Performance of an experimental ground-coupled heat pump system [600539]

Performance of an experimental ground-coupled heat pump system
for heating, cooling and domestic hot-water operation
Calin Sebarchievici, Ioan Sarbu*
Department of Building Services Engineering, Polytechnic University Timisoara, Piata Bisericii 4A, 300233 Timisoara, Romania
article info
Article history:
Received 3 June 2014Accepted 8 November 2014Available online
Keywords:
Geothermal energyGCHP system
Optimised adjustment
Experimental measurementsSystem performancesSimulation modelsabstract
The ground-coupled heat pump (GCHP) system is a type of renewable energy technology providing space
heating and cooling as well as domestic hot water. However, experimental studies on GCHP systems are
still insuf ficient. This paper first presents an energy-operational optimisation device for a GCHP system
involving insertion of a buffer tank between the heat pump unit and fan coil units and consumer supply
using quantitative adjustment with a variable speed circulating pump. Then, the experimental mea-
surements are used to test the performance of the GCHP system in different operating modes. The mainperformance parameters (energy ef ficiency and CO
2emissions) are obtained for one month of operation
using both classical and optimised adjustment of the GCHP system, and a comparative analysis of these
performances is performed. In addition, using TRNSYS (Transient Systems Simulation) software, two
simulation models of thermal energy consumption in heating, cooling and domestic hot-water operationare developed. Finally, the simulations obtained using TRNSYS are analysed and compared to experi-
mental data, resulting in good agreement and thus the simulation models are validated.
©2014 Elsevier Ltd. All rights reserved.
1. Introduction
Ground-coupled heat pump (GCHP) system is a type of renew-
able energy technology and is popular for air-conditioning and
domestic hot-water worldwide [1e3]. A number of GCHP systems
have been used in residential and commercial buildings worldwide
because of their noticeable high ef ficiency and environmental
friendliness. Due to the heat capacity of ground, ambient air tem-
perature variations are directly re flected only on the surface ground
temperature, their effect being reduced at deeper layers. According
to the reports of the 2010 World Geothermal Congress, the GCHP
systems have the largest energy use and installed capacity, ac-
counting for 69.7% and 49.0% of the worldwide capacity and use.
The installed capacity is 35,236 MWt and the annual energy use of
214,782 TJ/yr, with a capacity factor of 0.19 (in the heating mode).
Almost all of the installations occur in North American, Europe and
China, increasing from 26 countries in 2000, to 33 countries in
2005, to the present 43 countries. Sweden, Denmark, Switzerland,
Austria, and the United States are the leaders in this field[4]. The
number of installed GCHP systems has grown continuously by10e30% annually in recent decades [5e7]. Extrapolating currently
observed growth rates for Europe of 5.4 million heat pump units
per year leads to an expectation of 70 million installed units in
Europe by 2020 [8]. The use of GCHPs in the achievement of
adequate temperatures has been studied by several researchers
[9e12]
A GCHP system consists of a heat pump unit coupled with a
ground heat exchanger (GHE), usually a vertical borehole heat
exchanger (BHE) or, less commonly, horizontal loops [13]. A BHE is
commonly drilled to a depth between 20 and 300 m with a diam-
eter of 100 e200 mm. A closed single or double U-tube is often
inserted inside the borehole, and a heat carrier fluid is circulated in
theU-tube to exchange heat or cold with the surroundings. For
safety and stability reasons, a bentonite-cement suspension or an
enhanced-cement is used to back fill the space between the U-tube
and its surrounding soil/rock. A GCHP utilises the ground as a heat
source in heating and a heat sink in cooling mode operation. In the
heating mode, a GCHP absorbs heat from the ground and uses it to
heat the building. In the cooling mode, heat is absorbed from the
conditioned spaces and transferred to the earth through its GHE.
Most existing studies of GCHP systems concentrate on theo-
retical and simulation model research [7,14e16]or in situ moni-
toring of the heat transfer in BHE [5,6,17 e21]. Only a few
researchers have investigated the experimental operation*Corresponding author. Tel.: ț40 256403991; fax: ț40 256403987.
E-mail address: ioan.sarbu@ct.upt.ro (I. Sarbu).
Contents lists available at ScienceDirect
Renewable Energy
journal homepage: www.elsevier.com/locate/renene
http://dx.doi.org/10.1016/j.renene.2014.11.020
0960-1481/ ©2014 Elsevier Ltd. All rights reserved.Renewable Energy 76 (2015) 148 e159

performance of GCHP systems. Hwang et al. [19] presented the
actual cooling performance of a GCHP system installed in Korea for
1 day of operation. Pulat et al. [22] evaluated the performance of a
GCHP with a horizontal GHE installed in Turkey under winter cli-
matic conditions. The coef ficient of performance (COP) of the entire
system and the heat pump unit were found to be 2.46 e2.58 and
4.03e4.18, respectively. Yang et al. [23] reported the heat transfer of
a two-region vertical U-tube GHE after an experiment performed in
a solar geothermal multifunctional heat pump experimental sys-
tem. Lee et al. [24] conducted experiments on the thermal perfor-
mance of a GCHP integrated into a building foundation in summer.
