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A techno-economic comparison of ground-coupled and air-coupled heat pump system for space cooling
ARTICLE IN PRESS
Building and Environment 42 (2007) 1955–1965 www.elsevier.com/locate/buildenv
A techno-economic comparison of ground-coupled and air-coupled heat pump system for space cooling Hikmet Esena,, Mustafa Inallib, Mehmet Esena a
Department of Mechanical Education, Faculty of Technical Education, University of Firat, 23119 Elazig, Turkey b Department of Mechanical Engineering, Faculty of Engineering, University of Firat, 23119 Elazig, Turkey Received 26 January 2006; received in revised form 9 March 2006; accepted 3 April 2006
Abstract This paper reports a techno-economic comparison between a ground-coupled heat pump (GCHP) system and an air-coupled heat pump (ACHP) system. The systems connected to a test room in Firat University, Elazig (38.411N, 39.141E), Turkey, were designed and constructed for space cooling. The performances of the GCHP and the ACHP system were experimentally determined. The experimental results were obtained from June to September in cooling season of 2004. The average cooling performance coefficients (COPsys ) of the GCHP system for horizontal ground heat exchanger (HGHE) in the different trenches, at 1 and 2 m depths, were obtained to be 3.85 and 4.26, respectively and the COPsys of the ACHP system was determined to be 3.17. The test results indicate that system parameters can have an important effect on performance, and that GCHP systems are economically preferable to ACHP systems for the purpose of space cooling. r 2006 Elsevier Ltd. All rights reserved. Keywords: Heat pump; Ground-coupled; Air-coupled; Techno-economic analysis; Cooling; Turkey
1. Introduction A ground-coupled heat pump (GCHP), also called an Earth-coupled heat pump or a geothermal heat pump, operates much like the common an air-coupled heat pump (ACHP) by transferring heat, rather than creating it. Unlike an ACHP, a GCHP transfers heat to and from the Earth to provide space cooling and heating. The most conventional GCHP is the water-to-air heat pump which circulates water through a liquid-to-refrigerant heat exchanger in the ground, with an air-cooled condenser (heat pump) or air-heated evaporator (air conditioner) in the space to be conditioned. The liquid-to-refrigerant heat exchanger consists of a buried thermoplastic piping network which is the ground-coupled heat exchanger. Closed-loop GCHPs have demonstrated the potential to increase capacity and reduce total power consumption by Corresponding author. Tel.: +90 424 2370000/4228; fax: +90 424 2367064. E-mail addresses: hikmetesen@firat.edu.tr (H. Esen), minalli@firat.edu.tr (M. Inalli), mesen@firat.edu.tr (M. Esen).
0360-1323/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.buildenv.2006.04.007
providing a reduced condensation temperature in the summer and a higher evaporation temperature in the winter. Air temperatures which have large daily and monthly cyclic changes are damped or moderated by the Earth mass [1]. The GCHP have high installation costs that make it important to design the system to maximize performance. The GCHP exchange heat with the ground, and maintain a high level of performance even in colder climates for space heating. This results in more efficient use of energy. For this reason many community utilities support the use of GCHP and are active in an effort to persuade the heating ventilating and air conditioning (HVAC) industry to increase the number installed [2]. Heating and cooling heat pumps, providing both space heating and cooling. The most common type is the reversible air-to-air heat pump, which either operates in heating or cooling mode. However, the capacity and performance of ACHPs decrease rapidly with decreasing ambient temperature during heating season, and with increasing ambient temperature during cooling season. The technical and economic performance of a heat pump is closely related to the characteristics of the heat source/sink.
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H. Esen et al. / Building and Environment 42 (2007) 1955–1965
In particular, a regional heating and cooling system using a GCHP might be considerably more efficient than an ACHP system because the difference of soil temperature and room temperature is smaller than that of air temperature and room temperature. In comparison to systems based on other natural heat sinks such as rivers or sewers, there are no restrictions for the location of the GCHP system. Moreover, heat released underground in summer could be retrieved for the winter heat demand. Thus, year-round energy consumption for climate control might be reduced by the GCHP system. Genchi et al. [3] concluded that the regional heating-and-cooling system using a GCHP is a viable option for reducing CO2 emissions from the viewpoint of life-cycle CO2 emissions that include the emissions associated with the initial construction of the system. Since the idea of a GCHP system was advanced in the late 1940s, a great deal of theoretical and experimental works has been summarized in literature [4]. Healy and Ugursal [5] gave a description of a performance and economic feasibility of GCHP in cold climate. In their study, a comparative economic evaluation was carried out to assess the feasibility of using a GCHP in place of conventional heating/cooling systems and an ACHP system. It was found that all conventional heating systems have smaller coefficient of performance. Petit and Meyer [6–8] have made a techno-economic comparison of a GCHP system and an ACHP system for South Africa climatic conditions. From these studies, they were concluded that GCHP systems are more viable than ACHP systems. In another study, De Swardt and Meyer [9] compared the performance of a GCHP using municipality water as heat source/sink with the performance of an aircoupled heat pump. This comparison was achieved using simulation models that were experimentally validated. The comparison was conducted in both cooling and heating cycles. The results indicated that the utilization of municipality water reticulation systems as a heat source/ sink is a viable method of optimizing energy usage in the air conditioning industry, especially when used in the heating mode. The installations in Turkey on GCHP system are very limited although the temperature of ground is stable in most areas and suitable for GCHP applications [10]. Almost all of the heat pump systems used in Turkey utilize ambient air as the heat source and sink. The usage of other sources and sinks is very rare. Barriers still exist in Turkey regarding the usage of ground and other alternative sources in heat pumps; such sources are not well known, and they are notorious for troublemaking. Another barrier is the lack of designers, contractors and operators who can properly design, install and operate these systems economically and reliably. Esen et al. [11] has reported a detailed techno-economic analysis of a ground source heat pump system and six conventional heating systems for the climate conditions of Turkey in heating season of 2002–2003. In hot climates such as in Turkey, GCHPs represent a viable
alternative to ACHPs and conventional space cooling and heating systems because of their higher operating efficiency, especially during the cooling season. This article focuses on economic and performance comparison of horizontal GCHP system with ACHP system in Turkey. This comparison was achieved by using experimental studies conducted in cooling season. An experimental set-up, described in the next section, was constructed and tested for the first time on the basis of an academic study performed in Elazig, Turkey. 2. Description of the heat pump systems The designed and constructed heat pump systems that will be described are all based on the basic vapourcompression cycle. The schematic views of the constructed GCHP and ACHP system used for space cooling are shown in Fig. 1(a) and (b), respectively. The photographs of horizontal ground heat exchanger of GCHP system and condenser of ACHP system are seen in Figs. 2(a) and (b), respectively. The evaporator of the both systems is shown in Fig. 3. The ACHP system consists of only two circuits: the refrigerant (R-22) circuit and air circuit (condenser fan). The GCHP system mainly consists of three separate circuits: (a) the ground heat exchanger (GHE) circuit or water-antifreeze solution circuit consisting of a horizontal 100 m of pipe length divided into two GHEs, each 50 m length, 0.3 m of pipe distance, 0.016 m of nominal pipe diameter, and 15 m2 spacing were buried with in 1 and 2 m depths in different trenches, horizontal ground heat exchanger; (HGHE1) and horizontal ground heat exchanger; (HGHE2), (b) the refrigerant circuit and (c) the fancoil circuit or air circuit. Combined with Table 1, the main component specifications and characteristics of the ACHP and GCHP system are better understood for this work. In winter, the water-antifreeze solution in the pipes (HGHE1 and HGHE2) of the GCHP system extracts heat from the Earth and carries it into the room. In summer, the system reverses and takes heat from the room and stores it into the cooler ground, as seen in Fig. 1a. The performance evaluations of heating and cooling mode operation of the GCHP system were given in the other studies [12–14]. The heat transfer from Earth to the heat pump or from the heat pump to Earth is maintained with the fluid or waterantifreeze solution circulated through the GHE. In the cooling mode, the refrigerant, which circulates through the evaporator, evaporates by extracting heat from room air and then enters the hermetic compressor. The refrigerant is compressed by the compressor and then enters to the condenser, where it condenses by rejecting heat to the water-antifreeze solution. After the refrigerant leaves the condenser, the capillary tube provides almost 10 1C superheat that essentially gives a safety margin to reduce the risk of liquid droplets entering the compressor. A fan blows across the evaporator to move the cooled air of the
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
1957 Multi channel temperature scanner
Evaporator
:Thermocouples
Compressor Reversing valve
Capillary tubes Refrigerant
Wattmeter
circuit (R 22) Condenser
Tap collectors
Water-antifreeze
Circulating pump
solution circuit
Ground
HGHE2
HGHE1
ea
m
(a)
air
str
:Thermocouples
Ro
om
Multi-channel temperature scanner
Evaporator
Compressor
Refrigerant Capillary tubes
Reversing valve
circuit (R 22)
Wattmeter
Fl
(b)
fro
m
su
ui ds rro trea un m di ng s
Condenser
Fig. 1. Schematic views of the constructed (a) GCHP system and (b) ACHP system.
room. The air velocities were measured with a digital anemometer having a maximum error of 3%. The flow rate of the circulated water-antifreeze solution through the
closed loop HGHE was measured by using a rotameter and controlled by a hand-controlled tap mounted on the collector. The power consumption of the both heat pump
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Fig. 2. Photographs of condensers (a) of GCHP system and (b) of ACHP system.
