geo thermal heat exchanger

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Development of a thermodynamic performance-analysis program for

CO2 geothermal heat pump system

Young-Jae Kima,*, Keun-Sun Chang b

aDepartment of Bio-Chemical Engineering, Sun Moon University, Asan, Chungnam 337-840, Republic of KoreabDepartment of Mechanical Engineering, Sun Moon University, Asan, Chungnam 337-840, Republic of Korea

1. Introduction

In recent years, refrigeration and air-conditioning equipments

have

been

being

developed

for

more

efficient

and

compact

way

inorder to comply with the internationalmovement of energy saving

and regulation reinforcement on environmental protection. This

tendencyhaspromoted thedevelopmentof geothermalheatpump

systems which are able to correspond efficiently to the change of

cooling and heating loads. To develop new environmental-friendly

geothermal heat pump systems, a suitable refrigerant for each

usage must be selected [1].

The fullyhalogenated chlorofluorocarbons (CFCs)havebeen the

most commonly used until 1980s. However, there is currently a

worldwide trend to seek ozone safe alternative refrigerants to

conventional CFCs [1,2]. For the short-term replacement of CFCs,

HFCs have been being considered as zero ozone depletion (ODP)

refrigerants, but they cannot be free from their high potential of

globalwarming (GWP) [2]. In this reason, much attention has been

being paid on natural refrigerants such as carbon dioxide,

ammonia, air, water, and hydrocarbons as a long-term solution

of alternative refrigerantwith zero ozonedepletion and zero global

warming. Natural refrigerants shown in Table 1 are halogen-free

working fluids based on molecules that occur in nature and are

environmentally benign due to their very low or zero ODP and

GWP [3].

Currently, ammonia is widely used as a refrigerant for large-

scale freezers. For usage, it should be noted that ammonia has

toxicity though flammability is less. Propane and butane have

strong

flammability.

Air

is

used

extensively

as

a

refrigerant

inaircraft industry. Its advantages are to require fewer heat

exchangers, but its efficiency is quite poor. Water has the potential

to be a very efficient refrigerant, but it requires operation in a deep

vacuum. This leads to costly large-volume vacuum tanks thatmust

house all themachinery, such asheat exchangers and compressors.

Among natural refrigerants, CO2 is one of the most promising

alternatives, because it has outstanding thermodynamic, trans-

port, and other environmentally friendly properties. As a result of

continuous efforts to improve efficiency, two-stage CO2refrigera-

tor was developed in 1889 and the multiple-effect CO2cycle was

developed in 1905. However, in the early 19th century, CO2 was

replaced by CFCs due to their excellent characteristics as a

refrigerant. CO2 is now becoming attractive again as an environ-

mentally friendly refrigerant. CO2 has a very low global warming

potential compared to traditional CFCs and HCFCs [46].

The geothermal heat pump is known as a highly efficient

renewable device for heating and cooling houses and buildings as

well as for supplying warm water. During the winter it operates so

as to absorb heat from the underground and reject heat into the

building. Refrigerant is evaporated in coils placed underground

and the vapor is compressed for condensation by water, used to

heat the building, at temperatures above the required heating

level. The geothermal heat pump also serves as air conditioning

during the summer. The flow direction of refrigerant is simply

reversed, and heat is transferred out of the building and back into

the underground coils [7].

Journal of Industrial and Engineering Chemistry 19 (2013) 18271837

A R T I C L E I N F O

Article history:

Received 9 October 2012Accepted 23 February 2013

Available online 4 March 2013

Keywords:

CO2Geothermal heat pump systems

Cycle simulation program

Internal heat exchanger

A B S T R A C T

In this research, a steady-state cycle simulation program for thermodynamic performance analysis of

CO2 geothermal heat pumpsystemswasdeveloped.A seriesof case studies were conductedby changing

systemparametersand operationconditionsin order to investigate theeffect of various systemvariables

on the geothermal heat pump cycle including an internal heat exchanger (IHX). The simulation results

were validated by comparing them with experimental data. The mean deviations of the COPs, cooling

capacities, and compressor powers between experimental and simulation results are 4.5%, 3.8%, 6.5%,

respectively at the 5 8C superheated degree and 32% EEV opening.

2013 TheKorean Society of Industrial andEngineering Chemistry. Publishedby Elsevier B.V. All rights

reserved.

* Corresponding author at: Department of Bio-Chemical Engineering, Sun Moon

University, Asan, Chungnam 337-840, Republic of Korea. Tel.: +82 0415302372.

E-mail address: [email protected] (Y.-J. Kim).

Contents lists available at SciVerse ScienceDirect

Journal of Industrial and Engineering Chemistry

journ al homepage: www.elsev ier .co m/ locate / j iec

1226-086X/$ see front matter 2013 The Korean Society of Industrial and Engineering Chemistry. Published by Elsevier B.V. All rights reserved.

http://dx.doi.org/10.1016/j.jiec.2013.02.028
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A large number of studies have been carried out on heat pump

systems using natural refrigerants in the world. In the early 1990s,

Gustav Lorentzen and his colleagues revived research on the CO2refrigeration cycle in order to address the environment problems

of ozone depletion and global warming effect [8]. They have

concentrated on the experimental evaluation and thermodynamic

modeling of mobile air-conditioning systems and developed a

prototype CO2mobile air-conditioning system through successive

studies [9].

