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Solar-assisted heat pump – A sustainable system for low-temperature water heating applications
Energy Conversion and Management 77 (2014) 550–557
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Solar-assisted heat pump – A sustainable system for low-temperature water heating applications S.K. Chaturvedi a, V.D. Gagrani b, T.M. Abdel-Salam c,⇑ a
Department of Mechanical and Aerospace Engineering, Old Dominion University, Norfolk, VA, USA DNV KEMA Energy and Sustainability, Philadelphia, USA c Department of Engineering, East Carolina University, Greenville, NC, USA b
a r t i c l e
i n f o
Article history: Received 13 June 2013 Accepted 29 September 2013
Keywords: Solar-assisted Heat pump Water heating Life cycle cost Low temperature applications
a b s t r a c t Direct expansion solar assisted heat pump systems (DX-SAHP) have been widely used in many applications including water heating. In the DX-SAHP systems the solar collector and the heat pump evaporator are integrated into a single unit in order to transfer the solar energy to the refrigerant. The present work is aimed at studying the use of the DX-SAHP for low temperature water heating applications. The novel aspect of this paper involves a detailed long-term thermo-economic analysis of the energy conservation potential and economic viability of these systems. The thermal performance is simulated using a computer program that incorporates location dependent radiation, collector, economic, heat pump and load data. The economic analysis is performed using the life cycle cost (LCC) method. Results indicate that the DX-SAHP water heaters systems when compared to the conventional electrical water heaters are both economical as well as energy conserving. The analysis also reveals that the minimum value of the system life cycle cost is achieved at optimal values of the solar collector area as well as the compressor displacement capacity. Since the cost of SAHP system presents a barrier to mass scale commercialization, the results of the present study indicating that the SAHP life cycle cost can be minimized by optimizing the collector area would certainly be helpful in lowering, if not eliminating, the economic barrier to these systems. Also, at load temperatures higher than 70 °C, the performance of the single stage heat pump degrades to the extent that its cost and efficiency advantages over the electric only system are lost. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction In recent decades, global energy consumption has increased substantially. A major part of the global energy consumption comes from conventional energy sources such as fossil fuels. In 2004, the total worldwide energy consumption was 15 TW (=1.5 1013 W) with 86.5% from fossil fuels [1]. However, combustion of fossil fuels involves production of carbon dioxide and other gases that cause the greenhouse effect, leading to global warming. The atmospheric concentration of CO2 has increased by 31% above pre-industrial levels since 1750 [2]. Growing demand for petroleum-based fossil fuels has led to concerns that they may be depleted in the next two decades to a level that would cause a major disruption in the energy supply chain. As a result, finding alternative energy sources that are cleaner as well as economical has become a critical societal need which led to the development of renewable energy sources, such as solar and wind, in recent years. The United States is the world’s largest energy user, consuming 100 quadrillion BTU (29,000 TW h). About twenty-one percent of ⇑ Corresponding author. Tel.: +1 252 328 9649; fax: +1 252 737 1041. E-mail address: [email protected] (T.M. Abdel-Salam). 0196-8904/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.enconman.2013.09.050
this total energy is used in the residential sector [3]. Water heating alone accounts for about 20% of the energy consumed in a typical American household and about 2.6% of the total energy consumed in commercial building in the United States [4]. It is estimated that nearly 7% of the total US energy consumption is in the form of low temperature (<80 °C) water heating applications in which energy demand is met primarily through either natural gas or electrical resistance heaters. For instance, in the residential sector nearly 70% of water heating needs are met through natural gas heaters and about 20% through electrical resistance heaters [5]. However, the use of these conventional systems involves carbon emission, either directly or indirectly, into the earth’s atmosphere. For instance, a natural gas water heater typically contributes about 2 tons of CO2 annually to the atmosphere. Electric hot water systems are particularly deleterious to the environment since they are indirectly responsible for about three times the emission of CO2 for each kW h of electrical energy, compared to natural gas. So why are electric hot water heaters still in use? According to the US Energy Information Agency there are nearly 40 million households with electric hot water heaters [1]. In the United States the reason for their presence in the market place has to do with their very low capital cost compared to other alternatives. Even
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Nomenclature Ac solar collector surface area (m2) COPH coefficient of performance for heating COPeffective effective coefficient of performance DHW domestic hot water DX-SAHP direct expansion solar assisted heat pump fAux fraction of the auxiliary energy fsolar fraction of the solar energy F0 solar collector efficiency factor h specific enthalpy (J/kg) Icoll averaged instantaneous radiation in the collector plane (W/m2) K specific heat ratio KT monthly average clearance index LCC life cycle cost H average solar radiation on horizontal surface _ coll collector mass flow rate (kg/s) m P pressure (kPa) PER primary energy ratio PERelect primary electrical energy ratio PERSAHP primary solar assisted heat pump energy ratio PW present worth auxiliary thermal energy (W) Q_ Aux thermal energy delivered by the heat pump (W) Q_ H
