ground temperature in a 270 m deep borehole at Norra Frescati, Stockholm (Gehlin et al., 2016)

Finally, although the drilling of deeper boreholes has become easier recently due to the development of new drilling equipment and techniques, it is still an expensive solution. The cost per drilled meter is not linear and theoretically increases exponentially with depth (Gehlin et al., 2016). Indeed, the drilling compressor power is much more important, and risks of complications are higher, for instance clogging or drill bit change due to the discovery of another kind of rock. Besides, the deviation increases with depth and while the demand on precision is small in the countryside, it becomes more important in urban areas where risks for intersecting with other installations are higher, leading to an increase in costs. Deep boreholes also often need to be larger in diameter due to a higher required mass-flow rate. Consequently, setting up new deep BHEs usually requires to invest a large amount of money, which is conceivable only for some condominium or commercial buildings.

To conclude, drilling deeper and deeper is not always the best solution to get more heat power if no attentive investigation is carried on while designing the GSHP. If enough area is available, it may be more interesting to drill several shallow boreholes instead of a very deep one. The potential of using such deep boreholes in areas where the space is limited like cities is, however, important, and the focus should now be given to the optimization of that kind of systems. Drilling deep is expensive but more and more common, and the

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reconversion of very high depth abandoned boreholes, like the 2302 m deep borehole heat exchanger at Weggis in Switzerland studied by Kohl et al. (2002), is an interesting alternative both for heat generation and research purposes.

3 Ingemar 5, four deep BHEs in Stockholm 3.1 Presentation of the system

The GSHP system that is studied throughout this report belongs to the condominium “Ingemar 5” located in central Stockholm on Figure 8. It consists of two buildings of 29 apartments built in 1884. The system has been put into service during spring 2016 and is currently used for providing space heating and domestic hot water (DHW). Formerly, these services were supplied by a Scandinavian energy company through a connection to Stockholm’s district heating network. Today, the system is still connected in case the capacity of the GSHP is overcome. The main goal of this new heat pump was to decrease the energy costs for the residents.

Figure 8. Location of the condominium “Ingemar 5” (after SGUs Kartvisare)

According to Geological Survey of Sweden (“SGUs Kartvisare”), the ground in the area is mainly composed of sandstone (metagraywacke), graphite, and metamorphic rocks and sediments (mica schist, paragneiss, migmatite, quartzite…). Furthermore, several BHEs have already been drilled in the neighborhood. The closest ones are located no further than a street away from the condominium: two residences are equipped with three boreholes of 300 m and 8 boreholes of 270 m deep, respectively. In these circumstances, the installers decided to drill four boreholes of about 510 m deep to meet the needs of the condominium’s heating load.

The four BHEs are connected in parallel and contain U-tubes made of polyethylene. The external diameter of each pipe is 50 mm and their SDR (Standard Dimension Ratio, i.e. the ratio between the outside diameter and the thickness) is 17. The top 150 m of the upward-flowing pipe have a larger thickness to avoid thermal loses from the fluid. The SDR is consequently 11 over that portion. Furthermore, directional measurements have been performed by the division of Applied Thermodynamics and Refrigeration of KTH. It gives the exact position of the boreholes along their depth.

Ingemar 5

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Figure 9. Simplified hydraulic flowchart

The heat pump system is represented on Figure 9 (a more complete hydraulic flowchart is available in Appendix 1). It is composed of two units of 60 and 40 kW (respectively called HP1 and HP2), containing each two modules (hence two vapor-compression cycle) called M1 to M4. One module (M2) alternatively provides DHW and space heating, while the three others are only supplying space heating. Three tanks are used to store the DHW before being used. The system is also connected to the district heating network.

The ground loop secondary fluid is circulating thanks to a pump CPb3 of 4 kW and two smaller pumps CPb1 and CPb2. Each module M1 to M4 includes another small circulation pump, respectively called CPh1 to CPh4, for moving the hot water produced by the condenser.

The refrigerants used in the primary loop are R410A for HP1 and R407C for HP2. R410A is an almost azeotropic mixture, composed of R32 (50 wt%) and R125 (50 wt%). It has a low temperature glide (below 1 K) in the “vapor + liquid” region and almost behaves as a pure fluid. On the contrary, R407C is a zeotropic mixture of R32 (23 wt%), R125 (25 wt%) and R134a (52 wt%). Its temperature glide is about 5-7 K (Makhnatch, 2017). Note that the required operating pressures are higher for R410A. Both refrigerants have similar impact on the environment. They have been selected by the heat pump manufacturer, which is equipping its 60 kW units with R410A, and all its less powerful units with R407C for regulatory reasons.

