Optimisation Of Phase Change Material Thermal Energy Storage

OPTIMISATION OF PHASE CHANGE MATERIAL THERMAL ENERGY STORAGE
WITH A GROUND SOURCE HEAT PUMP
1
Paul McKenna1 and Donal P. Finn1
School of Mechanical and Materials Engineering, University College Dublin, Ireland
ABSTRACT
This paper presents the optimisation of a ground
source heat pump (GSHP) system with phase change
material (PCM) thermal energy storage (TES) for both
charging and discharging modes. The system is installed in a commercial building in Marseille, France,
and is used for cooling and dehumidification of ventilation air in an air handling unit (AHU). The charging
of the system was optimised for PCMs of 0 ◦ C and
8 ◦ C as well as for two different heat pumps – a small
heat pump with a 4.2 kW cooling capacity and a larger
heat pump with a 26 kW capacity. It was found that
the heat transfer fluid (HTF) mass flow rate through
the heat pump has a significant effect on the system
performance and the optimal flow rate is different for
each of the PCM melting temperatures and heat pump
sizes considered. The HTF flow rate through the AHU
was optimised for different weather conditions and different PCM temperatures. Finally, the monetary savings of using a TES system were calculated, showing
savings of between 14% and 30%.
INTRODUCTION
Energy storage is a focus of many electrical utilities
presently. The use of more renewable energy (especially solar and wind energy) can lead to an imbalance between demand and supply. This imbalance can
be managed by producing and storing thermal energy
when there is excess supply, and using the stored energy when there is high demand. Arce et al. (2011)
showed that the use of thermal energy storage (TES)
could lead to a potential electrical load reduction in
the EU of 1.16 GWth (7.5%) with an expected CO2
reduction of 258 MTCO2 (5.5%). Ice TES is commonly used in the USA. Braun (2007b) describes a
method of charging and discharging an ice TES system that realises savings within 2% of the theoretical
maximum possible savings. This leads to savings that
can be as high as 60% compared to chiller priority
control, with typical savings between 25% and 30%
(Braun, 2007a). Chaichana et al. (2001) showed cost
reductions of 55% with full ice storage compared to a
system without TES.
Ground source heat pumps (GSHPs) have an established market in Scandanavia, Germany, Austria and
Switzerland. For example, in Sweden, 18% of heating systems in family homes are GSHPs (Fors´en and
Nowak, 2010). More recently, there is evidence of
strong growth of this sector in several other European
markets, especially the UK and France (Fors´en and
Nowak, 2010). Although the use of GSHPs is increasing in northern France, it is still less common for them
to be used in southern France or indeed other Mediterranean climates. GSHPs have been shown to offer
many advantages compared to other space conditioning systems. Approximately 1.4% of residential heating in Europe is provided by GSHPs, saving around
0.7% of greenhouse gas emissions compared to a traditional heat mix defined as 50% gas, 30% oil, 10% solid
fuel and 10% electricity (Bayer et al., 2012). Blum
et al. (2011) calculated that one GSHP can reduce the
yearly CO2 emissions of an average person by 20%,
compared to a conventional heating mix of 53% natural gas, 42% heating oil, 4% electricity and 1% coal.
There are drawbacks to the installation of GSHPs, including high capital costs and invasive installation procedure as well as underperformance of heat pumps due
to incorrect system design (Boait et al., 2011). Despite the high capital cost, Bolling and Mathias (2008)
showed GHSPs to have the shortest payback compared
to three other heating and cooling systems in five different US cities. The other systems studied were: a
high efficiency furnace and electric air conditioner; an
absorption air conditioner and direct electrical heating; and a thermally driven heat pump. This shows the
advantage of GSHPs over other technologies in situations where heating and cooling loads co-exist.
The two technologies mentioned above – TES and
GSHPs – deal with the electricity storage issue, and
the need for efficient combined heating and cooling
technologies. However, there is little research in the
literature into integrated GSHP and TES systems or
into the optimisation of such a system. This highlights
a need for the study of integration and optimisation issues relating to these systems, which is the motivation
of the current research.