though their operating cost is higher compared to alternatives such as natural gas water heaters, some consumers still prefer them due to their low initial cost. However, due to implicit environmental cost for remediation or sequestering of CO2 emission from fossil fuel burners it is worthwhile to look at alternatives such as renewable energy sources that operate with reduced or very little CO2 emission. Ideally a sustainability consideration of these renewable energy systems should entail two broad elements: (a) reduction or elimination of primary energy (coal, natural gas, etc.) consumption through substitution of renewable energy sources to achieve reduced CO2 emission and (b) economic competitiveness demonstrated through a life cycle cost analysis. These aspects will be discussed in more detail in later sections of this paper. 2. The proposed system and scope of study In the present study, a direct expansion solar-assisted SAHP system is used for water heating applications as an alternative to electric or natural gas water heater. The theory and operational aspects of this system have been described in the literature and the readers can find details in Refs. [6–42]. The present study is focused on the long-term thermo-economic analysis of the SAHP system to evaluate both the energy conservation potential as well as the economic viability of these systems. Aspects pertaining to the system life cycle costs (LCC) and effective resource utilization as characterized by the primary energy ratio are analyzed. The direct expansion SAHP water heating system (DX-SAHP) is a combination of a solar collector and a heat pump in which the collector also acts as the heat pump evaporator. Such integration is not only cost effective due to elimination of a heat exchanger required in convention SAHP systems that use water-based collectors. It is noteworthy that the use of a refrigerant as both the working fluid of the heat pump and as the collector results in quenching of the collector. This causes the collector to operate at a temperature low enough to boost the solar energy collection efficiency, yet it is high enough to enhance the heat pump performance compared to an air source heat pump [9]. This has implications for both reduced capital cost of the solar collector as
Q_ load Q_ solar VD v s SAHP T T1 Ta T max T min UL _ Aux W Wcomp _ comp W TMY U
gcoll gcomp gth gtrans sa
thermal load (W) solar energy absorbed by the refrigerant (W) volume displacement (m3/s) specific volume (m3/kg) specific entropy (J/kg k) solar assisted heat pump temperature (°C) collector temperature (°C) ambient temperature (°C) average maximum air temperature (°C) average minimum temperature (°C) overall heat loss coefficient of the solar collector (W/m2) auxiliary power (W) work compressor (J/kg) compressor power (W) typical meteorological year latitude of location (°) collector efficiency compressor efficiency thermal efficiency electrical transmission efficiency transmittance-absorptance product
well as lowered heat pump operating costs due to enhanced heat pump coefficient of performance (COPH) compared to the conventional air source heat pump. A refrigerant with low boiling point, R-134a, is employed as the working fluid in the cycle shown schematically in Fig. 1 and on a T–s diagram in Fig. 2. The system consists of four main components employed in a typical vapor compression refrigeration cycle namely; an evaporator, compressor, condenser, and an expansion valve. In the present system the heat pump evaporator also doubles up as the solar collector. The fin-tube evaporator is exposed directly to the sun. The incident solar energy, ends-up being absorbed by the liquid refrigerant after making it’s way through the fin section between the tubes that carry the refrigerant. So process 4–1 involves constant pressure evaporation of sub-cooled liquid into saturated vapor at section 1. In the process 2–3, the refrigerant is compressed from evaporator pressure to condensing pressure. This is followed by the constant pressure process 3–4 in which the refrigerant is desuperheated and condensed to the saturation liquid state. During this process, the latent heat of condensation is transferred to water circulating through the condenser. After the condenser, the refrigerant passes through a throttle valve that leads to flashing of the refrigerant, creating a mixture of liquid and vapor states. The thermal analysis of the DX-SAHP system can be performed using different configurations of solar collectors. The bare solar collectors have relatively lower initial cost, but their performance is adversely impacted by higher energy losses [9]. Singlecover solar collectors have relatively higher initial cost, but the associated heat losses are relatively small compared to that from bare collectors. This translates into higher solar collector efficiency and higher overall system efficiency. In this study we have used a single cover flat plate solar collector. 3. Long-term thermal analysis Thermal performance of the SAHP system is primarily indicated by parameters such as heat pump coefficient of performance (COPH), the system effective coefficient of performance of (COPeffective), and primary energy ratio (PER). All these parameters
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Fig. 1. Schematic of solar-assisted heat pump system.
4. The compressor model
T
2
The compressor work, Wcomp, or the compressor power _ comp ð¼ m _ ref: W comp Þ for a given pressure ratio P2/P1 is determined W from the expression,
W comp ¼
3
4
1
s Fig. 2. T–s diagram for the solar assisted heat pump system.
are evaluated for specified geographical location based on chosen condensing temperature (T3). The thermal performance of the proposed system is analyzed for the city of Norfolk VA, USA. In order to predict the long-term averaged thermal performance of the system, monthly averaged instantaneous values of radiation in the collector plane ðIcoll Þ are calculated for each hour from the long-term averaged daily solar data for the Norfolk location. From the radiation values ðIcoll Þ, the hourly performance of the system is calculated for an entire day in the middle of the chosen month. The hourly performance parameters are integrated to yield daily, monthly and yearly performance data. The system analysis makes use of the typical meteorological year (TMY) data which includes for many locations, the long-term climatic mean conditions such as the monthly average daily solar radiation on a horizontal surface ðHÞ, the monthly average clearness index ðK T Þ, and the monthly average maximum and minimum air temperature ðT max and T min Þ and the latitude of the location (a) [43]. The air temperature variation during the day is calculated by assuming a sinusoidal function to represent the air temperature variation during the collector operating hours. The solar radiation model for calculation radiation in the collector plane ðIcoll Þ is described in some details in Ref. [43]. The radiation module takes into account the available solar energy that can be captured by the collector based on meteorological data obtained for the specified geographical location. Output of radiation model is in the form of available solar energy (Icoll) at a specified slope angle of the collector. The system data used in the analysis are given in Table 1.
