Ground source heat pump performance in case of high humidity soil and yearly balanced heat transfer

Energy Conversion and Management 76 (2013) 956–970

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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman

Ground source heat pump performance in case of high humidity soil and yearly balanced heat transfer Luigi Schibuola a, Chiara Tambani a, Angelo Zarrella b, Massimiliano Scarpa a,⇑ a b

University IUAV of Venice, Dorsoduro 2206, 30123 Venice, Italy University of Padua, Via Venezia 1, 30131 Padua, Italy

a r t i c l e

i n f o

Article history: Received 24 May 2013 Accepted 1 September 2013

Keywords: Ground source heat pumps High humidity soils HVAC Inverter-driven compressor Lagoon water Yearly balanced heat transfer

a b s t r a c t Ground source heat pump (GSHP) systems are spreading also in Southern Europe, due to their high energy efficiency both in heating and in cooling mode. Moreover, they are particularly suitable in historical cities because of difficulties in the integration of heating/cooling systems into buildings subjected to historical preservation regulations. In these cases, GSHP systems, especially the ones provided with borehole heat exchangers, are a suitable solution instead of gas boilers, air-cooled chillers or cooling towers. In humid soils, GSHP systems are even more interesting because of their enhanced performance due to higher values of soil thermal conductivity and capacity. In this paper, GSHP systems operating under these boundary conditions are analyzed through a specific case study set in Venice and related to the restoration of an historical building. With this analysis the relevant influences of soil thermal conductivity and yearly balanced heat transfer in the design of the borehole field are shown. In particular, the paper shows the possibility to achieve higher compactness of the borehole field footprint area when yearly balanced heat transfer in the borehole field is expected. Then, the second set of results contained in the paper shows how GSHP systems designed for high humidity soils and yearly balanced heat loads at the ground side, even if characterized by a compact footprint area, may still ensure better performance than other available and more common technologies such as boilers, air-cooled chillers, chillers coupled with cooling towers and heat pumps and chillers coupled with lagoon water. As a consequence, in this last comparison, the efficiency of GSHP systems is confirmed even when compact borehole installation areas are used. The analysis is performed in terms of both input energy consumption ratio and primary energy ratio and takes into account the energy consumption of auxiliary devices as well. The conclusions are drawn in terms of percentage, thus allowing the reader to extend the results achieved in this paper to systems operating in other contexts under similar boundary conditions. Ó 2013 Elsevier Ltd. All rights reserved.

1. Introduction Ground source heat pump systems are among the most interesting high energy efficiency technologies for building heating and cooling, as confirmed in many papers, such as in [1–4], and, as regards the case of Mediterranean climate, in [5]. In [5], an experimental assessment of GSHP system behavior was performed, whereas in the present paper mostly simulations are presented, thus allowing the authors to explore a wider range of parameter variations. In the present paper, the experimental basis consists in the Ground Response Test suitable for a consistent assessment of the equivalent thermal conductivity of the soil under examination.

⇑ Corresponding author. E-mail addresses: [email protected] (L. Schibuola), [email protected] (C. Tambani), [email protected] (A. Zarrella), [email protected] (M. Scarpa). 0196-8904/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.enconman.2013.09.002

Moreover, in [5], the comparison between GSHP systems and air– water heat pumps is not completely applicable to real systems due to the presence of a specially optimized water–water heat pump using propane (R290) as a refrigerant fluid, whereas the considered air–water heat pump was available on the market. In particular, horizontal and vertical ground heat exchangers constitute the most diffuse and versatile applications of GSHP, often due to the lower amount of required authorizations and operation limitations, when compared against ponds or ground water, usually constrained by limitations in water temperature raise and mass flow rates. But the definitive spread of these systems can happen only if ratio performance/costs is increased, because of current high installation costs. Design and energy estimation tools may help in making this technology more cost-efficient, for instance by avoiding oversizing and determining optimal design water supply temperatures, depending on the specific building intended use and

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Nomenclature RppA Abbreviations AHU Air Handling Unit CR Capacity Ratio GRT Ground Response Test GSHP ground source heat pump HVAC Heating, Ventilation and Air-Conditioning IWEC International Weather for Energy Calculations PCM Phase Change Materials PLF Part-Load Factor SHGC Solar Heat Gain Coefficient of windows U-value overall heat transfer coefficient of building transparent and opaque surfaces Symbols c specific heat (J/(kg K)) C heat capacity (J/K) COP coefficient of performance (–) E energy (kW h) i counter of geothermal field slices (–) j counter of geothermal field annular regions (–) L length or depth (m) m number of geothermal field slices (–) n number of geothermal field annular regions (–) Pu thermal capacity (kW) heat flow (W) Q_ r radial coordinate in the borehole discretization, starting from the axis of the borehole R thermal resistance (K/W) rm radial position of the thermal node in the borehole discretization, starting from the axis of the borehole ReadValue value read on the measurement device (variable) Rp0 thermal resistance between pipe and surrounding ground per length unit (m K/W)

configuration as well as on soil characteristics, thus achieving the maximum optimization of the borehole field, as shown in relevant studies such as [6–8]. Moreover, as regards the enhancement of the compactness of boreholes and hence their energo-economical optimization, the application of Phase Change Materials (PCM) inside boreholes may be a future viable means to reduce borehole length at peak conditions, as explained in [9], thanks to the exploitation of their high thermal inertia (consequent to the related latent heat of fusion). In particular, the context of high humidity soils and yearly balanced heat transfer at the ground side is studied in this paper and the related performance of GSHP systems is compared against various available choices in GSHP system parameters. The study by Leong et al. [10] showed the major influence of high humidity soil in the assessment of GSHP system performance. However, compared with [10], the present paper shows a higher degree in the variation of GSHP parameters, thus showing how each design choice can influence the final performance. Furthermore, the evaluations performed in the frame of this study take into account even the heat pump control mode, namely on–off strategy or inverter-driven compressor. This topic has been investigated in other papers, such as in [11–15], but with contrasting results. While Zhao et al., in [11], define variable speed compressors as the best choice in order to improve the performance of GSHP systems in part load conditions, Karlsson et al., in [12] and [13], show that inverter-driven compressors increase the efficiency of GSHP systems under part load conditions, but cannot improve the seasonal performance, because of the additional inefficiencies brought by the inverter and the connected motor. But in

