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Abstract—
In this paper, a detailed simulation model of a solar-
powered triple-effect LiBr–H
2
O absorption chiller is developed to
supply both cooling and heating demand of a large-scale building,
aiming to reduce the fossil fuel consumption and greenhouse gas
emissions in building sector. TRNSYS 17 is used to simulate the
performance of the system over a typical year. A combined energetic-
economic-environmental analysis is conducted to determine the
system annual primary energy consumption and the total cost, which
are considered as two conflicting objectives. A multi-objective
optimization of the system is performed using a genetic algorithm to
minimize these objectives simultaneously. The optimization results
show that the final optimal design of the proposed plant has a solar
fraction of 72% and leads to an annual primary energy saving of 0.69
GWh and annual CO
2
emissions reduction of ~166 tonnes, as
compared to a conventional HVAC system. The economics of this
design, however, is not appealing without public funding, which is
often the case for many renewable energy systems. The results show
that a good funding policy is required in order for these technologies
to achieve satisfactory payback periods within the lifetime of the
plant.
Keywords—
Economic, environmental, multi-objective
optimization, solar air-conditioning, triple-effect absorption chiller.
I. I
NTRODUCTION
HE primary energy use in buildings has been dominated
by conventional air-conditioning systems which
contribute to about 40% of the greenhouse gas emissions in
building sector [1], [2]. With the rising demand for indoor
comfort, rising concerns about climate change and depletion
of fossil fuel resources, finding an environmentally friendly
and energy-efficient alternative to conventional air-
conditioning systems is necessary [3]. Solar heating and
cooling (SHC) absorption chillers are considered as one
promising alternative to conventional air-conditioning systems
since much of the technology has already been proven at
commercial scales. Absorption chillers are mainly categorized
by the number of effects which refer to the number of times
the high temperature heat source is used by the chiller to
produce cooling. Moving toward a higher effect cycle leads to
higher chiller COPs, but in turn requires higher driving
A. Shirazi is with The University of New South Wales, Sydney, Australia
(phone: +61-413077896; e-mail: a.shirazi@unsw.edu.au).
R. A. Taylor and G. L. Morrison are with The University of New South
Wales, Sydney, Australia (e-mail: robert.taylor@unsw.edu.au,
g.morrison@unsw.edu.au).
S. D. White is with Commonwealth Scientific and Industrial Research
Organization (CSIRO) Energy Centre, Newcastle, Australia (e-mail:
Stephen.D.White@csiro.au).
temperature. Single-effect chillers operate in the temperature
range of 80 °C to 100 °C, achieving thermal COPs of up to
0.7-0.8 [4]. Double-effect chillers can achieve higher COPs up
to 1.4, but require significantly higher driving temperatures of
around 180 °C [5]. Having three cascading generators, triple-
effect absorption chillers can produce cooling at even higher
COPs of around 1.8, but require a very high-temperature heat
source of 210-240 °C [6]. Such temperatures can be supplied
by using high temperature concentrating collectors such as
parabolic troughs.
Most of solar absorption chillers installed around the world
are based on single-effect chillers and non-concentrating flat
plate or evacuated tube collectors [7]–[10]. The major
disadvantage to single-effect chillers is its low COP, which
leads to a large collector area to supply the thermal heat
demand of the chiller. Combining high-temperature solar
thermal collectors and multi-effect absorption chillers can be
more energy-efficient due to the higher COP of these chillers,
resulting in less solar thermal energy and potentially less
collector area to supply a given amount of cooling [11], [12].
Cabrera et al. [13] carried out a comprehensive literature
review on the use of parabolic trough collectors (PTCs) in
solar air-conditioning applications and summarized the
existing experiences and realizations for the potential
application of these collectors to drive double-effect
absorption chillers. They found that the yearly installation rate
of this type of systems is still low, but according to the market
potential, this rate is expected to increase in the near future.