P1 v 1
gcomp
k k1
# " k1 P2 k 1 P1
ð1Þ
where k is the ratio of specific heats, and for R-134a it has a value of 1.106. The assumption of ideal gas behavior during the compression process turns out to be fairly reasonable since the compressor work calculated in this manner is somewhat overestimated compared to the compressor work calculated directly from the refrigerant property tables. The extent of overestimation is roughly 5–7% for the range of pressure ratios considered in this study. Since the coefficient of performance of the heat pump (COPH) is inversely related to the compressor work, this leads to conservative or underestimated values of COPH. The thermodynamic properties at state points 1 and 3 are calculated from curve fitting the thermodynamic table data, using polynomial functions. 5. The collector model The collector model is used to determine the collector temperature T1 for given values of ambient temperature Ta and Icoll , collector parameters (sa, F0 , Ac and UL), refrigerant properties h1 and h2 and the heat pump parameters (the displacement volume rate VD). The steady state energy balance on the collector, expressed by Eq. (2), states that the net energy absorbed by the refrigerant circulating through the collector equals incident solar radiation minus the heat loss from the collector.
VD
v1
ðh1 h3 Þ ¼ F 0 Ac ½Icoll ðsaÞ U L ðT 1 T a Þ
ð2Þ
From the above equation one can solve for T1 as
T1 ¼ Ta þ
Icoll ðsaÞ VD ðh1 h4 Þ UL v 1 UL F 0 Ac
ð3Þ
Since h1 and v1 are functions of collector temperature T1, the procedure for solving for T1 is iterative. A computer program, incorporating location dependent radiation data, collector data, economic data, heat pump and load data, was developed and implemented. A flow chart for the program is given in Fig. 3.
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Table 1 Standard system parameters for thermal analysis.
1 2 3 4 5 6 7 8 9 10 11
Parameter
Value
Condensing temperature Collector area Collector slope Collector efficiency factor Transmittance-absorptance product Collector heat loss coefficient Compressor efficiency Compressor displacement Number of persons in house Ground reflectance Average hot water consumption per person
60 °C (variable) 3 m2 (variable) 36.9° 0.85 0.9 6 W/°C m2 0.8 0.0003518 m3/s (variable) 4 0.2 0.07571 m3/day (20 Gal/ day)
For a given location, all the data described above are provided as inputs. First the value of the collector temperature at a given time of the day is assumed and values of enthalpies at state points 1 and 4 are calculated from the polynomial fit for the refrigerant properties. From the collector model (Eq. (2)), a new value of T1 is calculated and compared with the previously assumed value of T1. This iterative procedure is continued until a converged value of T1 is obtained. Using the compressor, collector and thermal modules, monthly averaged instantaneous values of compressor work, solar energy absorbed and auxiliary energy are calculated. The values of these parameters are integrated over a given month, and eventually over all months to yield the annual values. The thermal energy delivered by the heat pump Q_ H , and the solar energy absorbed by the refrigerant, Q_ solar are calculated from the following equations
_ coll ðh2 h3 Þ Q_ H ¼ m
ð4Þ
_ coll ðh1 h4 Þ Q_ solar ¼ m
ð5Þ
The coefficient of performance of the heat pump (COPH) and the collector efficiency are calculated from Eqs. (6)–(8) listed below.
COPH ¼
QH _ comp W
ð6Þ
Fig. 3. Flow chart of thermal analysis computational algorithm.
where
_ coll ¼ m
gcoll ¼
6. Economic analysis
VD
ð7Þ
v1 Q_ solar
ð8Þ
Icoll Ac
Since the system requires auxiliary energy due to intermittency of available solar energy, the effective coefficient of performance is determined as
COPeffectiv e ¼
Q_ load _ _ Aux Þ ðW comp þ W
ð9Þ
_ Aux is the auxiliary energy required to meet the load in the where W absence of solar radiation. From the integrated monthly energy fluxes, one can calculate auxiliary energy ðQ_ Aux Þ and solar energy ðQ_ solar Þ as a fraction of thermal load ðQ_ load Þ
fsolar ¼
Q_ solar Q_ load
and f Aux ¼
Q_ Aux Q_ load
In the present study, economic analysis is performed using the life cycle cost (LCC) method. The life cycle cost is the sum of all costs associated with the SAHP system over its’ entire lifetime, in present cost, that takes into account the time value of money. Several factors such as initial cost, operating cost, and maintenance cost need to be taken into account. The method for calculating LCC is well established and readers are referred to Ref. [44]. Taking into account the utility inflation rate, and the monetary inflation rate and the present worth (PW) factors, the LCC can be expressed as
LCC ¼ PWSystem þ PWInstillation þ PWmaintenance þ PWEnergy usage PWtax
ð11Þ
Using LCC one can compare the economic sustainability of the proposed SAHP system with other alternatives such as the solar electric water heater. The system cost and other related economic parameters are shown in Table 3.
ð10Þ
Table 2 provides the monthly thermal performance data of the system for all months of the year for the City of Norfolk, Virginia.