RppB t t1

t2

u V_ kg,Eq

Dz Dh Ds h h

q

thermal resistance between adjacent borehole pipes per length unit (m K/W) thermal resistance between opposite borehole pipes per length unit (m K/W) time first reference time for the calculation of the equivalent thermal conductivity of the ground in the Ground Response Test (GRT) [s] second reference time for the calculation of the equivalent thermal conductivity of the ground in the Ground Response Test (GRT) [s] uncertainty (variable) volumetric flow rate (m3/s) equivalent thermal conductivity of the ground (W/(m K)) slice thickness (m) temperature difference (K) time-step duration (s) temperature (°C) average temperature (°C) density (kg/m3)

Subscripts B borehole Cond condenser g ground In inlet Input related to input energy consumption (gas or electricity) Needs related to energy needs Out outlet Primary related to primary energy w water Ds referred to the previous calculation time-step

this regard Cuevas and Lebrun [14] show energy losses due to the inverter are low, since its efficiency is between 95% and 98%, as well as the motor losses are negligible. Recently, Madani et al. [15] show that the improvement in energy performance of GSHP systems provided with inverter-driven heat pumps compared with on–off heat pumps depends on the ratio of the heat pump size to the design heating load. However, in [15], cycling losses are not taken into account, heat pumps are modeled as an optimized connection of components, and cold climate is applied, whereas, in the present paper, cycling compressor losses are considered, detailed performance data of commercially available heat pumps are used, and typical Mediterranean climate is considered. In the second set of results shown in this paper, compact GSHP systems have been further compared with other HVAC (Heating, Ventilation and Air Conditioning) systems suitable for historical coastal cities, such as heat pumps coupled with lagoon water, and with traditional HVAC systems, such as gas boilers, air–water heat pumps and chillers coupled with cooling towers. To sum up, this study is aimed at the evaluation of the most suitable GSHP system configurations in case of high humidity soils and yearly balanced heat transfer at the ground side. This analysis takes advantage of a case study sited in Venice and consisting in the restoration of an historical building. In this analysis the GSHP system performance is calculated both in heating and in cooling conditions via detailed simulation software developed by the authors and aimed at modeling the heat pump and the borehole heat exchanger. This software is applied to a probable profile of heat loads obtained through the simulation of the real building

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via building energy simulation software EnergyPlus. The analysis contained in the paper determines the influence of the main design parameters and the consequent system optimization. The second set of comparisons deals additionally with the performance achievable by the possible alternative systems by considering an air-source heat pump, a lagoon water source heat pump and the most traditional solution, consisting in condensing boilers and chillers coupled with cooling towers.

the building structure, and an Aula Magna, provided with air-conditioning by an AHU (Air Handling Unit). The thermal energy is generated by means of two invertible heat pumps operating in parallel and briefly described in Table 3. The heat pump detailed performance data rated in heating and in cooling conditions were given according with current Standards on heat pump performance rating in nominal conditions (i.e.: [16–19]) and part load operation [20]. In particular, nominal capacity and COP were rated by the manufacturer under the following conditions:

2. Materials and methods In this section the main boundary conditions, assumptions and calculation methods used in this paper are described. As a consequence, in the following subsections, the simulated building-plant system is described, as well as the comparisons performed among various GSHP configurations and other HVAC systems. Moreover, the software tool used to perform these comparisons is described. 2.1. The building-plant system The simulated HVAC system takes its basic characteristics and boundary conditions from a system serving an existing building sited in Venice. As a consequence, in the present subsection, this building and the related HVAC system are described. The building under restoration is the south part of a former convent named ‘‘Tolentini’’, now the main building of University IUAV of Venice. The building has a total volume of 9980 m3 and a total floor area of 1840 m2, and underwent a relevant restoration including the design of new HVAC systems, whereas building structures and external walls were not retrofitted because of historic preservation commitments. The related U-values (overall heat transfer coefficients of building transparent and opaque surfaces) are resumed in Table 1. Building energy simulations were performed considering people occupancy from 08:00 to 24:00 (library and reading rooms are open till midnight), from Monday to Friday, and the values of internal heat gains are resumed in Table 2. The weather conditions refer to IWEC (International Weather for Energy Calculations) file for Venice contained in the EnergyPlus weather file database. The building mainly consists in offices, provided with primary air ventilation system and fan-coil terminal units camouflaged in

Table 1 Main characteristics of the opaque and transparent building constructions. Type of building construction

U-value 2

(–)

(–)

(W/(m K))

Floors

Slab-on-grade Slab-above-grade Internal External Internal

1.30 0.26 1.68 0.51 1.91 0.45 1.80

Vertical walls Roofs Windows

Solar Heat Gain Coefficient (SHGC) (–)

Total area (m2)

– –

440 83 1320 2120 360 541 231

– – – 0.63

Table 2 Design levels for internal heat gains. Category of internal heat gain

People (people) Lighting (W/m2) Electrical appliances (W/m2)