Man et al. [25] performed an in situ operation performance test of a
GCHP system for cooling and heating provision in a temperate zone.
The experimental results indicate that the performance of the
GCHP system is affected by its intermittent or continuous operation
modes. Petit and Meyer [26] compared the thermal performances
of a GCHP with an air source air conditioner, finding that a hori-
zontal or vertical GCHP was more favourable in terms of economic
feasibility. Esen and Inalli [27] proposed using the in situ thermal
response test to determine the thermal property of the ground for
the GCHP applications in Turkey, and they found that the thermalconductivity and effective thermal resistance of the ground vary
slightly with depth.
The present paper is focused on the energy and environmental
analysis and modelling of a geothermal experimental plant from a
continental temperate climate, located in an institutional building
at the Polytechnic University of Timisoara, Romania. The system
consists of a reversible GCHP. One the main innovative contribution
of this study consists in the achievement and implementation of an
energy-operational optimisation device for the GCHP system using
quantitative adjustment with a buffer tank and a variable speed
circulating pump. The experimental measurements are used to test
the performances of the GCHP system at different operating modes.
The main performance parameters (energy ef ficiency and CO
2
emissions) are obtained for 1 month of operation using both clas-
sical and optimised adjustment of the GCHP system. A comparative
analysis of these performances for both heating and cooling and
domestic hot-water (DHW) with different operation modes is
performed. The second purpose of this paper is to develop two
simulation models of thermal energy consumption in heating/
cooling and DHW operation using TRNSYS software. Finally, the
simulations obtained using TRNSYS software are analysed and
compared to experimental measurements.2. Description of experimental laboratory
Experimental investigations of GCHP performance were con-
ducted in a laboratory ( Fig. 1 ) at the Polytechnic University of
Timisoara, Romania, located at the ground floor of the Civil Engi-
neering Faculty building with six floors and a heated basement. The
city has a continental temperature climate with four different
seasons. The heating season runs in Timisoara from 1 October to 30
April, and the cooling season runs from 1 May to 30 September. The
laboratory room has an area of 47 m2, and its height is 3.70 m. The
envelope (external walls) is made of 200 mm porous brick with a
100 mm thermal insulating layer and 20 mm lime mortar. The
thermal transmittances ( U-values) are as follows: walls 0.345 W/
m2K and double-glazed windows 2.22 W/m2K. The area of the
windows is 16 m2, and the area of the interior door is 2.1 m2. The
indoor air design temperature is 20/C14C for the heating season and
26/C14C for the cooling season. The outdoor air design temperature
is/C015/C14C for the heating season and 32.6/C14C for the cooling season.
The GCHP installed in this experimental laboratory heated and
cooled through a fan coil system. With the mentioned input data, a
heating load of 3.11 kW and a cooling load of 2.15 kW were ob-tained. The laboratory area was assimilated with a three-person
apartment area in Timisoara. Considering the DHW daily mean
consumption of 50 l/person, a tank hot-water temperature of 45
/C14C
and a cold water temperature of 20/C14C, a DHW load of 4.36 kW was
determined. Fig. 2 illustrates the monthly energy demand for lab-
oratory heating (positive values) and cooling (negative values).
3. Description of the experimental system
The GCHP experimental system consisted of a BHE, heat pump
unit, buffer tank, circulating water pumps, fan coil units, sink, data
acquisition instruments and auxiliary parts, as shown in Fig. 3 . The
heat carrier fluid can be delivered towards two fan coils units in two
flow rate adjustment modes:
/C0direct, by a recirculation pump connected inside of the heat
pump unit of the GCHP system (classical solution);
/C0indirect, by a fixed speed circulating pump connected to a buffer
tank. The GCHP automation can control the operation of the
circulating pump connected to the buffer tank by on/off switch-
ing. This assembly improves the entire system operation. The
buffer tank allows decreasing the GCHP on/off switching because
Fig. 1. Experimental laboratory.
Fig. 2. Monthly energy demand for laboratory heating/cooling.C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 149

of its thermal inertia, and thus, the energy ef ficiency increases.
The solution for heat carrier fluidflow rate adjustment was
optimised using an automatic control device of circulating pump
speed [28]. The main components of an automatic device for
pump speed control are shown in Fig. 4 [29] .
InFig. 5 , a schematic of the automatic control device of the
circulating pump speed according to heating/cooling demand of
the room is presented. In comparison with the classical solution, in
which the circulating pump on/off switching is controlled by theGCHP automation, the optimised solution assures both the on/off
switching and the speed control of the circulating pump.
The temperature difference between the inside and outside of
the heated/cooled space is measured by temperature sensors TS1
and TS2 connected to a programmable logic controller (PLC). In the
PLC internal memory, a computational algorithm of the frequency
converter dependent on the measured temperature difference is
implemented. The PLC sends to the frequency converter the cor-
responding frequency value to ensure the fluidflow rate according
to heating/cooling load at that time. In addition, the PLC allows on/
off switching control of the circulating pump. For simultaneous on/
off switching of the circulating pump and GCHP, common tem-
perature values of this process were established.