7. Primary energy resource utilization For analysis of primary energy utilization in competing systems for domestic hot water (DHW) applications, one can use the
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Table 2 Monthly thermal performance of SAHP system for the City of Norfolk, Virginia. Month
QLoad (kJ)
QH (kJ)
Wc (kJ)
QAux (kJ)
COPH
COPeffective
PERoverall
January February March April May June July August September October November December
1834249.00 1656741.10 1746604.40 1563606.90 1462840.20 1225478.40 1201085.20 1179075.90 1119741.80 1244711.10 1542687.60 1702978.90
980841.44 997399.10 1444560.50 1563606.90 1462840.20 1225478.40 1201085.20 1179075.90 1119741.80 1244711.10 1123494.90 988435.75
224066.16 207337.25 284766.20 278685.12 289305.78 280105.75 290158.25 289580.75 279336.97 233339.90 223628.38 224564.36
853407.56 659342.00 302043.88 0.00 0.00 0.00 0.00 0.00 0.00 0.00 419192.75 714543.10
4.38 4.81 5.07 5.61 5.06 4.38 4.14 4.07 4.01 5.33 5.02 4.40
1.70 1.91 2.98 5.61 5.06 4.38 4.14 4.07 4.01 5.33 2.40 1.81
1.05 1.14 1.41 1.85 1.67 1.44 1.37 1.34 1.32 1.76 1.28 1.08
primary energy ratio (PER), a parameter that characterizes the overall efficiency of a number of interconnected energy transformation processes. For instance, in the DHW applications, in order to determine the PER one must trace the energy flow all the way back to the primary energy source (coal, natural gas, oil, etc.) to determine how many kilo Joules of thermal energy are delivered for end-use for every kJ of primary energy utilized. The ratio of the delivered thermal energy and the primary energy is known as the primary energy ratio (PER). Fig. 4 shows the energy flow streams for two competing water heaters namely, an electric water heater, and a solar-assisted heat pump with electric resistor as a back-up system. In the case of an electric water heater, the efficiency of the thermal power plant (nth) and the electrical transmission efficiency (ntrans) must be considered to account for losses in those systems. As a result, the PER of the overall system, a combination of the power plant, transmission and electric water heater, is nthntrans if the heat losses from the water tank are neglected. For a coal fired power plant, with typical values of nth = 0.35 and ntrans = 0.90 yields a PER value of 0.315. This represents an effective yield of 31.5% with the remainder 68.5% of the primary energy being lost as waste heat in the power plant and the utility transmission lines. It should also be noted that even though electric water heater are marketed as environmentally clean devices, in reality there are emissions of global warming gases such as CO2 and other pollutants at the power plant that must be attributed to electric heater for each kW h of electrical energy used in it. For the solar- assisted heat pump system with and electric resistance back-up system, the chart showing energy flow from the primary energy source to the end-use (delivered) thermal energy is illustrated in Fig. 4. Using the first law of thermodynamics for energy balances across each sub-system, one can show that the PERSAHP, the primary energy ratio of the SAHP water heater system can be expressed as follows.
Table 3 System input parameters for economic analysis.
1 2 3 4 5 6 7 8 9 10 11 12
Parameter
Value
Life span-SAHP system Initial system cost -SAHP system Installation cost -SAHP system Maintenance cost per year-SAHP system Federal tax incentive-SAHP system (30% system cost) Life span-electric heater Initial system cost – electric heater Installation cost – electric heater Maintenance cost per year-electric heater Electricity cost per kW h Utility inflation rate Monetary inflation rate
20 years $3750 $2000 $40 $1725 20 years $500 $200 $20 $0.1 10% 7%
PERSAHP 1 i ¼h ðntrans nth Þ fAux þ ð1fAux Þ
ð12Þ
COPH
In terms of fsolar (the fraction of load met by solar energy), the above equation can be rewritten as,
PERSAHP 1 ¼ ðntrans nth Þ ð1 fsolar Þ
ð13Þ
where
fsolar ¼ ð1 fAux Þ 1 The term
1 COPH
1 1 COP H
ð14Þ
can be regarded as the limiting value for
fsolar since any value of fsolar above it will result in a negative value for fAux. This limiting value of fsolar increases with COPH. For instance, for COPH = 5.0, the limiting value of fsolar is 0.8 and for COPH = 3.0, it is 0.67. Since the PER of the electric hot water equals ntransnth, PERSAHP can be expressed as
PERSAHP 1 ¼ 1 fsolar PERelec
ð15Þ
The primary energy ratio (PER) is defined as the ratio of the thermal energy delivered to the primary energy consumed. The variation of PERSAHP/PERelec as a function of fsolar is shown in Fig. 5. It is noted that as fsolar approaches 0.0, fAux equals 1.0 and the SAHP system is 100% reliant on the auxiliary electric heat, and the PERSAHP is accordingly equal to PERelec. At the other extreme, as fsolar approaches 1, the SAHP system operates in the solar only system mode, and the ratio approaches infinity asymptotically. The latter limit is not technically feasible as it would require COPH to approach infinity. This is also not economically feasible as it would require a very large solar collector field to achieve 100% solar only limit. In practical systems, fsolar is generally limited to a value in the range of 0.50–0.70. Since PERSAHP/PERelec is always greater than 1.0, from the perspective of efficient primary resource utilization, the SAHP water heater is always a better alternative than the electric water heater. However, from an economic viewpoint, this may not always be the case because of the tradeoffs between higher capital cost (for the SAHP system) and higher operating cost (for the electric water heater). 8. Discussion of results-annualized system performance In this section, the effects of changes in the collector area, compressor displacement volume (VD) and the load temperature on
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Fig. 4. Energy flow diagram for the SAHP system.