The potential of triple-effect absorption systems coupled with
high-temperature solar thermal collectors have been studied by
a few researchers [14], [15]. Although triple-effect absorption
chillers have been around for a while, there has been no
system-level modeling and optimization of these chillers
coupled with high-temperature concentrating solar thermal
collectors for air-conditioning applications. Motivated by this
gap, this paper presents a detailed multi-objective optimization
of a SHC triple-effect LiBr–H
2
O absorption chiller, aiming to
determine the optimal performance of the system from
energetic, economic, and environmental perspective
simultaneously. A complete dynamic simulation model of this
system was developed in TRNSYS 17 [16], where its energy
performance was evaluated. A detailed economic analysis was
conducted to calculate the total levelized cost of the plant,
including the capital cost, operating and maintenance cost, the
fuel cost, and a penalty cost for CO
2
emissions. A multi-
objective optimization approach using a genetic algorithm was
Multi-Objective Optimization of a Solar-Powered
Triple-Effect Absorption Chiller for Air-Conditioning
Applications
Ali Shirazi, Robert A. Taylor, Stephen D. White, Graham L. Morrison
T
World Academy of Science, Engineering and Technology
International Journal of Energy and Power Engineering
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employed through coupling TRNSYS and MATLAB to
determine the optimal design of the system. The optimization
objectives were the primary energy consumption and the total
levelized cost of the system, which were both minimized.
Finally, a LINMAP decision-making method was used to
determine the final optimal design of the system, which was
compared to that of a reference conventional system.
II.S
YSTEM
D
ESCRIPTION
In this paper, PTCs were coupled with a triple-effect LiBr–
H
2
O absorption chiller in order to supply both cooling and
heating demand of a large reference building. This improves
the economics of the plant compared to those producing either
chilled or hot water alone. It should be noted that PTCs are the
only solar thermal collectors available on the market, which
are capable of delivering temperatures greater than 210 °C
with an acceptable thermal efficiency. As illustrated in Fig. 1,
a gas burner is employed as the backup system, delivering heat
to the chiller when solar-driven energy is not sufficient. The
rated cooling capacity of the absorption chiller is selected in
order to satisfy the maximum cooling load of the building.
Fig. 1 A generic layout of a solar-assisted absorption chiller with a
gas-fired backup
The incident solar radiation absorbed by the collector arrays
increases the temperature of the stratified hot water storage
tank. A feedback controller is used to adjust the pump flow
rate to achieve a fixed set-point temperature at the collector
outlet. A pressure relief valve is used to limit the fluid outlet
temperature to the maximum allowable value specified by the
user as an input to the model. During cooling periods when the
temperature at the top 75% of the storage tank is above the
chiller’s required hot water temperature, only the solar heat
source is used to drive the absorption chiller to produce chilled
water for space cooling. Once the temperature at the top of the
tank drops below the required value at the chiller’s generator,
the gas burner is switched on to supply the entire cooling
demand of the building. The chilled water produced by the
chiller is delivered to the cooling fan coil using a fixed flow
pump to satisfy the required building cooling load. The chiller
is also connected to a cooling tower loop to remove heat from
its absorber and condenser into the ambient. During heating
periods, the hot water stored in the tank is directly delivered to
the heating fan coil unit to cover the building heating load.
Similarly, if the temperature of the storage tank drops below
the required temperature for the distribution system, the gas
burner is turned on to satisfy the entire heating load.
A conventional air-conditioning system, consisting of a
vapor compression mechanical chiller and a gas-fired heater,
is considered here as a reference system to compare the
performance of the proposed solar triple-effect absorption
chiller from energetic, economic, and environmental
viewpoints.
III. S
YSTEM
S
IMULATION
The proposed SHC absorption chiller system was modeled
in TRNSYS 17 [16], which is commonly used for simulation
of transient thermal systems. The following details the
mathematical model of the key system components.
A. Parabolic Trough Solar Collector
The useful heat collected from a PTC module (
u,SC
Q
) can be
calculated as [17]:
u,SC w p, w SC, out SC,in
L
abbavga
n
SC
QmcT T
FU
AF K G T T
CR
(1)
L12avga
FU c c T T
(2)
where
n
F
is the collector zero loss efficiency at normal
incidence, K
θb
(θ) and G
b
denote the incidence angle modifier
for beam radiation and the solar beam irradiance on the
collector surface, F′ and U
L
are the collector efficiency factor
and overall heat loss coefficient, and A
a
and CR
SC
are the
collector aperture area and concentration ratio. The terms c
1
and c
2
are the first and second order heat loss coefficients, and
T
avg
and T
a
are the average temperature of the collector
working fluid and ambient temperature, respectively. A new
type (labelled ‘Type 237’) was developed to simulate the
performance of the PTC module in TRNSYS environment.