3.1. BHE
The GHE of this experimental GCHP consisted of a simple vertical
borehole that had a depth of 80 m. Antifreeze fluid (30% ethylene
glycol aqueous solution) circulates in a single polyethylene U-tube of
32 mm internal diameter, with a 60 mm separation between thereturn and supply tubes, buried in borehole. The borehole overall
diameter was 110 mm. The borehole was filled with sand and
finished with a bentonite layer at the top to avoid intrusion of pol-
lutants in the aquifers. The average temperature across the full
borehole depth tested was 15.1
/C14C. The ground characteristics are
based on measurements obtained from the Banat Water Resources
Management Agency [30]. The average thermal conductivity and
thermal diffusivity of the ground from the surface to 80 m deep
tested were 1.90 W/(m K) and 0.79 /C210/C06m2/s, respectively [28].
The boreholes were completely back filled with grout mixed with
drilling mud, cement and sand in speci ficp r o p o r t i o n s .T h et h e r m a l
conductivity and thermal diffusivity of the grout tested by manu-
facturer were 2.32 W/(m K) and 0.93 /C210/C06m2/s, respectively.
3.2. Heat pump unit
The heat pump unit was a reversible ground-to-water scroll
hermetic compressor unit with R410A as a refrigerant. The nominal
heating and cooling capacities were 6.5 kW (35/C14C supply/0/C14C re-
turn) and 3.8 kW (23/C14C return/15/C14C supply), respectively. The heat
pump unit was a compact type model having an inside refrigeration
system and DHW tank with a 175 L capacity. The operation of the
heat pump was governed by an electronic controller, which,
depending on the building water return temperature, switched the
heat pump compressor on or off. The heat source circulation pump
was controlled by the heat pump controller, which activates the
source pump 30 s before compressor activation.
The operation of the heat pump is characterised by the coef fi-
cient of performance (COP hp), which is de fined as the ratio between
the useful thermal energy Etand electrical energy consumption of
heat pump Eel:
COPhp¼Et
Eel(1)
The energy ef ficiency ratio (EER hp) is analogous to the COP hpbut
relates the cooling performance.
The coef ficient of performance of heat pump in cooling mode is
obtained by the following equation:
COPhp¼EERhp
3:412(2)
Fig. 3. Schematic of the experimental GCHP system with optimised flow rate adjustment.
Fig. 4. Schematic of a control device of the variable speed pump.C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 150

where value 3.412 is the transformation factor from Watts to Btu/h
(British Thermal Units per hour).
The COP hpcan be expressed also in terms of the temperature of
the hot environment ( Th) and the temperature of the cool envi-
ronment ( Tl)[13]:
COPhp¼1
Th
Tl/C01(3)
The COP of the GCHP system (COP sys)i sd e fined by Eq. (1), where
Eelis the energy consumption of the GCHP system, which includes
the energy consumption of the compressor of heat pump unit,
circulating pumps, fan coil units, frequency converter, and PLC.
The carbon dioxide (CO 2) emission CCO2of the heating system
during its operation is calculated with the following equation:
CCO2¼gelEel (4)
where gel¼0.547 kg CO 2/kWh is the speci ficC O 2emission factor
for electricity [31].
To obtain the COP hpor COP sysand CO 2emission, it is necessary
to measure the heating/cooling energy and electricity used by the
heat pump unit or the GCHP system.
3.3. Circulating water pumps
The water circulating loops of the GCHP consisted of a GCHP-
buffer tank water loop and buffer tank-fan coil unit water loop. Twocentrifugal pumps with rated flow of 2.8 m3/h and 5.5 m3/h were
chosen for the first and the second water circulating loop, respec-
tively. The first circulating pump ( fixed speed circulating pump con-
nected to heat pump unit) was controlled by the GCHP automation,
and the second pump (variable speed circulating pump connected to
buffer tank) was controlled by an automatic control device.
3.4. Fan coil units
Two parallel connected fan coil units were utilised as terminal
units of the GCHP. The total thermal power of these two fan coil
units was 3.2 kW.
3.5. GCHP data acquisition system
The GCHP data acquisition system consists of the indoor and out-
door air temperature, supply/return temperature, heat source tem-
perature (outlet BHE temperature), DHW temperature, relative air
humidity, and main operating parameters of the system components.
4. Measuring apparatus
A network of sensors was setup to allow monitoring of the most
relevant parameters of the system. Two thermal energy meters
were used to measure the thermal energy produced by the GCHP
and the extracted/injected thermal energy to the ground. A thermal
energy meter was built with a heat computer, two PT500
Fig. 5. Schematic of the automatic control device for the circulating pump.C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 151

temperature sensors and an ultrasonic mass flow meter. The two
PT500 wires temperature sensors with an accuracy of ±0.15/C14Cw e r e
used to measure the supply and return temperature for a hydraulic
circuit (the water-antifreeze solution circuit or the fan coil circuit).