40 35
PER solar /PER elec.
30 25 20 15 10 5 0
0
0.2
0.4
0.6
0.8
1
fSolar Fig. 5. Variation of PERSAHP/PERelec as a function of fsolar.
the long-term system performance are discussed. Fig. 6 shows the variation of the system life cycle cost (LCC) with the collector area for a specified load temperature and compressor displacement volume. The presented load temperature is 60 °C and the volume
displacement of the compressor is 0.0003518 m3/s. From Fig. 6, one observes that a minimum value of LCC occurs at a collector area of 3 m2. Further increase in collector area beyond the minimum point yields a higher LCC. This is the optimal configuration from the long-term economic perspective due to two counteracting effects, namely the capital cost and the operating cost of the system. The capital cost of the system for a fixed load temperature and compressor displacement volume increases with increasing collector area while the system operational cost decreases. Increased solar energy collection raises the collector/evaporator temperature, leading to a lower temperature lift in the heat pumping process. This translates into reduced compressor power input, and consequently lower operational cost. The variation of the LCC with the compressor displacement volume is shown in Fig. 7. For a fixed 3 m2 solar collector area, it can be seen that a minimum value of LCC is obtained at VD = 0.0003518 m3/s. The reason for this occurrence has to do with two opposing trends _ comp Þ, and the auxinvolving the variation of compressor power ðW _ Aux Þ, as the compressor volume is changed. iliary electrical powerðW Since the capital cost is fixed for a given collector area, the varia_ Aux and W _ comp . As higher tion in LCC is solely due to changes in W values of compressor displacement are chosen the refrigerant mass flow rate through the system also increases. This results in greater _ comp . Since W _ Aux is the differheat pump capacity ðQ_ H Þ and higher W ence between the fixed load the heat pump capacity ðQ_ H Þ, the net _ Aux . These two effects transpire to yield an result is a decrease in W optimal value of VD as shown in Fig. 7.
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14000
14000
12000
12000
10000
10000
LCC ($)
LCC ($)
556
8000 6000
8000 6000
4000
4000
2000
2000
0
0
2
0
6
4
o
Tc=50 o Tc=55 o Tc=60 o Tc=65 o Tc=70 2
3
4
5
6
Collector area (m2 )
2
Collector area (m ) Fig. 6. Variation of life cycle cost with collector area (VD = 35.18 105 m3/s, load temperature = 60 °C).
Fig. 8. Variation of life cycle cost (LCC) with collector area and load temperature.
3
14000
Ac=3 Ac=4 Ac=5 Electical heater
12000 10000
2
Annual PER
LCC ($)
C C C C C
8000 6000 4000
1
2000 0
0
0.0002
0.0004
0.0006
0.0008
0.001
VD (m 3/s)
0 40
50
60
70
80
Load temerature (oC)
Fig. 7. Variation of life cycle cost (LCC) with compressor displacement (VD) (Ac = 3.0 m2/s, load temperature = 60 °C).
Fig. 9. Variation of overall annual per with collector area and load temperature.
The variation of LCC as a function of the collector area and the load temperature is shown in Fig. 8. For comparison the LCC of the electric water heater is also displayed. One notes that as the load temperature increases the LCC increases for a given collector area (fixed capital cost) because the operating costs are higher due to the higher temperature lift during the heat pumping process. Also, for a given load temperature, the LCC value increases with area since the capital cost is greater than the reduction in the operating cost due to enhanced heat pump efficiency. Towards the higher end of the load temperature range, as seen in Fig. 8, the heat pump LCC becomes comparable to the LCC of the electric hot water system. Unlike the electric hot water heater whose thermal and economic performance is independent of the load temperature, the performance of the SAHP system depends strongly on the load temperature. Fig. 9 shows the variation of the annual PER of the SAHP system as a function of the collector area and the load temperature. For comparison, the PERelec of the electric hot water heater is also included. For all collector areas and load temperatures the PERSAHP values are higher than unity and also well above the PERelec value of 0.33. This indicates that the SAHP system utilizes primary energy far more efficiently compared to the electric hot water heater over the load temperature range considered in this study. The PERSAHP value decreases with the increasing load temperature
and this implies that the SAHP system is utilizing a primary energy less efficiently. For a fixed load temperature gain in PERSAHP are achieved. This is due to elevated collector temperature that results _ comp in a smaller temperature lift. The resultant reduced value of W implies a decrement in the primary energy resource demand for a given thermal load requirement. Fig. 9 also suggests that the SAHP system is best suited from the efficient energy resource utilization perspective, for water heating or preheating applications in the 50– 70 °C load temperature range. In this temperature range both the LCC and PER values are decidedly superior to the electric only system. 9. Conclusions The present study employed modeling and long-term averaged monthly solar data to simulate the long-term averaged thermal performance of a direct-expansion solar assisted heat pump. Results were obtained for a range of load temperature and for a number of collector areas and compressor displacement volumes. The thermo-economic analysis results of the SAHP system indicates that the system is an appropriate match for the temperature water heating applications. Since a large amount of primary energy is consumed in the USA in meeting the demands for low temperature
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(50–70 °C) domestic water heating applications, the SAHP water heaters are the ideal systems for such applications. These systems, as the present study reveals are both economical as well as energy conserving solutions compared to the electric-only hot water heaters. Reduced consumption of electricity and the capture of renewable solar energy through collectors should have a favorable environmental impact since reduced carbon emission would result if these systems are used as an alternative to the electric-only systems. The simulation of the system also indicates that its long-term economic performance (LCC) can be optimized with respect to the collector area and the compressor displacement volume for a specified load temperature. At load temperatures higher than 70 °C, the performance of the single stage heat pump considered in this study degrades to the extent that its cost and efficiency advantages over the electric only system are lost. In the higher load temperature range (>70 °C) a two stage heat pump system may be more appropriate [45] and its long-term thermo economic cost-efficiency benefit analysis should be considered in a future study.