The design parameters used for this type are representative of
NEP Polytrough 1800 [18], which are summarized in Table I.
B. Hot Water Storage Tank
A TRNSYS ‘Type534’ TESS model [19] was used to
simulate the performance of the stratified storage tank in this
study. The tank was assumed to have 10 thermal stratification
nodes, an aspect ratio of 3.5 [20], and a heat loss coefficient of
0.833W/m
2
K. More details about the mathematical model of
the tank can be found in TRNSYS TESS Library
documentation [19].
C. Solar-Gas Heat Source Controller
Combining the mathematical equations of ‘Type 6’ (an
auxiliary gas-fired heater) and ‘Type 11 (a tee-piece) from the
TRNSYS standard library with some control-logic necessary
to switch between solar and gas modes, a new TRNSYS type
(labelled ‘Type 223’) was developed to model the solar-gas
heat source component. The outlet temperature of the heater
World Academy of Science, Engineering and Technology
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can be determined by [21]:
AH loss ,AH
AH, out AH ,in
w,AH p, w
QQ
TT mc
(3)
where
AH
Q
and
loss, AH
Q
are the burner heating rate and heat
losses through the burner, respectively.
TABLE I
P
ERFORMANCE
P
ARAMETERS OF THE
PTC U
SED IN THI S
S
TUDY
[18]
Parameter Unit Value
Aperture area m
2
18.45
Max operating pressure bar 40
Test flow rate L/h m
2
75.9
Collector azimuth ° 0
Concentration ratio - 54
n
F -0.689
c
1
W/m
2
K0.36
c
2
W/m
2
K
2
0.0011
D. Absorption Chiller
The triple-effect absorption chiller was modeled using the
adapted characteristic equation method capable of predicting
the chiller’s performance by using two algebraic equations to
calculate the chiller cooling capacity and the driving heat input
as functions of a term called ∆∆T′ [22]. ∆∆T′ can be expressed
as:
GACE
TTaT eT
(4)
where T
G
, T
AC
, and T
E
are the average temperature of the
external heat carrier fluids at the generator, absorber-
condenser, and evaporator. The characteristic equations for the
cooling capacity and driving heat of the chiller are defined as:
EE E
Qs Tr
(5)
GG G
Qs Tr
(6)
The characteristic coefficients (i.e. a, e, s
E
, r
E
, s
G
, and r
G
)are
determined using multiple linear regression algorithms applied
to a set of performance data points given by the chiller
manufacturer. The heat removal rate from the absorber and
condenser and the chiller COP can be determined by:
AC E G aux
QQQQ
(7)
E
Gaux
Q
COP QQ
(8)
where aux
Q
is the energy consumption of the absorption
chiller pump. Based on this method, a TRNSYS type (labelled
‘Type 220’) was developed to simulate the performance of the
absorption chiller. The triple-effect chiller was selected from
Thermax Ltd [23], and its technical specifications at rated
conditions are summarized in Table II.
TABLE II
T
ECHNICAL
S
PECIFICATIONS OF THE
T
RIPLE
-E
FFECT
A
BSORPTION
C
HILLER
U
SED IN
T
HIS
S
TUDY
Parameter Unit Value
CHW temperature (inlet/outlet) °C 12/7
CW temperature (inlet/outlet) °C 30/37
HW temperature (inlet/outlet) °C 210/195
CHW flow rate m
3
/hr kW
c
0.172
CW flow rate m
3
/hr kW
c
0.191
HW flow rate m
3
/hr kW
c
0.032
COP - 1.81
E. Vapor Compression Chiller
A water-cooled vapor compression chiller was employed in
the reference conventional system to provide the building
cooling load. TRNSYS ‘Type 666’ was used to model the
performance of the chiller. The detailed mathematical model
of this type can be found in TRNSYS documentation [24].
F. Cooling Tower
An open circuit cooling tower was used to reject heat from
the chiller to the ambient. TRNSYS ‘Type51b’ was used to
simulate the performance of the cooling tower. More details
about the mathematical model of this type can be found in
TRNSYS documentation [21].