The ground loop secondary fluid is a mixture of bioethanol and water, and its mass concentration in alcohol is estimated to 24.5 %, which corresponds to a freezing temperature of -15°C (Melinder, 2007).

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3.2 Ground thermal conductivity estimation

Together with the directional measurements, a Distributed Thermal Response Test (DTRT) has been performed in one of the boreholes during spring 2016. This test, introduced by Acuña et al. (2009), consists in measuring the circulating fluid temperature along a borehole during several phases: undisturbed ground conditions, constant amount of heat injection, recovery after the injection is stopped. This method is used to determine the local thermal properties of the ground such as the thermal conductivity.

In this report, an average ground thermal conductivity is estimated using exclusively the undisturbed ground temperature profile. A benefit of that method is that the conductivity can be determined quickly without the need to necessarily perform a DTRT. However, the conductivity is not determined locally along the borehole, and directional measurements are required.

If we consider that the heat transfer in the ground is pure conduction (which is a regular assumption as showed by Spitler and Bernier (2016)), Fourier’s law says:

𝑞𝑞 = −𝑘𝑘𝑑𝑑𝑇𝑇₀

𝑑𝑑𝑧𝑧 (3)

with 𝑇𝑇0 the undisturbed ground temperature (K), 𝑧𝑧 the depth (m), 𝑞𝑞 the geothermal heat flux (W/m²) and 𝑘𝑘 the ground conductivity (W/mK).

Figure 10. Undisturbed temperature profile derived from directional measurements (Mazzotti, 2016)

Therefore, if 𝑞𝑞 can be estimated, plotting the ground temperature against the depth will give the conductivity. From Näslund et al. (2005) and the International Heat Flow Commission (2011), a geothermal heat flux between 0.05 and 0.06 W/m² is a realistic assessment for Stockholm region.

A recorded profile of the ground temperature, derived from the directional measurements, is given Figure 10. It shows that the ground temperature is influenced in the top 100 m, most likely by nearby buildings.

The temperature gradient is estimated between 215 and 480 m. At these depths, the temperature profile is not affected by the heat leakage. The following value has been found:

𝑑𝑑𝑇𝑇0

𝑑𝑑𝑧𝑧 = −1.512 ∙ 10⁻² K/m (4)

which allows to finally estimate an interval for the conductivity 𝑘𝑘:

𝑘𝑘 ∈ [3.31 − 3.97] W/mK (5)

This values may be explained by the composition of the ground in this area of Stockholm: thermal conductivities are generally high for greywackes (Banks, 2012) and metamorphic rocks (for example

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Sundberg (1988) found a mean value of 6.62 W/m∙K for the quartzite and 3.58 W/m∙K for metamorphic sediments), and very high for the graphite which reach at least 25 W/m∙K.

3.3 Data summary for system monitoring

In order to estimate the performances of the system, COPs and SPFs should be determined according to the definitions of section 2.2. Consequently, COPH1 (and SPFH1) only considers the four heat pump modules (condensers, compressors). COPH2 (and SPFH2) includes in addition the circulation pumps CPb1, CPb2 and CPb3 for the ground loop secondary fluid. The scopes of these COPs are depicted on Figure 11, where only one module of one heat pump is represented. Furthermore, one can calculate COPH3 by looking at the district heating energy consumption, and COPH4 by taking additionally into account the circulation pumps CPh1 to CPh4, which are located in the hot side secondary loop.

To determine these coefficients as well as collect more information about deep borehole heat exchangers, data needs to be recorded. In this study, two ways of recording and downloading data have been used: an online application from the heat pump manufacturer; and an acquisition card to which various sensors have been connected.

The first procedure provides values recorded automatically from the heat pump units. It mainly consists of temperatures at several locations indicated on Figure 11 that will be used to determine the thermodynamic cycles and thus the COPH1:

- Temperature of the vaporous refrigerant exiting the evaporator 𝑇𝑇_𝑎𝑎^′′. - Compressor inlet temperature 𝑇𝑇_𝑏𝑏.

- Compressor outlet temperature 𝑇𝑇𝑐𝑐. - Condenser outlet temperature 𝑇𝑇𝑑𝑑.

- Evaporator inlet, cold side secondary loop temperature 𝑇𝑇𝑏𝑏,𝑖𝑖𝑚𝑚. - Evaporator outlet, cold side secondary loop temperature 𝑇𝑇𝑏𝑏,𝑐𝑐𝑜𝑜𝑡𝑡. - Condenser inlet, hot side secondary loop temperature 𝑇𝑇_{ℎ,𝑖𝑖𝑚𝑚}. - Condenser outlet, hot side secondary loop temperature 𝑇𝑇ℎ,𝑐𝑐𝑜𝑜𝑡𝑡. - Circulation pump speeds (CPb1, CPb2, CPh1 to CPh4).