Design level Aula magna

Reading rooms

250 25 10

310 25 5

– Water–water heat pump: o Heating mode: user side outlet temperature: 45 °C; external side inlet temperature: 10 °C. o Cooling mode: user side outlet temperature: 7 °C; external side inlet temperature: 30 °C. – Air–water heat pump:  Heating mode: user side outlet temperature: 45 °C; external side inlet temperature: 7 °C (wet bulb temperature: 6 °C).  Cooling mode: user side outlet temperature: 7 °C; external side inlet temperature: 35 °C. Both heat pumps are provided with Scroll compressors and use R410A as a working fluid. In this paper the performance of the main heat pump, i.e. the water–water heat pump coupled with the borehole field, is simulated depending on borehole field layout, as well as against other kinds of HVAC systems. The water–water heat pump capacity is small compared with the air–water one because of constraints imposed by the local Government in the borehole heat exchanger installation (especially: limited length of the boreholes, around 40 m, in order to avoid the remote risk of connections between different underground water reservoirs) and the limited soil surface available for the installation. However, even if the ground source heat pump cannot provide the entire heating/cooling load, it is considered as the main heat pump, then supported by the air–water heat pump, used for back-up integration. As a matter of fact, large part of the total heating/cooling capacity is assigned to the Aula Magna (250 seats), that is seldom occupied. As a consequence, the fraction of the total heating/cooling energy provided by the water–water heat pump is far over the ratio of its maximum heating/cooling capacity to the total installed heating/cooling capacity. Fig. 1 resumes the general scheme of the HVAC system. The borehole field consists in 18 vertical ground heat exchangers 40 m deep placed as shown in Fig. 2, with average mutual distance equal to 4 m. The borehole tubes consist in polyethylene pipes PE 100 PN 16 with DN 32 diameter. Horizontal connection pipelines are provided with an insulation layer and the top of the borehole is at a depth of 1 m. The layers constituting the soil were determined via a geological report based on soil logs and correlations derived from areas close to the site and described in [21,22], then refined by the results of the soil log carried out during a Ground Response Test (GRT). Moreover, several ground water reservoirs are present, laying on the top of each other. The related water flow strongly depends on the tide level in lagoon and on marine intrusion, with no clear correlation, so the entity and direction of the water flow are very difficult to be foreseen. The influence of rain is very modest indeed [21]. The presence of flowing water is advantageous on the yearly energy balance, due to the reduction of seasonal thermal drifts and the increase of the total thermal capacity of the ground around the boreholes. As a consequence, the present simulations were performed neglecting the presence of flowing water, thus allowing the calculation results to be considered on the safe side.

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L. Schibuola et al. / Energy Conversion and Management 76 (2013) 956–970 Table 3 Main characteristics of the heat pumps used in the simulated system. Role (–)

Kind (–)

Refr. fluid (–)

Compr. type (–)

Volume flow rates 3

Int. (m /s) Main Back-up

Water–water, coupled with borehole heat exchangers Air–water

R410A

Scroll

0.0025

R410A

Scroll

0.0092

Nominal capacity 3

Ext. (m /s)

Heating (kW)

Heating (–)

Cooling (–)

50

41

4.16

4.38

183

174

3.21

2.84

0.0035 20.0

Nominal COP

Cooling (kW)

Fig. 1. Main scheme of the plant and related characteristics.

Fig. 2. Layout of the borehole heat exchangers.

In brief, the soil consists in the layers listed in Table 4. The physical properties resumed in Table 4 were assumed according with [23]. The resulting average thermal conductivity, weighted on the layer thicknesses, is 1.91 W/(m K), equal to the value of the equivalent thermal conductivity of the ground resulting from the performed Ground Response Test. In this regard, the assessment of the uncertainty in this Ground Response Test (GRT) measurement is briefly discussed. The value of the expanded (for a 95%

level of confidence in a normal probability distribution) combined uncertainty achieved for the current Ground Response Test is equal to ±0.25 W/(m K), thus around 13%, in agreement with [24,25]. In detail, the uncertainty analysis of this GRT has been performed starting from the following equation:

kg;Eq ¼

Q_ B lnðt2 Þ  lnðt 1 Þ  ; 4  p  LB hB ðt2 Þ  hB ðt1 Þ

ð1Þ

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Table 4 Data about the layers constituting the soil (achieved by drilling in the site of construction and by GRT). Number of layer (–)

Soil composition (–)

Thickness/depths (m)/(m) ? (m)

Thermal conductivity (W/(m K))

Density (kg/m3)

Specific heat (J/(kg K))

Presence of acquifer (Yes/No)

1 2 3 4 5 6 7 8

Filling material (pebbles and clays) Silty clay Grey clay and sand Loamy grey sand Sand, fine Grey clay loam Sand, fine Sand, compact

2/0 ? 2 13/2 ? 15 3/15 ? 18 3/18 ? 21 3/21 ? 24 2/24 ? 26 10/26 ? 36 4/36 ? 40

2.0 1.6 1.6 1.6 2.3 1.6 2.3 2.3

2700 2700 2700 2700 2600 2700 2600 2600

890 1260 1260 1260 1070 1260 1070 1070

Yes No No Yes Yes No Yes Yes

with

2.2. First analysis: the borehole field behavior in case of high humidity soil and yearly balanced heat transfer

hB ðtÞ ¼

hB;In ðtÞ þ hB;Out ðtÞ ; 2

ð2Þ

according with [26], and

Q_ B ðtÞ ¼ V_ w  qw  cw  ðhB;In ðtÞ  hB;Out ðtÞÞ:

