Theoretical analysis on ground source heat pump

Energy and Buildings 75 (2014) 447–455
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Energy and Buildings
journal homepage: www.elsevier.com/locate/enbuild
Theoretical analysis on ground source heat pump and air source heat
pump systems by the concepts of cool and warm exergy
Rongling Li a,∗ , Ryozo Ooka b , Masanori Shukuya c
a
b
c
Graduate School of Engineering, University of Tokyo, Japan
Institute of Industrial Science, University of Tokyo, Japan
Laboratory of Building Environment, Tokyo City University, Japan
a r t i c l e
i n f o
Article history:
Received 20 December 2013
Received in revised form 31 January 2014
Accepted 6 February 2014
Keywords:
Cool exergy
Warm exergy
Ground source heat pump
Air source heat pump
Exergy efﬁciency
a b s t r a c t
This study presents exergetic characteristics of both ground source heat pump systems (GSHPs) and air
source heat pump systems (ASHPs) based on the concepts of “cool exergy” and “warm exergy”. Quantitative example followed by theoretical analysis shows that GSHPs consume less exergy than ASHPs do.
This is because ﬁrstly “cool exergy” is obtained from the ground in GSHPs, whereas no “cool exergy” is
extracted from the environment by the ASHPs. Secondly, temperature difference between refrigerant
via cooling water and ground in GSHPs is smaller than that between refrigerant and air in ASHPs. In the
GSHP, cool exergy ﬂows into the cooling water from the ground and then enters the indoor air through
the refrigerant cycle. In the ASHP, the refrigerant cycle separates the electricity input of the compressor
into “cool exergy” and “warm exergy.” The “cool exergy” enters the indoor air and the “warm exergy” is
exhausted to the ambient environment. The analysis also shows that compressor requires largest exergy
input among the total exergy inputs, and the exergy consumption in the refrigerant cycle is the highest. Thus, the improvement of the compressor performance to reduce its electricity consumption was
conﬁrmed to be of vital in minimizing unnecessary exergy consumption.
© 2014 Elsevier B.V. All rights reserved.
1. Introduction
For evaluating an energy-related system, the quality of the
energy inﬂow and outﬂow at any point in the system can be determined using the concept of exergy. Exergy is a portion of energy
that can be utilized for work relative to a reference state condition,
in which the exergy value is zero. The exergy method shows the real
change in the work of the system, process by process. This is the
exergy concept presented by Ahern [1]. Exergy analysis has been
applied to many ﬁelds of engineering and science, such as mechanical engineering for optimization of power plants and cogeneration
stations, and food engineering for analyzing processing operations [2]. Several studies have demonstrated the applicability of
the exergy concept to heating and cooling systems [3–7]. These
studies have shown potential ways to improve system energy and
exergy performance, e.g., lowering supply air temperatures [3] and
∗ Corresponding author at: Graduate School of Engineering, University of Tokyo,
Komaba 4-6-1 Meguro-ku, Tokyo 153-8505, Japan. Tel.: +81 3 5452 6434;
fax: +81 3 5452 6432.
E-mail addresses: [email protected], [email protected] (R. Li).
http://dx.doi.org/10.1016/j.enbuild.2014.02.019
0378-7788/© 2014 Elsevier B.V. All rights reserved.
improving insulation of the building envelope [4] to increase the
exergy efﬁciency of the system.
Heat pump systems, especially GSHPs have been widely used on
account of their high energy performance, and the installed capacity has increased dramatically over the last 15 years [8,9]. Some
studies have applied the exergy concept to GSHPs [10–12]. However, these studies have not dealt with warmth and coolness in
the built environment, which are relative to “warm exergy” and
“cool exergy” [6,7]. In order to evaluate the system performance
and indoor thermal comfort, it is necessary to apply the “warm
exergy” and “cool exergy” concepts.
In this paper, based on “warm exergy” and “cool exergy”, exergy
ﬂow pattern from heat pump systems to indoor air is demonstrated for a better understanding of heat pump systems leading to
such a development of low exergy systems. On the basis of energy,
entropy, and exergy balance equations, the entropy and exergy processes of heat pump systems are presented, and a mathematical
model including exergy supply, exergy consumption, entropy generation and entropy disposal for each component is demonstrated.
Furthermore, a case study is presented, where this model is used
for both a GSHP and an ASHP. The exergy consumptions and exergy
efﬁciencies of these two systems are calculated, and the potential
for improvement is discussed.