The objective of the current paper is to describe sensitivity studies of an installed GSHP TES system using an integrated TRNSYS simulation model, which
is used to cool and dehumidify ventilation air during
summer months. The simulations of the TES charging and discharging are used to calculate the mass flow
rates that lead to the highest system performance. This
model can then be used to analyse the savings potential compared to a system without TES. The overall
research objective is to develop general control strate-
gies for buildings with GSHPs and PCM TES, and to
test these control strategies on installed systems.
SIMULATION MODEL
The system described in this paper is part of an installed system in a commercial building in Marseille,
France. The building is a single story building with office space of 250 m2 and a workshop area of 88 m2 . It
includes a GSHP with 4.2 kW cooling capacity, which
is connected to a borehole heat exchanger consisting
of 6 double U-tube vertical boreholes, each of 100 m
in length. This heat pump is used to supply cooling to
a tank containing spherically encapsulated PCM with a
melting temperature of 0 ◦ C. Nodules have a diameter
of 98 mm and occupy 59% of the 500 litre tank. The
thermophysical properties of water are used to represent the PCM in the simulation. The capacity of the
PCM tank (including a 20 ◦ C temperature change) is
117 MJ. This PCM is then used to supply cooling and
dehumidification to an air handling unit (AHU) during the summer. This system was retrofitted and completed in the last year. A schematic of the system can
be seen in Figure 1. The PCM loop (in green) is for
tank discharging via the AHU, while the charging of
the PCM tank by the GSHP is shown in orange. The
GSHP being used for the TES is of a smaller capacity than the main building heating/cooling heat pump
which has a capacity of 26 kW.
Figure 1: System schematic showing PCM charging
and discharging.
The TRNSYS system model is composed of individual custom TRNSYS components, including a GSHP,
a PCM TES tank and an AHU. McKenna and Finn
(2013) describe in detail the construction of the models for these components as well as the validation
and calibration of the models. The borehole heat exchanger is simulated using TRNSYS TYPE997, from
the TESS Geothermal Library.
METHODOLOGY
A model of the described system is set up using the
TRNSYS simulation software. A mathematical model
of each of the system components is used to form the
individual TRNSYS modules and these are connected
together to describe the system. The simulations are
run with a time step of one second to ensure accurate
representation of system transient behaviour.
Three different scenarios for charging the PCM tank
are considered. These are shown in Table 1. These
scenarios involve the use of either the 4.2 kW dehumidification heat pump or the 26 kW main building
heat pump to charge the PCM. Two PCMs with different melting points (0 ◦ C and 8 ◦ C) are also analysed.
The thermophysical properties are assumed to be the
same as the 0 ◦ C PCM for the purposes of this analysis. The 8 ◦ C PCM was initially considered for installation in the system, but after experimental tests performed on this PCM, it was decided that a 0 ◦ C PCM
would provide better performance characteristics.
Table 1: System configurations being examined
Scenario
.
1
2
3
Heat Pump
Capacity (kW)
4.2
4.2
26
PCM Melting
Temperature ( ◦ C)
0
8
8
The charging of the PCM tank is analysed for a range
of mass flow rates through the PCM tank loop and
through the ground loop. The mass flow rate through
the PCM tank loop is varied between 200 kg/h and
2017 kg/h, while the mass flow rate through the ground
loop is analysed between 100 kg/h and 1657 kg/h. The
pumps installed are variable speed pumps and the upper limit on the flow rates (2017 kg/h and 1657 kg/h)
is set by calculating the maximum flow rates that the
installed pumps can deliver through the respective circuits.
The discharging of the PCM tank to provide ventilation air cooling and dehumidification is also examined.
For the discharging, a range of HTF flow rates are
inspected to determine a preferential discharging rate
based on the ambient temperature and humidity supply conditions to the AHU. These simulations are performed for three different summer days. These days
are taken from the Meteonorm weather data files for
Marseille and represent the following: a very hot day
(Day 1) with a maximum temperature of 32.5 ◦ C, a hot
day (Day 2) with a maximum temperature of 27.7 ◦ C
and a warm day (Day 3) with a maximum temperature
of 23.75 ◦ C.