ð3Þ

As a consequence, the combined standard uncertainty of the Ground Response Test is calculated through Eq. (4), considering negligible uncertainty in the measure of time:

vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi    ffi u  u  1  lnðt2 Þlnðt1 Þ  u _ 2 þ  Q_ B  lnðt2 Þlnðt1 Þ  u 2 þ L;B u 4pLB hB ðt2 ÞhB ðt1 Þ Q;B 2 4pLB hB ðt 2 ÞhB ðt 1 Þ u  2 ¼u ;  Q_  t þ2  4pBLB  lnðt2 Þlnðt1 Þ 2  uh  ðhB ðt 2 ÞhB ðt 1 ÞÞ

ukg;Eq

ð4Þ

The first set of results shown in this paper exhibits the behavior of the borehole field in the conditions mentioned above against the variation in design parameters such as the borehole length (ranging from 30 m to 70 m) and mutual distance (ranging from 2 m to 5 m), as well as the tube configuration (single U-tube and double U-tube). The heat pump set-point temperatures for the water flowing in the HVAC system are set to 7 °C in cooling and to 45 °C in heating, according with typical supply temperatures to fan-coils. Moreover, two options for part-load control are considered: on–off and inverter-driven (down to 25% of the maximum speed) compressor. As a consequence, four main combinations of borehole heat exchanger and heat pump are determined: – Single U-tube borehole heat exchanger + on–off heat pump. – Double U-tube borehole heat exchanger + on–off heat pump. – Single U-tube borehole heat exchanger + inverter-driven heat pump. – Double U-tube borehole heat exchanger + inverter-driven heat pump.

with

uh B

sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi  2   pffiffiffi 1 1   ¼ 2    uhB  ¼ 2    uhB  2 2

ð5Þ

and

uQ_ B

rffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi  2  2   ¼ q  cw  ðhB;In ðtÞ  hB;Out ðtÞÞ  u _  þ 2  V_ w  q  cw  uh  w

V

w

B

ð6Þ Furthermore, the expanded uncertainty data resumed in Table 5 were assumed or found in catalogues for a 95% level of confidence. In the end, applying the previous equations and starting uncertainties to the values monitored during the current Ground Response Test, the expanded combined uncertainty (95% level of confidence) equal to ±0.25 W/(m K) mentioned above was achieved. An additional characteristic of the present paper consists in the peculiar yearly balanced heat transfer at the ground side encountered in the case under examination. In the next years, this will be a frequent condition indeed, due both to the relevant increase in building envelope heating performance and to the increasing need for comfort in summer.

When the double U-tube configuration is used, the mass flow rate is doubled, in order to avoid laminar flow regimes (and consequently low heat transfer coefficients) of the secondary fluid flowing in the borehole pipes. 2.3. Second analysis: compact ground source heat pumps against traditional HVAC systems The second set of results of this paper focuses on the comparison of ground source heat pump systems with more traditional types of HVAC systems, specified in Table 6, in order to estimate the advantage of GSHP systems against other technologies available in Mediterranean coastal cities when compact borehole installation areas are used. For this purpose, the reference borehole configuration used in the previous analysis (consisting in: double U-tubes; borehole length = 40 m; borehole spacing = 4 m) is used as a reference for the comparison with other HVAC systems.

Table 5 Uncertainties of the measurement devices used in the Ground Response Test. Device

Source of data

Expression of the expanded uncertainty for 95% level of confidence

Probability distribution

Thermometer – Pt100 Volume flow meter Meter

Catalogue Catalogue Assumption

0.2 + 0.05%  ReadValue (°C) 0.6%  ReadValue (m3/s) 1m

Normal Normal Normal

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L. Schibuola et al. / Energy Conversion and Management 76 (2013) 956–970 Table 6 Description of the plants considered in the present HVAC system comparison. Code

Plant description

A B C D E F

Invertible heat pump coupled with borehole heat exchangers and controlled by an on–off strategy Invertible heat pump coupled with borehole heat exchangers and provided with an inverter-driven compressor Invertible heat pump coupled with lagoon water and a cooling tower (for peak loads in summer) and provided with a double stage control Condensing boiler + Chiller coupled with a cooling tower and provided with a double stage control Invertible heat pump coupled with outdoor air and controlled by an on–off strategy Invertible heat pump coupled with outdoor air and provided with an inverter-driven compressor

The performance of the other HVAC systems involving heat pumps were calculated by means of the heat pump model developed by the authors and described in the next subsection. In case of heat pumps coupled with lagoon water, the values of the lagoon water temperature measured in year 2007 in the centre of Venice, at 1.5 m depth, were used. The performance of gas boilers was simulated by the interpolation of catalogue data of a 4-star level boiler available on the market, depending on hourly water outlet set-point temperature and Capacity Ratio. The calculations were performed taking into account the energy consumption due to auxiliary devices as well, assuming the following values: – Pump and fan efficiencies:  Hydronic pump average efficiency (motor + impeller): 55%.  Fan average efficiency (motor + impeller): 45%. – Pressure losses:  Total pressure drop along the ground source circuit (according with [27] and including the pressure drop in the heat exchanger at the external side of the heat pump): 70,000 Pa.  Total pressure drop along the lagoon water circuit (including 1.5 m water head for lagoon water suction and the pressure drop in the heat exchanger at the external side of the heat pump): 120,000 Pa.  Total pressure drop along the cooling tower circuit – water side (including the pressure drop in the heat exchanger at the external side of the heat pump): 90,000 Pa.  Total pressure drop along the cooling tower duct – air side: 100 Pa. 2.4. The heat pump model The simulations were performed by means of detailed software developed by the authors and aimed at modeling heat pumps and borehole heat exchangers. In the present subsection, the model developed for the prediction of the thermal behavior of heat pumps is described, whereas in the next subsection the one aimed at the simulation of borehole heat exchangers is explained. From here on the term heat pump will be used in a general way, to refer to heat pumps, chillers, and air-conditioners. The heat pump model used for the simulation of the performance of heat pumps is described in [28–30]. In brief, the developed model is aimed at the prediction of vapor compression based heat pumps, is implemented in Fortran and needs input data that may be found in specification sheets usually provided by manufacturers. It starts from catalogue data (in particular, performance rated under nominal conditions) to estimate parameters characterizing the heat pump. Then, exploiting these estimated parameters, the behavior of the heat pump is predicted per each time-step (usually 10 min or 1 h long) of the year-round simulation. The inaccuracy of the model is low (around 10%) [30], especially if one considers the low amount of data and specifications necessary to run a simulation. Therefore this model is very suitable for integration within