modes by reversing the refrigerant flow direction. All measured quantities were regularly recorded in every halfhour. A room of 16:24 m2 floor areas was used as a space to be cooled. 3. Energy performance and uncertainty analysis for the both systems The energy efficiencies of the GCHP and ACHP system were evaluated in terms of the COP, defined as the ratio of cooling energy output to the electrical energy input. 3.1. Analysis of the GCHP system The heat rejected from the unit in the cooling mode (HGHE load) Q_ conGCHP is calculated by the following equation. _ wa C p;wa ðT co;wa T ci;wa Þ. Q_ conGCHP ¼ m Fig. 3. Photograph of evaporator of the both systems.
systems, i.e. the electrical power input to the compressor, water-antifreeze circulating pump and evaporator/condenser fan, was measured by means of wattmeter. The inlet and outlet temperatures of the R-22 in the condenser, compressor and evaporator were measured with copperconstantan thermocouples, T-type. In addition, the inlet and outlet temperatures of the circulated water-antifreeze solution through the closed loop HGHE was measured by using the same kind of thermocouples. Also, the temperatures of the ground at various depths, 1 and 2 m, were measured with T-type thermocouples. All measured temperatures were determined by using a multi-channel temperature scanner. The ambient and indoor air temperatures were measured with thermometers. The inlet and outlet pressures of the compressor and evaporator were measured by using Bourdon-type manometers. The reversing valve is used to switch between heating and cooling
(1)
_ comp and the The power input to the compressor W _ evaf are calculated by the Eqs. (2) and (3), evaporator fan W respectively. _ comp ¼ I comp U comp cos j, W
(2)
_ evaf ¼ I evaf U evaf cos j. W
(3)
There is power input to the water-antifreeze circulating _ wacp , at the GCHP system as different from the pump; W ACHP system and it is calculated by _ wacp ¼ I wacp U wacp cos j. W
(4)
The evaporator load equals to the heat extracted from the indoor air stream, Q_ evaGCHP is calculated by Q_ evaGCHP ¼ rair V_ air C p;air ðT eo;air T ei;air Þ.
(5)
The cooling performance coefficient of the GCHP unit, COPhpGCHP , is estimated by COPhpGCHP ¼
Q_ evaGCHP , _ comp W
(6)
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Table 1 Main components specification and characteristics of the ACHP and GCHP studied
Ground-coupled heat pump (GCHP)
Main circuit
Element
Technical specification
Ground coupling circuit
HGHE1, HGHE2 (HGHE1; 50 m) (HGHE2; 50 m) Water-antifreeze solution circulating pump
Pipe distance: 0.3 m; pipe diameter: 0.016 m; HGHE1: piping depth 1 m; HGHE2: piping depth 2 m; material: polyethylene, PX-b Cross link. Manufacturer: Alarko; type: NPVO-26-P; power: 40, 62, 83 W
Fan circuit
Fan of evaporator
Refrigerant circuit
Heat exchanger (Evaporator for cooling) Heat exchanger (condenser) Capillary tube Compressor
Manufacturer: Friterm; Diameter: 350 mm; air volumetric flow rate: 2350 m3 =h, power: 145 W; Manufacturer: Friterm; type: HS 10; capacity: 3.77 kW; heat transfer surface: 10 m2
Air-coupled heat pump (ACHP)
Dryer Observe glass Fan circuit
Fan of condenser
and the average performance coefficients of the GCHP system, COPsysGCHP is calculated by COPsysGCHP ¼
Q_ evaGCHP . _ _ evaf þ W _ wacp W comp þ W
(7)
The cooling equipment systems used in residential and small commercial buildings often express cooling system efficiency in terms of the energy efficiency ratio (EER). This coefficient is defined as the ratio of net cooling capacity (in Btu/h) divided by the electric input in watts under standard rating conditions [15]: EER ¼ 3:412COPhp ,
(8)
where the units of EER are Btu W1 h1 .
Manufacturer: Tunc- Technique; type: TTE 3; capacity: 6.97 kW (with Freon 22); Copper capillary tube; 2 m long; inside diameter: 1.5 mm Manufacturer: Tecumseh Europe; type: hermetic; model: FH 5524 F; the rated power of electric motor driving: 2 HP (1.4 kW); R-22; single phase Manufacturer: Carly; model: DCY 083; connection: 3/8 in; Manufacturer: Carly; model: VCYL 13; connection: 3/8 in; Manufacturer: Friterm; Diameter: 350 mm; air volumetric flow rate: 2350 m3 =h, power: 145 W;
_ conf , at the ACHP system as different condenser fan; W from the GCHP system and it is calculated by _ conf ¼ I conf U conf cos j. W
(11)
The cooling performance coefficient of the ACHP unit, COPhpACHP , is estimated by COPhpACHP ¼
Q_ evaACHP , _ comp W
(12)
and the average performance coefficients of the ACHP system, COPsysACHP is calculated by COPsysACHP ¼
Q_ evaACHP . _ evaf þ W _ conf _ comp þ W W
(13)
3.2. Analysis of the ACHP system
3.3. Uncertainty analysis
The ACHP system cycle is used for space cooling. During cooling, the evaporator coil cools and removes moisture from the indoor air stream, Q_ evaACHP . The condenser is located outdoors and rejects Q_ conACHP (equal _ comp ) to the surroundings. These capato Q_ evaACHP þ W cities are given in Eqs. (9) and (10).
Uncertainty analysis (the analysis of uncertainties in experimental measurement and results) is a powerful tool, particularly when it is used in the planning and design of experiments. As we will see in literature [16], there are realistic, practical cases in which all the measurements in an experiment can be made with 1% uncertainty, yet the uncertainty in the final experimental result will be greater than 50%. Uncertainty analysis, when used in the initial planning phase of an experiment, can identify such situations and save the experimentalist much time, money, and, embarrassment. The experiment collects data of three types, temperature, flow rate, and input power. Total uncertainty (TU) in the measurement of the mass
Q_ evaACHP ¼ rair V_ air C p;air ðT eo;air T ei;air Þ,
(9)
Q_ conACHP ¼ rair V_ air C p;air ðT co;air T ci;air Þ.
(10)
_ comp and the The power input to the compressor W _ evaf are calculated by using Eqs. (2) and evaporator fan W (3), respectively. Also, there is power input to the
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
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flow rate wm_ may be calculated as follows [17]: wm_ ¼ ðw2ro þ w2sl þ w2td Þ1=2 , wm_ ¼ ð1:252 þ 2:52 þ 0:752 Þ1=2 ¼ 2:89%.
(14)
The result R is a given function in terms of the independent variables. Let wR be the uncertainty in the result and w1 ; w2 ; . . . ; wn be the uncertainties in the independent variables. The result R is a given function of the independent variables x1 ; x2 ; x3 ; . . . ; xn . If the uncertainties in the independent variables are all given with same odds, then uncertainty in the result having these odds is calculated by [18] " 2 2 2 #1=2 qR qR qR wR ¼ w1 þ w2 þ þ wn . qx1 qx2 qxn (15) It is important that all uncertainties used in Eq. (15) can be evaluated at the same confidence level. The uncertainty estimates in the COPhpACHP , COPhpGCHP , COPsysACHP and COPsysGCHP can be calculated from Eq. (15) as 2 6 6 wCOPhpACHP ¼ 6 6 4
wCOPhpGCHP
wCOPsysGCHP
4. Results and discussions 4.1. Ambient and ground conditions Annual air and ground temperatures were needed as boundary conditions for determining the year round performance of GCHP and an ACHP system. These temperatures, given in Fig. 4, were obtained from an experimental measurement. This figure is also show the average temperature of ground in the winter and summer conditions. The ‘‘air’’ temperature line represents the average air temperatures between the average monthly minimum and maximum temperatures. In the cooling season (June–
31=2 2 2 2 qCOPhpACHP qCOPhpACHP qCOPhpACHP wrair þ wV_ air þ wC p;air 7 qrair qC p;air 7 qV_ air 2 2 2 7 7 , qCOPhpACHP qCOPhpACHP qCOPhpACHP 5 þ wT eo;air þ wT ei;air þ wW comp qT eo;air qT ei;air qW comp
31=2 2 2 2 2 qCOPhpGCHP qCOPhpGCHP qCOPhpGCHP wrair þ wV_ air þ wC p;air 7 6 qrair qC p;air 7 6 qV_ air 6 ¼6 2 2 2 7 7 , qCOPhpGCHP qCOPhpGCHP qCOPhpGCHP 5 4 þ wT eo;air þ wT ei;air þ wW comp qT eo;air qT ei;air qW comp
(16)
(17)
31=2 2 2 2 qCOPsysACHP qCOPsysACHP qCOPsysACHP wrair þ wV_ air þ wC p;air 7 qrair qC p;air 7 qV_ air 2 2 2 7 7 qCOPsysACHP qCOPsysACHP qCOPsysACHP þ wT eo;air þ wT ei;air þ wW comp 7 7 , qT eo;air qT ei;air qW comp 7 7 2 2 7 qCOPsysACHP qCOPsysACHP 5 þ wW evaf þ wW conf qW evaf qW conf
(18)
31=2 2 2 2 2 qCOPsysGCHP qCOPsysGCHP qCOPsysGCHP wrair þ wV_ air þ wC p;air 7 6 qrair qC p;air qV_ air 7 6 6 2 2 2 7 7 6 qCOP qCOP qCOP sysGCHP sysGCHP sysGCHP 6 wT eo;air þ wT ei;air þ wW comp 7 ¼6 þ 7 . qT eo;air qT ei;air qW comp 7 6 7 6 2 2 7 6 qCOPsysGCHP qCOPsysGCHP 5 4 þ wW evaf þ wW wacp qW evaf qW wacp
(19)
2
wCOPsysACHP
collects data uncertainties on the experiments. The uncertainties are based on a limited number of tests, so the estimates may change with more testing. From the test data, it can be clearly seen that total uncertainties of two different systems is quite near in this study (see Table 3).