Bullock [10] carried out theoretical performance analysis of

carbon dioxide as a refrigerant in subcritical and transcritical

cycles in a vapor compression cycle. He concluded that the CO2heat pump system would require an efficient expander or

significantly improved compressor and heat exchangers because

it is less efficient than the HCFC-22 (CHClF2) system by 30% in the

cooling mode and 25% in the heating mode.

Hwang and Radermacher [11] theoretically evaluated carbon

dioxide refrigeration cycle by comparing the performance of CO2with HCFC-22 for water heating and chilling modes. They showed

that CO2 is amore desirable refrigerant in the case ofwaterheating

since its performance is about 10% better than HCFC-22.

Brown et al. [12], McEnaney et al. [13,14], and Preissner et al.

[15] explored experimental results for prototype CO2automotive

air conditioner (AC) and compared the results with conventional

HFC-134a (C2H2F4). Hermann and Rene [16] and Hanfner [17]

performed experimental study on water heating CO2 mobile airconditioning system. They studied the performance of the CO2cycle with an internal heat exchanger, and compared their results

with other refrigerant cycle. Adriansyah [18] theoretically and

experimentally investigated a combined air conditioning and tap

water heating plant using CO2. He concluded that the optimum

condition at which the system reaches the highest coefficient of

performance (COP) for cooling is determined by component

parameters such as gas cooler configuration and percentage of

heat recovery. The results showed the total COP of the combined

system is higher than that of the air conditioning system without

heat recovery.

CO2 is a refrigerant that operates at very high pressures in a

transcritical cycle in most operating conditions compared to HFC

refrigerants.

Therefore

the

piping

needs

to

be

25%

thicker

for

a

CO2refrigeration system than for an HFC system in order to withstand

the higher pressure. For the successful replacement and use of

natural refrigerants such as CO2, thermodynamic performance

evaluations of geothermal heat pump systems must be carried out

since CO2 has significantly different thermodynamic properties

from those of conventional refrigerants. For such evaluations, it is

important to develop a thermodynamic performance-analysis

program for predicting the performance of geothermal heat pump

systems. In addition, development of the geothermal heat pump

system requires complex experiments because it includes various

complex variables and their interactions. Therefore, a thermody-

namic performance-analysis program for geothermal heat pump

systems can be effectively used for saving time and reducing the

risk,

which

may

take

place

during

experiment

[19].

In this study, a thermodynamic performance-analysis program

to predict the steady-state performance of the CO2 geothermal

heat pump has been developed and was tested using a series of

case studies to validate the program accuracy. It can simulate the

thermodynamic performance parameters such as COP, cooling and

heating capacities of the indoor and outdoor heat exchangers,

compressor power consumption, etc. This program utilized Visual

Basic for the graphic user interface (GUI), consisted of pre-

processor for inputdata andpost-processor for the output data and

Digital Visual Fortran for the main analysis code. The National

Institute of Standards and Technology (NIST) REFPROP V6.01 was

used for estimating the CO2 thermodynamic and transportproperties and equilibrium behaviors.

2. Modeling of the CO2 geothermal heat pump cycle

The CO2geothermal heat pump system in this study is mainly

composed of thewater cooled indoor andoutdoorheatexchangers,

an internalheat exchanger, a compressor, an expansiondevice, and

a 4-way valve as shown in Fig. 1. The concept shown in Fig. 1

basically represents a vapor compression heat pump cycle.

Depending on the mode of operation (cooling or heating), either

heat exchanger can serve as the evaporator or gas cooler. The

indoor unit serves as an evaporator in cooling mode and as a gas

cooler in heating mode, but the outdoor unit serves as a gas cooler

in cooling mode and an evaporator in heating mode.The word of cooling mode in heat pumping implies a system

managing the indoor temperature below that of the surroundings.

This requires continuous absorption of heat from a low tempera-

ture level,usually accomplishedby evaporation of a refrigerant in a

steady-stateflow process. The vapor formed in the evaporatormay

be returned to its original liquid state for reevaporation. The

refrigerant vapor leaves the evaporator and enters the compressor

at the vaporizing temperature and pressure and it is simply

compressed and then cooled in the gas cooler without condensa-

tion in the case of the transcritical CO2 cycle as shown in Fig. 2. The

cooled liquid leaves the gas cooler and enters the expansiondevice.

The pressure of the liquid is reduced to the evaporating pressure as

the liquid passes through the expansion device. In the CO2 heat

pump

cycle

a

liquid

evaporating

at

constant

pressure

provides

ameans for heat absorption at constant temperature. Likewise,

cooling of the vapor in the transcritical state,after compression to a

higher pressure, provides for the rejection of heat. The liquid from

the gas cooler is returned to its original state by an expansion

process.

3. Heat exchangers

In the present study, a multi-tube heat exchanger, which

contains a number of parallel smaller tubes enclosed in a larger

tube, was used for the gas cooler, evaporator, and internal heat

exchanger. The high pressure CO2 flows through the inner tubes

and the low-pressure water flows through the annular space

between

the

inner

tubes

and

the

outer

tube.