building energy simulation software. Moreover, unlike many other heat pump models contained in building energy simulation software, the model is based on a physical description of the thermodynamic cycle of the heat pump, instead of a mere numerical one, hence it is reliable even at operation conditions far from the nominal ones, whereas curve-fit heat pump models are not reliable out of available data ranges. The input data needed for the parameter estimation phase are mostly available in usual catalogues and are summarized in the following list: – Secondary fluids, at the internal and external sides (water– water, water–air, air–water, air–air). – Refrigerant fluid (R134a, R407C, and R410A are currently available). – Coefficients of performance (COP) rated under heating and cooling nominal conditions. – Values of mass flow rates and inlet and outlet temperatures for secondary fluids under heating and cooling nominal conditions. – Power consumption in auxiliary devices rated under heating and cooling nominal conditions. – Evaporator and condenser heat transfer effectivenesses, by means of a heat exchange index ranging from 0 to 10: the higher the rank, the lower the temperature difference between the refrigerant fluid and the secondary fluids under nominal boundary conditions, basing on Table 7. Sensitivity analyses in [28,30] showed that the value of the effectiveness index may be used to further improve accuracy, but with no dramatic effect on results. As a first step in the frame of each simulation, the numerical model starts from the inputs listed above to calculate the basic parameters concerning the main components of the heat pump, through the calculation procedure summarized below and performed for each operation mode (heating and cooling): (1) Computation of the power applied to the cycle by the compressor, under nominal conditions. (2) Calculation of the evaporation and condensation temperatures, depending on the kind of heat pump and on the operation mode (heating or cooling), using the evaporator and condenser heat transfer effectivenesses mentioned above.

Table 7 Ranges for temperature differences between secondary fluids and refrigerant fluid at evaporator and condenser in heating and cooling. Air

Evaporator, cooling case Condenser, cooling case Evaporator, heating case Condenser, heating case

Water

Min DT (°C)

Max DT (°C)

Min DT (°C)

Max DT (°C)

11 6 6 18

20 20 15 27

4 5 5 4

8 12 12 8

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(6) Calculation of enthalpy at reference point 3 (Fig. 3). (7) Determination of the isentropic efficiency of the compressor at test conditions. After the determination of the main heat pump parameters, the performance at each time-step is calculated through a steady state approach in the description of the thermodynamic cycle and a quasi steady state approach in the part-load management. The performance calculated by the model at each time-step is mainly expressed in terms of COP, compressor power and thermal capacities and temperatures at the evaporator and condenser. For that purpose, at each calculation time-step, the model is supplied with the following data:

Fig. 3. Main thermodynamic points of the reference cycle.

(3) Calculation of the thermal effectivenesses of the evaporator and condenser, per each running mode. (4) Calculation of the reference points of the nominal thermodynamic inverse cycle, basing on the evaporation and condensation temperatures previously computed and by adoption of the thermo-physical properties of the specific refrigerant fluid. The thermodynamic points of the thermodynamic cycle drawn by the software are shown in Fig. 3. (5) Calculation of the refrigerant fluid volume flow rate, considered constant in any operating condition, since the present model is aimed at the description of the thermal behavior of heat pumps driven by volumetric compressors.

– Mass flow rates of the secondary fluids. – Secondary fluid temperatures at the evaporator and condenser inlet. – Average thermal capacity needed at the user side. In the prediction of the heat pump performance for each simulation time-step, an iterative calculation procedure is performed. The main steps of the calculation are summarized as follows: (1) The model assumes the nominal thermal capacities as starting values of the heat flows transferred at the condenser and evaporator. (2) Based on such thermal capacities, the model calculates the condensation and evaporation temperatures of the working fluid. (3) The whole thermodynamic inverse cycle is traced.

Fig. 4. General scheme of heat pump model input/output.

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(4) The software compares the thermal capacities transferred by the current thermodynamic cycle through both the evaporator and the condenser with the ones assumed at the beginning of the current iteration. If their difference is within a predefined tolerance, then the process stops and the iteration loop is not performed any longer. Otherwise, the process begins a new iteration, setting the just calculated condenser and evaporator capacities as starting values and performing the iteration from step 2. At the end of the iterative process, the user side and ambient heat flows and the coefficient of performance at full load are obtained. To sum up, the general calculation flow and main input/output variables for each calculation phase are summarized in Fig. 4. At this point, the COP at part load conditions is calculated through the approach described in EN 14825:2012 [20], using a Part-Load Factor (PLF) depending on the type of part-load control (on–off or modulating by inverter-driven compressor) and on the Capacity Ratio (CR, i.e. the ratio of the needed thermal capacity to the full load thermal capacity). 2.5. The borehole heat exchanger model In this subsection, the model used to simulate borehole heat exchangers is described, together with its integration with the heat pump model shown in the previous subsection, thus constituting