The combination of tank charging and discharging are
compared against the solution without any TES and
the off-peak electricity rate for economic advantage is
calculated in each case.
The system performance factor (SPF) of the various
tank charging scenarios is calculated. The SPF of a
system is a ratio of the thermal energy provided by
the system to the electricity consumption of the system over a certain period of time. The period of time
in this case is the total length of time for charging of
the PCM tank. There are two different SPFs considered in this system. These are defined as:
R t Q˙ hp (t)
SP F1 =
SP F2 =
0 Php (t)
dt
˙ hp (t)
Q
0 Php (t)+Pp,g (t)+Pp,t (t)
Rt
(1)
dt
(2)
SP F1 describes the performance of the heat pump
alone in supplying thermal energy to the PCM tank
and is equivalent to the instantaneous coefficient of
performance (COP) of the heat pump integrated over
the charging period (Equation 1), while SP F2 includes the electricity use of both pumps while charging
the tank (Equation 2).
The SPF can be seen to be lower when the HTF is supplied at a lower temperature. Also, the smaller heat
pump has a lower SPF than the larger heat pump.
CHARGING RESULTS
0 ◦ C PCM
Simulations were performed for tank charging at the
set up that is currently installed in the building. This is
the smaller heat pump of 4.2 kW capacity and a 0 ◦ C
PCM. The SPF for the charging of the tank can be seen
in Figures 3 and 4.
NON-TES SCENARIO
In order to facilitate comparison between the aforementioned scenarios and a base case, simulations were
performed without the presence of a PCM tank. This
case involves the cooling and dehumidification of the
inlet air to the AHU. For comparison purposes, the
cooling of this air was performed using the design
HTF flow rate of 900 kg/h through both the evaporator
and condenser of the heat pump. A buffer tank of 200 l
was inserted between the heat pump and the AHU and
the inlet temperature to the AHU was set at 2±1 ◦ C
for comparison with the 0 ◦ C PCM and at 9±1 ◦ C for
comparison with the 8 ◦ C PCM. This system can be
seen in Figure 2.
Figure 3: SP F1 for charging.
It can be seen in Figure 3 that the heat pump SPF is
highest when the flow rates through both the evaporator and the condenser are highest. These higher flow
rates lead to a higher heat transfer in the heat pump,
therefore increasing SP F1 .
Figure 4: SP F2 for charging.
Figure 2: System schematic for base case.
The SP F2 over the three days in question are seen in
Table 2.
Table 2: SP F2 for system without PCM.
Day 1
Day 2
Day 3
2 ◦C
2.73
2.60
2.53
9 ◦C
3.06
2.98
2.86
9 ◦ C Main HP
4.51
4.36
3.47
Comparing Figures 3 4, when the power consumption of the pumps are included in the SPF calculation
(SP F2 ), the SPF at higher ground mass flow rates and
tank mass flow rates decreases from 2.35 to 2.1. There
is little change in the SPF at lower mass flow rates.
The SP F2 for charging shows little variation with tank
mass flow rate as the higher heat transfer in the PCM
tank with an increased flow rate is almost exactly offset by the increased power consumption of the pump.
However, the SPF is significantly higher when a high
mass flow rate (circa 1600 kg/h) through the ground
loop is used. Therefore, when charging the PCM tank,
a mass flow rate through the ground loop of 1600 kg/h
should be used, while a mass flow rate through the
PCM tank of over 1400 kg/h leads to optimal charging
conditions. The energy consumption of the pumps at
the highest flow rates is low (2 MJ for the ground loop
and 3.5 MJ for the PCM tank loop) compared to the
energy consumption of the heat pump (50 MJ). This
means that the charging of the PCM tank is still more
efficient at higher flow rates even though the pump energy consumption has increased.
flow rates. Figure 6 shows that the lowest performance
is achieved at low ground mass flow rates and low tank
mass flow rates. However, the highest charging performance in this scenario is for a tank flow rate of 800
kg/h and a ground flow rate of 1600 kg/h. The difference between the 0 ◦ C PCM and the 8 ◦ C PCM is
that the optimal tank mass flow rate is lower for the
8 ◦ C PCM than it was for the 0 ◦ C case. The higher
SP F1 for the 8 ◦ C PCM means that SP F2 is influenced more by changes in pump energy consumption
than in the 0 ◦ C PCM case, which had a lower SP F1 .