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R. Li et al. / Energy and Buildings 75 (2014) 447–455
Nomenclature
List of symbols
compressor power [kW]
Ecomp
Efan,ia
indoor fan power [kW]
Epump
power of cooling water pump [kW]
Efan,oa
outdoor fan power [kW]
energy ﬂux between indoor air and refrigerant in
Qe
the evaporator [kW]
Qc
energy ﬂux between refrigerant and cooling water
or outdoor air [kW]
Qg
energy ﬂux between ground and cooling water [kW]
Qc,GS
energy ﬂux between refrigerant and cooling water
in GSHP systems [kW]
Qc, AS
energy ﬂux between refrigerant and outdoor air in
ASHP systems [kW]
To
outdoor temperature [K]
Tg
average ground temperature [K]
refrigerant evaporation temperature [K]
Te
refrigerant condensation temperature [K]
Tc
indoor air temperature [K]
Tia
Tia,sup
supply air temperature [K]
Tw
cooling water temperature [K]
Tw,re
return water temperature [K]
outlet air temperature of condenser heat exchanger
Toa,out
of ASHP [kW]
Tc,GS
refrigerant condensing temperature of GSHP [K]
Tc,AS
refrigerant condensing temperature of ASHP [K]
To−c,AS temperature difference between Tc,AS and the ambient temperature To [K]
Tia−e difference between refrigerant evaporating temperature and indoor air temperature [K]
Tw−g temperature difference between ground and cooling water [K]
Tc,GS−w difference between cooling water temperature and
refrigerant condensing temperature [K]
output exergy from the refrigerant to the indoor air
Xe
at the evaporator [kW]
output exergy from the refrigerant to the cooling
Xc,GS
water of GSHP systems [kW]
Xc,AS
output exergy from the refrigerant to the outdoor
air of ASHP systems [kW]
exergy extracted from ground and delivered to coolXg
ing water [kW]
Xia,sup
supply air exergy [kW]
Xia
return air exergy [kW]
cooling water exergy [kW]
Xw
Xw,re
return cooling water exergy [kW]
exergy contained by ambient air (=0) [kW]
Xo
Xoa,out
outlet air exergy of outdoor fan [kW]
Xrefcycle exergy consumed in the refrigerant cycle [kW]
Xevap
exergy consumed in the heat exchanging process
between indoor air and the refrigerant [kW]
Xcond,GS exergy consumed in the heat exchanging process
between the cooling water and the refrigerant [kW]
Xcond,AS exergy consumed in the heat exchanging process
between outdoor air and condenser [kW]
Xgex
exergy consumed in the heat exchanging process
between cooling water and ground [kW]
Srefcycle entropy generated in the refrigerant cycle [kW/K]
entropy generated in the heat exchanging process
Sevap
between indoor air and the refrigerant [kW/K]
Scond,GS
Scond,AS
Sgex
mia
mw
moa
ca
cw
k
l1
l2
U
entropy generated in the heat exchanging process between the cooling water and the refrigerant
[kW/K]
entropy generated in the heat exchanging process
between the outdoor air and the refrigerant [kW/K]
entropy generated in the heat exchanging process
between cooling water and ground [kW/K]
indoor fan airﬂow rate [kg/s]
cooling water ﬂow rate [kg/s]
outdoor airﬂow rate [kg/s]
speciﬁc heat capacity of air [kJ/kgK]
speciﬁc heat capacity of water [kJ/kgK]
irreversibility factor (the ratio of actual COP to theoretical COP)
circumference of pipe cross section [m]
pipe length [m]
overall heat-transfer coefﬁcient of underground
heat-exchanger pipe [W/m2 K]
2. Basic theory
According to Shukuya [6,7], in a system at a temperature higher
than its environment, exergy ﬂow can be considered as the ﬂow
of thermal energy contained by the system to disperse into the
environment. This exergy is called “warm exergy” ﬂow. It is shown
in Fig. 1(a). In the ﬁgure, the environment temperature To acts as
the cold reservoir and heat Q is extracted from the hot reservoir
with temperature T. The exergy ﬂow Ex is exactly the same as the
maximum amount of work Wmax to be obtained from an imaginary
reversible perfect heat engine.
Ex = Wmax =
1 − To
Q
T
(1)
If the system temperature is lower than the ambient temperature,
then the thermal energy contained by the system is smaller than
the environment. Because of this, heat ﬂows into the system from
the environment. The exergy ﬂow in this condition is “cool exergy”.
Fig. 1(b) illustrates the deﬁnition of “cool exergy” ﬂow. The equation
for “cool exergy” is
Ex = Wmax =
1 − To
(−Q ∗ )
T
(2)
Exergy balance equations are obtained from energy and entropy
balance equations [6,7]. First, following the laws of energy
conservation and entropy generation, energy balance equations
and entropy balance equations are set up in a general form
as
[energy input] = [energy stored] + [energy output]
Fig. 1. Deﬁnition of “warm exergy” and “cool exergy”.
(3)
R. Li et al. / Energy and Buildings 75 (2014) 447–455
449
compressor
expansion valve
indoor
fan
condenser
evaporator
U-tube heat exchanger
indoor
fan
cooling
water pump
condenser
evaporator
compressor
outdoor fan
expansion valve
(b) ASHP
(a) GSHP
Fig. 2. Components of heat pump systems.
vapor enters the condenser and releases thermal energy to the cooling ﬂuid, namely, cooling water in the GSHP and ambient air in the
ASHP. During this process, the refrigerant is condensed from vapor
into liquid. The liquid refrigerant goes through the expansion valve
where its pressure sharply decreases, causing vigorous evaporation
and a dramatic decrease in temperature. Next, the low temperature refrigerant enters the evaporator. In Fig. 2(a), which shows
the GSHP, the refrigerant releases heat into the cooling water in
the condenser. The water is cooled down by the ground through
the underground U-tube heat exchanger. In Fig. 2(b) ASHP, the
refrigerant releases heat into the ambient air.