The

heat

exchangers

Table 1

Characteristics of some natural refrigerants [4].

Refrigerant R744 R717 R290 R600 R600a R1270

Chemical formula CO2 NH3 C3H8 n-C4H10 i-C4H10 C3H6Molar mass 44.01 17.03 44.10 58.12 58.12 42.08

Critical temp. (8C) 30.98 132.25 96.68 152.0 134.67 92.40

Boiling point (8C) 78.40 33.33 42.09 0.5 11.67 47.7

Critical pres. (kPa) 7384 11,333 4247 3796 3640 4665

ODP 0 0 0 0 0 0

GWP 1

0

3

3

3

3Toxicity No Yes No No No No

Y.-J. Kim, K.-S. Chang/Journal of Industrial and Engineering Chemistry 19 (2013) 182718371828


3/11

were designedwith the counter-flow patternwhereCO2 andwater

streams areflowing in oppositedirections in order tomaximize the

heat transfer efficiency. In the case of the gas cooler, CO2 inthe transcritical state enters the inner tubes and then is cooled by

the countercurrentlyflowing coldwater through theannulusof the

gas cooler.

The section-by-section method [21,22] shown in Fig. 3 was

used for performance analysis of a countercurrent multi-tube heat

exchanger. Performance analysis using the section-by-sectionmethod can be applied to very complex refrigerant circuits

including superheated phase, two-phase, and subcooled region

as well as transcritical region. The energy balance equation that

describes the flow of CO2and water for each discretized node via

the section-by-section method may be written as

D Qn mcHc;j Hc;j1 mwHw;j Hw;j1 (1)

In thecase of thegas cooler, the CO2inlet temperature (Tc,1) of

the first section (n = 1) is known from the compressor outlet

conditions estimated through compressor simulation. Therefore,

Fig. 1. Schematic diagram of the CO2 geothermal heat pump system.

Fig. 2. Temperatureentropy diagram of the CO2 heat pump cycle [20]. Fig. 3. Control section for section-by-section method.

Y.-J. Kim, K.-S. Chang/Journal of Industrial and Engineering Chemistry 19 (2013) 18271837 1829


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if thewater inlettemperature (Tw,2) at thefirstsection is assumed

as the initial guess, the outlet temperaturesof CO2 andwater flow

are calculated through the above energy balance equation

including quantity of heat transfer calculated by e-Ntu method.

At every control section (nth) after the first section, the CO2inlet

temperature (Tc,j) is known from the previous simulation at the

(n 1)th control section and then, the water inlet temperature

(Tw,j+1) can be iteratively calculated by finding the value of

convergence to the target value (Tw,j) known from simulation at

the previous control section. In the last control section the

iteratively estimated water inlet temperature is compared to the

water inlet temperature given as input data. If the difference

between two values is not fall within the error limit, the water

inlet temperature (Tw,2) assumed at the first control section as an

initial guess is iteratively changed until the convergence is

reached.

In e-Ntu method, quantity of heat transfer ( Qn) in a control

section can be obtained using heat exchanger effectiveness (e) and

the number of transfer units (Ntu).

Qn eCminTc;in Tw;in (2)

NtuUnAn

Cmin(3)

where C is heat capacity mCp, Cminmeans the smaller value among

heat capacities of the water and CO2.

Heat transfer effectiveness (e) for the countercurrentflow in the

gas cooler without phase change is estimated with the number of

transfer units (Ntu) as follows:

e 1 expNtu1 Cr

1 CrexpNtu1 Cr (4)

CrCminCmax

If the CO2 flow in a control section at the evaporator is a two-phase

fluid,

Crbecomes

zero.

Therefore,

Eq.

(4)

can

be

written

as

Eq.

(5).e 1 expNtu (5)

The equation to calculate overall heat transfer coefficient, Un, of

each control section is given as follows:

1

Un

DAo

DAihn;c

DAolnro=ri

2pkDz

1

hn;w(6)

where k is the thermal conductivity of tube wall.

The heat transfer rate in a control section can be determined

from an energy balance on the CO2 and water flows and can be

expressed as:

Qn;CO2 mCO2 CpCO2 TCO2i TCO2j1 (7)

Qn;H2O mH2O CpH2O TH2Oj TH2Oj1 (8)

Therefore, the outlet temperatures of the CO2and the water flows

are determined to be

TCO2j1 TCO2j Qn;CO2

mCO2:CpCO2(9)

TH2Oj TH2Oj1 Qn;H2O

mH2O CpH2O(10)

The CO2 geothermal heatpump system incorporates an internal

heat exchanger (IHX) which is installed between the outlet of the

gas cooler and the evaporator. Thus, the CO2 flow at the high-

pressure

side

of

the

gas

cooler

is

liquefied

from

the

transcritical

state due to heat rejection through IHX and the CO2 flow at the

low-pressure side of the evaporator becomes a superheated gas

state by heat absorption. One purpose of the internal heat

exchanger is to further cool the CO2 flow from the gas cooler by

exchanging heat with the CO2 flowing out from the evaporator.