Also, an ultrasonic mass flow meter was used to measure the mass
flow rate for a hydraulic circuit. The thermal energy meters were
AEM meters, model LUXTERM, with a signal converter IP 67 and
accuracy <0.2%. A three-phase electronic electricity meter
measured the electrical energy consumed by the system (the heat
pump unit, the circulating pumps, a feeder 220 Vca/24 Vcc, a fre-
quency converter and a PLC) and another three-phase electronic
electricity meter measured the electrical energy consumed by the
heat pump compressor. The two three-phase electronic electricity
meters were multifunctional type from AEM, model ENERLUX-T,
with an accuracy grade in ±0.4% of the nominal value. The moni-
toring and recording of the experiments were performed using a
personal computer (PC). The indoor and outdoor air temperature
was measured by AFS sensors and supply/return, heat source and
DHW temperature was recorded by PTC immersion sensors, all
connected to the GCHP data acquisition system and having an ac-
curacy of ±0.2/C14C.
5. Laboratory experiment results
The system was monitored for two years. The experimental
measurements were performed for two cases of flow rate adjust-
ment of the heat carrier fluid in the system: case (1) eclassical
adjustment, by fixed speed circulating pump connected to heat
pump unit and case (2) /C0optimised adjustment, by variable speed
circulating pump connected to buffer tank.
In heating operation, the experiments were conducted for a
one-month period for each of the two analysed cases, from the22nd of January 2012 to the 20th of February 2012 and the 15th of
January 2013 to the 13th of February 2013. The outdoor tempera-
ture varied in the range of /C05.9 to 10.1/C14C. The monthly mean values
of the outdoor temperature during the two periods were almost
equal. In cooling operation, the experiments were conducted for a
one-month period for each of the two analysed cases, from the 21st
of May 2012 to 19th May 2012 and the 28th of May 2013 to the 26th
of June 2013. The outdoor temperature varied in the range of
15.2e34.8/C14C. The monthly mean values of the outdoor tempera-
ture during the two periods were almost equal.
5.1. The GCHP system performances in different operation modes
5.1.1. Energy performance and CO 2emissions of the GCHP system in
heating operation
The experimental parameters including indoor air temperature
(ti), outdoor air temperature ( te), and heat source temperature ( ts),
recorded for a one-month period, are plotted in Figs. 6 and 7 .I t
should be noted that in case (2) a reduced indoor air temperature
was obtained, around the set-point temperature of 22/C14C, leading to
better comfort in the room. In addition, the reduced oscillation of
the antifreeze fluid temperature leads to a lower heat source de-
mand. The heat source temperature in winter is up to 12 e13/C14C
higher than the outdoor air temperature, this increases the capacity
and the ef ficiency of GCHP system.
Table 1 presents the summary of the mean values of the tem-
peratures ( ti,te,ts), electricity consumption ( Eel), useful thermal
energy for heating ( Et), COP of the GCHP system (COP sys) and the
heat pump unit (COP hp), and CO 2emission ( CCO2). The COP hpvalues
of the heat pump unit for classical and optimised solution are 4.82
and 5.06, respectively. For case (2), the COP sys¼4.40 is 7.5% higher
and the CO 2emission level is 7% lower than in case (1). Due to the
properties of the climate, ground etc. of the place where the
measurements were conducted and a higher underground water
flow rate, the heat source temperature is increased and the COP hp
and COP sysare notable values for both solutions. Therefore when
these results were compared with results of the similar studies
reported in Refs. [9,32] but in other local geothermal conditions, it
is seen that the performance values obtained here are improved
noticeably.
Fig. 6. Recorded indoor and outdoor air temperature during heating operation.
Fig. 7. Heat source temperature evolution during the heating provision tests.Table 1
GCHP system performance for classical and optimised adjustment in heatingoperation.
Case t
i[/C14C]te[/C14C]ts[/C14C]Eel[kWh] Et[kWh] COP sysCOP hpCCO2
[kg]
(1) Classic 22.65 3.25 16.24 125.18 510.62 4.07 4.82 50.45
(2) Optimised 21.84 3.76 17.08 116.47 512.54 4.40 5.06 46.94
Fig. 8. Outdoor air temperature evolution during the cooling provision tests.C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 152

5.1.2. Energy performance and CO 2emissions of the GCHP system in
cooling operation
Figs. 8 and 9 show the variation in time of the outdoor air
temperature ( te), indoor air temperature ( ti), and heat source
temperature ( ts). In solution (2), a more reduced indoor air tem-
perature around a set-point temperature of 26/C14C was obtained,
leading to better comfort in the room. Table 2 summarises the
mean values of the temperatures ( ti,te,ts), electricity consumption
(Eel), useful thermal energy for cooling ( Et), EER of GCHP system
(EER sys) and CO 2emission ( CCO2). In case (2), the EER syswas 8%
higher, and the CO 2emission level was 8% lower than in case (1).