To assess the energy and exergy ﬂows in the systems process by
process, it is necessary to subdivide the systems into several subsystems [1,6,7,13]. The GSHP is subdivided into four subsystems,
which are shown in Fig. 3. These are (a) refrigerant, (b) inhaled
indoor air at the evaporator where heat is exchanged between the
air and refrigerant, (c) cooling water at the condenser where the
water exchanges thermal energy with the refrigerant, and (d) cooling water in the underground U-pipe heat exchanger where the
water exchanges thermal energy with the ground. The ASHP is subdivided into three subsystems, which are shown in Fig. 3(a), (b), and
(e) inhaled outdoor air at the condenser.
[entropy input] + [entropy generated]
= [entropy stored] + [entropy output]
(4)
Because energy − entropy · To = exergy the exergy balance equation
can be set up by combining Eqs. (3) and (4) as (3) − (4) · To . This
results in
[exergy input] − [exergy consumed]
= [exergy stored] + [exergy output],
(5)
where To is the common ambient temperature for all the components in the system to be analyzed with the unit of Kelvin. The
exergy efﬁciency is deﬁned as
[exergy efﬁciency] =
[exergy output]
[electricity input]
(6)
In the case of GSHPs, the natural exergy is extracted from the
ground. To evaluate the natural exergy use, natural exergy ratio
is deﬁned as
[natural exergy ratio] =
[natural exergy input]
[total exergy input]
(7)
Here, the total exergy input includes the total electricity input and
natural exergy input.
3.1. Refrigerant cycle
Fig. 3(a) shows the refrigerant subsystem in the cycle. According
to the fundamental Eqs. (3) and (4), energy and entropy balance
equations can be presented as Eqs. (8) and (9), respectively.
3. Modeling
Heat pumps can work in both heating and cooling modes by
employing a reversing valve to reverse the ﬂow of refrigerant in
the refrigerant cycle system. The analysis method for these two
working modes is the same. In this study, only the cooling mode is
chosen for exergy analysis.
Fig. 2 shows the heat pump systems operating in cooling mode.
Heat ﬂows into the circulation refrigerant from the indoor air at the
evaporator, and then the refrigerant enters the compressor, where
it is compressed at a constant entropy to convert it into vapor. The
Ecomp + Qe = Qc
(8)
Qe
Qc
+ Srefcycle =
Te
Tc
(9)
Here, Ecomp [kW] is the compressor power, and Qe [kW] is the energy
ﬂow into the refrigerant from the indoor air at the evaporator. Qc
[kW] is the energy ﬂow from the refrigerant toward the cooling
water or outdoor air. Te and Tc [K] are the evaporating temperature
Tw
Te
Qe
Ecomp
Tc,GS
Tia,sup
Tc
Te
Tia
Qe
Qc
(b) Inhaled indoor air
at the evaporator
Epump
Tw
(c) Cooling water at
the condenser
Fig. 3. Subsystems.
Toa,out
Tw,re
E2
Efan,ia
(a) Refrigerant
Qc,GS
Tw,re
Tc,AS
Qg
Qc,AS
Tg
(d) Cooling water in
the underground heat
exchanger
To
Efan,oa
(e) Inhaled
outdoor air at
the condenser
450
R. Li et al. / Energy and Buildings 75 (2014) 447–455
and condensing temperature of the refrigerant, respectively. Srefcycle
[kW/K] is the generated entropy within the refrigerant cycle.
According to the fundamental Eq. (5), exergy balance equations
can be set up by combining Eqs. (8) and (9) as (8) − (9) · To .
For the GSHP,
Ecomp + Xc,GS − Xrefcycle = Xe
Xc,GS =
Xe =
To
Tc,GS
1−
1−
(10)
(11)
(−Qc,GS )
To
Te
(−Qe )
(12)
For the ASHP,
Ecomp − Xrefcycle = Xe + Xc,AS
Xc,AS =
To
Tc,AS
1−
(13)
Qc,AS
(14)
Xrefcycle = Srefcycle To
(15)
As can be seen in Eq. (10), in the GSHP, there are two inputs of
exergy into the refrigerant cycle: Ecomp and Xc,GS , and the output
exergy is Xe . Xc,GS [kW] is the “cool exergy” ﬂowing from cooling
water to the refrigerant through the condenser’s heat exchanger,
and Xe [kW] is the “cool exergy” output from the refrigerant into the
indoor air at the evaporator. In Eq. (11), the negative sign of −Qc,GS
indicates that the direction of energy ﬂow is from the cooling water
to the refrigerant. Similarly, in Eq. (12) −Qe indicates that the direction of energy ﬂow is from the refrigerant to the indoor air. Eq. (13)
shows the exergy balance equation of the ASHP. This equation represents the refrigerant cycle splitting the input exergy Ecomp into
two parts: Xc,AS [kW], which is “warm exergy” ﬂowing from refrigerant to outdoor air through the condenser’s heat exchanger and
Xe [kW], which is “cool exergy”.
In the cooling season, Te is usually lower than To , and Xe can
thus be considered “cool exergy” according to the concept of “cool
exergy” and “warm exergy”. Furthermore, most of the time in the
cooling season Tc,AS is greater than To , and Tc,GS is lower than To ,
meaning that Xc,AS can be considered “warm exergy” and Xc,GS “cool
exergy.”