6 6 6 6 6 ¼6 6 6 6 4
It is also important to point out that the uncertainty analysis performed in this work only provided an approach of the effects that would result from the transport of experiment
September) cooling is required, and in the heating season (December–March) heating is required and the ‘‘Air’’ temperatures in Fig. 4 can also be used as external ambient
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
Air
Ground 1.0 m
Ground 2.0 m
1961
Relative humidity
30
80
25
70
Temperature (°C)
50 15 40 10 30
Relative humidity (%)
60
20
5 20 0 r be
r
em
be
ec
ov N
D
em
ob ct O
pt
em
be
er
r
t us Se
A
ug
Ju
ly
ne Ju
ay M
il A
pr
ch ar M
ua br
-5
Fe
Ja
nu
ar
y
ry
10 0
Months Fig. 4. Monthly ambient air and ground properties. Table 2 Climatic data’s of Elazig over cooling season of 2004 [19] Months
Average outdoor temperature (1C) Minimum outdoor temperature (1C) Maximum outdoor temperature (1C) Average relative humidity (%) Average wind velocity (m/s) Average solar radiation (Joule/cm2 day) Average soil temperature at 1 m depth (1C) Minimum surface soil temperature (1C)
June
July
Aug.
Sep.
22.4 14.5 29.5 42.4 2.7 2535 19.7 12.8
26.4 18.3 33.4 35.3 2.7 2522 23.5 16.0
27.1 19.1 34.9 33.8 2.4 2177 25.4 16.9
20.6 13.4 28.7 40.4 2.4 1819 24.8 11.4
conditions to simulate the ACHP system. Table 2 gives the climatic data of the cooling season [19]. The ground characteristics are based on measurements obtained from the Elazig State Meteorological Station (ESMS) [19]. The thermal diffusivity and the thermal conductivity for moist clay ground are aground ¼ 6:71 107 m2 =s and kground ¼ 2:2 W=mK, respectively [14]. Healy and Ugursal [5] have examined the effect of soil type on the energy consumption of GCHPs and found that the energy consumption of GCHPs was not significantly affected by soil type in Nova Scotia. Because of this, the effects of earth conditions (thermal properties) have not been examined in this study. 4.2. Experimental performance information Various experiments have been carried out on the GCHP and ACHP system during the cooling season of 2004. The
mean values of the measured data and calculated results in cooling season are given in Table 3. The heat pump systems were tested, and then compared for very similar cooling loads of the test room. Fig. 5 depicts the daily variation of space cooling rate of the both systems from 01 June to 16 September 2004. The daily mean heat rate (cooling capacity) extracted by the evaporator unit for GCHP-1m, GCHP-2m systems and ACHP system are found to be 5.49, 6.08 and 4.39 (TU ¼ 4:35) kW, respectively. The lowest cooling capacities for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 4.38, 5.07 and 3.24 ðTU ¼ 4:35Þ kW, respectively, in 14–16 August 2004. The highest cooling capacities for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 7.1, 7.5 and 6.26 ðTU ¼ 4:35Þ kW, respectively, in 01–03 June 2004. The cooling capacity increases as the heat sink’s temperature (air/water-antifreeze temperature) decreases, thus lowering the condensation temperature. The cooling
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Table 3 The average experimental results and mean total uncertainties of the two systems in cooling season Item
GCHP-1m
GCHP-2m
ACHP
Unit
Mean total uncertainty (%)
Measured parameters Evaporation pressure Condensation pressure Evaporating temperature Condensing temperature Temperature of water-antifreeze solution at HGHE inlet Temperature of water-antifreeze solution at HGHE outlet Soil temperature in depth of 1 and 2 m Outdoor air temperature Condenser: water-antifreeze solution mass flow rate Condenser: air mass flow rate Evaporator: air mass flow rate Compressor electric current Brine circulating pump electric current Condenser fan electric current Evaporator fan electric current Current of all systems Two-phase voltage Power factor
0.45 1.6 3.0 42.0 36.5 31.2 26 25 0.85 — 1800 6.2 0.38 — 0.69 7.27 220 0.92
0.46 1.53 2.0 40.0 35.1 29.2 23 25 0.85 — 1800 6.1 0.38 — 0.69 7.17 220 0.92
0.4 1.72 6.5 45.0 — — — 25 — 2000 1800 6.5 — 0.89 0.69 8.08 220 0.92
MPa MPa 1C 1C 1C 1C 1C 1C m3 =h m3 =h m3 =h A A A A A V —
2:72 2:72 2:75 2:75 1:38 1:38 1:38 1:38 2:89 2:89 2:89 3:00 3:00 3:00 3:00 3:00 3:00 1:00
Calculated parameters Power input to the compressor Powerinput to the circulating pump Power input to the condenser fan Power input to the evaporator fan Total power of systems The cooling coefficient performance of system The cooling coefficient performance of unit Cooling energy efficiency ratio of the system
1262 77 — 140 1479 3.8 4.5 15.3
1240 77 — 140 1457 4.2 4.9 16.7
1315 — 180 140 1635 3.1 3.8 13
W W W W W — —
4:35 4:35 4:35 4:35 4:35 5:28 5:28 8:09
8 GCHP 1 m
GCHP 2 m
ACHP
7
Capacity (kW)
6
5
4
01
-J 03 une -J -0 u 4 08 ne-J 04 u 14 ne-J 04 u 16 ne-J 04 u 22 ne-J 04 u 28 ne-J 04 u 30 ne-J 04 u 08 ne-J 04 u 14 ly-J 04 u 16 ly-J 04 u 22 ly-J 04 u 28 ly-J 04 u 30 ly-J 04 08 uly-A 04 14 ugt -A -04 16 ugt -A -04 22 ugt -A -04 28 ugt -A -04 30 ugt -A -04 08 ugt -S -04 14 ept-S 04 16 ept-S 04 ep t-0 4
3
Days Fig. 5. Daily cooling capacities for the both systems.
capacity (Fig. 5) of the GCHP system approximates the cooling capacity of the ACHP system since the air and ground temperatures are similar. The cooling capacity of
the GCHP system increases with an increase in depth. This increase in cooling capacity was less than 586 W for a temperature increase of 3 1C over a depth increase of 1 m.
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5.5 GCHP 1 m
GCHP 2 m
ACHP
5
COPsys
4.5
4
3.5
3
2.5
01
-Ju 03 ne-Ju 04 08 ne-Ju 04 14 ne-Ju 04 16 ne-Ju 04 22 ne-Ju 04 28 ne-Ju 04 30 ne-Ju 04 08 ne-Ju 04 14 ly-Ju 04 16 ly-Ju 04 22 ly-Ju 04 28 ly-Ju 04 30 ly-J 04 08 uly-A 04 14 ugt -A -04 16 ugt -A -04 22 ugt -A -04 28 ugt -A -04 30 ugt -A -04 08 ugt-S 04 30 ept-S 04 30 ept-S 04 ep t-0 4
2
Days Fig. 6. Daily cooling COPsys for the both systems.
Fig. 6 shows the cooling performance coefficient of the both systems from 01 June to 16 September 2004. The daily mean cooling performance coefficient (COPsys ) for GCHP1m, GCHP-2m systems and ACHP system are determined to be 3.8, 4.2 and 3.1 (TU ¼ 8:09), respectively. The lowest COPsys for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 3, 3.5 and 2.3 (TU ¼ 8:09), respectively, in 14–16 August 2004. The highest COPsys for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 4.8, 5.2 and 4.5 (TU ¼ 8:09), respectively, in 01–03 June 2004. The COPsys of the ACHP system is lower than the COPsys of the GCHP system since the power consumed by the fan is greater than the power consumed by the circulating pump. Also, outdoor air temperature is significantly higher than the ground temperature in summer. For the GCHP system both the input power and the cooling capacity increases with depth. The gain in cooling capacity was greater and therefore the COPsys increased with depth. The difference between COPsys of the GCHP-1m and COPsys of the GCHP-2m was less than 0.4. A considerable improvement in COPsys was obtained due to the increase in cooling capacity and decrease in input power. The main difference between the two heat pumps is the normal working temperature of the outside circuit. In summer, the water-antifreeze in the external circuit will have a lower temperature than the air that cools the ACHP system, while in winter it is warmer than the air used to carry heat to the ACHP system. Therefore the COPsys of the GCHP will be higher in both modes of operation. The seasonal energy efficiency ratio (SEER) is used to define the average annual cooling efficiency of air conditioning or heat pump system. Naturally, the SEER
for a unit will vary depending on the climatic conditions of the region where it is established. In this study, the SEER of the GCHP-1m, GCHP-2m systems and ACHP system are obtained as 15.3, 16.7 and 13 (TU ¼ 8:09), respectively. The USDOE rule sets efficiency standards for single-phase air-cooled air conditioners and heat pumps at SEER ratings of 12.0 for both split-system heat pumps and single package systems [20]. As a result, especially the SEER of the GCHP system is moderate at longer-term testing. In many applications, the ground temperature in winter can be up to 15–20 1C higher than the air temperature [21], this increases the capacity and the efficiency of a GCHP system. Depending on the geographic localization, heat pump systems with a GHE can show an improvement in the efficiency of the system of 35% in heating mode compared to the conventional ACHP systems. This value reaches up to 40–60% in cooling performance. The advantages of the use of GCHPs compared to conventional ACHPs were shown to be an energy saving technology in the European Mediterranean area [1]. In other geographical areas, such as, Hamburg and Tehessaloniki [22], Nova Scotia [5], South Africa [6–8] and also similar climatic areas, the same conclusion would be applicable to locations with environmental conditions similar to those of Elazig, Turkey. Consequently, in the design of a GCHP system, all parts of the system should be checked in terms of energy efficiency; therefore, it will be necessary to conduct a predesign analysis to determine optimal system parameters that will ensure minimum energy consumption and favourable costs.