8 ◦ C PCM
26 kW Heat Pump
Simulations were performed for tank charging using
the smaller, 4.2 kW heat pump and a PCM with a melting point of 8 ◦ C. The SPF for the charging of the tank
can be seen in Figures 5 and 6.
Analysis was performed for tank charging using a
higher capacity heat pump, which is used for the heating and cooling of the main building and a PCM with
a melting point of 8 ◦ C. The SPF for the charging of
the tank can be seen in Figures 7 and 8.
Figure 5: SP F1 for charging.
Similar to the case with the 0 ◦ C PCM, SP F1 is highest when the flow rate through both the evaporator and
the condenser are highest. However, the SPF is considerably higher across the range of flow rates than for the
previous case – in Figure 3 the SPF ranges from 1.952.35, while in Figure 5 it ranges from 2.6-3.1. This is
due to the higher temperature of the HTF at the evaporator of the heat pump during the charging process.
Figure 6: SP F2 for charging.
Similarly to the 0 ◦ C PCM case, seen in Figure 4, the
SP F2 decreases at high ground and PCM tank mass
Figure 7: SP F1 for charging.
Considering Figure 7, it should be noted that in the
plots for this system configuration, the x- and y-axes
have been inverted. Therefore, in direct contrast to the
previous calculations, the highest SP F1 is shown for
low ground and tank mass flow rates. This highlights
the complexity of the system that is under consideration. A lower mass flow rate through the tank will
increase the temperature of the HTF returning to the
evaporator of the heat pump, while a lower mass flow
rate through the ground heat exchanger will decrease
the temperature of the HTF returning to the condenser.
Both of these serve to increase the COP of the heat
pump. However, the lower mass flow rates through the
heat pump decrease the heat transfer in the heat pump
and therefore lead to a lower COP. The balancing of
these two effects results in a higher heat pump performance for higher mass flow rates with the 4.2 kW heat
pump, and a higher performance with lower mass flow
rates for the 26 kW heat pump.
The SP F2 of the system shows a similar trend with
the performance at higher ground mass flow rates degrading very slightly.
tank via the AHU. Analyses were performed using
constant mass flow rates between 200 kg/h and 3600
kg/h. For each of these mass flow rates studied, the
results at the end of the day can be seen. These results
include the state of charge of the PCM tank at the end
of the day (1 being fully charged and 0 being fully discharged), the total heat transfer from the air over the
course of the day and the total pump electricity consumption (multiplied by a factor of 100) for the day.
0 ◦ C PCM
The discharging maps for the 0 ◦ C PCM can be seen
in Figure 9 for the three different days discussed.
Figure 8: SP F2 for charging.
The SP F2 of the system shows a similar trend with
the performance at higher ground mass flow rates degrading very slightly.
Considering Figure 8, the highest tank charging performance can be found at low mass flow rates for both
the tank and the heat pump. The total charging time
in this case was 16 hours, due to the low heat transfer
rate in the PCM tank at these low mass flow rates. This
is too long to for practical operation as tank charging should be completed during the night-time offpeak electricity pricing period. In Marseille, where the
building is located, this corresponds to an off-peak period from 11 pm to 7 am. A tank mass flow rate of over
700 kg/h leads to a PCM tank charging time of under
8 hours; the ground mass flow rate has little effect on
charging time. Therefore, in this scenario, charging
should be scheduled with a tank mass flow rate of 700
kg/h and a ground mass flow rate as low as possible.
DISCHARGING MAPS
The discharging of the tank and cooling and dehumidification of ventilation air in the AHU is discussed in
this section. Three different ambient temperature and
humidity scenarios are studied. These are taken from
the Meteonorm weather data for Marseille as follows:
• Day 1 – a very hot day with a maximum temperature of 32.5 ◦ C.