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Figure 11. Detailed hydraulic flowchart with data measurement points

Additionally, the sensors connected to the acquisition card are listed on Figure 11. Their detailed references are given in Appendix 2. The following data is collected by those sensors:

- Energy-meter for space heating system: power 𝐸𝐸ℎ𝑠𝑠̇ and flow 𝑉𝑉_ℎ𝑠𝑠̇ . - Energy-meter for DHW: power 𝐸𝐸𝐷𝐷𝐷𝐷𝐷𝐷̇ and flow 𝑉𝑉_{𝐷𝐷𝐷𝐷𝐷𝐷}̇ .

- Energy-meter for the BHEs: power 𝐸𝐸𝐵𝐵𝐷𝐷𝐵𝐵̇ and flow 𝑉𝑉_𝑏𝑏̇ .

- Four thermometers at each borehole outlet: temperatures 𝑇𝑇1 to 𝑇𝑇4. - Differential pressure transducer for BHE: pressure drop ∆𝐶𝐶.

- Energy-meter for circulation pump CPb3: power 𝐸𝐸𝐶𝐶𝐶𝐶𝑏𝑏3̇ .

The first three sensors were already installed but not connected to any data reader, while the other ones were non-existing. Unfortunately, the differential pressure transducer and the DHW energy-meter have not been connected yet because of technical issues (details about the first one are given in section 4.1.2, while an extra adaptor is needed to extract the signal of the second one). Another problem came up with the energy meter for the space heating system: the signal provided is sometimes not detected (more information are given in section 7).

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4 Practical activities

A very important part of this internship consisted in finding a way to record the data and equipping the system with the appropriate sensors. It included meeting with different protagonists such as the heat pump installer and the condominium responsible.

One of the chosen solution was to set up an acquisition card which is converting analog signals (from the sensors) into digital signals readily recordable on a local computer. The setup is represented in Figure 12. A programming software is used to configure the data acquisition card in order to perform measurements.

Figure 12. Data acquisition system overview (after Agilent 34970A User’s Guide)

4.1 Set up of new sensors

4.1.1 Boreholes thermometers

Four RTDs (resistance temperature detectors) have been connected to the borehole outlet pipes as shown on Figure 13. These sensors are measuring the temperature by correlating it with the resistance of a platinum probe. They are “Pt1000” sensors which means that the probe resistance at 0°C is 1000 ohms.

Figure 13. Thermometers connected to the 4 boreholes exit

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To avoid the measurements to be perturbed by air in-between the sensors and the fluid, thermal paste has been added inside the pipes (cf. Figure 14-a), which have been sealed with silicone afterwards (cf. Figure 14-b). Finally, a plastic protection has been added around the cables to protect them against mice.

Figure 14-a and Figure 14-b. Thermal paste (left) and silicone (right) on a borehole thermometer

The outlet temperatures can be used to obtain information about each borehole as well as the four boreholes as a whole (borehole field). For instance, together with a flow-meter and the temperature sensor in the inlet pipe (equal for all boreholes), they provide the power extracted from the BHEs 𝐸𝐸_{𝐵𝐵𝐷𝐷𝐵𝐵}̇ .

4.1.2 Boreholes differential pressure transducer

A differential pressure transducer such as the one of Figure 15-a has been installed between the two pipes leading to the BHE to evaluate its pressure drop. Both ends are connected to the pipes in the way shown on Figure 15-b. Silicone elements are then converting the pressure difference into an electrical signal, by using the piezoresistive effect.

Figure 15-a and Figure 15-b. Differential pressure transducer (left) and connection to a pipe (right)

Since the sensor used was quite old, it was calibrated before being used. To do so, it has been connected to a manual pump together with a reference manometer at one of its ends. Several pressure differences have been measured both with the data acquisition unit (in V) and the reference manometer (in bar) within the operating range. Then, a linear regression between those two values led to:

∆𝐶𝐶 = −48.87 ∙ ∆𝑉𝑉 + 0.03 bar (6)

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with ∆𝑉𝑉 the voltage measured by the transducer and ∆𝐶𝐶 its corresponding pressure. Next, uncertainties of type A and B can be determined (JCGM, 2008). Uncertainties 𝑢𝑢𝑥𝑥 and 𝑢𝑢𝑦𝑦 over the coefficients of equation 6 (respectively -48.87 and 0.03) are determined with the “linear regression” tool of Excel, while uncertainty 𝑢𝑢∆𝑝𝑝 over ∆𝑉𝑉 can be determined from the experimental variance of the sample of measurements of ∆𝑉𝑉. The combined uncertainty formula applied to equation 6 gives the type A uncertainty over ∆𝐶𝐶:

𝑢𝑢_∆𝑐𝑐= �(∆𝑉𝑉 ∙ 𝑢𝑢𝑥𝑥)²+ (−48.87 ∙ 𝑢𝑢_∆𝑝𝑝)²+ 𝑢𝑢_𝑦𝑦² (7) Then, type B uncertainty is deduced from the accuracy of the transducer and of the reference manometer.