3.2. Indoor air at the evaporator
[K], and the airﬂow rate is mia [kg/s]. Furthermore, Efan,ia [kW] is
the indoor fan power; Xia and Xia,sup [kW] are the exergy contained
by the return air and that by supply air, respectively. The difference between Xia,sup and Xia is the net “cool exergy” obtained by
the inhaled indoor air that is eventually exhaled into the room to
keep the room air temperature at the set point value. ca [kJ/kgK]
is the speciﬁc heat capacity of air, Sevap [kW/K] is the generated
entropy, and Xevap [kW] is the exergy consumed in the process of
heat exchange between indoor air and the refrigerant.
3.3. Cooling water
The cooling water system is subdivided into two subsystems:
the cooling water at the condenser part and the underground heat
exchanger.
3.3.1. Cooling water at the condenser
The subsystem of cooling water at the condenser part is shown
in Fig. 3(c). The cooling water exchanges heat with the refrigerant.
Energy, entropy, and exergy balance equations of this subsystem
are presented as Eqs. (22), (23) and (24)–(27), respectively.
Epump + cw mw (Tw − To ) = cw mw (Tw,re − To ) + (−Qc,GS )
cw mw ln
Tw,re
Tw
+ Scond,GS = cw mw ln
To
To
−Qc,GS
+
Tc,GS
(23)
Epump + Xw − Xcond,GS = Xw,re + Xc,GS
(24)
Xcond,GS = Scond,GS To
(25)
Xw = cw mw
Xw,re = cw mw
(Tw − To ) − To ln
Tw
To
(Tw,re − To ) − To ln
Tw,re
To
(26)
(27)
The negative sign of −Qc,GS indicates that “cool energy” is ﬂowing
from the cooling water to the refrigerant, which is the same as
shown in Eq. (11). Xc,GS [kW] is the exergy ﬂowing out from the
cooling water to the refrigerant.
3.3.2. Cooling water in the underground heat exchanger
The subsystem with the underground cooling water is shown in
Fig. 3(d), and the balance equations are
cw mw (Tw,re − To ) + (−Qg ) = cw mw (Tw − To )
Fig. 3(b) shows the subsystem where indoor air exchanges heat
with the refrigerant at the evaporator. Energy, entropy, and exergy
balance equations of this subsystem are given by Eqs. (16), (17) and
(18)–(21), respectively, according to the fundamental Eqs. (3)–(5).
(22)
cw mw ln
−Qg
Tw,re
Tw
+
+ Sgex = cw mw ln
Tg
To
To
(28)
(29)
Xw,re + Xg − Xgex = Xw
(30)
(31)
Efan,ia + ca mia (Tia − To ) + (−Qe ) = ca mia (Tia,sup − To )
(16)
Xgex = Sgex To
Tia,sup
Tia
−Qe
+
+ Sevap = ca mia ln
To
Te
To
(17)
Xg =
Efan,ia + Xe − Xevap = Xia,sup − Xia
(18)
Xevap = Sevap To
(19)
In this model, Tg is the average temperature of the ground. −Qg is
considered as one other input energy of the subsystem. Here, the
negative sign indicates the direction is from the ground to the cooling water. Xg [kW] is the extracted “cool exergy” from the ground.
During the process of heat exchange between the cooling water and
the ground, entropy Sgex is generated and exergy Xgex is consumed.
Combining Eqs. (24) and (30), the exergy balance equation for
the cooling water system can be written as
ca mia ln
Xia = ca mia
(Tia − To ) − To ln
Xia,sup = ca mia
Tia
To
(Tia,sup − To ) − To ln
(20)
Tair,sup
To
(21)
The indoor air is assumed to be uniformly distributed and the
indoor air temperature to be the same as the inlet air temperature
of the evaporator Tia [K]. The negative sign of −Qe means that “cool
energy” is ﬂowing from the refrigerant to the indoor air, which is the
same as indicated in Eq. (12). The variable Xe [kW] represents input
exergy from the refrigerant. The supply air temperature is Tia,sup
1−
To
Tg
(−Qg )
Epump + Xg − (Xcond,GS + Xgex ) = Xc,GS
(32)
(33)
Here, Epump and Xg are input exergy, Xcond,GS and Xgex are consumed
exergy, and Xc,GS is the output exergy.
The relationship between the cooling water temperature,
return water temperature, and the parameters of the U-pipe heat
R. Li et al. / Energy and Buildings 75 (2014) 447–455
451
Table 1
Classiﬁcation of Xg .
Temperature relation
Carnot efﬁciency
Warm or cool exergy
Xg
Direction of exergy delivery
Tg < To < Tw,re
Tg < Tw,re < To
To < Tg < Tw,re
To < Tw,re < Tg
Tw,re < Tg < To
Tw,re < To <Tg
−
−
+
+
−
+
Cool
Cool
Warm
Warm
Cool
Warm
+
+
−
+
−
+
Toward cooling water
Toward cooling water
Toward ground
Toward cooling water
Toward ground
Toward cooling water
(34)
Here, l1 and l2 [m] are the circumference and pipe length, respectively. U [W/m2 K] is the overall heat transfer coefﬁcient.
Eq. (35) can be derived from Eqs. (28), (32) and (34).