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4.3. Economic comparison The energy requirements of the test room for space heating and cooling are 30.6 and 37.2 kWh, respectively, for the day that experiments were conducted. To study the economic feasibility of a system, different methods could be used to evaluate the different figures of merit of the systems. Some examples are: Net present value method, the internal rate of return method, the annual cost and the other methods [23]. The annual cost method has been used to compare the cost effectiveness of the GCHP and ACHP system. Using a lifetime of 20 years for GCHP and ACHP system, an interest rate of 8% and an annual fuel price escalation rate of 4%, the annualized life cycle cost each cooling system has been calculated for the current prices of the electricity in Turkey. For the calculation of the operating expenditures of GCHP and ACHP system, it is assumed that systems run an average 12 hours per day for 108 (01 June–16 September 2004) days. In this period, the price of the electricity consumption for cooling load of test room are calculated as 78, 70 and 95 Euro/seasonal, respectively, for GCHP-1m, GCHP-2m and ACHP systems. The payback periods of the GCHP-1m and GCHP2m systems were found as 3.7 and 4.0 years against an ACHP system. The capital cost of the ACHP system considered in this project is cheaper than that for GCHP systems. However, the running cost of the GCHP systems are less than that of the ACHP system. 5. Conclusions This study compared the performance of two distinct heat pump systems. There is a general review with some experimental work presented. It did mention about the pros and cons of using ground-coupled and air-coupled systems. An experimental study was conducted to assess the techno-economic performance of a GCHP system and an ACHP system in hot and arid climate. The GCHP system is consisting of two HGHE. The comparison revealed that the deeper the HGHE, the greater the improvement in capacities and COPsys . The daily mean COPsys for GCHP-1m, GCHP-2m and ACHP systems are obtained to be 3.8, 4.2 and 3.1 (TU ¼ 8:09), respectively. Also, the economic analysis clearly shows that GCHP systems are economically preferable to the ACHP system.
COPhp C p;air C p;wa
I evaf _ air m _ wa m Q_ conGCHP Q_ conACHP Q_ evaGCHP Q_ evaACHP T ei;air T eo;air T ci;wa T co;wa T ci;air T co;air V_ air U comp U wacp U evaf _ comp W _ wacp W _ evaf W _ conf W wm_ wro wsl wtd wCOPhp wCOPsystem rair cos j
Nomenclature
COPsys
I comp I wacp
cooling coefficient of performance of any system, dimensionless cooling coefficient of performance of heat pump, dimensionless specific heat of air (kJ/kg K) specific heat of water-antifreeze solution (kJ/kgK)
current of compressor (A) current of water-antifreeze solution circulating pump (A) current of evaporator fan (A) mass flow rate of air (kg/s) mass flow rate of water-antifreeze solution (kg/s) heat rejection rate from the ground (kW) heat rejection rate from the air (kW) space cooling load of the GCHP system (kW) space cooling load of the ACHP system (kW) average air temperature entering evaporator unit (1C) average air temperature leaving evaporator unit (1C) inlet average water-antifreeze solution temperature of HGHE (1C) outlet average water-antifreeze solution temperature of HGHE (1C) average air temperature entering condenser unit (1C) average air temperature leaving condenser unit (1C) volumetric flow rate of air (m3 =s) voltage of compressor (V) voltage of water-antifreeze solution circulating pump (V) voltage of evaporator fan (V) power input to compressor (kW) power input to circulating pump (kW) power input to evaporator fan (kW) power input to condenser fan (kW) total uncertainty in the measurement of the mass flow rate (%) uncertainty in the rotameter reading (%) uncertainty associated with the system leakages (%) uncertainty associated with the temperature differences (%) uncertainty associated with the cooling coefficient of the heat pump unit (%) uncertainty associated with the cooling coefficient of the system (%) density of air (kg=m3 ) power factor, dimensionless
Abbreviations
ACHP EER GCHP
air-coupled heat pump energy efficiency ratio ground-coupled heat pump
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
GHE HGHE SEER TU
ground heat exchanger horizontal ground heat exchanger seasonal energy efficiency ratio total uncertainty
Acknowledgement The authors gratefully acknowledge the financial support from the Scientific Research Projects Administration Unit of Firat University for this study performed under project with Grant no. 2001/616. References [1] Romero G, Urchueguia JF, Witte H, Cambien W, Magraner T. Comparative study between a geothermal heat pump system and airto-water heat pump system for heating and cooling in typical conditions of the European Mediterranean Coast, in: Proceedings World geothermal congress, Antalya, Turkey, 24–29 April, 2005. [2] Rottmayer SP. Simulation of ground coupled vertical U-tube heat exchangers. Master of Science, University of Wisconsin-Madison, 1997. [3] Genchi Y, Kikegawa Y, Inaba A. CO2 payback-time assessment of a regional-scale heating and cooling system using a ground coupled heat-pump in a high energy-consumption area in Tokyo. Applied Energy 2002;71(3):147–60. [4] Bi Y, Guo T, Zhang L, Chen L. Solar and ground coupled heat pump system. Applied Energy 2004;78(2):231–45. [5] Healy PF, Ugursal VI. Performance and economic feasibility of ground coupled heat pumps in cold climate. International Journal of Energy Research 1997;21(10):857–70. [6] Petit PJ, Meyer JP. A techno-economic analytical comparison of the performance of air-coupled and horizontal-ground coupled airconditioners in South Africa. International Journal of Energy Research 1997;21(11):1011–21. [7] Petit PJ, Meyer JP. Techno-economic analysis between the performances of heat coupled air conditioners in South Africa. Energy Conversion and Management 1998;39(7):661–9.
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[8] Petit PJ, Meyer JP. Economic potential of ground-coupled heat pumps compared to air-coupled air conditioners in South Africa. Energy 1998;23(2):137–43. [9] De Swardt CA, Meyer JP. A performance comparison between an air-coupled and a ground-coupled reversible heat pump. International Journal of Energy Research 2001;25(10):899–910. [10] Hepbasli A. Performance evaluation of a vertical ground-coupled heat pump system in Izmir, Turkey. International Journal of Energy Research 2002;26(13):1121–39. [11] Esen H, Inalli M, Esen M. Technoeconomic appraisal of a ground source heat pump system for a heating season in eastern Turkey. Energy Conversion and Management 2006;47(9–10):1281–97. [12] Inalli M, Esen H. Experimental thermal performance evaluation of a horizontal ground-coupled heat pump system. Applied Thermal Engineering 2004;24(14–15):2219–32. [13] Inalli M, Esen H. Seasonal cooling performance of a ground-coupled heat pump system in a hot and arid climate. Renewable Energy 2005;30(9):1411–24. [14] Esen H. Investigation of parameters effecting performance of a ground coupled heat pump. MS thesis, Department of Mechanical Engineering, University of Firat, Elazig, Turkey; 2002 [in Turkish]. [15] Glossary of Efficiency Terms, hhttp://www.cwlp.com/Energy_ services/Energy_systems_tech/btui. [16] Coleman HW, Steele WG. Experimentation and uncertainty analysis for engineers. 2nd ed. New York: Wiley; 1999. p. 4. [17] Asan H, Namli L. Uncertainty analysis of experimental studies of heat transfer and pressure drops. in: Proceedings of 11th National heat science congress, Edirne, Turkey; 1997, p. 369–78 [in Turkish]. [18] Holman JP. Experimental methods for engineers, 6th ed. Singapore: McGraw-Hill Book Co.; 1994. [19] Elazig State Meteorological Station, Records for weather data of Elazig, Turkey; 2004. [20] ASHRAE 90.1, Addendum Assist in Energy Use Reduction, www.hvacmall.com/news/article_01114.htm. [21] Closed-Loop/Ground-Source Heat Pump Systems, Installation Guide, International Ground Source Heat Pump Association, Oklahoma State University; 1988. [22] Zogou O, Stamatelos A. Effect of climatic conditions on the design optimization of heat pump systems for space heating and cooling. Energy Conversion and Management 1998;39(7):609–22. [23] Thuesen GJ, Fabrycky WJ. Engineering economy, 7th ed. Englewood Cliffs, NJ: Prentice-Hall International Editions; 1989.
