The following article was published in ASHRAE Journal, September 2006. © Copyright 2006 American Society of Heating, Refrigerating
and Air-Conditioning Engineers, Inc. It is presented for educational purposes only. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE.
Closed-Loop
Ground-Coupled
Heat Pump Systems
By Michel A. Bernier, Ph.D, Member ASHRAE
C
losed-loop ground-coupled heat pump systems offer several advantages over
conventional HVAC systems. For instance, they collect renewable ground
heat or recuperate building heat rejection that accumulated in the ground during
the cooling season. Furthermore, because of their relatively high coefficient of
performance (COP) in both heating and cooling they are, as noted by the U.S.
Department of Energy and the U.S. Environmental Protection Agency, among the
most energy-efficient and environment-friendly heating and cooling systems.
Finally, they emit less greenhouse gas than conventional HVAC systems given
the power generation mix found in most jurisdictions.
... some design engineers are still reluctant to specify these
systems [because they] do not fully understand the relatively
complex phenomena occurring in the borehole and in the
ground at peak conditions and during long periods of time.
The type of system under consideration is presented in Figure
1. This closed-loop system represents one of the most popular
configurations. A heat transfer fluid is pumped through a series
of vertical boreholes, where heat is collected (rejected) with a
corresponding fluid temperature increase (decrease). Borehole
depth is project dependent, but is usually in the 50 m – 150 m
(165ft – 495 ft) range.
As shown in the cross section, boreholes are usually filled
with a grout to facilitate heat transfer from the fluid to the
ground and to protect groundwater aquifers as required by
some state regulations. Fluid then returns to the building,
where heat pumps either collect (reject) heat in the fluid loop,
thereby decreasing (increasing) the fluid temperature. At any
given time, some heat pumps may operate in heating mode
while others might be in cooling mode. Thus, it is possible to
transfer energy from one section of the building to the other via
the fluid loop. Finally, in some situations it is advantageous to
design so-called hybrid systems where a supplementary heat
rejecter or extractor is added to reduce the length of the ground
heat exchanger.
Despite the environmental advantages, some design engineers
are still reluctant to specify these systems. Two main reasons
exist for this. First, the capital and maintenance costs of these
systems often are perceived to be higher than conventional
systems. In reality, cost data on installed systems1,2,3 do not
totally support this assertion. Furthermore, some cost containment options, such as hybrid systems, can reduce the length
(and cost) of the ground heat exchanger. The second reason,
which is the main focus of this article, is that some designers
do not fully understand the relatively complex phenomena occurring in the borehole and in the ground at peak conditions
and during long periods of time.
Performance Data on Heat Pumps
Before discussing the design of the ground heat exchanger,
it is important to review performance data on heat pumps.
The coefficient of performance (COP) and capacity of heat
pumps depend on several parameters such as fluid flow rate
and temperatures on the source and load sides. The COP is
September 2006
defined as the ratio of useful energy (either in cooling or heating) to the power input to the unit (used to run the compressor
and the fan). Note that the cooling COP is used here instead
of the usual EER values (cooling COP = EER/3.41) to make
a direct comparison between cooling and heating COP. Also,
note that pumping power usually is not included in the COP
values provided by manufacturers. Figure 2 presents values of
COP in heating and cooling for 10 commercially available 10.5
kW (3 ton) water-to-air extended range heat pumps. As shown
in Figure 2, heat pumps are not created equal and significant
differences exist among manufacturers. Also, some heat pumps
operate over a wider temperature range than others. It is also
important to note the strong dependency of the COP on the
inlet fluid temperature.
In cooling, the inlet fluid temperature should be as low as possible to reduce heat pump energy consumption. While in heating
mode the inlet fluid temperature should be as high as possible.
In other words, the temperature lift across the heat pump, i.e.,
the difference between the source and load temperatures (i.e.,
Tin,HP – TLoad in Figure 1) should be minimized. One way to
minimize the lift is to increase the length of the ground heat
exchanger so that Tin,HP tends towards the undisturbed ground
temperature, Tg. However, oversized ground heat exchangers
are not economically feasible. So, the design engineer must
find the right compromise between the length of the ground
heat exchanger that will give an acceptable Tin,HP to reduce
heat pump energy consumption as much as possible. In some
cases, this may imply the use of a hybrid system to reduce peak
ground loads and the length of the ground heat exchanger.