Combining those two and expanding the uncertainty with a coverage factor corresponding to a confidence level of 95 %, leads to an overall uncertainty for ∆𝐶𝐶:

𝑈𝑈 = 0.04 bar (95 %) (8)

which is constant for all the pressures within the operating range of the transducer.

Unfortunately however, the transducer did not lead to reasonable values of pressure when connected to the GSHP system. It may have been damaged during the transportation or by an excessive pressure inside the circuit.

4.1.3 Circulation pump CPb3 energy-meter

The energy-meter shown in Figure 16 has been connected in series to the 3-phase supply current of the pump CPb3. It can be configured via USB with a specific software, and several parameters such as the electrical power can be retrieved through two analog outputs.

Figure 16. CPb3 energy-meter

Since the speed of the pump may vary and since its characteristics are not available for all speeds, this measure is particularly useful. The pump power is significant for deep boreholes, because of the high mass flow-rate and the big pressure drop inside the BHEs.

4.2 Programming of the acquisition card

All the sensors have to be connected to the acquisition card through modules like the one pictured in Figure 17. Several types of modules have been used, each including different types of channels according to the kind of analog signals (voltage, current, pulse) provided by the sensors. Some channels can also provide the supply voltages needed by a few sensors (the differential pressure transducer for instance). All details for connection (and later, programming) are available in Appendix 3.

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Figure 17. Acquisition card module connected to a 4-20 mA signal (yellow cables)

Moreover, a computer program is required to set the data acquisition (an excerpt is given Figure 18). It aims to convert the electrical data received by the acquisition unit (mV, mA…) into the variables needed (kW, m³/h…), for a specific time step. For this monitoring, one set of measurements is taken every minute. Some new parameters are introduced, for example powers are calculated from temperatures and flow. A summary of all those variables is given in Appendix 3.

Furthermore, the program provides real-time evolution of the different measured parameters in form of graphs such as the ones displayed in Figure 19. All the recorded data are written and saved in a file, which can later be used as a basis for analysis.

Figure 18. Part of the dataflow program realized

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Figure 19. Example of system monitoring

5 Data analysis methodology

5.1 Determination of thermodynamic cycles

In order to derive their COPs and to get precise information about their operation, the thermodynamic cycles of each module are required. That means the points a, b, c and d of the p-h diagram presented Figure 3 have to be found. To do so, the software REFPROP (V7.0) and its Excel add-on have been used. Given two input parameters, they provide all the thermodynamic properties of the refrigerant, defining its state (hence its position on the diagram).

The data collected by the heat pump units, detailed in section 3.3, is used in this section. Unfortunately, some measurements lack to be able to determine the state of the refrigerant at all important locations in the system. Indeed, there is neither direct nor indirect measurement of the condensing high pressure of the cycle 𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ, through for instance any of the temperatures of c’ or c’’ of Figure 4: it will have to be approximated.

Nevertheless, the evaporating low pressure of the system 𝑃𝑃𝑙𝑙𝑐𝑐𝑙𝑙 can be readily determined from the dew curve temperature 𝑇𝑇_𝑎𝑎^′′, which is part of the heat pump data set. Then, enthalpy ℎ_𝑏𝑏 of point b is derived from 𝑃𝑃𝑙𝑙𝑐𝑐𝑙𝑙 and 𝑇𝑇𝑏𝑏.

The expansion process (step d-a in Figure 3) is considered isenthalpic, hence the enthalpy ℎ_𝑎𝑎 of point a is equal to the enthalpy ℎ_𝑑𝑑 of point d. This last enthalpy is determined assuming that the isotherms are merged with the isenthalpic lines in the liquid zone (as explained in section 2.1): in this way, the intersection between the isotherm 𝑇𝑇𝑑𝑑 and the boiling curve will provide the enthalpy sought. However, 𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ cannot be derived from 𝑇𝑇𝑑𝑑 and ℎ𝑑𝑑.