The comparison of the GCHP system experimental perfor-
mances in the heating and cooling operation ( Tables 1 and 2 )
indicate that the system performance in heating and cooling
operation was almost equal. This is because the heating load was
higher than the cooling load, and in addition, the electricity con-
sumption in heating operation was higher than the electricity
consumption in cooling operation.
The COP values of the GCHP system were compared to the
existing COP values applied in GCHP research. The research of Man
et al. [25] revealed a COP of GCHP systems of 4.19 e4.57 in the
winter season and 3.9 e4.53 in the summer season. In addition, the
summer research of Michopoulos et al. [18], which used a vertical
heat exchanger at a depth of 80 m, reported a COP sysof 4.4e4.5. It is
seen that the performance values obtained here are similar.
5.1.3. The GCHP system performance in heating and DHW operation
For the comparative study of the two flow rate adjustment
cases, the same DHW volume was used, Vdhw¼1.22 m3. The
experimental parameters including indoor air temperature ( ti),
outdoor air temperature ( te), DHW temperature ( tdhw) and heat
source temperature ( ts), recorded for a week period for each of the
two analysed cases, from 12th of January 2012 to 18th of January
2012 and 7th of January 2013 to 13th of January 2013, are plotted in
Figs. 10 and 11 . The heat source temperature in winter is approxi-
mately 15/C14C higher than the outdoor air temperature.
Table 3 presents the summary of the mean values of the tem-
peratures ( ti,te,tdhw,ts), DHW volume ( Vdhw), electricity con-
sumption ( Eel), useful thermal energy ( Et), COP of the GCHP system
(COP sys) and the heat pump unit (COP hp), and CO 2emission ( CCO2).Although the COP sysresulted in almost equal values for the two
cases, the experimental results indicate that when using an auto-
matic control device for the circulating pump speed, an electricity
savings of 3% and a CO 2emission level decrease of 3% were
obtained.
Analysing the experimental data ( Tables 1 and 3 ) results indi-
cate that the COP sysof the GCHP system operating in heating and
DHW mode in comparison with heating operation mode decreased
significantly in the range of 20.6 e23.9% in comparison to the two
cases, from 4.07 e4.40 to 3.23 e3.35, respectively. The COP hpvalues
of the heat pump unit for cases (1) and (2) are 3.81 and 3.95,
respectively.
5.1.4. The GCHP system performance in cooling and DHW operation
To determine the GCHP system performance in the summer
season, experimental measurements over a week period for each of
the two analysed cases were performed, from 27th of June 2012 to
3rd of July 2012 and 24th of June 2013 to 30th of June 2013. During
the measurements, both the cooling and DHW load for a family
using a DHW volume Vdhw¼1.36 m3were assured. Fig. 12 illus-
trates the evolution of indoor air temperature ( ti) and outdoor air
temperature ( te), and Fig. 13 illustrates the evolution of the DHW
temperature ( tdhw) and heat source temperature ( ts).
Table 4 presents a summary of the mean values of the temper-
atures ( ti,te,tdhw,ts), DHW volume ( Vdhw), electricity consumption
(Eel), useful thermal energy ( Et), COP of the GCHP system (COP sys),
and CO 2emission ( CCO2).
Fig. 10. Recorded indoor and outdoor air temperature during heating and DHW
operation.
Fig. 9. Recorded indoor air and heat source temperature during cooling operation.
Table 2
GCHP system performance for classical and optimised adjustment in coolingoperation.
Case t
i[/C14C]te[/C14C]ts[/C14C]Eel
[kWh]Et[kWh] EER sys
[Btu/Wh]COP sysCCO2
[kg]
(1) Classic 26.22 24.54 20.50 70.99 287.21 13.79 4.04 28.60
(2) Optimised 25.71 24.88 20.65 65.16 288.45 15.09 4.42 26.25
Fig. 11. DHW and heat source temperature evolution during heating and DHW pro-
vision tests.C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 153

Although the COP syswas almost equal for the two cases, the
experimental results indicate that when using the automatic con-
trol device for the circulating pump speed, an electricity savings of
5% and a CO 2emission level decrease of 5% were obtained.
The experimental data ( Tables 2 and 4 ) results of the COP sysof
the GCHP system operating in cooling and DHW mode in com-
parison with cooling operation mode indicate that there was a
decrease only in the range of 3.2 e8.6% compared with the two
cases, from 4.04 e4.42 to 3.91 e4.04, respectively.
InFigs. 14 and 15 are summarised the performances of the GCHP
system in the different operation modes to show the experimental
measurement results of the COP sysand CO 2emission ( CCO2).
5.2. The GCHP performance in DHW operation
5.2.1. The DHW production with different temperatures
To analyse the heat pump unit performances of the GCHP sys-
tem that produces the DHW for a three person family, a mean daily
consumption of approximately 50 l/person at a DHW set-point
temperature ( tdhw-set ) of 40, 45, 50, and 60/C14C was considered.