References [1] World consumption of primary energy by energy type and selected country groups. Energy Information Adminstration, U.S. Department of Energy; 2006. [2] http://en.wikipedia.org/wiki/Global_warming. [3] Annual energy outlook. U.S. Department of Energy; 2006. [4] www.eere.energy.gov/consumer/your_home/water_heating/index.cfm/ mytopic=12760. [5] www.solarhome.org/newssolarwaterheaters.html. [6] Kush BA. Performance of heat pump at elevated evaporating temperatures with application to solar input. Sol Energy 1980;102:203–10. [7] Franklin J, Saaski F, Yamagiwa A. A high efficiency direct expansion solar panel. In: Flat plate solar collector conference, Orlando, Florida, USA; 1977. p. 187–95. [8] Sporn P, Ambrosc E. The heat pump and solar energy. In: The world symposium on applied solar energy, Phoenix, Arizona, USA; 1955. p. 1–5. [9] Chaturvedi SK, Chiang YF, Roberts AS. Analysis of two phase flow collectors with applications to heat pumps. Sol Energy Eng 1982;104:358–64. [10] Abou-Ziyan HZ, Ahmed MF, Metwally MN, Abd El-Hameed HM. Solar-assisted R22 and R134a heat pump systems for low-temperature applications. Appl Therm Eng 1997;17:455–69. [11] Chaturvedi SK, Chen DT, Kheireddine A. Thermal performance of a variable capacity direct expansion solar-assisted heat pump. Energy Convers Manage 1998;39:181–91. [12] Chaturvedi SK, Abdel-Salam TM, Sreedharan S. Two-stage direct expansion solar-assisted heat pump for high temperature applications. Appl Therm Eng 2009;29:2093–9. [13] Aziz W, Chaturvedi SK, Kheireddine A. Thermodynamic analysis of twocomponent, two-phase flow in solar collectors with application to a directexpansion solar-assisted heat pump. Energy 1999;24:247–59. [14] Li YW, Wang RZ, Wu JY, Xu YX. Experimental performance analysis on a directexpansion solar-assisted heat pump water heater. Appl Therm Eng 2007;27:2858–68. [15] Chaturvedi SK, Abazeri M. Transient simulation of a capacity modulated, direct-expansion. Sol Energy 1987;39:421–8. [16] Chaturvedi SK, Roberts AS, Mei V. Solar collector as heat pump evaporator. In: 13th Intersociety energy conversion conference, Boston, USA; 1979. p. 4–9. [17] Kuang YH, Sumathy K, Wang RZ. Study on a direct-expansion solar-assisted heat pump water heating system. Int J Energy Res 2003;27:531–48.
557
[18] Huang BJ, Chyng JP. Performance characteristic of integral type solar-assisted heat pump. Sol Energy 2001;71:403–14. [19] Chaturvedi SK, Mei V. Thermal performance of SAHP system with combined collector/evaporator. In: AIAA terrestrial energy systems conference, Orlando, Florida, USA; 1979. [20] Kuang YH, Wang RZ, Yu LQ. Experimental study on solar assisted heat pump system for heat supply. Energy Convers Manage 2003;44:1089–98. [21] Kuang YH, Wang RZ. Performance of a multi-functional direct expansion solar assisted heat pump system. Sol Energy 2006;80:795–803. [22] Huang BJ, Chyng JP. Integral-type solar-assisted heat pump water heater. Renew Energy 1999;16:731–4. [23] Huang BJ, Lee JP, Chyng JP. Heat-pipe enhanced solar-assisted heat pump water heater. Sol Energy 2005;78:375–81. [24] Yumrutas R, Unsal M. Analysis of solar aided heat pump system with seasonal thermal energy storage in surface tanks. Energy 2000;25:1231–43. [25] Krakow KL, Lin SA. Solar source heat pump with refrigerant cooled solar collectors for cold climates. Trans ASHRAE 1982;88:417–39. [26] O’Dell MP, Mitchell JW, Beckman WA. Design method and performance of heat pump with refrigerant-filled solar collectors. Sol Energy Eng 1984;106:159–64. [27] Matsuki K, Shinobu Y, Yoshikawa M. A prototype direct-expansion solar heat pump system. Trans ASHRAE 1987;93. [28] Ito S, Miura N, Wang K. Performance of a heat pump using direct expansion solar collectors. Sol Energy 1999;65:189–96. [29] O’Dell MP, Mitchell JW, Beckman WA. Beckman, solar heat pump systems with refrigerant-filled collectors. Trans ASHRAE 1983;89:519–25. [30] Ito S, Miura N, Wang K. Performance of a heat pump using direct expansion solar collectors. Sol Energy 1998;65:189–96. [31] Hawlader MN, Chou SK, Ullah MZ. The performance of a solar assisted heat pump water heating system. Appl Therm Eng 2001;21:1049–65. [32] Ito S, Miura N, Takano Y. Studies of heat pumps using direct expansion type solar collectors. Sol Energy Eng 2005;127:60–4. [33] Chyng JP, Lee CP, Huang BJ. Performance analysis of a solar assisted heat pump water heater. Sol Energy 2003;74:33–44. [34] Gorozabel Chata FB, Chaturvedi SK, Almogbel A. Analysis of a direct expansion solar assisted heat pump using different refrigerants. Energy Convers Manage 2005;46:2614–24. [35] Fernández-Seara J, Piñeiro C, Dopazo JA, Fernandes F, Sousa PXB. Experimental analysis of a direct expansion solar assisted heat pump with integral storage tank for domestic water heating under zero solar radiation conditions. Energy Convers Manage 2012;59:1–8. [36] Kara O, Ulgen K, Hepbasli A. Exergetic assessment of direct-expansion solarassisted heat pump systems: review and modeling. Renew Sustain Energy Rev 2008;12:1383–401. [37] Kong XQ, Zhang D, Li Y, Yang QM. Thermal performance analysis of a directexpansion solar-assisted heat pump water heater. Energy 2011;36:6830–8. [38] Kuang YH, Wang RZ. Performance of a multi-functional direct-expansion solar assisted heat pump system. Sol Energy 2006;80:795–803. [39] Ozgener O, Hepbasli