In order to determine 𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ, let us investigate at the heat transfer inside the condenser. Variation of temperatures of a pure refrigerant and of a secondary fluid (in this case, water) are represented Figure 20.

The idea of the following assumption is that the temperature of the refrigerant at the dew point 𝑇𝑇_𝑐𝑐^′′ can be

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estimated thanks to the temperature of the water leaving the condenser 𝑇𝑇ℎ,𝑐𝑐𝑜𝑜𝑡𝑡, which is known. Indeed, since the temperature of the superheated refrigerant at point c (compressor exit) is very high (it may be superior to 90 °C), there is heat transfer at the entrance of the condenser that may be enough to warm the exiting secondary fluid up to the condensing temperature 𝑇𝑇_𝑐𝑐^′′. A first approximation would be:

𝑇𝑇_{ℎ,𝑐𝑐𝑜𝑜𝑡𝑡}≈ 𝑇𝑇_𝑐𝑐^′′ (9)

Figure 20. Temperatures in the condenser for refrigerant (blue) and water (black) (from Lundqvist, 2009)

However, the temperature profile of the refrigerant is modified in the case of a zeotropic mixture like R407C. Hence, this hypothesis will be adapted to the two different heat pumps working with R410A and R407C in the next section.

After 𝑇𝑇_𝑐𝑐^′′ has been determined, the point c’’ belonging to the dew curve leads to 𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ. Finally, the enthalpy ℎ_𝑐𝑐 of point c is derived from 𝑃𝑃_{ℎ𝑖𝑖𝑖𝑖ℎ} and 𝑇𝑇_𝑐𝑐. The enthalpy ℎ_𝑑𝑑 of point d can be calculated in a similar way, and its value is always very close to ℎ_𝑎𝑎 which is consistent with an isenthalpic expansion model.

In the end, the whole thermodynamic cycle of the refrigerant has been determined, and the COPH1 of the module 𝑚𝑚 can be calculated with the following enthalpies:

𝐶𝐶𝐶𝐶𝑃𝑃_{𝐷𝐷1,𝑘𝑘}^𝑐𝑐 =ℎ𝑐𝑐− ℎ𝑎𝑎

ℎ_𝑐𝑐− ℎ_𝑏𝑏 (10)

Note that these COPs are not taking into account the compressor extra power consumption due to the heat losses to the surroundings (mechanical, electrical), that are usually around 5 % (Granryd et al., 2009). Hence, one can define a corresponding efficiency 𝜂𝜂_𝑐𝑐 that is assumed to be 0.95, and the total COP for each module will be:

𝐶𝐶𝐶𝐶𝑃𝑃_𝐷𝐷1^𝑐𝑐 =ℎ_𝑐𝑐− ℎ_𝑎𝑎

ℎ_𝑐𝑐− ℎ_𝑏𝑏∙ 𝜂𝜂_𝑐𝑐 = 𝐶𝐶𝐶𝐶𝑃𝑃_𝐷𝐷1^𝑐𝑐 ∙ 𝜂𝜂_𝑐𝑐 (11) Observe that data is recorded throughout time. Therefore, for every time interval one can determine a cycle and a COPH1.

Temperature (T)

Length over condenser 𝑇𝑇_{ℎ,𝑐𝑐𝑜𝑜𝑡𝑡}

𝑇𝑇ℎ,𝑖𝑖𝑚𝑚

𝑇𝑇_𝑐𝑐

𝑇𝑇𝑑𝑑

𝑇𝑇_𝑐𝑐^′′

𝑇𝑇_𝑐𝑐^′

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5.2 Consistency and verification of the model

The consistency of the assumption taken as expressed in equation 9 can be checked. Indeed, with the position of point b (compressor inlet) and the temperatures of points c (compressor outlet) and d (condenser outlet), one can define a region of the p-h diagram where point c must be located, as represented on Figure 21. Obviously, this region is delimited to a line that is a part of the isotherm 𝑇𝑇𝑐𝑐, which is measured. This line is delimited at the top by the isentropic line 𝑠𝑠𝑏𝑏 which corresponds to the case of an ideal compression:

the enthalpy difference between points b and c cannot be smaller. On the other hand, point d must belong to the liquid zone to have subcooling so that the heat pump works properly. Note that it may happen that point d ends up within the two phase region, for instance when the refrigerant charge is too small, but it is assumed that it is not the case here. Hence, the high pressure of the cycle cannot be smaller than the pressure 𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ,𝑐𝑐𝑖𝑖𝑚𝑚 for which 𝑇𝑇_𝑑𝑑 crosses the bubble point curve, which is delimiting the region at the bottom.