This increases the amount of CO2 in liquid phase (lower quality)

flowing into the evaporator, and thus increases the cooling

performance, which in turn results in increase of the COP of the

geothermal heat pump system. In this research, a countercurrent

multi-tube heat exchanger was used as an internal heat exchanger

and the performance of the internal heat exchanger was also

analyzed by section-by-section method.

The overall heat transfer coefficient (U) shown in Eq. (6) was

calculated based on the waterside and CO2-side heat transfer

coefficients. The Gnielinski [23] or Petukhov [24] equations were

used for estimating the CO2-side heat transfer coefficient at

transcritical region in the gas cooler. The Gnielinski correlation

shown in Eq. (11) is used for 2300 Re 104.

hi f =2Re 1000Pr

1 12:7ffiffiffiffiffiffiffiffiffif=2

p Pr2=3 1

kiDi

(11)

where f 1:58lnRe 3:282.

The

Petukhov

correlation

for

104

Re

5

106

is

expressed

as

hi f =8RePr

1:07 12:7ffiffiffiffiffiffiffiffiffif =8

p Pr2=3 1

k iDi

(12)

where f = (0.79 ln(Re) 1.64)2.

The waterside heat transfer coefficient for the gas cooler was

calculated by using the DittusBoelter correlation [25] for as

shown in Eq. (13).

ho 0:023R0:8e Pr

0:4koDh

heating (13)

The CO2-side heat transfer coefficient in two-phase region at

the evaporator was estimated with the Gungor and Winterton

correlation

[26]

and

the

waterside

heat

transfer

coefficient

wascalculated by using DittusBoelter correlation expressed in

Eq. (14).

ho 0:023R0:8e Pr

0:3koDh

cooling (14)

In the case of the internal heat exchanger, the CO2-side heat

transfer coefficient in high pressure was calculated with the

Gnielinski or Petukhov equations and that in low pressure was

estimated by using the DittusBoelter correlation for heating.

4. Compressor

The compressor simulation was carried out on the basis of the

loss

and

efficiency-based

compressor

model

[27]. The

loss

andefficiency-based compressor model estimates the internal energy

balances in a compressor fromdesign, internalefficiency,andheat-

loss values specified by user. A schematic diagram of the loss and

efficiency compressor model is represented in Fig. 4. Ten unknown

variables shown in Fig. 4 are: (1) CO2 mass flow rate ( mc), (2)

enthalpy at the suction port (hsuction port), (3) enthalpy at the

discharge port (hdischarge port), (4) enthalpy at the shell outlet

(houtlet), (5) work done on the CO2( Wc), (6) work done on the shaft

( Ws), (7) work input to the compressor ( Wcm), (8) rate of heat loss

due to cooling of compressor and motor ( Qcooling), (9) rate of

compressor shellheat loss ( Qcan),and (10) rate ofheat transfer from

the discharge gas to the suction gas ( Qhillo). These ten unknowns

are iteratively calculated from the following 10 independent

equations:



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energy balance equations

(1) energy balance between the compressor shell inlet and

suction port

input = output

mchsuctionport hinlet Qhilo Qcooling Qcan (15)

(2) energy balance between the suction port and discharge port

mchdischargeport hsuctionport Wc (16)

(3) energy balance between the discharge port and compressor

shell outlet

mchoutlet hdischargeport Qhilo (17)

seven defining equations

(4)

Qhilo ahilo Wcm; actual (18)

where ahilois the fraction of compressor power consump-

tion transferred from the discharge line to the suction line

(specified by user or 0.03 as a default value)

(5)

Qcan acan Wcm; actual (19)

where acan is the fraction of the compressor power

consumption which is rejected from the shell to the

ambient air (specified by user or 0.9 (1.0 hmotorhmech))

(6)

Qcooling 1 hmotor hmech Wcm (20)

(7) energy balance between the suction port and discharge port

Wc mchisen;discharge port hsuction port

3413 hisen(21)

where hisenis the isentropic efficiency based on the suction

port

(8)

hmechWcWs

; Ws Wc

hmech(22)

(9)

hmotorWsWcm

; Wcm Ws

hmotor(23)

(10) hvol,suction port: volumetric efficiency based on the suction

port

mc hvol; suction port D Soper=ysuction port (24)

hvol; suction port mc;actualysuction port

D Soper

where D is the total compressor displacement [in3]; Soper is the

actual compressor motor speed [rpm]; ysuction port is the

specific volume at suction port.

Fig. 4. Schematic diagram of compressor energy balance.



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5. Expansion device

One important device in the geothermal heat pump cycle is the

expansion device. The purpose of the expansion device is to reduce

the pressure of the refrigerant and to control refrigerant mass flow

rate in the system. The common expansion devices widely utilized

in geothermal heat pump systems are capillary tubes, short tube

orifices, thermostatic expansion valve (TXV), and electronic

expansion valve (EEV). Among the expansion devices, the EEV

genders attention recently due to a wide range of operating

condition and high capability of load control.