Ground Heat Exchanger Length
Determining the required length of a vertical geothermal heat
exchanger is not a trivial task and should not be undertaken
lightly. “Back of the envelope” calculations usually lead to either
conservative, uneconomical systems or to undersized ground
About the Author
Michel A. Bernier, Ph.D., is a professor in the département de génie
mécanique at École Polytechnique de Montréal in Montréal.
ASHRAE Journal
13
heat exchangers with potential operational problems linked to
out-of-range inlet temperatures to the heat pumps. Software to
determine the length is a must and, in some cases, simulation
tools are required to ascertain the impact of advanced systems
such as hybrid designs.
Ground heat exchangers usually are designed for the worst
conditions by considering that the ground heat exchanger needs
to handle three consecutive thermal pulses of various magnitude
and duration.4 The magnitude corresponds to the yearly average
ground load (qy), the highest monthly ground load (qm), and
the peak hourly load (qh). The corresponding durations are usually 20 years, one month, and six hours. The required borehole
length to exchange heat at these conditions is given by:
ASHRAE Journal
Tin,HP
To Other Boreholes
Grout
Borehole
Pipe
Tg
A
TW
A – A
A
40°F 60°F 80°F 100°F 120°F
1
where L is the total borehole length required. The values of
R20y, R1m , and R6h represent effective ground thermal resistances for 20 years, one month, and six hours thermal pulses.
The borehole wall temperature is represented by Tw (Figure 1).
Borehole thermal interference between adjacent boreholes is
accounted for by introducing a temperature penalty, Tp. This
article does not address borehole thermal interference. For a
more detailed discussion on this subject, the reader is referred
to the book by Kavanaugh and Rafferty.5
The effective ground thermal resistances depend mainly on
ground thermal conductivity and, to a lesser extent, on borehole
diameter and ground thermal diffusivity. Table 1 presents some
typical values for three different ground thermal conductivities.5
As shown, the effective ground thermal resistances decrease
with an increase in the ground thermal conductivity (kground).
The variations of R20y, R1m and R6h are almost proportional to
the variation of 1/ kground. For example, when kground varies from
1.2 to 3.1 W/m · K, a reduction of 1/ kground of 60% from 0.833 to
0.323, the values of R20, R1m and R6h are also reduced by about
60%. Thus, in this particular example, an increase in ground
thermal conductivity from 1.2 to 3.1 W/m · K would translate,
according to Equation 1, into a 60% reduction in the required
length of the heat exchanger.
This significant impact of ground thermal conductivity is
somewhat mitigated by the borehole thermal resistance and
borehole interference. Nonetheless, this shows the importance of knowing as precisely as possible the ground thermal
conductivity. This is why many engineers perform a ground
thermal conductivity test prior to their design. The cost of
such a test varies from one region to the other (from $2,500 to
$7,500 excluding the borehole itself). It usually pays for itself
by avoiding oversized (or undersized) loops. For more details
on the impact of thermal conductivity error on the length of the
ground heat exchanger, the readers are referred to the article
by Bernier6 and the “Outside the Loop” newsletter.7
The temperature difference in the borehole, i.e., between
the mean fluid temperature, Tm (= (Tin,HP+Tout,HP)/2) and Tw,
is given by:
14
Tload
Optional
Supplementary
Heat Rejector (Hybrid
System)
Figure 1: Closed-loop ground-coupled heat pump system.
qyR20y + qmR1m + qhR6h
Tw – (Tg + Tp)
Tout, HP
8
20°F 40°F
Cooling
8
6
COP
L=
Water-to-Air Heat Pump
COP
Other Heat Pumps
4
60°F
80°F
Heating
6
4
2
2
0°C 10°C 20°C 30°C 40°C 50°C
Inlet Fluid Temp.
–10°C 0°C 10°C 20°C 30°C
Inlet Fluid Temp.
Figure 2: COP values in heating and cooling for 10 commercially
available 10.5 kW (3 ton) water-to-air extended range heat pumps
for standard entering air conditions and a nominal water flow rate
of 0.57 L/s (9 gpm).
(Tm – Tw) =
qhRb
L
2
The value of Rb is the effective borehole thermal resistance.
It depends on borehole diameter, pipe diameter, separating distance between pipes, grout thermal conductivity, pipe thermal
conductivity, and fluid flow rate. Table 2 presents some typical
values of Rb for two commonly used configurations and two
grout thermal conductivities. These values of Rb were obtained
by assuming turbulent flow. At peak load, laminar flow is undesirable in a borehole as it increases borehole thermal resistance.