Figure 21. Positions of a p-h diagram where point c must be located

The range obtained for point c is more or less narrow; for example, data recorded in June 2017 lead to high pressure ranging from 7 to 11 bars for HP1, and from 3 to 6 bars for HP2. This is however sufficient to adapt the hypothesis expressed in equation 9. It showed that with this assumption, the isentropic efficiency had a tendency to be excessively close to 1, leading to a high pressure very close to the upper limit for HP1, while for HP2 it was sometimes lower than the minimum pressure required 𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ,𝑐𝑐𝑖𝑖𝑚𝑚.

It was mentioned in previous sections that the temperature profiles in the condensers of both heat pumps are different. At constant pressure, the phase change temperature of the zeotropic refrigerant R407C is decreasing during condensation, while it is constant for the azeotropic mixture R410A as shown on Figure 20. Hence, for both heat pumps, the performances of the condensers may vary and the hypothesis made equation 9 may differ slightly.

Knowing that, in order to get a balanced system and reasonable values for the high pressure and the isentropic efficiency, equation 9 has been reconsidered, distinguishing the cases of both heat pumps:

� 𝑇𝑇ℎ,𝑐𝑐𝑜𝑜𝑡𝑡≈ 𝑇𝑇_𝑐𝑐^′′+ 3 K for HP1

𝑇𝑇ℎ,𝑐𝑐𝑜𝑜𝑡𝑡≈ 𝑇𝑇_𝑐𝑐^′′− 2 K for HP2 (12)

This new hypothesis may not always lead to an exact position for point c, as this temperature difference may vary depending on the external conditions. However, it is sufficient to get a first estimation of the performances of the system that will be used in the following parts of this report.

Enthalpy (h)

Pressure (p)

𝑇𝑇𝑑𝑑

cis

𝑇𝑇_𝑐𝑐

𝑠𝑠𝑏𝑏

𝑃𝑃ℎ𝑖𝑖𝑖𝑖ℎ,𝑐𝑐𝑖𝑖𝑚𝑚

Vapor + Liquid Vapor

Liquid

cmin

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To get an idea of the accuracy of this method, let us look at the range of COPs derived from the extrema positions cis and cmin where point c must be located on Figure 21 (that is, the difference between the maximum COPH1 in the case the compression would be isentropic, and the minimum COPH1 in the case the point d would be located on the dew line). Calculations lead to average differences of 1.35 and 0.85 for HP1 and HP2 respectively during the periods when data has been recorded. It means that the COPs derived in the following parts of this report belong to an interval of uncertainty of 1.35 and 0.85 for HP1 and HP2 respectively. However, the compression is most likely not isentropic, and this large interval could be reduced by performing a more detailed analysis, for example if more information about the heat exchange process inside the condenser is available.

Finally, this method of calculating the COPH1 through an estimation of the high pressure cycle could be checked in several ways. Indeed, if all the hot side secondary loop energy-meters were connected to the data acquisition unit, it would be possible to perform an energy-balance over the heat pump units and then to derive the COPs without any thermodynamic cycle analysis. Besides, one could also install high pressure sensors inside the heat pump units.

5.3 Flow distribution in the cold side secondary loop

The flow of secondary refrigerant going through each evaporator module (see Figure 9) is required in the following steps of the analysis. Recall from section 3.3 that the total flow 𝑉𝑉𝑏𝑏̇ in the secondary ground loop is recorded with the data acquisition unit. Flow distribution in-between the modules depends on the circulation pumps CPb1 and CPb2. We assume that if only either one of these pumps is working, all the flow 𝑉𝑉𝑏𝑏̇ is going through the branch of this pump.

However, if both pumps are working, the flow is divided according to their characteristic diagrams given Figure 22. Furthermore, the circulation pump speeds are variable but are most of the time either 0 or 100

% of the maximum speed. For the sake of simplicity, only these two cases will be considered. Since both pumps are connected in parallel through two similar circuits, their resulting characteristic is obtained by adding both flows at a given pressure as it is shown on Figure 23.