A. A review on the energy and exergy analysis of solar assisted heat pump systems. Renew Sustain Energy Rev 2007;11:482–96. [40] Tagliafico L, Scarpa F, Tagliafico G, Valsuani F. An approach to energy saving assessment of solar assisted heat pumps for swimming pool water heating. Energy Build 2012;55:833–40. [41] Chaturvedi SK, Mohieldin TO, Chen DT. Second-law analysis of solar-assisted heat pumps. Energy 1991;16:941–9. [42] Cervantes GJ, Torres-Reyes E. Experiments on a solar-assisted heat pump and an exergy analysis of the system. Appl Therm Eng 2002;22:1289–97. [43] Duffie JA, Beckman WA. Solar engineering of thermal processes. John Wiley & Sons Inc.; 1991. [44] Dhillon BS. Life cycle costing for engineers. Boca Raton, Florida, USA: CRC Press, Taylor & Francis Group; 2010. [45] Gagrani VD. Simulation and thermo-economic analysis of a solar-assisted heat pump system for domestic water heating application. Mechanical engineering. Norfolk, Virginia, USA: Old Dominion University; 2007.
Contents lists available at ScienceDirect
Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Solar-assisted heat pump – A sustainable system for low-temperature water heating applications S.K. Chaturvedi a, V.D. Gagrani b, T.M. Abdel-Salam c,⇑ a
Department of Mechanical and Aerospace Engineering, Old Dominion University, Norfolk, VA, USA DNV KEMA Energy and Sustainability, Philadelphia, USA c Department of Engineering, East Carolina University, Greenville, NC, USA b
a r t i c l e
i n f o
Article history: Received 13 June 2013 Accepted 29 September 2013
Keywords: Solar-assisted Heat pump Water heating Life cycle cost Low temperature applications
a b s t r a c t Direct expansion solar assisted heat pump systems (DX-SAHP) have been widely used in many applications including water heating. In the DX-SAHP systems the solar collector and the heat pump evaporator are integrated into a single unit in order to transfer the solar energy to the refrigerant. The present work is aimed at studying the use of the DX-SAHP for low temperature water heating applications. The novel aspect of this paper involves a detailed long-term thermo-economic analysis of the energy conservation potential and economic viability of these systems. The thermal performance is simulated using a computer program that incorporates location dependent radiation, collector, economic, heat pump and load data. The economic analysis is performed using the life cycle cost (LCC) method. Results indicate that the DX-SAHP water heaters systems when compared to the conventional electrical water heaters are both economical as well as energy conserving. The analysis also reveals that the minimum value of the system life cycle cost is achieved at optimal values of the solar collector area as well as the compressor displacement capacity. Since the cost of SAHP system presents a barrier to mass scale commercialization, the results of the present study indicating that the SAHP life cycle cost can be minimized by optimizing the collector area would certainly be helpful in lowering, if not eliminating, the economic barrier to these systems. Also, at load temperatures higher than 70 °C, the performance of the single stage heat pump degrades to the extent that its cost and efficiency advantages over the electric only system are lost. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction In recent decades, global energy consumption has increased substantially. A major part of the global energy consumption comes from conventional energy sources such as fossil fuels. In 2004, the total worldwide energy consumption was 15 TW (=1.5 1013 W) with 86.5% from fossil fuels [1]. However, combustion of fossil fuels involves production of carbon dioxide and other gases that cause the greenhouse effect, leading to global warming. The atmospheric concentration of CO2 has increased by 31% above pre-industrial levels since 1750 [2]. Growing demand for petroleum-based fossil fuels has led to concerns that they may be depleted in the next two decades to a level that would cause a major disruption in the energy supply chain. As a result, finding alternative energy sources that are cleaner as well as economical has become a critical societal need which led to the development of renewable energy sources, such as solar and wind, in recent years. The United States is the world’s largest energy user, consuming 100 quadrillion BTU (29,000 TW h). About twenty-one percent of ⇑ Corresponding author. Tel.: +1 252 328 9649; fax: +1 252 737 1041. E-mail address: [email protected] (T.M. Abdel-Salam). 0196-8904/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.enconman.2013.09.050