G. Building Load
The building considered in this paper represents a large
hotel consisting of 6 floors and a basement with total floor
area of 11,346 m
2
. The building has an average window-to-
wall ratio of 30.2% in total. The building walls and fabric
were designed to meet minimum requirements provided by the
Building Code of Australia [25]. The thermal behavior of the
building was simulated using multi-zone building model
‘Type 56’ and the weather data module ‘Type 15-2’ in
TRNSYS.
IV. E
NERGETIC
A
NALYSIS
The energy performance of the proposed SHC triple-effect
chiller is analyzed by the primary energy consumption of the
plant (PEC
SHC
):
SHC SHC, E SHC,NG
PEC PEC PEC
(9)
SHC,E E E
PEC PEF E
(10)
SHC, NG NG NG
PEC PEF E
(11)
E
E
and E
NG
are the annual energy consumed by the electric
equipment and the auxiliary gas burner. PEF
E
and PEF
NG
,
respectively, denote the primary energy conversion factors for
electricity grid and natural gas, which are presented in Table
III.
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V. E
CONOMIC
A
NALYSIS
The levelized annual total cost of the plant (C
tot,L
) consists
of four main elements: the capital investment (CI
L
), the
operating and maintenance cost (OMC
L
), the fuel cost (FC
L
),
and the pollution cost due to CO
2
emissions (CDEC
L
):
tot, L L L L L
CCIOMCFCCDEC
(12)
The input parameters used for the economic analysis of the
proposed SHC absorption system are summarized in Tables III
and IV [26]–[28].
TABLE III
I
NPUT
P
ARAMETERS
U
SED FOR THE
E
NERGETIC
,E
CONOMIC AND
E
NVIRONMENTAL
A
NALYSES OF THE
SHC A
BSORPTION
C
HILLER
P
LANT IN
T
HIS
S
TUDY
[26]
Parameter Unit Value
Inflation rate % 2.9
Interest rate % 6
Electricity price $/kWh 0.27
Natural gas price $/GJ 19
Carbon tax $/tonne CO
2
-e 25.4
Lifetime of the system Year 20
Maintenance cost of the system % 1.25
Installation, integration, and piping cost of the
plant % of capital cost 150
CO
2
emission factor for electricity grid kg CO
2
/MWh 756
CO
2
emission factor for natural gas kg CO
2
/MWh 184
Primary energy factor for electricity grid
(PEF
E
)kWh
PE
/kWh
E
3.07
Primary energy factor for natural gas (PEF
NG
)kWh
PE
/kWh
NG
1.22
The SHC absorption chiller proposed in this study impose
additional expenses (mainly in terms of capital costs)
compared to the reference conventional system. This
additional cost, however, can be compensated over time due to
total cumulative savings from the reduction in the system non-
renewable energy consumption. The payback period (PBP) of
the additional costs compared to the reference conventional
system can be estimated as [29]:
SHC
LLL
CI
PBP CES CDERC OMRC
(13)
where
SHC
CI
is the additional cost associated with the solar
absorption plant, while CES
L
, CDERC
L
and OMRC
L
are the
levelized annual cost of energy saving, CO
2
emissions
reduction, and annual O&M cost reduction, respectively.
VI. E
NVIRONMENTAL
A
NALYSIS
Increasing environmental concerns necessitate considering
the environmental impacts of energy systems in the design
stage. As such, in this study, the amount of CO
2
emissions
released by the proposed SHC plant was considered as an
important environmental factor to identify the carbon emission
offset relative to the reference conventional system. The
annual carbon dioxide emissions from the SHC plant
(CDE
SHC
) can be estimated as:
SHC SHC,E SHC, NG
CDE CDE CDE
(14)
2
SHC,E E CO , E
CDE E EF
(15)
2
SHC, NG NG CO , NG
CDE E EF
(16)
where 2
CO ,E
EF
and 2
CO ,NG
EF
are the CO
2
emission factors for
grid electricity and natural gas, which are listed in Table III.