In this research, the electronic expansion valvewasused and its

performance was evaluated with the equation given by Hwang

et al. [28]. The mass flow rate for CO2at the EEV can be estimated

with following equations:

p1 c1p2c2 p3

c3p4c4p5

c5 (25)

Five dimensionless p-numbers and the coefficients used in this

correlation are shown in Table 2. The parameters in Table 2 are

defined as follows:

m, mass flow rate, At,m, minimum area of orifice, rin, density at

the

EEV

inlet, Dp,

pressure

drop

across

expansion

valve,

L,effective orifice length, Dm, minimum orifice diameter, Do,

orifice diameter, pin, pressure at the EEV inlet, pc, critical

pressure of CO2, Tin, temperature at the EEV inlet, Tc, critical

temperature of CO2. It was reported that this empirical

correlation derived from Buckingham p-theorem predicts the

CO2 mass flow rate through the EEV within 5.4% errors.

6. Thermodynamic performance analysis for the CO2geothermal heat pump cycle

6.1. Procedures for thermodynamic performance analysis

Fig. 5 shows the flow diagram for the thermodynamic

performance-analysis

program

developed

in

the

present

study.The input data required to start the program are as follows: tube

diameter and length for outer and inner tubes, number of inner

tubes,flow rateand inlet temperature ofwater asa secondaryfluid,

expansion device specifications, and the degree of superheat at the

inlet of the compressor. Furthermore, the performance-analysis

program requires the geometric dimensions (diameter and length)

of thepipes in the cycle in order topredict thepressuredrops in the

geothermal heat pump cycle. Thermodynamic performance

analysis is based on the following assumptions:

(1) steady state operation;

(2) countercurrent flow in all type of heat exchangers;

(3) neglecting the heat loss through the heat exchangers and

expansion

devices;

(4) neglecting the change of kinetic and potential energy;

(5) neglecting pressure drop of thewater flow as a secondary fluid.

The compressor suction and discharge pressure are assumed to

be the main iterative variables as shown in Fig. 5. Based on

assumed compressor suction anddischargepressures, theCO2 flow

rate at the compressor and the conditions at compressor outlet are

estimated using the compressor module. The compressor inlet

conditions are also estimated by the degree of superheat given in

input datum. The gas cooler inlet conditions are calculated from

the compressor outlet conditions by considering the pressure drop

between the compressor outlet and the gas cooler inlet. And then,

the outlet conditions of the gas cooler including heat duty in the

gas cooler are estimated by using the gas cooler module on the

basis of the conditions atgas cooler inlet.Based on gas cooler outlet

and compressor inlet conditions estimated earlier, the inlet

conditions of the expansion device and the outlet conditions of

the evaporator are calculated using the internal heat exchanger

module. CO2flow rate in the expansion device is computed using

expansiondevicemodule and it is compared with theCO2 flow rate

estimated at the compressor module. The compressor discharge

pressure is iteratively adjusted using Secant method until the

difference between the CO2 flow rate at the compressor and the

CO2 flow rate in the expansion device is within a prescribedtolerance. After simulating the evaporator on the basis of outlet

conditions at the expansion device and evaporator specifications

given as input data, the compressor suction pressure is iteratively

adjusted using Secant method until the difference between the

evaporator outlet enthalpy computed at evaporator simulation

and enthalpy estimated at internal heat exchanger simulation is

within a prescribed limit.

The output of the thermodynamic-performance program

includes the COP, the CO2flow rate, compressor power consump-

tion, cooling capacity in the evaporator, heating capacity in the gas

cooler, line pressure drops, etc.

7. Experimental apparatus

The experimental apparatus as shown in Fig. 6 is comprised of a

compressor, indoor and outdoor heat exchangers, an internal heat

exchanger, an expansion device, an oil separator recovering the oil

from the compressor, and accumulator located at the exit of the

evaporator. In addition, aby-pass linewas installed to carry out the

comparative experimentsaccording to the existence of the internal

heat exchanger. The specifications for the CO2 geothermal heat

pump system are summarized in Table 3. A 4-way valve was also

equipped to choose the operation mode such as cooling and

heating. As shown in Table 3, all the equipment in the cycle was

designed to withstand 40 MPa pressure and all instruments and

fittings are able to safely operate at more than 20 MPa pressure.

The temperatures, pressures, flow rates, and power consump-

tions

at

the

important

points

of

the

cycle

were

measured

using

T-type thermocouple probes, pressure transducers, and a mass flow

meter. The uncertainties for the instruments are estimated as

0.2% for the pressure measurements, 0.2 8C for the temperature

measurements, 0.2% for the flow measurements, and 0.01% for the

integrating W-m.

Before operating the experiment, the system was vacuumed

first by the vacuum pump, and then proper amount of CO2 was

charged. Optimum amount of refrigerant charge was determined

at the highest COP, and found to be 2200 g without the internal

heat exchanger and 2400 g with the internal heat exchanger. All

data were collected when they reached the steady state. After

receiving sufficient amount of data, the EEV opening was changed

for the next test condition. A set of experiments for various EEV

openings

were

performed,

in

order

to

analyze

the

thermodynamic

Table 2

Five dimensionless p-numbers and coefficients in the correlation.

p1 p2 p3 p4 p5

m

At;m

ffiffiffiffiffiffiffiffiffiffiffiffi ffiffirinDp

q LDm

DmD0

pinpe

TinTc

Constant Value

C1 1.17100

C2 3.99102

C3 7.27

102

C4 3.86101

C5 4.55100

Units: m (kg/s),At,m (m2), rin (kg/m

3), Dp (Pa), L (m), Dm (m), Do (m), pin(Pa),pc(Pa),

Tin(K), Tc (K).



