Closed-Loop Ground-Coupled Heat Pump Systems

Closed-Loop

Ground-Coupled

Heat Pump Systems

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.

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 electroni-cally or in paper form without permission of ASHRAE.

About the Author

Michel A. Bernier, Ph.D., is a professor in the département de génie mécanique at École Polytechnique de Montréal in Montréal.

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 contain-ment 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 oc-curring 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

defined as the ratio of useful energy (either in cooling or heat-ing) 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 pos-sible 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.,

T_in,HP – T_Load in Figure 1) should be minimized. One way to minimize the lift is to increase the length of the ground heat exchanger so that T_in,HP tends towards the undisturbed ground

temperature, T_g. 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 T_in,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

... 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.

Optional Supplementary

Heat Rejec-tor (Hybrid System)

Other Heat Pumps

Tout, HP Tin,HP TW Tload Tg A A A – A To Other Boreholes

Water-to-Air Heat Pump

Grout Borehole Pipe

Figure 1: Closed-loop ground-coupled heat pump system.

8 6 4 2 COP 8 6 4 2 COP 40°F 60°F 80°F 100°F 120°F 20°F 40°F 60°F 80°F 0°C 10°C 20°C 30°C 40°C 50°C

Inlet Fluid Temp. –10°C 0°C 10°C 20°C 30°CInlet Fluid Temp.

Cooling Heating

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).

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 (q_y), the highest monthly ground load (q_m), and the peak hourly load (q_h). The corresponding durations are usu-ally 20 years, one month, and six hours. The required borehole length to exchange heat at these conditions is given by:

q_yR_20y + q_mR_1m + q_hR_6h L =

T_w – (T_g + T_{p) 1}

where L is the total borehole length required. The values of

R_20y, R_1m , and R_6h represent effective ground thermal resis-tances for 20 years, one month, and six hours thermal pulses. The borehole wall temperature is represented by T_w (Figure 1). Borehole thermal interference between adjacent boreholes is accounted for by introducing a temperature penalty, T_p. 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 (k_ground). 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, R_1m and R_6h 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 impor-tance 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, T_m

(

/

)

q_hR_b

(T_m – T_w) =

L 2

The value of R_b is the effective borehole thermal resistance.

It depends on borehole diameter, pipe diameter, separating dis-tance between pipes, grout thermal conductivity, pipe thermal conductivity, and fluid flow rate. Table 2 presents some typical values of R_b for two commonly used configurations and two

grout thermal conductivities. These values of R_b were obtained by assuming turbulent flow. At peak load, laminar flow is unde-sirable 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 R_b will be high but q_h 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, mechani-cal 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.

Combining Equations 1 and 2, one gives Equation 3, the design length equation:

q_hR_b + q_yR_10y + q_mR_1m + q_hR_6h L =

T_m – (T_g + T_{p ) 3} Although different in appearance, this equation is basically the same as the one contained in the 2003 ASHRAE

Hand-book—HVAC Systems and Equipment.4 _{The design length should}

be calculated with the worst conditions in cooling and in heat-ing. 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.

150 100 50 0 –50 –100 –150 4e + 5 2e + 5 0 –2e + 5 –4e + 5 Hourly Building Loads, kW _{Hourly Building Loads, Btu/h}

0 2,000 4,000 6,000 8,000 Hour of the Year

50 000 40 000 30 000 20 000 10 000 0 Annual CO 2 Emissions, kg CO 2 0 200 400 600 800 1000 1200 Emissions, kg CO2 per MWh of Electricity Produced

Qu ébec Vermont Canada A vg. Natural Gas P ower Plant USA A vg.

Georgia Alberta Coal P

ower Plant North Dakota a c d b

Figure 3: Hourly building loads for the example building. 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.