TABLE IV
C
APITAL
C
OST
F
UNCTION OF THE
M
AIN
S
YSTEM
C
OMPONENTS
(
IN
AUD)
[26]–[28]
System component Capital cost function
Solar collector
aSC
Z 520A
Storage tank
SST T
Z2,50V0
Auxiliary heater
AH AH,rate d
Z 102Q
Absorption chiller
ACH ACH, rated
QZ 855
Vapor compression chiller
WCH WCH,rated
Z 480Q
Cooling tower
CT CT,rated
Z 12,966 Q 103,603ln
Cooling/Heating coils
0.4162
CC/HCCC/HC
29,817 AZ
Pump
0.71
P
P
P
0.2
1,062 W 1 1
Z
Variable Frequency Drive
PVFD
Z91W
Controllers
CTRL
Z 6,750
VII. S
YSTEM
O
PTIMIZATION
Multi-objective optimization is a realistic approach to
handle real-world engineering problems dealing with
conflicting objectives which must be addressed
simultaneously. In a multi-objective optimization problem, a
set of non-dominated solutions, known as Pareto optimal
solutions, is obtained, which represents a hierarchy of best
possible trade-offs between the considered objective functions
[30]. Genetic algorithms have been proven to provide a robust
and efficient approach to achieve a set of reliable global
optimal solutions to a multi-objective optimization problem
[31]. As such, TRNSYS was coupled with MATLAB in the
present study, using its genetic algorithm optimization toolbox
to perform the multi-objective optimization of the system. The
primary energy consumption and the total annual cost of the
SHC plant were selected as two conflicting objective functions
to be minimized simultaneously. Five design parameters were
selected for the system optimization, which are presented in
Table V. It should be noted that the lower and upper bounds of
the collector specific area (A
SC,min
and A
SC,max
) were selected
in a way to limit the total solar fraction of the plant between
~25% and ~90%.
VIII.R
ESULTS AND
D
ISCUSSION
To demonstrate the modeling and optimization
methodology described above, a case study was developed for
the proposed SHC plant integrated into the modeled reference
hotel building in Sydney, Australia, a location with relative
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sunny climate with annual global horizontal and beam
irradiance of about 1608.5 kWh/m
2
and 853.9 kWh/m
2
,
respectively [32]. The annual cooling and heating load profile
of the building under Sydney’s climate is shown in Fig. 2.
TABLE V
D
ESIGN
P
ARAMETERS
C
ONSIDERED FOR
S
YSTEM
O
PTIMIZ ATION
Design parameter Unit Range of variation
Solar collector specific area (A
SC
)m
2
/kW
c
A
SC,min
< A
SC
<A
SC,max
Storage tank specific volume (V
ST
)L/m
2
10 < V
ST
<100
Solar pump nominal flow rate (
P1,nominal
m
)L/hr m
2
49 <
P1,nominal
m
<195
Collector set-point temperature in summer
(T
SP,s
)°C 190 < T
SP,s
< 220
Collector set-point temperature in winter
(T
SP,w
)°C 70 < T
SP,w
< 100
The maximum cooling and heating demands of the building
are 965 kW and 520 kW, respectively. Accordingly, a triple-
effect absorption chiller with a nominal cooling capacity of
1163 kW was selected from Thermax Ltd. The characteristic
coefficients obtained for this chiller are:
a = −2.14, e = 3.29, s
E
= 14.6, r
E
= −1171.7, s
G
= 7.05, and
r
G
= −504.2
Fig. 2 Monthly load demand of the reference hotel building modeled
in this study under Sydney’s climate
The Pareto front of optimal solutions obtained from multi-
objective optimization of the system is shown in Fig. 3. As
mentioned earlier, the conflicting relation between the two
objective functions is evident in this figure. The lowest
primary energy consumption (and thus the minimum
environmental impact) is achieved at design point A, while the
total levelized cost has its highest value at this point. The
highest primary energy consumption occurs at design point B,
where the system total cost stands at its minimum. If the
primary energy consumption was considered as the sole
objective function, point A then would represent the optimal
design point of the system. In other words, point A shows an
extreme design where the solar energy source is most
weighted to contribute to meeting the load requirements of the
buildings. Were the plant levelized cost to be the sole
objective in the optimization process, then point B would be
preferred as the optimum design. Fig. 3 also shows the
levelized cost and the primary energy consumption
corresponding to the reference conventional system operating
under the same conditions. As can be seen in this figure, all
optimal solutions corresponding to SHC plant outperform the
conventional system, proving the acceptable energy efficiency
of a gas-fired backup when used with a high-COP triple-effect
chiller-based design.