Figure 22-a and Figure 22-b. Pressure gain against flow for CPb1 (left) and CPb2 (right) (from NIBE F1345 manual)

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Figure 23. Resulting available pressure against flow for CPb1 and CPb2 in parallel (100 %)

The curves are smooth enough to be estimated with a polynomial, hence the pumps available pressure 𝑃𝑃𝑎𝑎𝑎𝑎

can be determined from 𝑉𝑉𝑏𝑏̇ :

�𝑃𝑃^{𝑎𝑎𝑎𝑎} = −0.378 ∙ 𝑉𝑉𝑏𝑏̇ ²− 7.3218 ∙ 𝑉𝑉𝑏𝑏̇ + 168.39 for 𝑉𝑉𝑏𝑏̇ < 6.05 m³/h

𝑃𝑃𝑎𝑎𝑎𝑎 = −0.2028 ∙ 𝑉𝑉𝑏𝑏̇²+ 2.802 ∙ 𝑉𝑉𝑏𝑏̇ + 100.28 for 𝑉𝑉𝑏𝑏̇ ≥ 6.05 m³/h (13) Next, 𝑃𝑃_{𝑎𝑎𝑎𝑎} leads to the flows 𝑉𝑉_{𝐶𝐶𝐶𝐶𝑏𝑏1}̇ and 𝑉𝑉_{𝐶𝐶𝐶𝐶𝑏𝑏2}̇ through the circulation pumps CPb1 and CPb2 by solving the following equations:

𝑃𝑃_{𝑎𝑎𝑎𝑎} = −0.378 ∙ 𝑉𝑉_{𝐶𝐶𝐶𝐶𝑏𝑏2}̇ ²− 7.3218 ∙ 𝑉𝑉_{𝐶𝐶𝐶𝐶𝑏𝑏2}̇ + 168.39 (14)

𝑉𝑉_{𝐶𝐶𝐶𝐶𝑏𝑏1}̇ = 𝑉𝑉_𝑏𝑏̇ − 𝑉𝑉_{𝐶𝐶𝐶𝐶𝑏𝑏2}̇ (15)

Finally, since each branch is subdivided into two modules, these flows are divided by two to obtain the flow in the evaporator of each module, leading to 𝑉𝑉_{𝑏𝑏,𝑐𝑐}̇ (with 𝑚𝑚 the module index).

5.4 Performance analysis of the whole system

In order to consider the circulation pumps in the performance analysis and to determine COPs for the whole system, additional data, such as the pumps electrical power and the heat provided by the condensers, is required.

While the electrical power of CPb3 is directly measured with the energy-meter introduced section 4.1.3, the powers of the other circulation pumps are derived from graphs provided by the manufacturer that are given Figure 24. These show the powers of the pumps CPb1 and CPb2 as a function of the flow.

0 20 40 60 80 100 120 140 160 180

0 5 10 15 20 25

External available pressure (kPa)

Flow (m3/s)

CPb1 CPb2 Parallel

(29)

-29-

Figure 24-a and Figure 24-b. Power against flow for CPb1 (left) and CPb2 (right) (from NIBE F1345 Installer manual)

As specified above, the speed of the circulation pumps (except CPb3) is either 0 or 100 %. Furthermore, the power 𝐸𝐸𝐶𝐶𝐶𝐶𝑏𝑏2̇ of the pump CPb2 is almost constant with the flow as shown on Figure 24-b, hence we assume:

𝐸𝐸_{𝐶𝐶𝐶𝐶𝑏𝑏2}̇ = 710 W (16)

However, as shown on Figure 24, the power of CPb1 is varying too much with the flow to make any other simplification. Consequently, it is determined through this graph by estimating the curve with a polynomial regression, and with the flow 𝑉𝑉𝐶𝐶𝐶𝐶𝑏𝑏1̇ determined in the previous section.

In a second step, the evaporating power 𝑄𝑄₂̇ of every module is determined with the aim to compute the condensing power 𝑄𝑄₁̇ and consequently, obtain the compressor power transferred to the refrigerant 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑘𝑘̇ . 𝑄𝑄₂̇ can be evaluated by:

𝑄𝑄₂̇ = 𝑉𝑉𝑏𝑏,𝑐𝑐̇ ∙ 𝜌𝜌_𝑏𝑏∙ 𝐶𝐶𝐶𝐶_𝑏𝑏(𝑇𝑇_{𝑏𝑏,𝑖𝑖𝑚𝑚}− 𝑇𝑇_{𝑏𝑏,𝑐𝑐𝑜𝑜𝑡𝑡}) (17) with 𝑇𝑇_{𝑏𝑏,𝑖𝑖𝑚𝑚}− 𝑇𝑇_{𝑏𝑏,𝑐𝑐𝑜𝑜𝑡𝑡} the temperature difference of the ground loop secondary fluid between the inlet and the outlet of the evaporator, and 𝜌𝜌𝑏𝑏∙ 𝐶𝐶𝐶𝐶𝑏𝑏 the product of its density by its specific heat capacity, that is, for a 5°C mixture, 4166 kJ/m³K (Melinder, 2007).