The experimental measurements were conducted for a one-weekperiod for each DHW set-point temperature from 1st of April
2013 to 28th of April 2013. The experimental parameters recorded
for a DHW set-point temperature of 60/C14C, including flow tem-
perature ( tf), return temperature ( tr), DHW temperature ( tdhw), and
heat source temperature ( ts), are plotted in Fig. 16 .
Table 5 presents the summary of the mean values of the tem-
peratures ( tdhw,ts), DHW volume ( Vdhw), electricity consumption
(Eel), useful thermal energy ( Et), COP of the heat pump unit (COP hp)
and CO 2emission ( CCO2). It should be noted that the heat pump's
COP hp, when operated in DHW mode, decreased from 2.06 to 1.61 if
the DHW temperature increased from 40/C14Ct o6 0/C14C, respectively.
InFig. 17 is illustrated the variation of COP hpdepending on the
DHW temperature.
5.2.2. In fluence of water temperature increment in the DHW tank
The performance of the heat pump unit of the GCHP system is
influenced by instantaneous consumed hot-water volume, which
influences CO 2emissions. Table 6 presents the summary of the
mean values of the measured temperatures ( tf,tr,tdhw,ts), electricity
consumption ( Eel), useful thermal energy ( Et), COP of the heat pump
unit (COP hp), and CO 2emission ( CCO2) for six experiments at
different water temperature increments ( Dt) in the DHW tank.
Fig. 18 illustrates the evolution of heat carrier fluid temperatures
(tf,tr) produced by GCHP, DHW temperature ( tdhw), and heat source
temperature ( ts) for water heating in DHW tank with Dt¼25/C14C.
The data in Table 6 indicates that with higher instantaneous DHW
consumption (increased Dt), the COP hpdecreased. This COP hp
decrease could reach a level of 23% when the water temperature
increment in the DHW tank was 25/C14C.
5.3. Uncertainty analysis
Uncertainty analysis (the analysis of uncertainties in experi-
mental measurement and results) is necessary to evaluate the
experimental data [33,34] . An uncertainty analysis was performed
using the method described by Holman [33]. A result Zis a given
function of the independent variables x1,x2,x3,…,xn. If the un-
certainties in the independent variables w1,w2,w3,…,wnare all
given with same odds, then uncertainty in the result wZhaving
these odds is calculated by the following equation [33]:
wZ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi/C18vZ
vx1w1/C192
ț/C18vZ
vx2w2/C192
ț :::ț/C18vZ
vxnwn/C192s
(5)Table 3
GCHP system performance for heating and DHW provision tests.
Case ti[/C14C] te[/C14C] ts[/C14C] tdhw[/C14C] Vdhw[m3] Eel[kWh] Et[kWh] COP sys COP hp CCO2[kg]
(1) Classic 21.27 1.03 16.79 42.63 1.22 82.66 266.99 3.23 3.81 33.31
(2) Optimised 21.61 0.93 16.27 42.64 80.12 269.13 3.35 3.95 32.28
Fig. 12. Recorded indoor and outdoor air temperature during cooling and DHW
operation.
Fig. 13. DHW and heat source temperature evolution during cooling and DHW pro-
vision tests.
Fig. 14. Variation of COP sysin different operation modes of GCHP system.C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 154

In the present study, the temperatures, thermal energy and
electrical energy were measured with appropriate instruments
explained previously. Error analysis for estimating the maximum
uncertainty in the experimental results was performed using Eq.
(5). It was found that the maximum uncertainty in the results is in
the COP sys, with an acceptable uncertainty range of 1.31 e1.69% in
heating operation mode and of 2.29 e3.38% in cooling operation
mode. The uncertainty of the COP hpwas estimated between 1.82
and 2.33% in heating mode and between 2.90 and 5.60% in DHW
mode.
6. Numerical simulation of useful thermal energy using
TRNSYS software
TRNSYS (Transient System Simulation) software [35] is one of
the most flexible modelling and simulation tools and can solve very
complex problems from the decomposition of the model in various
interconnected model components. The basic principle of TRNSYS
program is the implementation of algebraic and first-order ordi-
nary differential equations describing physical components into
software subroutines (called types) with a standard interface. STEC
library is based on steady state energy conservation formulated in
thermodynamic quantities (temperature, pressure, and enthalpy).
In order to consider transient effects like start-up, a “capacity ”
model has been developed that can be linked to each of the pre-
viously mentioned components and easily tuned to match empir-
ical data. This permits thermal capacity to be considered only in
those components where it has a large impact and avoids the huge
gain in computational complexity of a full transient model. One ofthe main advantages of TRNSYS for the modelling and design of
ground source heat pumps is that it includes components for the
calculation of building thermal loads, speci fic components for
heating/cooling, ventilating and air-conditioning (HVAC), heat
pumps and circulating pumps, modules for borehole heat ex-
changers and thermal storage, as well as climatic data files, which
make it a very suitable tool to model a complete air-conditioning/
heat pump installation to provide heating and cooling to a building.
Some statistical methods, such as the root-mean squared ( RMS),
the coef ficient of variation ( c
v), the coef ficient of multiple de-
terminations ( R2) and percentage difference may be used to
compare simulated and actual values for model validation.