7/11

performance parameters of the geothermal CO2 heat pump system

such as COP, compressor power, and the CO2 flow rate.

Experimental test conditions are presented in Table 4. The

experimental data were collected by averaging data measured

at every 10 s for 5 min, and decided to be valid when the

uncertainties were simultaneously maintainedwithin the limits of

range for temperatures with 0.1 8C, pressures with 5 kPa, and

flow rate with 0.2 g/s error bound.

8.

Results

and

discussion

In the present study, thermodynamic performance character-

istics of theCO2 geothermalheatpumpwere investigatedusing the

computer simulation. A number of case studies were carried out to

validate the simulation program. System performance character-

istics such as COP, cooling and heating capacities, CO2flow rates,

and compressor powers were expressed for various EEV openings

and for over a range of compressor frequencies.

In Fig. 7, simulation results are compared with experimental

data for the COP, cooling capacity, and compressor power in

coolingmode tovalidate the thermodynamicperformance analysis

of the simulation program. The mean deviations of the COPs,

cooling capacities, and compressor powers between experimental

and

simulation

results

are

4.5%,

3.8%,

6.5%,

respectively

at

the

5 8C

superheated degree and 32% EEV opening. The comparison

indicated that the simulation results were in good agreement

with those from the experiment with reasonable accuracy.

Variations of COP and cooling capacity with respect to

compressor frequencies ranged from 30 Hz to 55 Hz for 4 different

EEV opening positions at 5 8C of superheated degree in cooling

mode are shown in Fig. 8, and for compressor power in Fig. 9. As

shown in Figs. 8 and 9, the cooling capacity and compressor power

increase as compressor frequency increases as generally expected.

However,

the

consistent

linearity

of

the

increasing

rate

indicatesthe stable operability of the heat pump system. On the contrary,

the COP decreases upon increasing compressor frequency. In

general, the increase of compressor frequency induces an increase

of the refrigerant flow rate and pressure difference or compression

ratio. Increasing rate of high-pressure side due to frequency

increase is more noticeable than the decreasing rate of the low-

pressure side. The refrigerant flow rate increase leads to cooling

capacity increase, whereas the increase of pressure difference or

compression ratio results in required compressor power increase.

Since the increasing rate of compressorpower ishigher than thatof

cooling capacity, the consequent calculation of COP becomes

smaller with the compressor frequency increase in most cases.

In the case of EEV opening, itwas observed that theCOP, cooling

capacity,

and

compressor

power

all

decrease

with

increasing

EEV

Fig. 5. Flow chart of the thermodynamic-performance analysis program.



8/11

opening. The CO2 inlet temperature at the evaporator increases

with increasing EEV opening and the heat exchanging rate is

reduced due to the decrease in the inlet temperature difference

between CO2and secondary fluid of water. On the other hand, CO2flow rate increases with the increase of EEV opening, and

consequently

cooling

capacity

increases

with

the

increase

of

flow

rate. For the just given operating conditions in this study, these

combined effects result in the decrease of the COP and cooling

capacity with increasing EEV opening.

Fig. 10 shows variations of CO2 flow rate with respect to

compressor frequency and EEV opening at the 5 8C superheated

degree

in

cooling

mode.

As

seen

in

Fig.

10, CO2flow

rate

increaseswith increasing compressor frequency and EEV opening. As

explained earlier, CO2 flow rates increase with the increase of

compressor frequency and EEV opening due to expansion of the

throat area of EEV.

Variation of COP with respect to superheated degree and

compressor frequency at 20% EEV opening in cooling mode is

represented in Fig. 11. It is indicated that the COP decreases

linearlywith respect to the increment of the superheateddegree in

the range of39 8Cwhen EEV opening is kept at a value of20%. This

can be interpreted from the fact that COPs of the cycle are reduced

because compressor power increases more rapidly than cooling

capacity with an increase in the superheated degree.

Fig. 12 represents the comparison ofCOPs of the systemwith an

IHX

(internal

heat

exchanger)

and

a

basic

cycle

without

the

IHX.

Fig. 6. Schematic diagram of experimental apparatus.

Table 3

Specifications of the geothermal heat pump system.

Compressor Rotary type

Maximum pressure: 45MPa

Power requirement: 10.5kW

Frequency response: 3070Hz

Expansion valve Maximum pressure: 40MPa

Allowable flow rate: 0100g/s4-way valve Maximum pressure: 40MPa

Heat exchangers Types Outer tube Inner tube Path

OD/thickness (mm) OD/thickness (mm) No. No. L (m)

Outdoor exchanger Multitube 25.4/1.2 4/0.5 8 2 1.7

Indoor exchanger Multitube 19.05/1.2 4/0.5 8 2 1.65

Internal exchanger Multitube 19.05/1.2 4/0.5 8 2 1

Table 4

Test conditions for the geothermal heat pump system.