• Day 2 – a hot day with a maximum temperature
of 27.7 ◦ C.
• Day 3 – a warm day with a maximum temperature of 23.7 ◦ C.
The ventilation flow rate through the AHU is a constant 600 m3 /h (1 ACH) throughout the day and the
PCM tank is discharged between 8 am and 5 pm, or
until the ambient air temperature falls below the AHU
HTF inlet temperature.
The analyses begin with a fully charged PCM tank at a
uniform temperature of -10 ◦ C for the 0 ◦ C PCM and
at a temperature of -2 ◦ C for the 8 ◦ C PCM. These
temperatures are same as the final temperatures of the
PCM tank after charging.
For the discharging analysis (Figures 9 and 10), a constant mass flow rate was used while discharging the
Figure 9: Tank discharging for 0 ◦ C PCM.
It should be noted that the scale on the secondary yaxis is different for the three days. This axis shows the
state of charge of the PCM tank at 5 pm, with 0 being
fully discharged. On Day 1, the tank can be fully discharged over the course of the day with a mass flow
rate of 800 kg/h. This leads to a pump energy consumption of 1.4 MJ for heat transfer to the air of 120
MJ. However, to fully discharge the tank on Days 2
and 3, mass flow rates of at least 1600 kg/h and 3200
kg/h respectively are required. This leads to a higher
pump energy consumption over the course of the day
and a lower overall SPF. An efficient control solution
would run the pump at a flow rate that fully discharges
the tank without using a flow rate that is too high. As
can be seen in Figure 9, there is little increase in heat
transfer to the air with an increased mass flow rate
once the tank is fully discharged over the course of
the day.
8 ◦ C PCM
The discharging maps can be seen in Figure 10 for the
three days discussed.
Figure 11: Sensible and Latent Heat Transfer.
As can be seen in Figure 11, although there is not
much difference in total heat transfer to the air, a lot
more dehumidification is provided by the 0 ◦ C PCM
than by the 8 ◦ C PCM. This has implications for the
thermal comfort in the building. When the 0 ◦ C PCM
is in use, the HTF and therefore the coils in the AHU
are at a lower average temperature than with an 8 ◦ C
PCM. This leads to the higher ratio of latent to sensible heat transfer for the 0 ◦ C PCM shown in Figure
11.
ENERGY COMPARISON
Figure 10: Tank discharging for 8 ◦ C PCM.
As with the previous case, the scales on the secondary
y-axis are different for each of these graphs. Full discharging of a tank with an 8 ◦ C PCM requires a higher
flow rate than for a 0 ◦ C PCM. This is because the inlet
temperature to the AHU will be higher and therefore,
for the same heat transfer, a higher flow rate is needed.
For Day 1 and 2, flow rates of 2400 kg/h and 3600 kg/h
respectively are required, and for Day 3, even with the
maximum flow rate of 3600 kg/h, there is still a 25%
charge left in the tank at the end of the day.
From these calculations, it can be seen that control of
the flow rate during the day is critical to ensure the full
discharging of the PCM tank, and that a tank with an
8 ◦ C PCM requires more energy to discharge than one
with a 0 ◦ C PCM.
Also, as the dehumidification of the air is a primary objective of the AHU, the lower HTF temperatures provided by the 0 ◦ C PCM tank lead to a higher ratio of
latent to sensible cooling. This is seen in Figure 11.
In this section, the energy consumption in the three
scenarios mentioned is compared for each of the three
days under examination. For these calculations, the
optimum flowrates associated with the highest charging/discharging SPFs are used. These are then compared to the energy use for the base case without PCM
TES as calculated previously (Table 2).