However, when the load on the loop is small, laminar flow
can be acceptable. In this situation Rb will be high but qh will
be low, resulting in acceptable temperature differences in the
borehole (Equation 2).
Configuration B is usually assumed to prevail when pipes
are freely inserted in a borehole. In configuration C, mechanical spacers are required (usually every 3 m [10 ft] or so) to
spread the pipes so that they are nearer to the ground. It is
advantageous to have the smallest effective borehole thermal
resistance in order to have the smallest temperature difference
in the borehole. As shown in Table 2, this can be accomplished
by increasing grout thermal conductivity and/or pipe spacing.
In some regions, high thermal conductivity grouts have become
the norm. As for mechanical spacers, some installers do not like
to use them because tube insertion can become difficult.
a s h r a e . o r g September 2006
10 000
–100
–4e + 5
–150
0
2,000
4,000
6,000
Hour of the Year
8,000
Figure 3: Hourly building loads for the example building.
Combining Equations 1 and 2, one gives Equation 3, the
design length equation:
qhRb + qyR10y + qmR1m + qhR6h
L=
Tm – (Tg + Tp )
3
Although different in appearance, this equation is basically
the same as the one contained in the 2003 ASHRAE Handbook—HVAC Systems and Equipment.4 The design length should
0
0
North Dakota
20 000
Coal Power Plant
–2e + 5
d
b
Alberta
–50
30 000
a
c
USA Avg.
Georgia
0
40 000
Québec
Vermont
0
Annual CO2 Emissions, kg CO2
2e + 5
50
Hourly Building Loads, Btu/h
Hourly Building Loads, kW
100
Canada Avg.
50 000
4e + 5
Natural Gas Power Plant
150
200
400
600
800
1000 1200
Emissions, kg CO2 per MWh of Electricity Produced
Figure 4: Annual CO2 emissions as a function of CO2 emitted per
MWh of electricity produced for the example building. Lines a and
b correspond to Cases 1 and 2, respectively ( Table 3). Lines c and
d are for a system with a gas boiler (η = 80%) and a chiller with a
COP of 4 and 5, respectively.
be calculated with the worst conditions in cooling and in heating. The maximum of these two lengths is the required borehole
length. A number of commercially available software products
have implemented Equation 3 or an equivalent form. However,
Equation 3 has its limits, especially when hybrid systems are used.
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September 2006
ASHRAE Journal
15
In these cases, multiyear hourly simulation tools
are required to model the complex interaction between the ground heat exchanger, the heat pumps
and the supplementary fluid heater or cooler and
determine the optimum length. Moreover, the
simulation lead to a more precise evaluation of
heat pump energy consumption.
Example
Ground Type
kground
Ground Thermal Conductivity
W/m · K
(Btu/h · ft · °F)
R20y
m · K/W
(°F · ft · h/Btu)
R1m
R6h
m · K/W
m · K/W
(°F · ft · h/Btu) (°F · ft · h/Btu)
Clay
1.2 0.36
0.287
0.14
(0.69)
(0.62)
(0.50)
(0.25)
Wet Shale
1.9
0.23
0.19
0.10
(1.1)
(0.40)
(0.32)
(0.18)
In this section, the required length of a ground
3.1
0.14
0.12
0.07
Limestone
heat exchanger is determined for a given build
(1.8)
(0.24)
(0.20)
(0.12)
ing and four design options. A life-cycle cost
analysis of each option and an estimate of their Table 1: Typical effective ground thermal resistances. (Calculated for a bore diameter
CO2 emissions are also provided. The building of 150 mm [6 in.].)
used is this example is part of the TESS library
of TRNSYS.8 It has an area of 1486 m2 (16,000
ft2), and it is assumed to be located in Atlanta.