Building and Environment 42 (2007) 1955–1965 www.elsevier.com/locate/buildenv
A techno-economic comparison of ground-coupled and air-coupled heat pump system for space cooling Hikmet Esena,, Mustafa Inallib, Mehmet Esena a
Department of Mechanical Education, Faculty of Technical Education, University of Firat, 23119 Elazig, Turkey b Department of Mechanical Engineering, Faculty of Engineering, University of Firat, 23119 Elazig, Turkey Received 26 January 2006; received in revised form 9 March 2006; accepted 3 April 2006
Abstract This paper reports a techno-economic comparison between a ground-coupled heat pump (GCHP) system and an air-coupled heat pump (ACHP) system. The systems connected to a test room in Firat University, Elazig (38.411N, 39.141E), Turkey, were designed and constructed for space cooling. The performances of the GCHP and the ACHP system were experimentally determined. The experimental results were obtained from June to September in cooling season of 2004. The average cooling performance coefficients (COPsys ) of the GCHP system for horizontal ground heat exchanger (HGHE) in the different trenches, at 1 and 2 m depths, were obtained to be 3.85 and 4.26, respectively and the COPsys of the ACHP system was determined to be 3.17. The test results indicate that system parameters can have an important effect on performance, and that GCHP systems are economically preferable to ACHP systems for the purpose of space cooling. r 2006 Elsevier Ltd. All rights reserved. Keywords: Heat pump; Ground-coupled; Air-coupled; Techno-economic analysis; Cooling; Turkey
1. Introduction A ground-coupled heat pump (GCHP), also called an Earth-coupled heat pump or a geothermal heat pump, operates much like the common an air-coupled heat pump (ACHP) by transferring heat, rather than creating it. Unlike an ACHP, a GCHP transfers heat to and from the Earth to provide space cooling and heating. The most conventional GCHP is the water-to-air heat pump which circulates water through a liquid-to-refrigerant heat exchanger in the ground, with an air-cooled condenser (heat pump) or air-heated evaporator (air conditioner) in the space to be conditioned. The liquid-to-refrigerant heat exchanger consists of a buried thermoplastic piping network which is the ground-coupled heat exchanger. Closed-loop GCHPs have demonstrated the potential to increase capacity and reduce total power consumption by Corresponding author. Tel.: +90 424 2370000/4228; fax: +90 424 2367064. E-mail addresses: hikmetesen@firat.edu.tr (H. Esen), minalli@firat.edu.tr (M. Inalli), mesen@firat.edu.tr (M. Esen).
0360-1323/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.buildenv.2006.04.007
providing a reduced condensation temperature in the summer and a higher evaporation temperature in the winter. Air temperatures which have large daily and monthly cyclic changes are damped or moderated by the Earth mass [1]. The GCHP have high installation costs that make it important to design the system to maximize performance. The GCHP exchange heat with the ground, and maintain a high level of performance even in colder climates for space heating. This results in more efficient use of energy. For this reason many community utilities support the use of GCHP and are active in an effort to persuade the heating ventilating and air conditioning (HVAC) industry to increase the number installed [2]. Heating and cooling heat pumps, providing both space heating and cooling. The most common type is the reversible air-to-air heat pump, which either operates in heating or cooling mode. However, the capacity and performance of ACHPs decrease rapidly with decreasing ambient temperature during heating season, and with increasing ambient temperature during cooling season. The technical and economic performance of a heat pump is closely related to the characteristics of the heat source/sink.
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In particular, a regional heating and cooling system using a GCHP might be considerably more efficient than an ACHP system because the difference of soil temperature and room temperature is smaller than that of air temperature and room temperature. In comparison to systems based on other natural heat sinks such as rivers or sewers, there are no restrictions for the location of the GCHP system. Moreover, heat released underground in summer could be retrieved for the winter heat demand. Thus, year-round energy consumption for climate control might be reduced by the GCHP system. Genchi et al. [3] concluded that the regional heating-and-cooling system using a GCHP is a viable option for reducing CO2 emissions from the viewpoint of life-cycle CO2 emissions that include the emissions associated with the initial construction of the system. Since the idea of a GCHP system was advanced in the late 1940s, a great deal of theoretical and experimental works has been summarized in literature [4]. Healy and Ugursal [5] gave a description of a performance and economic feasibility of GCHP in cold climate. In their study, a comparative economic evaluation was carried out to assess the feasibility of using a GCHP in place of conventional heating/cooling systems and an ACHP system. It was found that all conventional heating systems have smaller coefficient of performance. Petit and Meyer [6–8] have made a techno-economic comparison of a GCHP system and an ACHP system for South Africa climatic conditions. From these studies, they were concluded that GCHP systems are more viable than ACHP systems. In another study, De Swardt and Meyer [9] compared the performance of a GCHP using municipality water as heat source/sink with the performance of an aircoupled heat pump. This comparison was achieved using simulation models that were experimentally validated. The comparison was conducted in both cooling and heating cycles. The results indicated that the utilization of municipality water reticulation systems as a heat source/ sink is a viable method of optimizing energy usage in the air conditioning industry, especially when used in the heating mode. The installations in Turkey on GCHP system are very limited although the temperature of ground is stable in most areas and suitable for GCHP applications [10]. Almost all of the heat pump systems used in Turkey utilize ambient air as the heat source and sink. The usage of other sources and sinks is very rare. Barriers still exist in Turkey regarding the usage of ground and other alternative sources in heat pumps; such sources are not well known, and they are notorious for troublemaking. Another barrier is the lack of designers, contractors and operators who can properly design, install and operate these systems economically and reliably. Esen et al. [11] has reported a detailed techno-economic analysis of a ground source heat pump system and six conventional heating systems for the climate conditions of Turkey in heating season of 2002–2003. In hot climates such as in Turkey, GCHPs represent a viable
alternative to ACHPs and conventional space cooling and heating systems because of their higher operating efficiency, especially during the cooling season. This article focuses on economic and performance comparison of horizontal GCHP system with ACHP system in Turkey. This comparison was achieved by using experimental studies conducted in cooling season. An experimental set-up, described in the next section, was constructed and tested for the first time on the basis of an academic study performed in Elazig, Turkey. 2. Description of the heat pump systems The designed and constructed heat pump systems that will be described are all based on the basic vapourcompression cycle. The schematic views of the constructed GCHP and ACHP system used for space cooling are shown in Fig. 1(a) and (b), respectively. The photographs of horizontal ground heat exchanger of GCHP system and condenser of ACHP system are seen in Figs. 2(a) and (b), respectively. The evaporator of the both systems is shown in Fig. 3. The ACHP system consists of only two circuits: the refrigerant (R-22) circuit and air circuit (condenser fan). The GCHP system mainly consists of three separate circuits: (a) the ground heat exchanger (GHE) circuit or water-antifreeze solution circuit consisting of a horizontal 100 m of pipe length divided into two GHEs, each 50 m length, 0.3 m of pipe distance, 0.016 m of nominal pipe diameter, and 15 m2 spacing were buried with in 1 and 2 m depths in different trenches, horizontal ground heat exchanger; (HGHE1) and horizontal ground heat exchanger; (HGHE2), (b) the refrigerant circuit and (c) the fancoil circuit or air circuit. Combined with Table 1, the main component specifications and characteristics of the ACHP and GCHP system are better understood for this work. In winter, the water-antifreeze solution in the pipes (HGHE1 and HGHE2) of the GCHP system extracts heat from the Earth and carries it into the room. In summer, the system reverses and takes heat from the room and stores it into the cooler ground, as seen in Fig. 1a. The performance evaluations of heating and cooling mode operation of the GCHP system were given in the other studies [12–14]. The heat transfer from Earth to the heat pump or from the heat pump to Earth is maintained with the fluid or waterantifreeze solution circulated through the GHE. In the cooling mode, the refrigerant, which circulates through the evaporator, evaporates by extracting heat from room air and then enters the hermetic compressor. The refrigerant is compressed by the compressor and then enters to the condenser, where it condenses by rejecting heat to the water-antifreeze solution. After the refrigerant leaves the condenser, the capillary tube provides almost 10 1C superheat that essentially gives a safety margin to reduce the risk of liquid droplets entering the compressor. A fan blows across the evaporator to move the cooled air of the
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
1957 Multi channel temperature scanner
Evaporator
:Thermocouples
Compressor Reversing valve
Capillary tubes Refrigerant
Wattmeter
circuit (R 22) Condenser
Tap collectors
Water-antifreeze
Circulating pump
solution circuit
Ground
HGHE2
HGHE1
ea
m
(a)
air
str
:Thermocouples
Ro
om
Multi-channel temperature scanner
Evaporator
Compressor
Refrigerant Capillary tubes
Reversing valve
circuit (R 22)
Wattmeter
Fl
(b)
fro
m
su
ui ds rro trea un m di ng s
Condenser
Fig. 1. Schematic views of the constructed (a) GCHP system and (b) ACHP system.
room. The air velocities were measured with a digital anemometer having a maximum error of 3%. The flow rate of the circulated water-antifreeze solution through the
closed loop HGHE was measured by using a rotameter and controlled by a hand-controlled tap mounted on the collector. The power consumption of the both heat pump
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Fig. 2. Photographs of condensers (a) of GCHP system and (b) of ACHP system.