963

the calculation tool used to calculate the performance of GSHP systems in this paper. Model CaRM (Capacity Resistance Model), described in [31,32], makes it possible to simulate the behavior of single U-tube, double U-tube, concentric pipes and energy pile heat exchangers. The solution of the temperature field is obtained via the electric analogy. The details of the model are described in [31,32], so in this subsection just a brief description of the model takes place. The heat transfer in the soil is considered in terms of heat conduction, whereas the effect of relevant underground water flow is not modeled, because consequent to phenomena of various kinds as well as difficult to predict. However, such a model is valid also for slow underground water flows [33], allowing the designer to be on the safe side, since the presence of underground water flow rates usually improves the performance of ground source heat pumps, due to the decrease of thermal drifts. Fig. 5 shows the general approach used in CaRM: the borehole heat exchanger is split into m slices along the borehole length. Each slice is then subdivided into n annular regions along the radius. For the j-th slice and the i-th annular region the following energy balance is considered:

hðj; i  1Þ  hðj; iÞ hðj; i þ 1Þ  hðj; iÞ þ Rðj; i  1Þ Rðj; iÞ ¼ Cðj; iÞ 

hðj; iÞ  hDs ðj; iÞ Ds

ð7Þ

(a)

(b)

(c)

Fig. 5. Scheme of the discretization of the thermal domain into thermal nodes in model CaRM, through the vertical section (a) and the horizontal one (b), and related symbols (c).

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where C is the thermal capacity of the specific soil subdivision (thermal node), R is the thermal resistance between two adjacent annular slices, h is the thermal node temperature at the end of the current calculation time-step and hDs is the temperature at the end of the previous calculation time-step. In this release the model does not take into account the axial thermal conduction in the ground because of its low influence in the present case, due to balanced thermal exchange at the ground side and short mutual distances of the borehole heat exchangers, according with [34]. Eq. (7) takes into account thermal diffusion in the radial direction and cylindrical symmetry is considered, so that the pipe wall temperature is considered uniform. The variation of the fluid temperature between inlet and outlet (around 4–5 °C) implies local deviations from the cylindrical thermal symmetry, indeed, but the related error was evaluated and considered as negligible [35]. The most external annular region of each slice is bounded by soil at undisturbed soil temperature (hg). For each slice the proper undisturbed temperature can be assigned, so that a user-defined temperature profile can be used as an undisturbed soil thermal condition, useful for sites where geothermal anomalies take place. The heat exchange between the fluid flowing in the borehole and the soil depends on the tube configuration (single U-tube, double U-tube or concentric pipes), on the pipe and grout shapes and dimensions, as well as on the thermal conductivity (and thermal capacity) of the ground and pipe. Each tube configuration implies specific thermal connections between the pipes, the grout, and the ground, able to resume the heat flows taking place on the horizontal section of the borehole. In the borehole heat exchanger model these connections are considered by means of specific thermal resistance networks, thus making the program able to simulate single U-tube heat exchangers as well as double U-tubes and concentric pipes. The specific values of the thermal resistances constituting each available configuration may be calculated by means of formulae available in literature, such as in [35,36], in case of typical characteristics, or, in case of special geometries and shapes, via finite volume or finite element software aimed at the calculation of 2-D thermal fields consequent to heat conduction, such as [37,38], and as performed in [32]. In Fig. 6 the model scheme for double U-tube boreholes is shown. The thermal resistances between the pipe and the external wall of the borehole are assumed as constant and they are input values for CaRM. Also the thermal convection within the tube might play an important role in the global behavior of the borehole heat exchanger, especially when laminar flow regimes take place. For this reason a detailed calculation of the convection heat transfer coefficient

Fig. 6. Scheme of the thermal connections modeled within the borehole heat exchanger (double U-tube): thermal resistances between adjacent tubes (RppA), thermal resistances between opposite tubes (RppB),and thermal resistances between tubes and ground (Rp0).

Fig. 7. Example of discretization of the thermal domain in case of contiguous boreholes.

Fig. 8. Flow chart of the calculation loop involving the borehole heat exchanger and the heat pump.

is carried out as well, depending on the current flow regime within the tubes (laminar, transition, or turbulent). The model allows the user to analyze the thermal behavior of the borehole field, here including the installation layout, taking into account the mutual thermal influence of adjacent boreholes. In this case study the most recurrent boundary condition consists in the adiabatic behavior at the interface placed between each borehole and the preceding and/or next one (Fig. 7). The thermal capacity of the borehole heat exchanger, due both to ground and to heat carrying fluid, is important when short timestep simulations are executed (i.e.: shorter than the borehole response time), thus the model was further improved, as shown in [39,40]. CaRM was integrated with the heat pump model described in the previous subsection, in order to get the complete simulation of the GSHP system, through an iterative calculation loop so that the tight interaction between the ground and the heat pump is analyzed. In Fig. 8, the flow chart of the resulting program is shown.

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Double U - On-Off

Single U - Inverter

Double U - Inverter

7.5 7.0 6.5

COP [-]

6.0 5.5 5.0 4.5 4.0 3.5 3.0 2.5 First

Second

Third

Fourth

Fifth

Sixth

Seventh

Year Fig. 9. Examples of trend of the seasonal heating COP along a seven year simulation.

Single U - On-Off

Double U - On-Off

Single U - Inverter

Double U - Inverter

7.5 7.0 6.5

COP [-]

6.0 5.5 5.0 4.5 4.0 3.5 3.0 2.5 First

Second

Third

Fourth

Fifth

Sixth

Seventh

Year Fig. 10. Examples of trend of the seasonal cooling COP along a seven year simulation.

3. Results and discussion In the two following subsections, the first analysis, regarding the borehole field behavior in case of high humidity soil and yearly balanced heat transfer, and the second analysis, pertaining the comparison of ground source heat pumps against traditional HVAC systems, are presented respectively.