In these cases, multiyear hourly simulation tools are required to model the complex interaction be-tween 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

In this section, the required length of a ground heat exchanger is determined for a given build-ing and four design options. A life-cycle cost analysis of each option and an estimate of their CO₂ emissions are also provided. The building used is this example is part of the TESS library of TRNSYS.8 _{It has an area of 1486 m}2 _(16,000

ft2_{), and it is assumed to be located in Atlanta.}

As shown in Equation 3, the correct deter-mination of the ground heat exchanger length depends on precise evaluations of ground loads, which depend on the building loads and heat pump COP. Peak hourly and monthly ground loads, q_h and q_m, and the yearly average power rejection/collection in the ground, q_a, need to be evaluated. Particular attention should be given to peak loads as, in most cases, more than 70% of the ground heat exchanger length will be required to handle the peak load. In this article, building loads are evaluated hourly using the TRNSYS simulation package.9

Alternatively, designers who do not have ac-cess to simulation programs may evaluate peak

cooling/heating loads using available load calculation software and calculate monthly and yearly loads using tabulated values of equivalent full load hours of operation (see table 9, page 32.20 in the 2003 ASHRAE Handbook).

The hourly cooling and heating loads for the example building are shown in Figure 3. The peak building cooling load is 111 kW (31.6 tons). The building is assumed to be equipped with fifteen 10.5 kW (3 ton) extended range heat pumps. The total annual building heating and cooling loads are 87 000 MJ (8.25 × 107 _{Btu) and 552 000 MJ (5.23 × 10}8 _{Btu), respectively.}

Bore-holes have a 150 mm (6 in.) diameter and include two 1 in. (25 mm) HDPE SDR-9 pipes (which can withstand a hydrostatic pressure of 140 m [450 ft] of water). The borehole-to-borehole distance is set to 8 m (26 ft). A wet shale ground composition is assumed (see Table 1 for corresponding resistance values) and the undisturbed ground temperature is 15°C (59°F).

The ground heat exchanger is sized according to Equation 3. Considering that the cooling loads are much greater than the heat-ing loads, the ground heat exchanger length is determined based on the cooling loads. It is assumed that the maximum acceptable inlet fluid temperature to the heat pump is 38°C (100°F). Finally, TRNSYS simulations are used to evaluate heat pump energy con-sumption every hour over 20 years of operation. Table 4 presents

Ground Type k_ground R_20y R_1m R_6h

Ground Thermal m · K/W m · K/W m · K/W Conductivity (°F · ft · h/Btu) (°F · ft · h/Btu) (°F · ft · h/Btu)

W/m · K (Btu/h · ft · °F) 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) Limestone 3.1 0.14 0.12 0.07 (1.8) (0.24) (0.20) (0.12)

Table 1: Typical effective ground thermal resistances. (Calculated for a bore diameter of 150 mm [6 in.].)

Table 2: Typical equivalent borehole thermal resistances for two pipe spacing and two grout thermal conductivities.

B C

Borehole Diameter 150 mm (6 in.) 150 mm (6 in.)

Pipe Nominal Diameter 1 in. (25 mm) DR-9 1 in. (25 mm) DR-9 Center-to-Center Pipe Distance 8.3 cm (3.3 in.) 11.7 cm (4.6 in.)

R_b,Equivalent Borehole Thermal Resistance, m · K/W (°F · ft · h/Btu)

Regular Grout—Thermal Conductivity 0.20 0.15

0.69 W/m · K (0.40 Btu/h · ft · °F) (0.34) (0.25) Thermally Enhanced Grout—Thermal Conductivity 0.10 0.09 2.1 W/m · K (1.2 Btu/h · ft · °F) (0.17) (0.15)

results obtained for four design options. The length determination results will be examined first followed by an analysis of energy consumption, life-cycle costs and CO₂ emissions.

Length

Case 1 uses low-efficiency heat pumps (bottom performance curve in Figure 2) and a B configuration borehole with a low ther-mal conductivity grout. The resulting total borehole length is 3165 m (10,400 ft) with 25 boreholes (5 × 5 configuration) each with a depth of 126.6 m (416 ft). It is assumed that the proper pipe thick-ness has been selected to resist the hydrostatic pressure resulting from the use of such a relatively deep borehole. The annual thermal imbalance, q_y, is relatively high at 21.4 kW (7.3 × 104 _Btu/h),

which leads to a relatively high borehole thermal interference with a resulting temperature penalty of 7.9K (14.2°F). The temperature difference between the mean fluid temperature in the borehole and the borehole wall, T_m – T_w, is also high at 9.3 K (16.7°F) due mainly to the use of a low thermal conductivity grout.