Fig. 3 Pareto front of optimal solutions obtained from multi-objective
optimization of the proposed SHC triple-effect absorption chiller
To comprehensively analyze the energetic/environmental
and economic performance of the proposed SHC plant, a final
optimal solution should be selected from the Pareto front
presented in Fig. 3. Since the dimension of the objective
functions considered in this study is not the same (i.e. PEC in
GWh/year and C
tot,L
in M$/year), all objectives must be non-
dimensionalized first. The Euclidian technique was therefore
used to non-dimensionalize the vectors of objectives in the
present work. More details about this method can be found in
[33].
Assuming equal weights assigned to both objective
functions, a LINMAP decision-making method was employed
to determine the final optimal design point of the SHC plants.
In this method, an ideal point is defined at which both
objectives are at their optimal values independent of the one
another [34]. The solution with a minimum spatial distance
from the ideal point is selected as the desired optimal solution,
which is marked in Fig. 3. It should be noted that in general no
decision-making method has superiority over the other; thus,
selection of the final optimal design point merely depends on
the importance of each objective function to the designer
under certain circumstances. The optimal values of design
parameters corresponding to the final optimal point as wells as
points A and B are presented in Table VI. As shown in this
table, the final optimal point selected by LINMAP decision-
maker has reached an equal trade-off between the energetic,
economic and environmental performance of the system.
The capital cost breakdown of the SHC plant at the final
optimal design point is shown in Fig. 4. According to this
figure, the solar collector array has the highest share of the
capital cost, accounting for about 43% of the total capital cost
of the plant. The absorption chiller is the second most
expensive component in each configuration, comprising 25%
of the total capital cost of the plant. The next highest capital
costs can be attributed to the storage tank, cooling coils,
cooling tower, and heating coils.
World Academy of Science, Engineering and Technology
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TABLE VI
O
PTIMAL
D
ESIGN
P
ARAMETERS OF THE
P
ROPOSED
SHC P
LANT AT
T
HREE
I
NDICATIVE
O
PTIMAL
D
ESIGNS
Design parameter Point A Point B Selected by LINMAP
A
SC
(m
2
/kW
c
) 5.2 0.8 2.9
V
ST
(L/m
2
) 63.3 30.3 49.8
P1,nominal
m
(L/hr m
2
)194.6 50.3 189.6
T
SP,s
(°C) 214.8 203.1 206.5
T
SP,w
(°C) 79.9 73.2 74.5
Fig. 4 Capital cost breakdown (in AUD) of the SHC plant at the final
optimal design point
The performance-related results of the proposed SHC
system at its extreme designs and final optimum design are
presented in Table VII. As shown in this table, the highest
value of solar fraction, and thus the lowest primary energy
consumption and CO
2
emissions are achieved at point A,
while the least energy-efficient and environmentally friendly
design is obtained at point B. The multi-objective optimization
approach, however, leads to a reasonable trade-off between
the energetic, economic and environmental performance of the
system.
TABLE VII
P
ERFORMANCE
-R
ELATED
A
NNUAL
R
ESULTS OF THE
SHC S
YSTEM AT
T
HREE
I
NDICATIVE
O
PTIMAL
D
ESIGNS
Parameter Unit Point A Point B Selected by LINMAP
A
SC
m
2
6021.7 947.8 3426.2
V
ST
m
3
381 28.7 170.6
SF % 0.88 0.24 0.72
COP
avg
- 1.62 1.62 1.62
PEC GWh/year 0.49 1.42 0.77
CDE tonne/ year 107.9 240.7 146.5
PES GWh/ year 0.99 0.11 0.69
CDER tonne/ year 205 72 166
C
tot,L
M$/ year 1.33 0.66 0.94
PBP year 105.8 47.6 63.8
The results presented in Table VII show that the proposed
SHC plant at its final optimal design point has a solar fraction
of 72%, resulting in an annual primary energy saving of 0.69
GWh and reducing the annual CO
2
emissions by ~166 tonnes
as compared to the reference conventional system.