3.1. First analysis: the borehole field behavior in case of high humidity soil and yearly balanced heat transfer In this subsection, the behavior of GSHPs in case of high humidity soil and yearly balanced heat transfer at the ground side is simulated for various configurations and parameters, in order to determine the degree of variation of the performance ensured by GSHPs under such boundary conditions. The main set of characteristics used for these simulations is described in detail in Sections 2.1 and 2.2. The results of system simulations are shown in Figs. 9 and 10 for mean seasonal COP in heating and in cooling respectively, along a 7-year period, thus showing possible thermal drifts. From Figs. 9

Fig. 11. Example of thermal energy exchange at the user and borehole sides in heating and cooling seasons.

and 10 it is clear that in the present context the system performance is rather constant year after year. In a few words, the soil undergoes a limited thermal drift and the warm-up period is really

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short (1–2 years). That derives from the peculiar even heat balance at the ground side characterizing the present case study and paper. In a few words, the yearly balance of the heat removed from and supplied to the soil in winter and in summer respectively is almost even. As a matter of fact, the building yearly heating and cooling needs are not balanced, but the corresponding amounts of heat exchanged with the soil are balanced, because of the compressor contribution, as can be seen from Fig. 11 (referring in particular to the following configuration: borehole kind = double U; borehole length = 40 m; borehole spacing = 4 m; heat pump control mode: inverter-driven compressor). Moreover, the relevant thermal conductivity of the soil due to its high humidity level allows the borehole to quickly discharge and redistribute thermal energy. The results from the performed simulations allow the HVAC planner to estimate the relevant benefits consequent to the use of the most efficient borehole configuration, consisting in double U-tube borehole heat exchangers. In winter, the mean COP value is between 4.0 (single U-tube + on–off control mode) and 5.3 (double U-tube + inverter-driven compressor), whereas in summer the

mean COP is in the range from 3.6 (single U-tube + on–off control mode) and 5.4 (double U-tube + inverter-driven compressor). Considering the configuration with single U-tube borehole coupled with on–off heat pumps as the reference, the use of double U-tube shows COP 11% and 28% higher in heating and in cooling respectively. On the other hand, the use of inverter-driven heat pumps may increase the COP of about 15% (in detail: 15% and 12.5% in heating and cooling respectively in case of single U-tube borehole heat exchangers, and about 18% both in heating and in cooling in case of double U-tube). As an example of the results achieved through the present simulation tool, the performance expected during July is shown in Figs. 12 and 13 in case of an on–off heat pump coupled with double U-tube heat exchangers (borehole length = 40 m; borehole spacing = 4 m). The profile of COP during the day reaches high values, in the first hours of operation in the day, with thermally discharged soil and at full load capacity, with a peak around 5. Then, after some hours, the soil thermally saturates and part-load control takes place too. These events make the COP lower. Then, during

COP

7.5

Pu

50

7.0

45

6.5

40

6.0

35

5.5

30

5.0

25

4.5

20

4.0

15

3.5

10

3.0

5

2.5

Pu [kW]

COP [-]

Double U - On-off (July)

0 1

3

5

9

11

16

18

22

24

29

31

Days [-] Fig. 12. Expected values of COP and thermal capacity during July.

T_Inlet

45.0

T_Outlet

T_Outdoor

Double U - On-off (July)

40.0

Temperature [°C]

35.0

30.0

25.0

20.0

15.0

10.0 1

2

3

4

5

8

9

10

11 15

16

17

18

19

22

23

24

25

29

30

31

Days [-] Fig. 13. Outdoor air temperature and expected values of water inlet and outlet temperatures at the borehole side during July.

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Double U -On-Off

Single U -Inverter

Double U -Inverter

7.5 7.0 6.5

COP [-]

6.0 5.5 5.0 4.5 4.0 3.5 3.0 2.5 30

40

50

60

70

Length [m] Fig. 14. Seasonal heating COP versus borehole length.

Single U - On-Off

Double U - On-Off

Single U - Inverter

Double U - Inverter

7.5 7.0 6.5

COP [-]

6.0 5.5 5.0 4.5 4.0 3.5 3.0 2.5 30

40

50

60

70

Length [m] Fig. 15. Seasonal cooling COP versus borehole length.

Heating

7.5

Cooling

7.0 6.5

COP [-]

6.0 5.5 5.0 4.5 4.0 3.5 3.0 2.5 2

3

4

Borehole distance [m] Fig. 16. Seasonal heating and cooling COP versus borehole mutual distance.

5

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the night, the soil thermally discharges and reaches again good conditions for heat rejection in the following day. Outlet temperatures from the borehole heat exchanger between 30 °C and 35 °C are encountered in the most critical days. Moreover, the sudden decrease in water inlet/outlet temperatures taking place on the 11th day comes from the sudden decrease in outdoor air temperatures (Fig. 13) and the consequent decrease in the cooling demand (Fig. 12). As a matter of fact, the chiller is requested to deliver lower cooling capacity and the water temperatures along the borehole heat exchanger get closer to the ground temperature, since lower temperature differences between the borehole and the ground are needed in order to discharge the heat supplied by the chiller to the borehole heat exchanger loop. Figs. 14 and 15 show the variation on performance consequent to borehole length variation. Longer borehole heat exchangers imply higher COP values, of course, and the increase in performance is even more relevant when the most performing configuration is

considered (i.e.: double U-tube borehole heat exchanger + inverter-driven heat pump; borehole length = 40 m; borehole spacing = 4 m), resulting in an increase of 17% and 74% in COP in heating and in cooling respectively for an increase in borehole length from 30 m up to 70 m. The large increase in the average cooling COP is given by the relevant occurrence of cooling peak values during the summer operation. This implies a fast saturation of the soil surrounding the borehole when using short borehole lengths, whereas longer borehole lengths avoid such a thermal saturation and allow the heat pump to keep high values of the cooling COP for the entire day. In the end, Fig. 16 shows the variation in COP consequent to various choices of borehole mutual distance, in the case of on–off heat pump coupled with double U-tube borehole heat exchangers. The figure shows that the increase of the borehole mutual distance does not influence relevantly the performance, because of the yearround even heat transfer balance.