Then, from the first law of thermodynamics together with the definition of the COPH1 and its value determined in section 5.1, 𝑄𝑄2̇ and 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑘𝑘̇ can simply be determined from the following equations:

𝑄𝑄1̇ − 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑘𝑘̇ − 𝑄𝑄2̇ = 0 (18) 𝐶𝐶𝐶𝐶𝑃𝑃𝐷𝐷1,𝑘𝑘= 𝑄𝑄₁̇

𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑘𝑘̇ (19)

Note that in this case, 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑘𝑘̇ is the power transferred to the refrigerant, but does not take into account the mechanical, electrical and heat losses of the compressor (see section 5.1). The total power consumed by the compressor 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡̇ is approximated with the efficiency 𝜂𝜂_𝑐𝑐:

𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡̇ =𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑘𝑘̇

𝜂𝜂_𝑐𝑐 (20)

(30)

-30-

Finally, one can define a global COPH1 for the whole system that takes into account all the modules running at the same time (the upper indexes represent the number of each module):

𝐶𝐶𝐶𝐶𝑃𝑃_𝐷𝐷1= 𝑄𝑄₁¹̇ + 𝑄𝑄₁²̇ + 𝑄𝑄₁³̇ + 𝑄𝑄₁⁴̇

𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡1 ̇ + 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡2 ̇ + 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡3 ̇ + 𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡4 ̇ (21) Similarly, the COPH2, which takes additionally into account all the circulation pump powers computed until now, is calculated by:

𝐶𝐶𝐶𝐶𝑃𝑃_𝐷𝐷2= ∑⁴_𝑐𝑐=1𝑄𝑄₁^𝑐𝑐̇

∑⁴_𝑐𝑐=1𝐸𝐸𝑐𝑐𝑐𝑐𝑐𝑐𝑐𝑐,𝑡𝑡𝑐𝑐̇ + ∑²_{𝐷𝐷𝐶𝐶=1}𝐸𝐸_{𝐶𝐶𝐶𝐶𝑏𝑏}^{𝐷𝐷𝐶𝐶}̇ + 𝐸𝐸_{𝐶𝐶𝐶𝐶𝑏𝑏3}̇ (22) Note that, very often, all the modules are not running at the same time. Those that are not working are not influencing the calculation of the global COPs since their condensing power are nil. Furthermore, recall that these values are instantaneous. In order to get an overview of the performances over a period of time, one can integrate term by term the powers in equations 21 and 22 to end up with a ratio of energies denoted 𝑆𝑆𝑃𝑃𝐹𝐹𝑐𝑐𝑝𝑝𝑝𝑝𝑖𝑖𝑐𝑐𝑑𝑑. Observe that integrating those values over a full year of operation will provide the SPF as defined in equation 2.

6 Results and discussion 6.1 System’s seasonal working

The system studied is designed to provide DHW and, to a larger extent, space heating. Logically, it will run close to its full capacity during winter, when the space heating demand is high. On the contrary, during summer, one module out of four will generally be sufficient to satisfy the low demand. As an example, Figure 25 shows the modules percentage of operation time over three different periods of the year with different average outdoor temperatures 𝑇𝑇𝑐𝑐𝑜𝑜𝑡𝑡 (i.e. at 100 % the module is working all the time over the period). While several modules are often running simultaneously in the first case (𝑇𝑇𝑐𝑐𝑜𝑜𝑡𝑡 = 7.7 °C), the space heating demand is almost nil for the last recording (𝑇𝑇𝑐𝑐𝑜𝑜𝑡𝑡= 17.9 °C). In this case, none module is working except M2 which is the only one able to satisfy the DHW.

Figure 25. Modules percentage of operation time for three periods of several days (2017) 0

10 20 30 40 50 60

M1 M2 M3 M4

Percentage of operation time (%)

T_out = 7.7 °C T_out = 11.5 °C T_out = 17.9 °C

Monitoring of a heat pump system using deep borehole heat exchangers

Monitoring of a heat pump system using deep borehole heat exchangers

Baptiste Delattre

Abstract

Sammanfattning

Acknowledgements

Nomenclature

Table of Figures

Table of Contents

1 Introduction

2 Theory about ground-source heat pumps 2.1 Introduction to the technology

2.2 Performance metrics

2.3 Deep boreholes considerations

3 Ingemar 5, four deep BHEs in Stockholm 3.1 Presentation of the system

3.2 Ground thermal conductivity estimation

3.3 Data summary for system monitoring

4 Practical activities

4.1 Set up of new sensors

4.2 Programming of the acquisition card

5 Data analysis methodology

5.1 Determination of thermodynamic cycles

5.2 Consistency and verification of the model

5.3 Flow distribution in the cold side secondary loop

5.4 Performance analysis of the whole system

6 Results and discussion 6.1 System’s seasonal working