Xg =
To
1−
Tg
l1 l2 U
cw mw (1 − e− cw mw )(Tg − Tw,re )
(35)
Here, (1 − To /Tg ) is the Carnot efﬁciency. If its value is positive, the
extracted exergy Xg turns out to be “warm exergy,” whereas on
the other hand, if the value is negative, Xg turns out to be “cool
exergy.” Furthermore, whether the value of Xg is positive or negative determines the direction of exergy to be delimited: a positive
value indicates that Xg ﬂows toward the cooling water, whereas a
negative value indicates that Xg ﬂows into the ground. On the basis
of the characteristics of the above-mentioned Carnot efﬁciency and
the sign of the extracted exergy Xg , Xg can be classiﬁed, as listed in
Table 1. This table can help us understand the different states of Xg .
For instance, if Tg < To < Tw,re , Xg is “cool exergy” ﬂowing from the
ground into the cooling water.
3.4. Outdoor air at the condenser
In ASHP systems, the outdoor air exchanges heat with the refrigerant at the condenser (see Fig. 3(e)). The exergy balance equations
of this subsystem can be derived in the same way as for the cooling
water subsystem described in Section 3.3.
Efan,oa + Xc,AS − Xcond,AS = Xoa,out − Xo
Xoa,out = ca moa
(Toa,out − To ) − To ln
(36)
Toa,out
To
Xo = 0
(Ecomp + Efan,ia + Efan,oa ) − (Xrefcycle + Xevap + Xcond,AS + Xoa,out )
= Xia,sup − Xia
The characteristic difference between Eqs. (39) and (40) is that
there is “cool exergy” input from the underground in Eq. (39).
3.6. Modeling coefﬁcient of performance
The theoretical coefﬁcient of performance (COP) of the refrigerant cycle of GSHPs can be written as
COPGS =
(42)
Tia−e = Tia,sup − Te
(43)
Using these temperature differences, COP can be expressed as
COPGS = k
Tia,sup
w
Te
ΔTw-g
ΔTia-e
Te
Ecomp
irreversible
cycle
Tc,GS
Tw
Tc,AS
ww
ΔTc,GS-w
w
(39)
irreversible
cycle
w
= Xia,sup − Xia
Tia,sup
ΔTia-e
ww
ww
(Ecomp + Efan,ia + Epump ) + Xg − (Xrefcycle + Xevap + Xcond,GS + Xgex )
Tia − Tia−e
Te
=k
Tc,GS − Te
(Tg − Tia ) + (Tc,GS−w + Tw−g + Tia−e )
(44)
Here, k is the irreversibility factor of the refrigerant cycles, which
is the ratio of the actual COP value over the theoretical value.
Ecomp
Exergy equations for the entire GSHP and ASHP systems can
be presented on the basis of the equations of all the subsystems
described above. By substituting the Xe of Eq. (18) and the Xc,GS of
Eq. (33) into Eq. (10), the exergy equation for a GSHP can be written
as
(41)
Tc,GS−w + Tw−g = Tc,GS − Tg
Here, the condenser heat exchanger’s inlet air temperature is the
same as the ambient air temperature To , and the exergy of ambient
air Xo is 0. Xc,AS [kW] is the input “warm exergy” ﬂowing from the
refrigerant through the condenser heat exchanger, as explained in
Section 3.1. Xoa,out is exergy exhausted to the environment, and the
difference between Xoa,out and Xo is the net “warm exergy” delivered to the outdoor environment, which is totally consumed in the
course of heat transfer.
3.5. Exergy equations for the GSHP and ASHP
Qe
Qe
Te
=
<
Ecomp
Qc,GS − Qe
Tc,GS − Te
Because of the irreversibility of the refrigerant cycle,
Qe /(Qc,GS − Qe ) < Te /(Tc,GS − Te ). For actual numerical calculation, we assume some temperature differences corresponding to
unavoidable irreversibility, as shown in Fig. 4. In the case of a
GSHP, the temperature differences are deﬁned as Tw−g , Tc,GS−w ,
Tia−e : Tw−g is the difference between the ground temperature
Tg and the cooling water temperature Tw , Tc,GS−w is the difference
between Tw and the refrigerant condensing temperature Tc,GS ,
and Tia−e is the difference between refrigerant evaporating
temperature Te and supply air temperature Tia,sup .
(37)
(38)
(40)
w
l1 l2
Tw = Tg + (Tw,re − Tg )e− cw mw U
Eqs. (18), (36) and (13) give the exergy equation for ASHPs as
w
exchanger can be expressed as an exponential decrease in temperature owing to the distance of ﬂow.
ΔTo-c,AS
To
Tg
(a) GSHP
Fig. 4. Refrigerant cycle models.
(b) ASHP
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R. Li et al. / Energy and Buildings 75 (2014) 447–455
Table 2
Temperatures.
Unit
Table 4
COP.
Tc,AS
317
44
K
◦
C
Tc,GS
300
27
To
Tw
305
32
329
20
Tw,re
297
24
Tg
Tia
Case
AC
GC
AD
GD
292
19
300
27
COP
4.3
11.6
3.0
5.6
than that in cases AC and GC, i.e., 12 ◦ C. The irreversibility factor of
the refrigerant cycle is 0.4 for all four cases.
Table 3
Conﬁguration of investigated cases.