INPUT ENERGY CONSUMPTION RATIO [-]

Reference level of INPUT ENERGY CONSUMPTION RATIO [-]

PRIMARY ENERGY CONSUMPTION RATIO [-]

Reference level of PRIMARY ENERGY CONSUMPTION RATIO (auxiliary devices included) [-]

PRIMARY ENERGY CONSUMPTION RATIO (auxiliary devices included) [-]

6.00 5.00

Ratio [-]

4.00 3.00 2.00 1.00 0.00

A

B

C

D

E

F

Plant Fig. 17. Input energy consumption ratio and primary energy consumption ratio in the heating season.

INPUT ENERGY CONSUMPTION RATIO [-]

Reference level of INPUT ENERGY CONSUMPTION RATIO [-] PRIMARY ENERGY CONSUMPTION RATIO [-]

Reference level of PRIMARY ENERGY CONSUMPTION RATIO (auxiliary devices included) [-]

PRIMARY ENERGY CONSUMPTION RATIO (auxiliary devices included) [-]

6.00 5.00

Ratio [-]

4.00 3.00 2.00 1.00 0.00

A

B

C

D

E

F

Plant Fig. 18. Input energy consumption ratio and primary energy consumption ratio in the cooling season.

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3.2. Second analysis: compact ground source heat pumps against traditional HVAC systems In this subsection, the behavior of GSHPs designed for high humidity soil and yearly balanced heat transfer at the ground side are compared against other available and more common HVAC systems, in order to determine whether GSHPs maintain their own advantage in energy efficiency even in case of more compact footprint area. The main set of characteristics used for these simulations is described in detail in Sections 2.1 and 2.3. In Figs. 17 and 18 the values of input energy consumption ratio (ENeeds/EInput) and primary energy consumption ratio (ENeeds/EPrimary, with primary energy conversion efficiency equal to 0.4 in case of electricity as an input energy) are shown for each type of system and for heating and cooling purposes: the higher the value, the better the plant efficiency. In particular, the primary energy consumption ratio is calculated in two ways: taking into account just the primary energy behind the delivered thermal energy and taking into account even the primary energy consumption due to auxiliary devices. The results achieved in the reference HVAC system (system D, i.e. the most traditional HVAC system) are highlighted by means of a broken line and used as a benchmark, referring to input energy consumption ratio and primary energy consumption ratio (thus including auxiliary devices). Fig. 17 shows that the best performance in heating is achieved by system B, whereas system A and system C score at the same level, with primary energy consumption ratios 7.5% and 12.5% lower respectively. In particular, in case of system B, the primary energy consumption ratio is double compared with heating by traditional boilers (system D). At an intermediate stage, the air–water invertible heat pumps show energy efficiencies higher than system D, but about 30% lower than in system B. In particular, on–off air– water heat pumps (system E) scores a value of primary energy ratio 30% lower than system B, whereas inverter-driven air–water heat pumps (system F) touches a primary energy ratio about 25% lower than system B. Fig. 18 shows more even results than in the case of heating. As a matter of fact, system B scores the best performance (primary energy ratio 50% higher than in system D), whereas systems A and C perform close to air–water inverter-driven heat pumps, resulting around 13% less efficient than system B 4. Conclusions The study was aimed to estimate the performance expectable from ground source heat pumps in case of high humidity soils and yearly balanced heat transfer at the ground side, taking advantage of an available case-study in Venice. In this paper a parametrical analysis was carried out, varying the main design parameters, in order to assess the sensitivity in performance in this context. In particular, the analysis determined the combination of double Utube borehole heat exchangers coupled with inverter-driven heat pumps as the best performing configuration among the simulated ones, with an increase of 14% and 15% in heating and cooling COP respectively comparing with the second best configuration. Moreover, the sensitivity of results due to borehole length and distance was calculated, showing that in case of yearly even heat transfer balance at the ground side negligible thermal drift takes place, hence the mutual borehole distance may be varied with no relevant variation in performance. The increase in borehole length has an amplifying effect that implies that larger variations in COP take place in case of the best GSHP configuration, with an increase of 17% and 74% in heating and cooling COP respectively, ranging from 30 m up to 70 m borehole length. In the second part of the paper, the reference GSHP configuration simulated in the first part of the paper was compared against

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other HVAC systems, expressing the consequent results in terms of primary energy consumption, including auxiliary devices. As a result, inverter-driven heat pumps coupled with borehole heat exchangers achieved the best performance, both in heating and in cooling. On–off heat pumps coupled with borehole heat exchangers and heat pumps coupled with lagoon water show good performance in winter (in terms of primary energy consumption ratio, around 7.5% and 12.5% respectively lower than the best GSHP configuration, including auxiliary devices), whereas the summer performance is comparable with inverter-driven air–water heat pumps, showing a 15% lower performance in terms of primary energy consumption ratio, when compared with inverter-driven heat pumps coupled with double U-tube borehole heat exchangers. In the heating period, the worst performance is achieved by the traditional heating plant, consisting in condensing boilers. In fact, that reaches double primary energy consumption compared with the most efficient systems, even if the electricity consumed by the auxiliary devices serving the borehole circuit is taken into account.

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