As shown in Equation 3, the correct determination of the ground heat exchanger length
B
C
depends on precise evaluations of ground loads,
Borehole Diameter
150 mm (6 in.) 150 mm (6 in.)
which depend on the building loads and heat
Pipe Nominal Diameter
1 in. (25 mm) DR-9 1 in. (25 mm) DR-9
pump COP. Peak hourly and monthly ground
Center-to-Center Pipe Distance
8.3 cm (3.3 in.) 11.7 cm (4.6 in.)
loads, qh and qm, and the yearly average power
rejection/collection in the ground, qa, need to be
Rb, Equivalent Borehole Thermal Resistance, m · K/W (°F · ft · h/Btu)
evaluated. Particular attention should be given
Regular Grout—Thermal Conductivity
0.20
0.15
to peak loads as, in most cases, more than 70%
0.69 W/m · K (0.40 Btu/h · ft · °F)
(0.34)
(0.25)
of the ground heat exchanger length will be
Thermally Enhanced Grout—Thermal Conductivity 0.10
0.09
required to handle the peak load. In this article,
2.1 W/m · K (1.2 Btu/h · ft · °F)
(0.17)
(0.15)
building loads are evaluated hourly using the
TRNSYS simulation package.9
Table 2: Typical equivalent borehole thermal resistances for two pipe spacing and
Alternatively, designers who do not have ac- two grout thermal conductivities.
cess to simulation programs may evaluate peak
cooling/heating loads using available load calculation software results obtained for four design options. The length determination
and calculate monthly and yearly loads using tabulated values results will be examined first followed by an analysis of energy
of equivalent full load hours of operation (see table 9, page consumption, life-cycle costs and CO2 emissions.
32.20 in the 2003 ASHRAE Handbook).
The hourly cooling and heating loads for the example building Length
are shown in Figure 3. The peak building cooling load is 111
Case 1 uses low-efficiency heat pumps (bottom performance
kW (31.6 tons). The building is assumed to be equipped with curve in Figure 2) and a B configuration borehole with a low therfifteen 10.5 kW (3 ton) extended range heat pumps. The total mal conductivity grout. The resulting total borehole length is 3165
annual building heating and cooling loads are 87 000 MJ (8.25 m (10,400 ft) with 25 boreholes (5 × 5 configuration) each with a
× 107 Btu) and 552 000 MJ (5.23 × 108 Btu), respectively. Bore- depth of 126.6 m (416 ft). It is assumed that the proper pipe thickholes have a 150 mm (6 in.) diameter and include two 1 in. (25 ness has been selected to resist the hydrostatic pressure resulting
mm) HDPE SDR-9 pipes (which can withstand a hydrostatic from the use of such a relatively deep borehole. The annual thermal
pressure of 140 m [450 ft] of water). The borehole-to-borehole imbalance, qy, is relatively high at 21.4 kW (7.3 × 104 Btu/h),
distance is set to 8 m (26 ft). A wet shale ground composition is which leads to a relatively high borehole thermal interference with
assumed (see Table 1 for corresponding resistance values) and a resulting temperature penalty of 7.9 K (14.2°F). The temperature
the undisturbed ground temperature is 15°C (59°F).
difference between the mean fluid temperature in the borehole and
The ground heat exchanger is sized according to Equation 3. the borehole wall, Tm – Tw, is also high at 9.3 K (16.7°F) due mainly
Considering that the cooling loads are much greater than the heat- to the use of a low thermal conductivity grout.
ing loads, the ground heat exchanger length is determined based
Case 2 is similar to Case 1 except that high-efficiency heat
on the cooling loads. It is assumed that the maximum acceptable pumps are used (top curve on Figure 2). The use of high-efinlet fluid temperature to the heat pump is 38°C (100°F). Finally, ficiency heat pumps diminishes the peak ground load, from
TRNSYS simulations are used to evaluate heat pump energy con- 147.5 kW (5.0 × 105 Btu/h) to 139.2 kW (4.7 × 105 Btu/h), and
sumption every hour over 20 years of operation. Table 4 presents the annual thermal imbalance, from 21.4 kW (7.3 × 104 Btu/h)
16
ASHRAE Journal
a s h r a e . o r g September 2006
Case 1
Case 2
Case 3
Case 4