Case
System
Mode
Tia,sup
Te
Tia-e
Tc,GS-w + Tw-g
To-c,AS
AC
GC
AD
GD
ASHP
GSHP
ASHP
GSHP
Cooling
Cooling
Dehumidifying
Dehumidifying
295
295
285
285
290
290
280
280
10
10
20
20
–
8
–
8
12
–
12
–
The COP model for ASHPs is derived by referring to Fig. 4(b), in
the same way as the model for the GSHP.
COPAS = k
Tia − Tia−e
Te
=k
Tc,AS − Te
(To − Tia ) + (To−c,AS + Tia−e )
To−c,AS = To − Tc,AS ,
(45)
(46)
where Tc,AS is the refrigerant condensing temperature of an ASHP,
and To−c,AS is the temperature difference between Tc,AS and the
ambient temperature To .
4. Case study
4.1. Temperature conditions assumed for four cases investigated
For a case study, typical ground and air thermal conditions in
summer in the Tokyo area and an ofﬁce room space with a thermal
“energy” cooling load of 35 kW are assumed. The results obtained
with these assumptions are expected to be useful as a reference at
least for this climate zone. The indoor air temperature is assumed
to be controlled ideally at a ﬁxed value. The same applies to the
water temperature. The compressors of both the GSHP and ASHP
are controlled by an inverter according to the temperature difference between the indoor air and the supply air.
For the present analysis, the following simpliﬁcations are
assumed.
(1) The speciﬁc heat capacities of water and air are constants.
(2) The chemical exergy of humid air caused by condensation is
neglected.
(3) The refrigerant temperature in condenser is constant.
(4) The ground temperature surrounding U-tube heat exchanger
with cooling water is constant.
Table 2 lists the temperatures assumed in this calculation.
Table 3 lists the four cases to be investigated. Supply air temperature Tia,sup is assumed to be 295 K (22 ◦ C) for both the ASHP in
cooling mode (case AC) and the GSHP in cooling mode (case GC).
For dehumidifying mode in the cases of ASHP (case AD) and GSHP
(case GD), the supply air temperature is assumed to be much lower
4.2. Results of calculation
4.2.1. COP
Table 4 lists the COP values of the four cases. The value of COP is
the highest in the GC case because the total temperature difference
is assumed to be the smallest in this case. Furthermore, the COP values in the two GSHP cases are higher than in the ASHP cases because
of the ground temperature Tg being lower than room temperature
Tia and much lower than ambient air temperature To .
4.2.2. Exergy values obtained
From Table 2, we ﬁnd the temperatures are in the relation of
Tg < Tw,re < To . As listed in Table 1, this relation means that the
exergy Xg is “cool exergy” ﬂowing toward the cooling water from
the ground.
To make our analysis method simpler to understand, two cases
are presented here by following the calculation models presented
in Section 3 with numerical values.
Table 5 lists the exergy balance of the GSHP for case GC. First,
we calculated the exergy balance of the cooling water subsystem
with Eq. (33). In this equation, the exergy inputs are Epump and Xg .
The value of Epump is selected according to pump speciﬁcations,
whereas Xg is derived from Eq. (32). By balancing the exergy, Xc,GS
is calculated according to Eq. (11). In this cooling water subsystem,
the consumed exergy is Xgex and Xcond,GS , which are derived from
Eq. (30) and (24), respectively. Xe , Xrefcycle , Xia,sup − Xia and Xevap can
be calculated in the same manner for the two other subsystems:
refrigerant cycle and indoor air. In the refrigerant cycle, the power
of the compressor is calculated from Ecomp = Qe /COP.
Following the same process, all of the exergy values for ASHPs
were calculated. Table 6 lists the results of the ASHP for case AC.
After the exergy balance of the entire system was calculated, the
exergy efﬁciency and the natural exergy ratio were calculated. The
whole procedure of calculation described above was also applied
to cases GD and AD.
4.2.3. Exergy ﬂow, exergy efﬁciency, and natural exergy ratio
The exergy ﬂows though the GSHP and the ASHP are shown
schematically in Fig. 5. Fig. 5(a) shows how in the GSHP, the exergy
ﬂows from the ground to the cooling water before going through the
refrigerant cycle and into the indoor air. The input exergy in this
ﬂow includes not only electricity Epump , Ecomp , and Efan,ia but also
natural exergy Xg . Fig. 5(b) shows how the refrigerant cycle separates electricity Ecomp into “cool exergy” Xe and “warm exergy” Xc,AS .
The “cool exergy” enters the indoor air while the “warm exergy” is
exhausted to the ambient environment. In the ASHP, no natural
exergy is used.
Table 5
Exergy balance of the GSHP for case GC.
System
Exergy input
Exergy output
Exergy consumption
Cooling water Eq. (33)
Xc,GS = 0.63 Eq. (11)
Refrigerant cycle Eq. (10)
Indoor air Eq. (16)
Epump = 1.01
Xg = 1.74 Eq. (32)
Ecomp = 3.02 Xc,GS = 0.63
Efan,ia = 0.60 Xe = 1.81
Xe = 1.81 Eq. (12)
Xia,sup − Xia = 0.88 Eqs. (20), (21)
GSHP Eq. (39)
Epump + Ecomp + Efan,ia = 4.63Xg = 1.74
Xia,sup − Xia = 0.88
Xgex = 0.45 Eq. (30)
Xcond,GS = 1.67 Eq. (24)
Xrefcycle = 1.84 Eq. (10)
Xevap = 1.53 Eq. (16)
Xgex + Xcond,GS + Xrefcycle + Xevap
= 5.49
R. Li et al. / Energy and Buildings 75 (2014) 447–455
453
Table 6
Exergy balance of the ASHP for case AC.