modes by reversing the refrigerant flow direction. All measured quantities were regularly recorded in every halfhour. A room of 16:24 m2 floor areas was used as a space to be cooled. 3. Energy performance and uncertainty analysis for the both systems The energy efficiencies of the GCHP and ACHP system were evaluated in terms of the COP, defined as the ratio of cooling energy output to the electrical energy input. 3.1. Analysis of the GCHP system The heat rejected from the unit in the cooling mode (HGHE load) Q_ conGCHP is calculated by the following equation. _ wa C p;wa ðT co;wa T ci;wa Þ. Q_ conGCHP ¼ m Fig. 3. Photograph of evaporator of the both systems.
systems, i.e. the electrical power input to the compressor, water-antifreeze circulating pump and evaporator/condenser fan, was measured by means of wattmeter. The inlet and outlet temperatures of the R-22 in the condenser, compressor and evaporator were measured with copperconstantan thermocouples, T-type. In addition, the inlet and outlet temperatures of the circulated water-antifreeze solution through the closed loop HGHE was measured by using the same kind of thermocouples. Also, the temperatures of the ground at various depths, 1 and 2 m, were measured with T-type thermocouples. All measured temperatures were determined by using a multi-channel temperature scanner. The ambient and indoor air temperatures were measured with thermometers. The inlet and outlet pressures of the compressor and evaporator were measured by using Bourdon-type manometers. The reversing valve is used to switch between heating and cooling
(1)
_ comp and the The power input to the compressor W _ evaf are calculated by the Eqs. (2) and (3), evaporator fan W respectively. _ comp ¼ I comp U comp cos j, W
(2)
_ evaf ¼ I evaf U evaf cos j. W
(3)
There is power input to the water-antifreeze circulating _ wacp , at the GCHP system as different from the pump; W ACHP system and it is calculated by _ wacp ¼ I wacp U wacp cos j. W
(4)
The evaporator load equals to the heat extracted from the indoor air stream, Q_ evaGCHP is calculated by Q_ evaGCHP ¼ rair V_ air C p;air ðT eo;air T ei;air Þ.
(5)
The cooling performance coefficient of the GCHP unit, COPhpGCHP , is estimated by COPhpGCHP ¼
Q_ evaGCHP , _ comp W
(6)
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Table 1 Main components specification and characteristics of the ACHP and GCHP studied
Ground-coupled heat pump (GCHP)
Main circuit
Element
Technical specification
Ground coupling circuit
HGHE1, HGHE2 (HGHE1; 50 m) (HGHE2; 50 m) Water-antifreeze solution circulating pump
Pipe distance: 0.3 m; pipe diameter: 0.016 m; HGHE1: piping depth 1 m; HGHE2: piping depth 2 m; material: polyethylene, PX-b Cross link. Manufacturer: Alarko; type: NPVO-26-P; power: 40, 62, 83 W
Fan circuit
Fan of evaporator
Refrigerant circuit
Heat exchanger (Evaporator for cooling) Heat exchanger (condenser) Capillary tube Compressor
Manufacturer: Friterm; Diameter: 350 mm; air volumetric flow rate: 2350 m3 =h, power: 145 W; Manufacturer: Friterm; type: HS 10; capacity: 3.77 kW; heat transfer surface: 10 m2
Air-coupled heat pump (ACHP)
Dryer Observe glass Fan circuit
Fan of condenser
and the average performance coefficients of the GCHP system, COPsysGCHP is calculated by COPsysGCHP ¼
Q_ evaGCHP . _ _ evaf þ W _ wacp W comp þ W
(7)
The cooling equipment systems used in residential and small commercial buildings often express cooling system efficiency in terms of the energy efficiency ratio (EER). This coefficient is defined as the ratio of net cooling capacity (in Btu/h) divided by the electric input in watts under standard rating conditions [15]: EER ¼ 3:412COPhp ,
(8)
where the units of EER are Btu W1 h1 .
Manufacturer: Tunc- Technique; type: TTE 3; capacity: 6.97 kW (with Freon 22); Copper capillary tube; 2 m long; inside diameter: 1.5 mm Manufacturer: Tecumseh Europe; type: hermetic; model: FH 5524 F; the rated power of electric motor driving: 2 HP (1.4 kW); R-22; single phase Manufacturer: Carly; model: DCY 083; connection: 3/8 in; Manufacturer: Carly; model: VCYL 13; connection: 3/8 in; Manufacturer: Friterm; Diameter: 350 mm; air volumetric flow rate: 2350 m3 =h, power: 145 W;
_ conf , at the ACHP system as different condenser fan; W from the GCHP system and it is calculated by _ conf ¼ I conf U conf cos j. W
(11)
The cooling performance coefficient of the ACHP unit, COPhpACHP , is estimated by COPhpACHP ¼
Q_ evaACHP , _ comp W
(12)
and the average performance coefficients of the ACHP system, COPsysACHP is calculated by COPsysACHP ¼
Q_ evaACHP . _ evaf þ W _ conf _ comp þ W W
(13)
3.2. Analysis of the ACHP system
3.3. Uncertainty analysis
The ACHP system cycle is used for space cooling. During cooling, the evaporator coil cools and removes moisture from the indoor air stream, Q_ evaACHP . The condenser is located outdoors and rejects Q_ conACHP (equal _ comp ) to the surroundings. These capato Q_ evaACHP þ W cities are given in Eqs. (9) and (10).
Uncertainty analysis (the analysis of uncertainties in experimental measurement and results) is a powerful tool, particularly when it is used in the planning and design of experiments. As we will see in literature [16], there are realistic, practical cases in which all the measurements in an experiment can be made with 1% uncertainty, yet the uncertainty in the final experimental result will be greater than 50%. Uncertainty analysis, when used in the initial planning phase of an experiment, can identify such situations and save the experimentalist much time, money, and, embarrassment. The experiment collects data of three types, temperature, flow rate, and input power. Total uncertainty (TU) in the measurement of the mass
Q_ evaACHP ¼ rair V_ air C p;air ðT eo;air T ei;air Þ,
(9)
Q_ conACHP ¼ rair V_ air C p;air ðT co;air T ci;air Þ.
(10)
_ comp and the The power input to the compressor W _ evaf are calculated by using Eqs. (2) and evaporator fan W (3), respectively. Also, there is power input to the
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flow rate wm_ may be calculated as follows [17]: wm_ ¼ ðw2ro þ w2sl þ w2td Þ1=2 , wm_ ¼ ð1:252 þ 2:52 þ 0:752 Þ1=2 ¼ 2:89%.
(14)
The result R is a given function in terms of the independent variables. Let wR be the uncertainty in the result and w1 ; w2 ; . . . ; wn be the uncertainties in the independent variables. The result R is a given function of the independent variables x1 ; x2 ; x3 ; . . . ; xn . If the uncertainties in the independent variables are all given with same odds, then uncertainty in the result having these odds is calculated by [18] " 2 2 2 #1=2 qR qR qR wR ¼ w1 þ w2 þ þ wn . qx1 qx2 qxn (15) It is important that all uncertainties used in Eq. (15) can be evaluated at the same confidence level. The uncertainty estimates in the COPhpACHP , COPhpGCHP , COPsysACHP and COPsysGCHP can be calculated from Eq. (15) as 2 6 6 wCOPhpACHP ¼ 6 6 4
wCOPhpGCHP
wCOPsysGCHP
4. Results and discussions 4.1. Ambient and ground conditions Annual air and ground temperatures were needed as boundary conditions for determining the year round performance of GCHP and an ACHP system. These temperatures, given in Fig. 4, were obtained from an experimental measurement. This figure is also show the average temperature of ground in the winter and summer conditions. The ‘‘air’’ temperature line represents the average air temperatures between the average monthly minimum and maximum temperatures. In the cooling season (June–
31=2 2 2 2 qCOPhpACHP qCOPhpACHP qCOPhpACHP wrair þ wV_ air þ wC p;air 7 qrair qC p;air 7 qV_ air 2 2 2 7 7 , qCOPhpACHP qCOPhpACHP qCOPhpACHP 5 þ wT eo;air þ wT ei;air þ wW comp qT eo;air qT ei;air qW comp
31=2 2 2 2 2 qCOPhpGCHP qCOPhpGCHP qCOPhpGCHP wrair þ wV_ air þ wC p;air 7 6 qrair qC p;air 7 6 qV_ air 6 ¼6 2 2 2 7 7 , qCOPhpGCHP qCOPhpGCHP qCOPhpGCHP 5 4 þ wT eo;air þ wT ei;air þ wW comp qT eo;air qT ei;air qW comp
(16)
(17)
31=2 2 2 2 qCOPsysACHP qCOPsysACHP qCOPsysACHP wrair þ wV_ air þ wC p;air 7 qrair qC p;air 7 qV_ air 2 2 2 7 7 qCOPsysACHP qCOPsysACHP qCOPsysACHP þ wT eo;air þ wT ei;air þ wW comp 7 7 , qT eo;air qT ei;air qW comp 7 7 2 2 7 qCOPsysACHP qCOPsysACHP 5 þ wW evaf þ wW conf qW evaf qW conf
(18)
31=2 2 2 2 2 qCOPsysGCHP qCOPsysGCHP qCOPsysGCHP wrair þ wV_ air þ wC p;air 7 6 qrair qC p;air qV_ air 7 6 6 2 2 2 7 7 6 qCOP qCOP qCOP sysGCHP sysGCHP sysGCHP 6 wT eo;air þ wT ei;air þ wW comp 7 ¼6 þ 7 . qT eo;air qT ei;air qW comp 7 6 7 6 2 2 7 6 qCOPsysGCHP qCOPsysGCHP 5 4 þ wW evaf þ wW wacp qW evaf qW wacp
(19)
2
wCOPsysACHP
collects data uncertainties on the experiments. The uncertainties are based on a limited number of tests, so the estimates may change with more testing. From the test data, it can be clearly seen that total uncertainties of two different systems is quite near in this study (see Table 3).