Test conditions Operation

Cooling Heating

Compressor freq. (Hz) 3060

Compressor (rpm) 18003600

CO2charge (g) 18002400

EEV opening (%) 1050

Water inlet temperature at the evaporator (8C) 17 12

Water flow rate at the evaporator (kg/s) 0.283 0.217

Water inlet temperature at the gas cooler (8C) 25 30

Water flow rate at the gas cooler (kg/s) 0.333 0.283



9/11


10/11

inlet temperature at the evaporator just after the EEV increases

with increasing EEV opening, so that the heating capacity is

decreased due to the increase of the inlet temperature difference

between CO2 and water. However, in the case of heating mode,compressor power was kept almost constant as the EEV opening

increased. As a result, the COPs were reduced as the EEV opening

increased.

The CO2 flow rate as a function of various EEV openings and

compressor frequencies at the 5 8C superheated degree in heating

mode is represented in Fig. 15. As shown in Fig. 15, the CO2flow

rate has a tendency to increase as the compressor frequency andEEV opening increase.

Fig. 16 shows variation of COPs with respect to superheated

degree and compressor frequency at the 16% EEV opening in

heating mode. The COPs in heating mode decrease as the

superheated degree increases in the range of 39 8C due to the

same reason mentioned for cooling mode. However, COPs in

heatingmode more steeply decrease than the case of coolingmode

with the increase of compressor frequencies.

Fig. 17 indicates the comparison of COPs as a function of

compressor frequency for the cycle with IHX (internal heat

exchanger) and a basic cycle without IHX in heating mode. The

22% EEV opening was determined to be optimum value for the

basic cycle and 16% EEV opening for the cycle with the IHX in

heating mode. In contrast to the cooling mode, the COPs of theheating cyclewith the IHXwere slightly smaller than those of basic

cycle at all compressor frequencies (1.32.4% smaller with IHX in

the frequency range of 3055 Hz). With the use of IHX, the heating

capacity increases due to the decrease of CO2 quality at the

evaporator inlet resulting from the lowered temperature of CO2leaving from the gas cooler. On the contrary, the heating capacity

2.5

2.7

2.9

3.1

3.3

3.5

3.7

3.9

4.1

4.3

4.5

25 30 35 40 45 50 55 60

COP

cooling

Compressor frequency[Hz]

no internal heat exchanger

with internal heat exchanger

(EEV opening : 32%)

(EEV opening : 20%)

Fig. 12. Cooling performance with compressor frequency for IHX.

1

2

3

4

5

6

7

8

9

1.8

2.0

2.2

2.4

2.6

2.8

3.0

3.2

3.4

3.6

3.8

25 30 35 40 45 50 55 60

COPheating


Heatingcapacity&Po

wer[kW]

COP_heating

Heating capacity

Compressor power

EEV opening: 14%

Superheated degree: 5C

Fig. 13. Variation of COPheating, heating capacity and compressor power for

compressor frequency (heating, superheated degree: 5 8C, EEV opening: 14%).

1

2

3

4

5

6

7

8

9

2.5

2.6

2.7

2.8

2.9

3.0

25 30 35 40 45 50 55 60

COPheating


Heatingcapacity&Power[kW]

COP_heating

Heating capacity

Compressor power

EEV opening 20%


Fig. 14. Variation of COPheating, heating capacity and compressor power for

compressor frequency (heating, superheated degree: 5 8C, EEV opening: 20%).

0.02

0.025

0.03

0.035

0.04

0.045

0.05

25 30 35 40 45 50 55 60

CO2flowrate[Kg/s]


EEV 16%

EEV 20%

EEV 24%

EEV 28%


Fig. 15. CO2 flow rate with EEV opening and compressor frequency (heating,

superheated degree: 5 8C).

2.7

2.8

2.9

3

3.1

3.2

3.3

25 30 35 40 45 50 55 60

COPheating


3C

5C

7C

9C

EEV opening: 16%

Superheated degree

Fig. 16. Variation of COPheatingfor superheated degree and compressor frequency

(EEV opening: 16%).



11/11

decreaseswith thedecreaseofCO2flow rate due to the reduction of

the EEV opening from22%without IHX to16%with IHX. As a result,

the heating capacity increases by about 8.7% with the use of IHX,

whereas the required compression power increasesby about10.6%due to the increase of the compressor outletpressure resulted from

the reduction of EEV opening. In overall, it was founded that the

use of IHX in heatingmode is not helpful to increase the COP of the

CO2geothermal heat pump system.

9. Conclusions

In the present study, a thermodynamic-performance analysis

program for CO2 geothermal heat pump system incorporating the

internal heat exchanger was developed. It consists of several

subroutines for modeling an evaporator, gas cooler, internal heat

exchanger, compressor, and electronic expansion valve and

estimating the thermodynamic and transport properties of CO2

and water. This program can be used to simulate the steady-statethermodynamic performances ofCO2geothermal heat pumpsystem

such as COP, heating and cooling capacity, power consumption, etc.