Case 2 is similar to Case 1 except that high-efficiency heat pumps are used (top curve on Figure 2). The use of high-ef-ficiency heat pumps diminishes the peak ground load, from 147.5 kW (5.0 × 105 _{Btu/h) to 139.2 kW (4.7 × 10}5 _{Btu/h), and}

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 (T_m – T_w) signifi-cantly, 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

Table 3: Results for the example building for four design options.

Case 1 Case 2 Case 3 Case 4

Length Determination Type of Heat Pump

(Low or High Efficiency) Low High High High

Borehole Thermal Resistance 0.20 m · K/W 0.20 m · K/W 0.09 m · K/W 0.09 m · K/W

Hybrid System No No No Yes1

Peak Ground Load (q_h) 147.5 kW 139.2 kW 139.2 kW 100.1 kW

Annual Thermal Imbalance (q_y) 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 (T_p) 7.9 °C 7.7 °C 8.8 °C 5.7 °C

Temperature Difference _{9.3 K 9.3 K 5.4 K 5.9 K}

In Borehole (T_m – T_w)

Total Ground Heat Exchanger Length 3165 m 2980 m 2280 m 1500 m

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

Heat Pumps 47 730 kWh 34 440 kWh 34 760 kWh 37 580 kWh Fluid Cooler — — — 420 kWh Cost Analysis Boreholes2 _{$103,812 $97,777 $82,190 $54,120} Heat Pumps3 _{$36,000 $49,500 $49,500 $49,500} Fluid Cooler4 _{— — — $10,500} Total—First Costs $139,812 $147,277 $131,690 $114,120

First Year Energy Cost5 _{$3,820 $2,755 $2,780 $3,005}

Present Value of 20 Years of Operation6 _{$50,904 $36,728 $37,068 $40,075}

Present Value of Total Costs $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%

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

other three cases. Cases 2 and 3 have similar energy consump-tion while the hybrid system consumes about 10% more energy than Cases 2 and 3. The fluid cooler of the hybrid system oper-ates an average of 125 hours per year with an average annual energy consumption of 420 kWh.

Cost Analysis

A life-cycle cost analysis is presented in Table 3. National average borehole costs are assumed to prevail. These costs are equal to $32.80/m ($10/ft) for the low thermal conductivity grout (Cases 1 and 2) and $36.60/m ($11/ft) for the high thermal con-ductivity grout (Cases 3 and 4). These numbers were obtained by adjusting for inflation the data obtained by Cane et al. in 19981

who reported average borehole costs of $29/m ($8.84/ft) for nine installations and on an analysis reported in a newsletter10 _{on grout}

costs. Heat pump costs, $2,400 and $3,300 for low- and high- efficiency heat pumps, are based on prices provided by a leading heat pump manufacturer. The

fluid cooler cost is assumed to be $10,500 based on a unit cost of $99.50 per kW ($350 per ton) reported by Yavuzturk and Spitler.11 _{The cost of electricity}

is assumed to be $0.08/kWh. Finally, the present value of 20 years of operation is based on a fuel escalation rate of 3%, in line with the predicted 2.7% average annual inflation rate12

and a discount rate of 7%. Results show that Case 4 has the lowest life-cycle cost followed by Case 3. The main difference between these two cases has to do with bore-hole costs, which differ by $28,070. This difference is greater than the capital cost

of the fluid cooler estimated at $10,500. Case 2 has the lowest energy consumption followed closely by Case 3. The present value of 20 years of operation of low-efficiency heat pumps (Case 1) is much higher than the three other cases that use high-efficiency heat pumps.

As with any cost analysis, its accuracy depends on the as-sumptions used. For example, if borehole costs are lower and electricity costs are higher than the ones assumed here, then the hybrid system might not be the lowest cost system.

CO₂ Emissions

The total amount of CO₂ emitted by a cooling/heating appa-ratus is often evaluated as the sum of the direct (equivalent CO₂ emissions linked to refrigerant leakage) and indirect (equivalent CO₂ emissions due to the energy consumption of the apparatus) effects.13 _{In what follows, only the indirect effect is considered}

as in most cases it is the most important.

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.