Nevertheless, based on today’s market the economic
performance of the proposed SHC plant does not seem
satisfactory without government subsidies, which is often the
case for many of renewable energy systems. The low
competitive economics of the proposed SHC system relative
to conventional HVAC systems is mainly due to the high
capital costs of concentrating high-temperature solar
collectors. In addition, the lower solar gain of concentrating
PTCs due to capturing only beam radiation leads the plant to
have larger collector areas to compensate for the loss of
diffuse component, thereby increasing the capital cost of the
solar field.
If 50% of the capital cost of the plant was financed by
subsidies, the payback period of the SHC plant at the final
optimal design presented in Table VI would drop just above
the lifetime of the plant (~21 year), which is still unacceptable
from an economic perspective. Increasing the financial
incentive up to 75%, the proposed SHC plant can achieve a
satisfactory payback time of about 6 years. Since such high
rates of financial incentives may not always be available from
public funding, another approach to improve the economics of
the proposed SHC system with less reliance on public
subsidies is to size the solar field to achieve lower solar
fractions. In other words, the solar collector field can be
(under) sized for high base load utilization, while the backup
heating and cooling energy source covers the remaining more
variable part of the overall building thermal load. As a case in
point, for an optimal design point which has a solar fraction of
about 40%, only 32% subsidies on the capital costs (of the
innovative components) is required in order for the additional
capital costs (compared to the reference conventional system)
to be recovered in ~9 years. Thus, based on the current market
the economically viable design of the proposed SHC plant is
better understood as a gas-driven system with solar
enhancement rather than a solar-powered system with gas
cooling and heating back-up. Overall, with increasing
maturity, the SHC absorption technology will likely become
more competitive with conventional systems for air-
conditioning applications.
IX. C
ONCLUSIONS
A comprehensive energetic, economic, and environmental
analyses and multi-objective optimization of a novel solar-
powered triple-effect LiBr–H
2
O absorption chiller was carried
out to evaluate the techno-economic potential of such systems
for air-conditioning applications. PTCs were employed to
drive the triple-effect chillers, while a gas burner was used as
backup when solar heat was not sufficient. A complete
dynamic simulation model of the proposed SHC plant was
developed in TRNSYS program, which was coupled with
MATLAB to perform the multi-objective optimization of the
system, where the primary energy consumption and levelized
annual total cost of the plant were minimized simultaneously.
Overall, the multi-objective approach of this study satisfied
both primary energy and economic and performance
objectives better than single-objective optimizations alone.
The optimization results indicated that the final optimum
design of the proposed SHC system results in a solar fraction
of 72%, achieving an annual primary energy saving of 0.69
GWh and saving ~166 tonnes of CO
2
emission as compared to
World Academy of Science, Engineering and Technology
International Journal of Energy and Power Engineering
Vol:10, No:10, 2016
1309International Scholarly and Scientific Research & Innovation 10(10) 2016 ISNI:0000000091950263
Open Science Index, Energy and Power Engineering Vol:10, No:10, 2016 waset.org/Publication/10005471
a reference conventional system. The economic performance
of the system, however, was not appealing, mainly due to the
high capital cost of concentrating PTCs required to drive the
triple-effect chiller. It was found that ~75% public funding
was required in order for the proposed SHC system to achieve
a satisfactory payback period of ~6 years. Finally, our results
showed that if some small amount of government subsidies is
available, the solar field can be sized to achieve lower solar
fractions to improve the economic feasibility of such systems.
This means that the most economically viable design (today)
is one in which the SHC absorption chillers are used as an
enhancement to the conventional system, rather than
attempting to rely mostly on solar-derived heat. Overall, the
challenge for the solar industry is to lower the cost of high
temperature collectors, thereby paving the way for
implementation of highly-efficient triple-effect chiller systems
integrated with solar thermal energy.
A
CKNOWLEDGMENT
This research is funded by the CRC for Low Carbon Living
Ltd (Project ID (RP1002)) supported by the Cooperative
Research Centres program, an Australian Government
Initiative.
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World Academy of Science, Engineering and Technology
International Journal of Energy and Power Engineering
Vol:10, No:10, 2016
1310International Scholarly and Scientific Research & Innovation 10(10) 2016 ISNI:0000000091950263
Open Science Index, Energy and Power Engineering Vol:10, No:10, 2016 waset.org/Publication/10005471