The energy use is calculated by dividing the capacity
of the PCM tank by the SPF in question. This is done
because the total heat transfer varied slightly depending on the final temperatures within the system. For
example, with a low flow rate through the system, the
final temperature of the PCM and HTF was slightly
higher. Because of these differing values of total heat
transfer, a more accurate comparison between the energy use in the different scenarios is made by dividing
the PCM tank capacity (117 MJ) by the SPF, which
yields the total energy use for that scenario if exactly
117 MJ heat transfer had occurred.
In Tables 3, 4 and 5, the total energy use can be compared for TES and non-TES scenarios. Echarge represents the total energy use in charging the PCM tank,
Edischarge represents the total electricity use in discharging the PCM tank and Ebase represents the total
electricity use without TES. The break-even price ratio
is included in these tables. This is the ratio of off-peak
to on-peak electricity rates which leads to the same
running cost for the TES and non-TES scenarios. An
electrical utility price ratio lower than this would lead
to economic savings in the use of PCM TES, while a
ratio higher than the break even price ratio would mean
that it would not be economically advantageous to use
TES. The percentage savings for an off-peak price of
64% of the on-peak rate (as is the case in Marseille)
are also shown.
0 ◦ C PCM
Table 3: Energy savings with 0 ◦ C PCM
Day 1
Day 2
Day 3
Echarge
(MJ)
55.7
55.7
55.7
Edischarge
(MJ)
1.17
1.80
3.41
Ebase
(MJ)
42.86
46.10
48.11
Price
Ratio
0.75
0.79
0.80
Savings
14%
19%
19%
For all three days, the break even point is for an off
peak electricity price between 0.75 and 0.8 of the peak
price. Savings of between 14% and 19% are seen for
off-peak to on-peak ratio of 0.64.
8 ◦ C PCM
Table 4: Energy savings with 8 ◦ C PCM
Day 1
Day 2
Day 3
Echarge
(MJ)
40.91
40.91
40.91
Edischarge
(MJ)
1.74
2.13
3.74
Ebase
(MJ)
38.23
39.17
40.89
Price
Ratio
0.89
0.90
0.91
Savings
27%
28%
27%
The use of an 8 ◦ C PCM leads to a break even offpeak electricity pricing of between 89% and 91% of
the on-peak price, and savings of up to 28% when the
off-peak tariff for Marseille is considered. Although
the savings are higher for this case, it is worth noting
that the latent heat transfer in this case was not as high
as that for the 0 ◦ C PCM. This could have an effect
on thermal comfort in the building during the day, as
additional sensible cooling can be supplied via the fan
coil units in the building, but their capacity for latent
cooling is lower than the AHU; the cooling coil in the
AHU is unfinned to provide maximum dehumidification, while the fan coil units have finned coils for a
higher heat transfer.
Main Heat Pump
Table 5 shows a comparison between a case with thermal energy storage and a case without thermal energy
storage, using the main heat pump.
Table 5: Energy savings with 8 ◦ C PCM and main heat
pump
Day 1
Day 2
Day 3
Echarge
(MJ)
31.2
31.2
31.2
Edischarge
(MJ)
1.74
2.13
3.74
Ebase
(MJ)
25.94
26.82
33.72
Price
Ratio
0.77
0.79
0.96
Savings
16%
18%
30%
The use of the main building heat pump leads to
a minimum economically advantageous off-peak:onpeak electricity price ratio of 0.77. Savings of between
16% and 30% are also observed with a price ratio of
0.64. As with the previous case, the dehumidification
of the air is lower than with the 0 ◦ C PCM.
CONCLUSION
This paper shows the importance of control strategies
when integrating a GSHP with PCM TES in a building. The flow rate on both the ground side and the
tank side of the heat pump affect the SPF of the charging process. For the heat pump that is currently in use
for the PCM loop, it was found that high flow rates on
both the ground and tank loops lead to a higher system
performance. Conversely, lower flow rates were found
to be advantageous if the main building heat pump is
being used for charging the PCM tank.
Tank discharging was performed through the AHU to
cool and dehumidify ventilation air. It was shown that
the mass flow rate to fully discharge the PCM tank
needs to be higher for a cooler day. However, a flow
rate that is higher than the optimum leads to a higher
electricity use, thus decreasing the SPF of the system.