Length Determination
Type of Heat Pump
(Low or High Efficiency)
Low
High
High
High
0.20 m · K/W
0.20 m · K/W
0.09 m · K/W
0.09 m · K/W
No
No
No
Yes1
Peak Ground Load (qh)
147.5 kW
139.2 kW
139.2 kW
100.1 kW
Annual Thermal Imbalance (qy)
21.4 kW
19.9 kW
19.9 kW
15.1 kW
Borefield Configuration
5 × 5
5 ×5
5 × 4
5 × 4
Borehole Thermal Interference (Tp) 7.9 °C
7.7 °C
8.8 °C
5.7 °C
Temperature Difference
In Borehole (Tm – Tw)
9.3 K
9.3 K
5.4 K
5.9 K
3165 m
2980 m
2280 m
1500 m
Borehole Thermal Resistance
Hybrid System
Total Ground Heat Exchanger Length
Annual Energy Consumption
Average Annual Cooling COP
3.86 5.44 5.35 4.89
Average Annual Heating COP
4.03 5.65 5.74 5.8
47 730 kWh
34 440 kWh
34 760 kWh
37 580 kWh
—
—
420 kWh
$103,812 $97,777
$82,190
$54,120
$36,000 $49,500
$49,500
$49,500
Heat Pumps Fluid Cooler
—
Cost Analysis
Boreholes2
Heat
Pumps3
Fluid
Cooler4
Total—First Costs First Year Energy Cost5
Present Value of 20 Years of
Operation6
Present Value of Total Costs
—
—
—
$10,500
$139,812
$147,277
$131,690
$114,120
$3,820
$2,755
$2,780
$3,005
$50,904
$36,728
$37,068
$40,075
$190,716
$184,005
$168,758
$154,195
1. 30 ton fluid cooler
2. $32.80/m ($10/ft) for Cases 1 and 2 and $36.60/m ($11/ft) for Cases 3 and 4
3. $2,400 and $3,300 for low- and high-efficiency heat pumps, respectively
4. $350/ton
5. Energy costs of $0.08/kWh are assumed
6. Assuming an electricity escalation rate of 3% and a discount rate of 7%
Table 3: Results for the example building for four design options.
to 19.9 kW (6.8×104 Btu/h). Both of these factors contribute in
reducing the required length down to 2980 m (9,770 ft).
The borehole thermal resistance has been lowered in Case 3
by using a high thermal conductivity grout and spreading the
pipes against the borehole wall. This reduces (Tm – Tw) significantly, and consequently the required length, which is reduced
to 2280 m (7,480 ft), a 23% drop when compared to Case 2.
In Case 4, the ground heat exchanger length has been reduced to
1500 m (4,920 ft). With a shorter length, the ground heat exchanger
only can transfer the required amount of heat at peak conditions
by raising the mean fluid temperature. However, the resulting inlet
temperature to the heat pumps at peak loads is higher than the upper
temperature limits of the heat pumps. To alleviate this problem, a
105 kW (30 ton) closed-circuit fluid cooler is used in the fluid loop
(positioned as indicated in Figure 1). It operates to maintain the
inlet fluid temperature to the heat pumps at 38°C (100°C) at peak
September 2006
conditions. The tower fan and pump require 3.38 kW (4.5 hp). As
shown in Table 4, the use of the fluid cooler reduces both the peak
ground load (down to 100.1 kW [3.4 × 105 Btu/h]) and the annual
thermal imbalance (down to 15.1 kW [5.2 × 104 Btu/h]).
Annual Energy Consumption
As expected, the average annual COP for the low efficiency
heat pumps (Case 1) are lower than the other three cases, while
the COP for Cases 2 and 3 are very similar. With the hybrid
system, the ground temperatures, and consequently the inlet
fluid temperature to the heat pumps, are higher than for Cases
2 and 3 on average. Consequently, the cooling COP for Case 4
differs from the ones observed for Cases 2 and 3 even though
the same high-efficiency heat pumps are used.
In terms of annual energy consumption, the low efficiency
heat pumps (Case 1) consume about 30% more energy than the
ASHRAE Journal
17
other three cases. Cases 2 and 3 have similar energy consumption while the hybrid system consumes about 10% more energy
than Cases 2 and 3. The fluid cooler of the hybrid system operates an average of 125 hours per year with an average annual
energy consumption of 420 kWh.
The CO2 emissions of the closed-loop ground-coupled heat
pump system considered previously will be compared with a
system that uses a gas boiler to provide heat and a conventional
chiller for cooling. Thus, in the former case, only electricity is
used for heating and cooling while in the latter case electricity
is used for cooling and natural gas for heating.