System
Exergy input
Exergy output
Exergy consumption
Refrigerant cycle Eq. (13)
Ecomp = 8.14
Xrefcycle = 4.70 Eq. (13)
Outdoor air Eq. (36)
Efan,oa = 0.74
Xc,AS = 1.63
Efan,ia = 0.60,
Xe = 1.81
Efan,oa + Ecomp + Efan,ia = 9.48
Xe = 1.80 Eq. (12)
Xc,AS = 1.63 Eq. (14)
Xoa,out − Xo = 0.35 Eqs.
(37), (38)
Xia,sup − Xia = 0.88
Xia,sup − Xia = 0.88
Xcond,AS + Xoa,out + Xrefcycle + Xevap = 8.60
Indoor air Eq. (16)
ASHP Eq. (40)
Next, Fig. 6 shows the exergy input and output for all four cases.
Cases AD and GD are compared with cases AC and GC, and the electricity input to the compressors, pumps and fans are calculated on
the basis of the assumption that these components are well controlled. For example, in cases AD and GD, the supply air temperature
is lower than in cases AC and GC, whereas the latent heat load in
all the cases is the same. Therefore, in cases AD and GD, the supply
air ﬂow rate is decreased, and consequently, the fan power input is
reduced proportionally.
In case AC, the total input electricity of the ASHP is 9.48 kW,
and no natural exergy is used. The net exergy output of the system
Xia,sup − Xia is 0.88 kW, which is consumed to meet the demand of
space cooling. Case GC has the same Xia,sup − Xia value, because the
cooling load and supply air temperature in both case GC and case AC
are the same. The total electricity input in case GC is lower than in
case AC because of the higher COP and the use of the natural exergy
Xg of 1.74 kW. In cases AD and GD, the systems are operated in
dehumidiﬁcation mode resulting in the low supply air temperature.
Therefore, the input exergy needed in these cases is higher than in
cases AC and GC.
The exergy efﬁciency of the heat pump systems was calculated
according to Eq. (6). The exergy efﬁciency in all the cases is between
9 and 20%. Both GSHP and ASHP have lower exergy input in cooling
mode than when operating in dehumidifying mode. For GSHP and
ASHP operating the same mode, the GSHP performs better than the
ASHP.
Xcond,AS = 2.02 Eq. (36)
Xevap = 1.53
The natural exergy ratio deﬁned as shown in Eq. (7). In case GC,
the GSHP extracted 1.74 kW natural exergy Xg from the ground, and
the natural exergy ratio was 27.3%. In case GD, the natural exergy
ratio was 20.0% because of the larger exergy demand in the room
space for dehumidiﬁcation.
Furthermore, Fig. 6 shows clearly that the highest electricity
input comes from the compressor Ecomp in all four cases. This means
that both exergy efﬁciency and natural exergy ratio are sensitive to
Ecomp . Thus, reducing the compressors electricity input, i.e., minimizing the exergy consumption for the refrigerant cycle, should
lead to better exergy performance.
4.2.4. Exergy consumption
Fig. 7 shows the calculated exergy consumption for each component. In both GSHP and ASHP systems, Xrefcycle has the highest
value, which means the biggest exergy loss to be avoided exists in
the refrigerant cycle. In cases AC and AD, the values of Xrefcycle are
much higher than those in cases GC and GD. The reason is that the
higher COP of GSHPs requires less electricity input than ASHPs. The
temperature differences between the condensing temperature and
the evaporating temperature, i.e., Tc − Te are 10, 27, 20, and 37 K in
cases GC, AC, GD, and AD, respectively. By comparing Tc − Te and
Xrefcycle for all cases, it is clear that the refrigerant cycle with higher
refrigerant temperature difference results in higher exergy consumption. Thus, a smaller temperature difference inside the
refrigerant cycle should improve its performance.
Xia,sup -Xia
Xia,sup -Xia
Efan,ia
Efan,ia
indoor
air
indoor
air
Xe
Xe
Ecomp
Ecomp
refrigerant
cycle
refrigerant
cycle
Xc,AS
Xc,GS
Efan,oa
Epump
cooling
water
Xg
(a) GSHP
outdoor
air
Xoa,out -Xo
(b) ASHP
Fig. 5. Exergy ﬂow through heat pump systems.
input/output exergy
subsystems
454
R. Li et al. / Energy and Buildings 75 (2014) 447–455
temperature difference between the ambient air and the refrigerant
in an ASHP.
5. Conclusion
In this study, a procedure for exergy analysis of both GSHPs
and ASHPs is presented. On the basis of “cool exergy” and “warm
exergy”, exergy ﬂow pattern of heat pumps is clariﬁed. Quantitative
examples are given for both GSHPs and ASHPs to demonstrate the
presented modeling. The major ﬁndings are summarized as follows.