The error can be estimated by the RMS defined as [36]:RMS¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
Pn
m¼1/C16
ysim ;m/C0ymea ;m/C172
nvuut
(6)
In addition, the coef ficient of variation cv, in percent and the
coefficient of multiple determinations R2are de fined as follows:
cv¼RMS/C12/C12ymea ;m/C12/C12100 (7)
R2¼1/C0Pn
m¼1/C16
ysim ;m/C0ymea ;m/C172
Pn
m¼1y2mea ;m(8)
where nis the number of measured data in the independent data
set; ymea, mis the measured value of one data point m;ysim,min-
dicates the simulated value; ymea ;mis the mean value of all
measured data points.
The percentage difference is calculated as follows: (extracting
the measured value from the simulated value)/(measured
value) /C2100.
6.1. Simulation of thermal energy used for laboratory heating and
cooling
6.1.1. De finition of the operation scheme
To simulate the thermal energy used to cover the heating/
cooling load of the experimental laboratory, the operationalTable 4
GCHP system performance for cooling and DHW provision tests.
Case ti[/C14C] te[/C14C] ts[/C14C] tdhw[/C14C] Vdhw[m3] Eel[kWh] Et[kWh] COP sys CCO2[kg]
(1) Classic 25.40 24.96 20.62 38.92 1.36 50.80 198.62 3.91 20.47
(2) Optimised 25.45 25.28 20.60 39.12 48.22 195.06 4.04 19.43
Fig. 15. Variation of CO 2emission in different operation modes of GCHP system.
Fig. 16. Recorded operation parameters of the GCHP during DHW provision tests for
DHW set-point temperature of 60/C14C.
Table 5GCHP performance during DHW provision tests.
No. t
dhw-set [/C14C]tdhw[/C14C]ts[/C14C]Vdhw
[m3]Eel[kWh] Et[kWh] COP hpCCO2
[kg]
1 40 39.60 15.76 0.968 15.18 31.29 2.06 6.11
2 45 44.55 14.25 18.31 35.90 1.96 6.823 50 49.39 14.95 22.94 42.68 1.86 9.244 60 59.47 15.20 32.67 52.78 1.61 13.16C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 155

connections were established between the building and all internal
and external factors.
Fig. 19 presents the operational scheme built in TRNSYS, where
the building thermal behaviour was modelled using a “Type 56 ”
subroutine. This subroutine was processed with the TRNBuild
interface by introducing the main construction elements, their
orientation and surface, shadow factors, and indoor activity type.
Weather data for the Timisoara were obtained from the Meteonorm
data base [37] and the weather data reader “Type 109 ”and “Type
89d ”were used to convert the data in a form readable from TRNSYS.
The simulation model took into account the outdoor air in fil-
tration, heat/cold source type, and interior gains. Also, for a good
approach of the model were de fined light thresholds, cooling,
shading and light integrators. To extract the results, an online
plotter ( “Type 65 ”) is used.
6.1.2. Simulation results and comparison with experimental data
Performing simulations for a one-year period (8760 h), thermal
energy used for heating and cooling were obtained and are pre-
sented beside the measured values in Table 7 . Statistical values such
asRMS,cvand R2are given in Table 8 for the GCHP system in
different operation modes.
There was a maximum percentage difference between the
TRNSYS simulated and measured values for the heating period of
approximately ț1.59% and for the cooling period of
approximately /C01.64%, which is very acceptable. The RMS and cv
values in heating mode are 2.722 and 0.0141, respectively and in
cooling mode are 3.080 and 0.0238, respectively. The R2-values in
the two operation mode are about 0.9998, which can be considered
as very satisfactory. Thus, the simulation model was validated by
the experimental data.6.2. Simulation of thermal energy used to produce DHW
6.2.1. De finition of the operation scheme
For simulation of the DHW production the operational scheme
built in TRNSYS from Fig. 20 was utilised. The assembly of GCHP
system consists of the standard TRNSYS weather data readers “Type
15-6 ”, a GCHP model “Type 919 ”, a BHE “Type 557a ”and a DHW
storage tank “Type 4 ”, with a capacity of 175 L. Also, in the simu-
lation model were de fined single speed circulating pumps “Type
114”for the antifreeze fluid in the BHE and “Type 3d ”for heat
carrier fluid of the DHW coil. A “Type 14 ”for the load pro file and a
daily load subroutine were created, this approach improving
significantly the numerical convergence of the model. Finally, two
model integrators ( “Type 25 ”and “Type 24 ”) were used to calculate
daily and total results for thermal energy produced.
6.2.2. Simulation results and comparison with experimental data
Useful thermal energy simulations for the assurance of the DHW
thermal load were performed for four hot-water temperatures: 40,
45, 50, and 60/C14C. The results of the simulation program are pre-
sented beside the experimental data in Table 9 . Statistical values
such as RMS and cvare given in Table 8 . A comparative analysis of
these results indicates that the thermal energy values for DHW
production simulated with TRNSYS were only 0.21 e0.62% lower
than the measured values in all four cases. The COP hpvalues are in
the range 1.56 e2.00, approached to measured values ( Table 5 ). The
R2-value about 0.9999 is very satisfactory and the maximum per-
centage difference of 0.62% is acceptable and thus the simulation
model is validated experimentally.