Cost Analysis
In the case of natural gas, the amount of CO2 emitted is taken
A life-cycle cost analysis is presented in Table 3. National as 1891 g of CO2 per m3 (0.1182 lb of CO2 per ft3) of natural gas
average borehole costs are assumed to prevail. These costs are or 51.11 g CO2 per MJ (0.4056 lb of CO2 per kWh). For electricequal to $32.80/m ($10/ft) for the low thermal conductivity grout ity, as indicated in Table 4, the amount of CO2 emitted per MWh
(Cases 1 and 2) and $36.60/m ($11/ft) for the high thermal con- of electricity varies significantly from one region to the other
ductivity grout (Cases 3 and 4). These numbers were obtained by depending on the fuel mix used in power plants for that region.
adjusting for inflation the data obtained by Cane et al. in 19981 Regions that rely on hydroelectricity, such as in the province of
who reported average borehole costs of $29/m ($8.84/ft) for nine Québec (Canada), have low CO2 emissions. Coal-based production
installations and on an analysis reported in a newsletter10 on grout regions, such as in North Dakota, present high CO2 emissions.
costs. Heat pump costs, $2,400 and $3,300 for low- and highFigure 4 presents the amount of CO2 emitted by these two
efficiency heat pumps, are based on prices provided by a leading systems for the example building used earlier. The x-axis is the
heat pump manufacturer. The
amount of CO2 emitted per
fluid cooler cost is assumed to
MWh of electricity produced.
Region
CO2 Emissions, kg/MWh (lb/MWh)
be $10,500 based on a unit cost
Values given in Table 4 are
Canada
of $99.50 per kW ($350 per
represented as vertical lines on
Québec
9 (19.8)
ton) reported by Yavuzturk and
the graph. The y-axis is the total
Alberta
910 (2,002)
Spitler.11 The cost of electricity
amount of CO2 emitted to heat
is assumed to be $0.08/kWh.
and cool the example building.
National Average
422 (928)
Finally, the present value of 20
In the case of the geothermal
USA
years of operation is based on
system, Cases 1 and 2 are conVermont
26 (56.9)
a fuel escalation rate of 3%, in
sidered (represented by Lines
North Dakota
1085 (2,393)
line with the predicted 2.7%
a and b on Figure 4). For the
average annual inflation rate12
gas boiler-chiller system, a gas
Georgia
642 (1,414)
and a discount rate of 7%.
boiler efficiency of 80% is asNational Average
631 (1,388)
Results show that Case 4
sumed, and two chiller’s COP (4
Power Plants By Fuel Type
has the lowest life-cycle cost
and 5) are considered (they are
Natural Gas 503 (1,107)
followed by Case 3. The main
represented by Lines c and d in
difference between these two
Figure 4). As mentioned earlier,
Coal 960 (2,112)
cases has to do with bore- Table 4: Representative CO emissions from electricity production in the example building is located
2
hole costs, which differ by some North American regions. Sources: www.epa.gov/cleanenergy/ in Atlanta. In that region, the
$28,070. This difference is egrid/pdfs/state.pdf and www.ec.gc.ca/pdb/ghg/inventory_e.cfm.
lowest CO2 emissions are obgreater than the capital cost
tained with high efficiency
of the fluid cooler estimated at $10,500. Case 2 has the lowest heat pumps with 21,100 kg (46,500) of annual CO2 emissions.
energy consumption followed closely by Case 3. The present The system with the gas boiler and low efficiency chiller has the
value of 20 years of operation of low-efficiency heat pumps highest CO2 emissions at 30,000 kg (66,140 lb) per year.
(Case 1) is much higher than the three other cases that use
Figure 4 also presents a number of interesting results parhigh-efficiency heat pumps.
ticularly when lines intersect each other. For example, Lines a
As with any cost analysis, its accuracy depends on the as- and d intersect at 360 kg (800 lb) of CO2 emissions per MWh
sumptions used. For example, if borehole costs are lower and of electricity produced. Thus, if the example building was
electricity costs are higher than the ones assumed here, then the located in a region with CO2 emissions higher than this value,
hybrid system might not be the lowest cost system.
then the low-efficiency heat pumps will emit more CO2 than
the gas boiler and high-efficiency chiller system. A similar
CO2 Emissions
behaviour occurs when Lines a and c intersect at 730 kg (1610
The total amount of CO2 emitted by a cooling/heating appa- lb) of CO2 per MWh of electricity produced. In that case the
ratus is often evaluated as the sum of the direct (equivalent CO2 low-efficiency heat pumps emit more CO2 than the gas boiler
emissions linked to refrigerant leakage) and indirect (equivalent and the low-efficiency chiller system.