A discharging algorithm should be used for discharging the PCM tank that would vary the flow rate through
the PCM tank depending on predicted load and ambient temperature. It was also shown that TES with an
8 ◦ C PCM can lead to the same cooling of the inlet air,
but a higher pump flow rate, and therefore increased
electrical energy consumption, is needed. Also, the
amount of dehumidification provided in this case is
significantly lower than with a 0 ◦ C PCM.
The combination of charging and discharging the
PCM tank was compared to direct cooling of the ventilation air using the GSHP. Although the total energy
use with TES was higher in all cases, it was shown
that monetary savings could always be made if an offpeak to on-peak electricity price ratio of less than 0.75
was in use. In Marseille, where the building is located,
this ratio is 0.64, which leads to savings in the cost of
electricity from 14% to 30% for the scenarios studied.
FUTURE WORK
As the highest SPF for tank charging with the installed
heat pump occurred at the highest flow rates that the
pumps in use can accommodate, analysis will be performed using higher capacity pumps, to see if significant savings could be made by using even higher flow
rates.
During less warm summer days, there is less demand
for cooling and dehumidification of ventilation air.
Also, the high HTF flow rate that is needed to fully
discharge the PCM tank during these days leads to a
higher pump energy consumption. It is proposed that
incomplete charging of the PCM tank during the offpeak period could lead to higher system performance
by increasing the heat pump COP during charging and
decreasing the pump electricity use during discharging.
The climate of the location directly affects the system performance. The effect of climate on the system
performance will be evaluated by analysing a similar
building at locations with various temperature ranges
and humidity levels.
An economic analysis of the system will be performed
over a complete season and the payback period such
an installation will be calculated.
NOMENCLATURE
Acronyms
ACH air changes per hour
AHU air handling unit
COP coefficient of performance
GSHP ground source heat pump
HTF heat transfer fluid
PCM phase change material
SPF
system performance factor
TES
thermal energy storage
Symbols
Q˙
heat transfer
E
total energy use
P
power
Subscripts
hp
heat pump
p,g
ground loop pump
p,t
PCM loop pump
for cool storage systems with dynamic electric rates.
HVAC&R Research, 13 (4):557 – 580.
Chaichana, C., Charters, W. W. S., and Aye, L. 2001.
An ice thermal storage computer model. Applied
Thermal Engineering, 21(17):1769 – 1778.
Fors´en, M. and Nowak, T. 2010. Outlook 2010: European Heat Pump Statistics. European Heat Pump
Association EEIG (EHPA).
kW
MJ
kW
ACKNOWLEDGEMENTS
This research has been supported by the Irish Research Council’s Embark Initiative and by the FP7
GroundMed project (DG TREN/FP7EN/218895).
REFERENCES
Arce, P., Medrano, M., Gil, A., Or´o, E., and Cabeza,
L. F. 2011. Overview of thermal energy storage
(TES) potential energy savings and climate change
mitigation in Spain and Europe. Applied Energy,
88(8):2764 – 2774.
Bayer, P., Saner, D., Bolay, S., Rybach, L., and
Blum, P. 2012. Greenhouse gas emission savings of
ground source heat pump systems in Europe: A review. Renewable and Sustainable Energy Reviews,
16(2):1256 – 1267.
Blum, P., Campillo, G., and Klbel, T. 2011. Technoeconomic and spatial analysis of vertical ground
source heat pump systems in Germany. Energy,
36(5):3002 – 3011.
Boait, P., Fan, D., and Stafford, A. 2011. Performance and control of domestic ground-source heat
pumps in retrofit installations. Energy and Buildings, 43(8):1968 – 1976.
Bolling, A. L. and Mathias, J. A. 2008. Investigation
of optimal heating and cooling systems in residential buildings. ASHRAE Transactions, 114(1):128 –
139.
Braun, J. E. 2007a. Impact of control on operating
costs for cool storage with dynamic electric rates.
ASHRAE Transactions, 113 (2):343 – 354.
Braun, J. E. 2007b. A near-optimal control strategy
McKenna, P. and Finn, D. P. 2013. A TRNSYS
model of a building HVAC system with GSHP and
PCM thermal energy storage - component modelling and validation. In 13th Conference of International Building Performance Simulation Association, Chambry, France.

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