6 6 6 6 6 ¼6 6 6 6 4
It is also important to point out that the uncertainty analysis performed in this work only provided an approach of the effects that would result from the transport of experiment
September) cooling is required, and in the heating season (December–March) heating is required and the ‘‘Air’’ temperatures in Fig. 4 can also be used as external ambient
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
Air
Ground 1.0 m
Ground 2.0 m
1961
Relative humidity
30
80
25
70
Temperature (°C)
50 15 40 10 30
Relative humidity (%)
60
20
5 20 0 r be
r
em
be
ec
ov N
D
em
ob ct O
pt
em
be
er
r
t us Se
A
ug
Ju
ly
ne Ju
ay M
il A
pr
ch ar M
ua br
-5
Fe
Ja
nu
ar
y
ry
10 0
Months Fig. 4. Monthly ambient air and ground properties. Table 2 Climatic data’s of Elazig over cooling season of 2004 [19] Months
Average outdoor temperature (1C) Minimum outdoor temperature (1C) Maximum outdoor temperature (1C) Average relative humidity (%) Average wind velocity (m/s) Average solar radiation (Joule/cm2 day) Average soil temperature at 1 m depth (1C) Minimum surface soil temperature (1C)
June
July
Aug.
Sep.
22.4 14.5 29.5 42.4 2.7 2535 19.7 12.8
26.4 18.3 33.4 35.3 2.7 2522 23.5 16.0
27.1 19.1 34.9 33.8 2.4 2177 25.4 16.9
20.6 13.4 28.7 40.4 2.4 1819 24.8 11.4
conditions to simulate the ACHP system. Table 2 gives the climatic data of the cooling season [19]. The ground characteristics are based on measurements obtained from the Elazig State Meteorological Station (ESMS) [19]. The thermal diffusivity and the thermal conductivity for moist clay ground are aground ¼ 6:71 107 m2 =s and kground ¼ 2:2 W=mK, respectively [14]. Healy and Ugursal [5] have examined the effect of soil type on the energy consumption of GCHPs and found that the energy consumption of GCHPs was not significantly affected by soil type in Nova Scotia. Because of this, the effects of earth conditions (thermal properties) have not been examined in this study. 4.2. Experimental performance information Various experiments have been carried out on the GCHP and ACHP system during the cooling season of 2004. The
mean values of the measured data and calculated results in cooling season are given in Table 3. The heat pump systems were tested, and then compared for very similar cooling loads of the test room. Fig. 5 depicts the daily variation of space cooling rate of the both systems from 01 June to 16 September 2004. The daily mean heat rate (cooling capacity) extracted by the evaporator unit for GCHP-1m, GCHP-2m systems and ACHP system are found to be 5.49, 6.08 and 4.39 (TU ¼ 4:35) kW, respectively. The lowest cooling capacities for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 4.38, 5.07 and 3.24 ðTU ¼ 4:35Þ kW, respectively, in 14–16 August 2004. The highest cooling capacities for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 7.1, 7.5 and 6.26 ðTU ¼ 4:35Þ kW, respectively, in 01–03 June 2004. The cooling capacity increases as the heat sink’s temperature (air/water-antifreeze temperature) decreases, thus lowering the condensation temperature. The cooling
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
1962
Table 3 The average experimental results and mean total uncertainties of the two systems in cooling season Item
GCHP-1m
GCHP-2m
ACHP
Unit
Mean total uncertainty (%)
Measured parameters Evaporation pressure Condensation pressure Evaporating temperature Condensing temperature Temperature of water-antifreeze solution at HGHE inlet Temperature of water-antifreeze solution at HGHE outlet Soil temperature in depth of 1 and 2 m Outdoor air temperature Condenser: water-antifreeze solution mass flow rate Condenser: air mass flow rate Evaporator: air mass flow rate Compressor electric current Brine circulating pump electric current Condenser fan electric current Evaporator fan electric current Current of all systems Two-phase voltage Power factor
0.45 1.6 3.0 42.0 36.5 31.2 26 25 0.85 — 1800 6.2 0.38 — 0.69 7.27 220 0.92
0.46 1.53 2.0 40.0 35.1 29.2 23 25 0.85 — 1800 6.1 0.38 — 0.69 7.17 220 0.92
0.4 1.72 6.5 45.0 — — — 25 — 2000 1800 6.5 — 0.89 0.69 8.08 220 0.92
MPa MPa 1C 1C 1C 1C 1C 1C m3 =h m3 =h m3 =h A A A A A V —
2:72 2:72 2:75 2:75 1:38 1:38 1:38 1:38 2:89 2:89 2:89 3:00 3:00 3:00 3:00 3:00 3:00 1:00
Calculated parameters Power input to the compressor Powerinput to the circulating pump Power input to the condenser fan Power input to the evaporator fan Total power of systems The cooling coefficient performance of system The cooling coefficient performance of unit Cooling energy efficiency ratio of the system
1262 77 — 140 1479 3.8 4.5 15.3
1240 77 — 140 1457 4.2 4.9 16.7
1315 — 180 140 1635 3.1 3.8 13
W W W W W — —
4:35 4:35 4:35 4:35 4:35 5:28 5:28 8:09
8 GCHP 1 m
GCHP 2 m
ACHP
7
Capacity (kW)
6
5
4
01
-J 03 une -J -0 u 4 08 ne-J 04 u 14 ne-J 04 u 16 ne-J 04 u 22 ne-J 04 u 28 ne-J 04 u 30 ne-J 04 u 08 ne-J 04 u 14 ly-J 04 u 16 ly-J 04 u 22 ly-J 04 u 28 ly-J 04 u 30 ly-J 04 08 uly-A 04 14 ugt -A -04 16 ugt -A -04 22 ugt -A -04 28 ugt -A -04 30 ugt -A -04 08 ugt -S -04 14 ept-S 04 16 ept-S 04 ep t-0 4
3
Days Fig. 5. Daily cooling capacities for the both systems.
capacity (Fig. 5) of the GCHP system approximates the cooling capacity of the ACHP system since the air and ground temperatures are similar. The cooling capacity of
the GCHP system increases with an increase in depth. This increase in cooling capacity was less than 586 W for a temperature increase of 3 1C over a depth increase of 1 m.
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
1963
5.5 GCHP 1 m
GCHP 2 m
ACHP
5
COPsys
4.5
4
3.5
3
2.5
01
-Ju 03 ne-Ju 04 08 ne-Ju 04 14 ne-Ju 04 16 ne-Ju 04 22 ne-Ju 04 28 ne-Ju 04 30 ne-Ju 04 08 ne-Ju 04 14 ly-Ju 04 16 ly-Ju 04 22 ly-Ju 04 28 ly-Ju 04 30 ly-J 04 08 uly-A 04 14 ugt -A -04 16 ugt -A -04 22 ugt -A -04 28 ugt -A -04 30 ugt -A -04 08 ugt-S 04 30 ept-S 04 30 ept-S 04 ep t-0 4
2
Days Fig. 6. Daily cooling COPsys for the both systems.
Fig. 6 shows the cooling performance coefficient of the both systems from 01 June to 16 September 2004. The daily mean cooling performance coefficient (COPsys ) for GCHP1m, GCHP-2m systems and ACHP system are determined to be 3.8, 4.2 and 3.1 (TU ¼ 8:09), respectively. The lowest COPsys for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 3, 3.5 and 2.3 (TU ¼ 8:09), respectively, in 14–16 August 2004. The highest COPsys for GCHP-1m, GCHP-2m systems and ACHP system are obtained as 4.8, 5.2 and 4.5 (TU ¼ 8:09), respectively, in 01–03 June 2004. The COPsys of the ACHP system is lower than the COPsys of the GCHP system since the power consumed by the fan is greater than the power consumed by the circulating pump. Also, outdoor air temperature is significantly higher than the ground temperature in summer. For the GCHP system both the input power and the cooling capacity increases with depth. The gain in cooling capacity was greater and therefore the COPsys increased with depth. The difference between COPsys of the GCHP-1m and COPsys of the GCHP-2m was less than 0.4. A considerable improvement in COPsys was obtained due to the increase in cooling capacity and decrease in input power. The main difference between the two heat pumps is the normal working temperature of the outside circuit. In summer, the water-antifreeze in the external circuit will have a lower temperature than the air that cools the ACHP system, while in winter it is warmer than the air used to carry heat to the ACHP system. Therefore the COPsys of the GCHP will be higher in both modes of operation. The seasonal energy efficiency ratio (SEER) is used to define the average annual cooling efficiency of air conditioning or heat pump system. Naturally, the SEER
for a unit will vary depending on the climatic conditions of the region where it is established. In this study, the SEER of the GCHP-1m, GCHP-2m systems and ACHP system are obtained as 15.3, 16.7 and 13 (TU ¼ 8:09), respectively. The USDOE rule sets efficiency standards for single-phase air-cooled air conditioners and heat pumps at SEER ratings of 12.0 for both split-system heat pumps and single package systems [20]. As a result, especially the SEER of the GCHP system is moderate at longer-term testing. In many applications, the ground temperature in winter can be up to 15–20 1C higher than the air temperature [21], this increases the capacity and the efficiency of a GCHP system. Depending on the geographic localization, heat pump systems with a GHE can show an improvement in the efficiency of the system of 35% in heating mode compared to the conventional ACHP systems. This value reaches up to 40–60% in cooling performance. The advantages of the use of GCHPs compared to conventional ACHPs were shown to be an energy saving technology in the European Mediterranean area [1]. In other geographical areas, such as, Hamburg and Tehessaloniki [22], Nova Scotia [5], South Africa [6–8] and also similar climatic areas, the same conclusion would be applicable to locations with environmental conditions similar to those of Elazig, Turkey. Consequently, in the design of a GCHP system, all parts of the system should be checked in terms of energy efficiency; therefore, it will be necessary to conduct a predesign analysis to determine optimal system parameters that will ensure minimum energy consumption and favourable costs.