In the case of natural gas, the amount of CO₂ emitted is taken as 1891 g of CO₂ per m3 _{(0.1182 lb of CO}

2 per ft3) of natural gas

or 51.11 g CO₂ per MJ (0.4056 lb of CO₂ per kWh). For electric-ity, as indicated in Table 4, the amount of CO₂ emitted per MWh of electricity varies significantly from one region to the other depending on the fuel mix used in power plants for that region. Regions that rely on hydroelectricity, such as in the province of Québec (Canada), have low CO₂ emissions. Coal-based production regions, such as in North Dakota, present high CO₂ emissions.

Figure 4 presents the amount of CO₂ emitted by these two systems for the example building used earlier. The x-axis is the amount of CO₂ emitted per MWh of electricity produced. Values given in Table 4 are represented as vertical lines on the graph. The y-axis is the total amount of CO₂ emitted to heat and cool the example building. In the case of the geothermal system, Cases 1 and 2 are con-sidered (represented by Lines a and b on Figure 4). For the gas boiler-chiller system, a gas boiler efficiency of 80% is as-sumed, and two chiller’s COP (4 and 5) are considered (they are represented by Lines c and d in

Figure 4). As mentioned earlier,

the example building is located in Atlanta. In that region, the lowest CO₂ emissions are ob-tained with high efficiency heat pumps with 21,100 kg (46,500) of annual CO₂ emissions. The system with the gas boiler and low efficiency chiller has the highest CO₂ emissions at 30,000 kg (66,140 lb) per year.

Figure 4 also presents a number of interesting results

par-ticularly when lines intersect each other. For example, Lines a and d intersect at 360 kg (800 lb) of CO₂ emissions per MWh of electricity produced. Thus, if the example building was located in a region with CO₂ emissions higher than this value, then the low-efficiency heat pumps will emit more CO₂ than the gas boiler and high-efficiency chiller system. A similar behaviour occurs when Lines a and c intersect at 730 kg (1610 lb) of CO₂ per MWh of electricity produced. In that case the low-efficiency heat pumps emit more CO₂ than the gas boiler and the low-efficiency chiller system.

The last observation regarding Figure 4 has to do with Line b, which is always lower than the other three. This indicates that the operation of high-efficiency heat pumps leads to the

Region CO₂ Emissions, kg/MWh (lb/MWh) Canada Québec 9 (19.8) Alberta 910 (2,002) National Average 422 (928) USA Vermont 26 (56.9) North Dakota 1085 (2,393) Georgia 642 (1,414) National Average 631 (1,388)

Power Plants By Fuel Type

Natural Gas 503 (1,107)

Coal 960 (2,112)

Table 4: Representative CO₂ emissions from electricity production in some North American regions. Sources: www.epa.gov/cleanenergy/ egrid/pdfs/state.pdf and www.ec.gc.ca/pdb/ghg/inventory_e.cfm.

least amount of annual CO₂ emissions even in regions that rely on coal for elec-tricity 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 ther-mal 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 trans-fer 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 de-sign 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 con-ductivity 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.

References

1. Cane, et al. 1998. Operating Experi-ences with Commercial Ground-Source Heat Pump Systems, Atlanta: ASHRAE.

2. Cane, D., Garnet, J.M. 2000. “Update on maintenance and service costs of com-mercial 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 ac-tions and total maintenance costs.” ASHRAE Transactions 106(1):408– 423.

4. 2003 ASHRAE Handbook—Applica-tions, 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 Newsletter for Geothermal Heat Pump De-signers and Installers.” 2(3). http://geoheat. 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 System Simulation Program – Documenta-tion Set (9 Volumes). Version 16.00.0038. Solar Energy Laboratory, University of Wis-consin-Madison. Madison, Wis. http://sel.http://sel. me.wisc.edu/trnsys.

10. 2000. “Grout thermal conducti-vity—bigger is not always better.” “Outside the Loop: A Newsletter for Geo-thermal Heat Pump Designers and Ins-tallers.” 3(2). http://geoheat.oit.edu/otl/http://geoheat.oit.edu/otl/ otl03-02.pdf.

11. Yavuzturk, C., Spitler, J.D. 2000. “Comparative study of operating and con-trol 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.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.