A number of case studies were carried out in order to validate the

thermodynamic analysis program and the simulation results were

compared with the experimental results. It was found that this

programmaybehighly advantageous to save time and to reduce the

risk, whichmaytakeplace inexperiment. This simulation program is

intended to serve as auseful tool fora thermodynamic performance

analysis when optimizing complex system variables and establish-

ing efficient operating conditions in the CO2 geothermal heat pump

systems. In the future the capabilities of this program will be more

improved for geothermal heat pump systems using natural

refrigerants such as ammonia and hydrocarbons. The following

important

results

were

obtained

in

this

study:

(1) e-Ntumethodwasused for simulating theheat exchangers such

as the evaporator, gas cooler, and internal heat exchanger.

(2) The compressor simulation was carried out on the basis of the

loss and efficiency-based compressor model which estimates

the internal energy balances in a compressor from design,

internal efficiency, and heat-loss values specified by user.

(3) The performance of the electronic expansion valve was

evaluated with the empirical equation derived from Bucking-

ham p-theorem.

(4) The mean deviations of the COPs, cooling capacities, and

compressor powers between experimental and simulation

results are 4.5%, 3.8%, 6.5%, respectively. The comparison

indicated that the simulation results were in agreement with

those from the experiment with reasonable accuracy.

(5) It was found that the COPs of the cycle with the IHX were

improved than those of the basic cycle by 26% in cooling

mode.

(6) In contrast with the cooling mode, the COPs of the heating

cycle with the IHX were slightly smaller than those of basic

cycle. It indicated that the use of IHX in heating mode has no

effect to increase the COP of the CO2 geothermal heat pump

system.

References

[1] R.J. Dossat, Principles of Refrigeration, 4th edition, Prentice-Hall, 1997.[2] M.P. Desai, Design Optimization and Development of an Energy-Efficiency Vapor

Compression Cooling System, 1997, p. 1.[3] S.B.Riffat, C.F.Afonso,A.C. Oliveira,D.A. Reay,AppliedThermalEngineering 17 (1)(1997) 33.

[4] N. Kagawa, International Journal of Air-Conditioning and Refrigeration 15 (4)(2007) 182.

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[6] Y. Hwang, M. Ohadi, R. Radermacher, Feature Article, The American Society ofMechanical Engineers, 1988.

[7] J.M. Smith, H.C. Vaness, M.M. Abbott, Introduction to Chemical EngineeringThermodynamics, 5th edition, McGraw-Hill, 1996,, p. 307.

[8] G. Lorentzen, J. Pettersen, in: Proceedings of the IIR International Symposium onRefrigeration, Energy, and Environment, Trondheim, Norway, (1992), p. 147.

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burg, MD, (1997), p. 20.[11] Y. Hwang, R. Radermacher, HVAC&R Research 4 (3) (1998) 245.[12] J.S. Brown, F. Samuel, Y. Motta, P.A. Domanski, International Journal of Refrigera-

tion 25 (2002) 19.

[13] R.P. McEnaney, D.E. Boewe, J.M. Yin, Y.V. Park, in: Proceedings of the SeventhInternational Refrigeration Conference at Purdue University, West Lafayette,Indiana, (1998), p. 145.

[14] R.P. McEnaney, Y.C. Park, J.M. Yin, P.S. Hrnjak, SAE International Congress andExposition, Detroit, Michigan, 1999 (Paper No. 1999-01-0872).

[15] M. Preissner, B. Cutler, S. Singanamalla, Y. Hwang, R. Radermacher, in: Proceed-ings of4th IIR-GustavLorentzenConferenceonNaturalWorkingFluids at Purdue,West Lafayette, IN, (2000), p. 185.

[16] H. Hermann, R. Rene, 4th IIR-Gustav Lorentzen Conference, 2000, p. 43.[17] A. Hanfner, 4th IIR-Gustav Lorentzen Conference, 2000, p. 177.[18] W. Adriansyah, Energy and Buildings 36 (2004) 690.[19] Y.J. Kim, K.S. Chang,H.H. Kim, Journal of Industrial Engineeringand Chemistry 13

(5) (2007) 674.[20] S.G. Kim, M.S. Kim, Korean Journal of Air-conditioning and Refrigeration Engi-

neering 15 (6) (2003) 471.[21] P.A. Domanski, NISTIR 89-4133, NIST, 1989.[22] B. Youn, H.Y. Park, Y.S. Kim, Proceedings of the SAREK 1996 25 (2) (1996) 151.[23] V. Gnielinski, International Chemical Engineering 16 (2) (1976) 359.[24] B.S. Petukhov, Advances in Heat Transfer 6 (1970) 503.[25] P.W. Dittus, L.M.K. Boelter, University of California Publication in Engineering 2

(13) (1930) 443.[26] K. Gungor, R. Winterton, International Journal of Heat and Mass Transfer 29

(1986) 351.[27] S.K. Fischer, C.K. Rice, ORNL/CON-80/R1, Oak Ridge National Lab, 1980.[28] Y.W. Hwang, O.J . Kim, in: Proceedings of the SAREK 2007 Summer Annual

Conference, 2007, p. 1237.

2.5

2.6

2.7

2.8

2.9

3

3.1

3.2

3.3

3.4

3.5

25 30 35 40 45 50 55 60

COPheating


no internal heat exchanger

with internal heat exchanger

(EEV opening : 22%)

(EEV opening :

Fig. 17. Heating performance with compressor frequency for IHX.


geo thermal heat exchanger

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