7. Conclusions
The use of heat pumps in modern buildings with improved
thermal insulation and reduced thermal load is a good alternative
to classical heating/cooling and DHW solutions. It is recommended
whenever possible to use consumers capable of covering both
heating and cooling demand. Indoor air conditioning can be aproduct of these coupled processes. This paper presents the eval-
uation of the performances of a GCHP system providing heating/
cooling and DHW to an experimental laboratory. Some main con-
clusions can be deduced from this study:
(1) The performed experimental research demonstrated higher
performances of the GCHP system for the flow rate adjust-
ment case using a buffer tank and an automatic control de-
vice for circulating pump speed versus a classical adjustment
Fig. 17. Variation of COP hpdepending on the DHW temperature.
Table 6
GCHP performance depending on water temperature increment in DHW tank.
No. Dt[/C14C]tt[/C14C]ts[/C14C]tdhw[/C14C]Eel[kWh] Et[kWh] COP hpCCO2[kg]
1 3 34.10 12.78 17.79 0.170 0.556 3.33 0.068
2 5 40.29 13.83 29.49 0.310 1.011 3.26 0.124
3 10 37.13 10.46 26.10 0.720 2.264 3.14 0.2904 15 38.66 11.53 26.76 0.980 2.952 3.01 0.3945 20 42.34 9.53 30.01 1.440 4.044 2.81 0.5806 25 43.72 10.12 31.50 2.010 5.128 2.55 0.810
Fig. 18. Recorded operation parameters of GCHP for water heating in the DHW tank
with 25/C14C.C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 156

case (COP sys7e8% higher and 7.5 e8% lower CO 2emission
level).
(2) The GCHP system, operating in heating or cooling mode, had
a COP sys>4, and the GCHP system operating in heating or
cooling and DHW mode had a 3 <COP sys<4 for both cases.
(3) In classical and optimised adjustment cases the COP hpvalues
for heating and DHW provision tests were 3.81 and 3.95,
respectively and for heating operation tests were 4.82 and
5.06, respectively.
(4) When using the circulating pump speed control, electricity
savings and a reduction of the CO 2emission of 3% for labo-
ratory heating and 5% for laboratory cooling were obtained at
the same time with DHW production.
(5) If the GCHP is used to produce only DHW for a family at
different temperatures between 40/C14C and 60/C14C, then the
COP hpwould decrease to approximately 2, and the CO 2
emission level would vary between 6.11 kg and 13.16 kg.(6) For an instantaneously consumed hot-water volume, the
energy performance of the GCHP can be decreased by up to
23% when the hot-water temperature in DHW tank must be
increased to 25/C14C.
(7) The developed TRNSYS simulation models can be used as a
tool to determine the GCHP performance in different oper-
ation modes to optimise the system energy ef ficiency and
ensure the user's comfort throughout the year.
Fig. 19. Scheme of the system model built in TRNSYS to simulate the thermal energy consumption.
Table 7
Thermal energy used for laboratory heating and cooling.
Month Heating energy [kWh] Percentage difference [%] Cooling energy [kWh] Percentage difference [%]
Simulated Measured Simulated Measured
January 505.00 512.48 /C01.45 0.00 0.00 0.00
February 391.40 390.12 ț0.32 0.00 0.00 0.00
March 303.21 300.87 ț0.77 0.00 0.00 0.00
April 99.45 97.89 ț1.59 0.00 0.00 0.00
May 0.00 0.00 0.00 208.00 211.23 /C01.52
June 0.00 0.00 0.00 242.00 242.82 /C00.33
July 0.00 0.00 0.00 337.90 333.12 ț1.43
August 0.00 0.00 0.00 437.80 445.14 /C01.64
September 0.00 0.00 0.00 324.30 319.20 ț1.59
October 189.70 191.31 /C00.84 0.00 0.00 0.00
November 348.90 345.23 ț1.06 0.00 0.00 0.00
December 477.50 480.21 /C00.56 0.00 0.00 0.00Table 8
Statistical values of useful thermal energy of GCHP system.
Operation mode RMS c v R2
Heating 2.72187 0.01409 0.99990075
Cooling 3.08003 0.02382 0.99977802
DHW production 8.50000 0.00464 0.99997906C. Sebarchievici, I. Sarbu / Renewable Energy 76 (2015) 148 e159 157

Abbreviations
AFS air flow sensor
BHE borehole heat exchanger
COP coef ficient of performance
DHW domestic hot-water
EER energy ef ficiency ratio
GCHP ground-coupled heat pump
GHE ground heat exchanger
HVAC heating, ventilating and air-conditioning
PC personal computer
PLC programmable logic controller
PTC positive temperature coef ficient
TRNSYS Transient Systems Simulation
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