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
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4.3. Economic comparison The energy requirements of the test room for space heating and cooling are 30.6 and 37.2 kWh, respectively, for the day that experiments were conducted. To study the economic feasibility of a system, different methods could be used to evaluate the different figures of merit of the systems. Some examples are: Net present value method, the internal rate of return method, the annual cost and the other methods [23]. The annual cost method has been used to compare the cost effectiveness of the GCHP and ACHP system. Using a lifetime of 20 years for GCHP and ACHP system, an interest rate of 8% and an annual fuel price escalation rate of 4%, the annualized life cycle cost each cooling system has been calculated for the current prices of the electricity in Turkey. For the calculation of the operating expenditures of GCHP and ACHP system, it is assumed that systems run an average 12 hours per day for 108 (01 June–16 September 2004) days. In this period, the price of the electricity consumption for cooling load of test room are calculated as 78, 70 and 95 Euro/seasonal, respectively, for GCHP-1m, GCHP-2m and ACHP systems. The payback periods of the GCHP-1m and GCHP2m systems were found as 3.7 and 4.0 years against an ACHP system. The capital cost of the ACHP system considered in this project is cheaper than that for GCHP systems. However, the running cost of the GCHP systems are less than that of the ACHP system. 5. Conclusions This study compared the performance of two distinct heat pump systems. There is a general review with some experimental work presented. It did mention about the pros and cons of using ground-coupled and air-coupled systems. An experimental study was conducted to assess the techno-economic performance of a GCHP system and an ACHP system in hot and arid climate. The GCHP system is consisting of two HGHE. The comparison revealed that the deeper the HGHE, the greater the improvement in capacities and COPsys . The daily mean COPsys for GCHP-1m, GCHP-2m and ACHP systems are obtained to be 3.8, 4.2 and 3.1 (TU ¼ 8:09), respectively. Also, the economic analysis clearly shows that GCHP systems are economically preferable to the ACHP system.
COPhp C p;air C p;wa
I evaf _ air m _ wa m Q_ conGCHP Q_ conACHP Q_ evaGCHP Q_ evaACHP T ei;air T eo;air T ci;wa T co;wa T ci;air T co;air V_ air U comp U wacp U evaf _ comp W _ wacp W _ evaf W _ conf W wm_ wro wsl wtd wCOPhp wCOPsystem rair cos j
Nomenclature
COPsys
I comp I wacp
cooling coefficient of performance of any system, dimensionless cooling coefficient of performance of heat pump, dimensionless specific heat of air (kJ/kg K) specific heat of water-antifreeze solution (kJ/kgK)
current of compressor (A) current of water-antifreeze solution circulating pump (A) current of evaporator fan (A) mass flow rate of air (kg/s) mass flow rate of water-antifreeze solution (kg/s) heat rejection rate from the ground (kW) heat rejection rate from the air (kW) space cooling load of the GCHP system (kW) space cooling load of the ACHP system (kW) average air temperature entering evaporator unit (1C) average air temperature leaving evaporator unit (1C) inlet average water-antifreeze solution temperature of HGHE (1C) outlet average water-antifreeze solution temperature of HGHE (1C) average air temperature entering condenser unit (1C) average air temperature leaving condenser unit (1C) volumetric flow rate of air (m3 =s) voltage of compressor (V) voltage of water-antifreeze solution circulating pump (V) voltage of evaporator fan (V) power input to compressor (kW) power input to circulating pump (kW) power input to evaporator fan (kW) power input to condenser fan (kW) total uncertainty in the measurement of the mass flow rate (%) uncertainty in the rotameter reading (%) uncertainty associated with the system leakages (%) uncertainty associated with the temperature differences (%) uncertainty associated with the cooling coefficient of the heat pump unit (%) uncertainty associated with the cooling coefficient of the system (%) density of air (kg=m3 ) power factor, dimensionless
Abbreviations
ACHP EER GCHP
air-coupled heat pump energy efficiency ratio ground-coupled heat pump
ARTICLE IN PRESS H. Esen et al. / Building and Environment 42 (2007) 1955–1965
GHE HGHE SEER TU
ground heat exchanger horizontal ground heat exchanger seasonal energy efficiency ratio total uncertainty
Acknowledgement The authors gratefully acknowledge the financial support from the Scientific Research Projects Administration Unit of Firat University for this study performed under project with Grant no. 2001/616. References [1] Romero G, Urchueguia JF, Witte H, Cambien W, Magraner T. Comparative study between a geothermal heat pump system and airto-water heat pump system for heating and cooling in typical conditions of the European Mediterranean Coast, in: Proceedings World geothermal congress, Antalya, Turkey, 24–29 April, 2005. [2] Rottmayer SP. Simulation of ground coupled vertical U-tube heat exchangers. Master of Science, University of Wisconsin-Madison, 1997. [3] Genchi Y, Kikegawa Y, Inaba A. CO2 payback-time assessment of a regional-scale heating and cooling system using a ground coupled heat-pump in a high energy-consumption area in Tokyo. Applied Energy 2002;71(3):147–60. [4] Bi Y, Guo T, Zhang L, Chen L. Solar and ground coupled heat pump system. Applied Energy 2004;78(2):231–45. [5] Healy PF, Ugursal VI. Performance and economic feasibility of ground coupled heat pumps in cold climate. International Journal of Energy Research 1997;21(10):857–70. [6] Petit PJ, Meyer JP. A techno-economic analytical comparison of the performance of air-coupled and horizontal-ground coupled airconditioners in South Africa. International Journal of Energy Research 1997;21(11):1011–21. [7] Petit PJ, Meyer JP. Techno-economic analysis between the performances of heat coupled air conditioners in South Africa. Energy Conversion and Management 1998;39(7):661–9.
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[8] Petit PJ, Meyer JP. Economic potential of ground-coupled heat pumps compared to air-coupled air conditioners in South Africa. Energy 1998;23(2):137–43. [9] De Swardt CA, Meyer JP. A performance comparison between an air-coupled and a ground-coupled reversible heat pump. International Journal of Energy Research 2001;25(10):899–910. [10] Hepbasli A. Performance evaluation of a vertical ground-coupled heat pump system in Izmir, Turkey. International Journal of Energy Research 2002;26(13):1121–39. [11] Esen H, Inalli M, Esen M. Technoeconomic appraisal of a ground source heat pump system for a heating season in eastern Turkey. Energy Conversion and Management 2006;47(9–10):1281–97. [12] Inalli M, Esen H. Experimental thermal performance evaluation of a horizontal ground-coupled heat pump system. Applied Thermal Engineering 2004;24(14–15):2219–32. [13] Inalli M, Esen H. Seasonal cooling performance of a ground-coupled heat pump system in a hot and arid climate. Renewable Energy 2005;30(9):1411–24. [14] Esen H. Investigation of parameters effecting performance of a ground coupled heat pump. MS thesis, Department of Mechanical Engineering, University of Firat, Elazig, Turkey; 2002 [in Turkish]. [15] Glossary of Efficiency Terms, hhttp://www.cwlp.com/Energy_ services/Energy_systems_tech/btui. [16] Coleman HW, Steele WG. Experimentation and uncertainty analysis for engineers. 2nd ed. New York: Wiley; 1999. p. 4. [17] Asan H, Namli L. Uncertainty analysis of experimental studies of heat transfer and pressure drops. in: Proceedings of 11th National heat science congress, Edirne, Turkey; 1997, p. 369–78 [in Turkish]. [18] Holman JP. Experimental methods for engineers, 6th ed. Singapore: McGraw-Hill Book Co.; 1994. [19] Elazig State Meteorological Station, Records for weather data of Elazig, Turkey; 2004. [20] ASHRAE 90.1, Addendum Assist in Energy Use Reduction, www.hvacmall.com/news/article_01114.htm. [21] Closed-Loop/Ground-Source Heat Pump Systems, Installation Guide, International Ground Source Heat Pump Association, Oklahoma State University; 1988. [22] Zogou O, Stamatelos A. Effect of climatic conditions on the design optimization of heat pump systems for space heating and cooling. Energy Conversion and Management 1998;39(7):609–22. [23] Thuesen GJ, Fabrycky WJ. Engineering economy, 7th ed. Englewood Cliffs, NJ: Prentice-Hall International Editions; 1989.