CO2 emissions due to the energy consumption of the apparatus)
The last observation regarding Figure 4 has to do with Line
effects.13 In what follows, only the indirect effect is considered b, which is always lower than the other three. This indicates
as in most cases it is the most important.
that the operation of high-efficiency heat pumps leads to the
18
ASHRAE Journal
a s h r a e . o r g September 2006
least amount of annual CO2 emissions
even in regions that rely on coal for electricity production. Thus, high-efficiency
heat pumps offer a clear environmental
advantage in this example.
Conclusions
This article reviewed the process
involved in calculating the required
length of a closed-loop vertical ground
heat exchanger linked to geothermal
heat pumps. As shown in Equation 3,
the determination of the required length
relies on an accurate determination of
a number of parameters. First, ground
loads should be determined as precisely
as possible including the annual thermal imbalance in the ground. Second,
ground conditions should be known.
Without knowledge of ground thermal
conductivity and ground temperature, a
proper evaluation of ground heat transfer cannot be made. Third, the effective
borehole thermal resistance should be
determined (Table 2).
In the second part of this article a
design example in a cooling-dominated
climate (Atlanta) is given, and four design options are considered (see results
in Table 3). A number of conclusions can
be drawn from this example.
First, the use of high-efficiency heat
pumps decreases peak ground loads and
the annual thermal imbalance both of
which contribute in reducing the length
of the ground heat exchanger.
Second, using a high thermal conductivity grout reduces the ground heat
exchanger length significantly (a drop of
23% in this particular case).
Finally, the use of a hybrid system
with a supplementary closed-circuit fluid
cooler is examined. Its use reduces the
required length and the corresponding
cost of the ground heat exchanger to the
point where this option has the lowest
life-cycle cost.
In terms of CO2 emissions, it is shown
that, for the example building, the use of
high-efficiency heat pumps leads to the
lowest CO2 emissions not only in Atlanta,
where the example building is located,
but even in jurisdictions that rely on coal
for electricity production.
September 2006
References
1. Cane, et al. 1998. Operating Experiences with Commercial Ground-Source Heat
Pump Systems, Atlanta: ASHRAE.
2. Cane, D., Garnet, J.M. 2000. “Update
on maintenance and service costs of commercial building ground-source heat pump
systems.” ASHRAE Transactions 106(1):
399 – 407.
3. Martin, M,A., Madgett, M.G., Hughes,
P.J. 2000. “Comparing maintenance costs of
geothermal heat pump systems with other
HVAC systems: preventive maintenance actions and total maintenance costs.” ASHRAE
Transactions 106(1):408– 423.
4. 2003 ASHRAE Handbook—Applications, Chap. 32.
5. Kavanaugh, S.P., Rafferty, K. 1997.
Ground-Source Heat Pumps—Design of
Geothermal System for Commercial and
Institutional Buildings. Atlanta: ASHRAE.
6. Bernier, M. 2002. “Uncertainty in the
design length calculation for vertical ground
heat exchangers.” ASHRAE Transactions.
108(1):939 – 944.
7. 1999. “Impact of conductivity error
on design results.” “Outside the Loop: A
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oit.edu/otl/otl02-03.pdf.
8. TESS. 2005 TESS Libraries Version
2.02, Reference Manuals (13 Volumes).
Thermal Energy Systems Specialists. http://
tess-inc.com.
9. SEL. 2005. TRNSYS 16—A Transient
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Solar Energy Laboratory, University of Wisconsin-Madison. Madison, Wis. �����������
http://sel.
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10. 2000. “Grout thermal conductiv i t y — b i g g e r i s n o t a lway s b e t t e r.”
“Outside the Loop: A Newsletter for Geothermal Heat Pump Designers and Installers.” 3(2). http://geoheat.oit.edu/otl/
���������������������������
otl03-02.pdf.
11. Yavuzturk, C., Spitler, J.D. 2000.
“Comparative study of operating and control strategies for hybrid ground-source
heat pump systems using a short time step
simulation model.” ASHRAE Transactions
106(2):192 – 209.
12. EIA. 2006. Annual Energy Outlook
2006 with Projection to 2030. http://www.
�����������
eia.doe.gov/oiaf/aeo/economic.html.
13. Sand, J.R., Fischer, S.K., Baxter
V.D. 1999. “Comparison of TEWI for
fluorocarbon alternative refrigerants and
technologies in residential heat pumps and
air-conditioners.” ASHRAE Transactions
105(1):1209 – 1218.
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