Fig. 6. Exergy input, output, exergy efﬁciency and natural exergy ratio.
In cases AC and GC, the exergy consumption at the evaporators
Xevap is slightly lower than in cases AD and GD. This is because in
cases AC and GC, the difference between room air temperature Tia
and refrigerant evaporating temperature Te is smaller. Similarly, we
ﬁnd that Xcond in GSHP systems is much lower than that in ASHP
systems. This is because the temperature difference between the
cooling water and the refrigerant in a GSHP is smaller than the
1. In the GSHP, the natural exergy ﬂows from the ground to the
cooling water and then enters the refrigerant cycle. Finally, it
becomes part of the “cool exergy” that acts to cool the indoor
air. In the ASHP, the refrigerant cycle separates the electricity
input from the compressor into the “cool exergy” and “warm
exergy.” The “cool exergy” enters the indoor air and the “warm
exergy” is exhausted to the ambient environment.
2. In general, the exergy consumption in the GSHP is much lower
than that in the ASHP. This is because in GSHPs, the temperature
difference between the heat source and heat sink is smaller than
that in ASHPs, and the heat source temperature in GSHPs is the
ground temperature, which is generally much lower than that in
ASHPs.
3. We ﬁnd that the exergy efﬁciency in all the cases is between
9 and 20%. GSHPs have higher exergy efﬁciency than ASHPs.
Furthermore, the exergy consumption is lower and the exergy
efﬁciency is higher when a system is operated in cooling mode
than when it is operated in dehumidiﬁcation mode. The exergy
efﬁciency of the GSHP is 19.1 and 19.9% when the system is operated in cooling mode and dehumidiﬁcation mode, respectively;
in ASHPs, the exergy efﬁciency is respectively 9.3 and 11.9%.
4. No “cool exergy” from the environment is extracted in ASHPs,
whereas in GSHPs, “cool exergy” is exploited from the ground.
The natural exergy ratio is 27.3% in cooling mode and 20.0% in
dehumidiﬁcation mode.
5. Because the highest electricity input comes from the compressor, and the highest exergy loss exists in the refrigerant cycle,
improving the compressor and reducing the electricity consumption of the refrigerant cycle should reduce the exergy loss.
Furthermore, improving the heat exchangers in the condenser
and evaporator should reduce the temperature difference inside
the refrigerant and decrease the exergy loss. A similar recommendation is given by Akpinar and Hepbasli [12]. Increasing the
supply air temperature in summer, lowering it in winter, and
using low-temperature heating and high-temperature cooling
[14] should reduce the exergy loss in evaporators and condensers.
References
Fig. 7. Comparison of exergy consumption.
[1] J.E. Ahern, Exergy Method of Energy Systems Analysis, John Wiley and Sons,
Hoboken, 1980.
[2] R.P. Singh, D.R. Heldman, Introduction to Food Engineering, 5th ed., Elsevier,
Amsterdam, 2014, pp. 1–64.
[3] H. Torio, D. Schmidt, Development of system concept for improving the performance of a waste heat district heating network with exergy analysis, Energy
and Buildings 42 (2010) 1601–1609.
[4] M. Dovjak, M. Shukuya, B.W. Olesen, A. Krainer, Analysis on exergy consumption patterns for space heating in Slovenian buildings, Energy Policy 38 (2010)
2998–3007.
[5] M. Schweiker, M. Shukuya, Comparative effects of building envelope improvements and occupant behavioural changes on the exergy consumption for
heating and cooling, Energy Policy 38 (2010) 2976–2986.
[6] M. Shukuya, Exergy concept and its application to the built environment, Building and Environment 44 (2009) 1545–1550.
[7] M. Shukuya, Exergy Theory and Applications in the Built Environment, Springer,
New York, 2013.
[8] A.J. Self, B.V. Reddy, M.A. Rosen, Geothermal heat pump systems: status
review and comparison with other heating options, Applied Energy 101 (2013)
341–348.
R. Li et al. / Energy and Buildings 75 (2014) 447–455
[9] J.W. Lund, D.H. Freeston, T.L. Boyd, Direct utilization of geothermal energy 2010
worldwide review, in: Proceedings of the World Geothermal Congress 2010,
Indonesia, 2010.
[10] L. Ozgener, A. Hepbasli, I. Dincer, Exergy analysis of two geothermal district
heating systems for building applications, Energy Conversion and Management
48 (2007) 1185–1192.
[11] A. Hepbasli, M.T. Balta, A study on modeling and performance assessment of
a heat pump system for utilizing low temperature geothermal resources in
buildings, Building and Environment 42 (2007) 3747–3756.
455
[12] E.K. Akpinar, A. Hepbasli, A comparative study on exergetic assessment of two
ground-source (geothermal) heat pump systems for residential applications,
Building and Environment 42 (2007) 2004–2013.
[13] A. Hideo, S. Masanori, Exergy-entropy process of electric lighting systems using
ﬂuorescent lamp, Journal of Architecture, Planning and Environmental Engineering, AIJ 483 (1996) 91–100.
[14] J. Babiak, B.W. Olesen, D. Petras, Low temperature heating and high temperature cooling, embedded water based surface heating and cooling system.
REHVA Technical Task Force 3, 2009.

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