this total energy is used in the residential sector [3]. Water heating alone accounts for about 20% of the energy consumed in a typical American household and about 2.6% of the total energy consumed in commercial building in the United States [4]. It is estimated that nearly 7% of the total US energy consumption is in the form of low temperature (<80 °C) water heating applications in which energy demand is met primarily through either natural gas or electrical resistance heaters. For instance, in the residential sector nearly 70% of water heating needs are met through natural gas heaters and about 20% through electrical resistance heaters [5]. However, the use of these conventional systems involves carbon emission, either directly or indirectly, into the earth’s atmosphere. For instance, a natural gas water heater typically contributes about 2 tons of CO2 annually to the atmosphere. Electric hot water systems are particularly deleterious to the environment since they are indirectly responsible for about three times the emission of CO2 for each kW h of electrical energy, compared to natural gas. So why are electric hot water heaters still in use? According to the US Energy Information Agency there are nearly 40 million households with electric hot water heaters [1]. In the United States the reason for their presence in the market place has to do with their very low capital cost compared to other alternatives. Even
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Nomenclature Ac solar collector surface area (m2) COPH coefficient of performance for heating COPeffective effective coefficient of performance DHW domestic hot water DX-SAHP direct expansion solar assisted heat pump fAux fraction of the auxiliary energy fsolar fraction of the solar energy F0 solar collector efficiency factor h specific enthalpy (J/kg) Icoll averaged instantaneous radiation in the collector plane (W/m2) K specific heat ratio KT monthly average clearance index LCC life cycle cost H average solar radiation on horizontal surface _ coll collector mass flow rate (kg/s) m P pressure (kPa) PER primary energy ratio PERelect primary electrical energy ratio PERSAHP primary solar assisted heat pump energy ratio PW present worth auxiliary thermal energy (W) Q_ Aux thermal energy delivered by the heat pump (W) Q_ H
though their operating cost is higher compared to alternatives such as natural gas water heaters, some consumers still prefer them due to their low initial cost. However, due to implicit environmental cost for remediation or sequestering of CO2 emission from fossil fuel burners it is worthwhile to look at alternatives such as renewable energy sources that operate with reduced or very little CO2 emission. Ideally a sustainability consideration of these renewable energy systems should entail two broad elements: (a) reduction or elimination of primary energy (coal, natural gas, etc.) consumption through substitution of renewable energy sources to achieve reduced CO2 emission and (b) economic competitiveness demonstrated through a life cycle cost analysis. These aspects will be discussed in more detail in later sections of this paper. 2. The proposed system and scope of study In the present study, a direct expansion solar-assisted SAHP system is used for water heating applications as an alternative to electric or natural gas water heater. The theory and operational aspects of this system have been described in the literature and the readers can find details in Refs. [6–42]. The present study is focused on the long-term thermo-economic analysis of the SAHP system to evaluate both the energy conservation potential as well as the economic viability of these systems. Aspects pertaining to the system life cycle costs (LCC) and effective resource utilization as characterized by the primary energy ratio are analyzed. The direct expansion SAHP water heating system (DX-SAHP) is a combination of a solar collector and a heat pump in which the collector also acts as the heat pump evaporator. Such integration is not only cost effective due to elimination of a heat exchanger required in convention SAHP systems that use water-based collectors. It is noteworthy that the use of a refrigerant as both the working fluid of the heat pump and as the collector results in quenching of the collector. This causes the collector to operate at a temperature low enough to boost the solar energy collection efficiency, yet it is high enough to enhance the heat pump performance compared to an air source heat pump [9]. This has implications for both reduced capital cost of the solar collector as
Q_ load Q_ solar VD v s SAHP T T1 Ta T max T min UL _ Aux W Wcomp _ comp W TMY U
gcoll gcomp gth gtrans sa
thermal load (W) solar energy absorbed by the refrigerant (W) volume displacement (m3/s) specific volume (m3/kg) specific entropy (J/kg k) solar assisted heat pump temperature (°C) collector temperature (°C) ambient temperature (°C) average maximum air temperature (°C) average minimum temperature (°C) overall heat loss coefficient of the solar collector (W/m2) auxiliary power (W) work compressor (J/kg) compressor power (W) typical meteorological year latitude of location (°) collector efficiency compressor efficiency thermal efficiency electrical transmission efficiency transmittance-absorptance product
well as lowered heat pump operating costs due to enhanced heat pump coefficient of performance (COPH) compared to the conventional air source heat pump. A refrigerant with low boiling point, R-134a, is employed as the working fluid in the cycle shown schematically in Fig. 1 and on a T–s diagram in Fig. 2. The system consists of four main components employed in a typical vapor compression refrigeration cycle namely; an evaporator, compressor, condenser, and an expansion valve. In the present system the heat pump evaporator also doubles up as the solar collector. The fin-tube evaporator is exposed directly to the sun. The incident solar energy, ends-up being absorbed by the liquid refrigerant after making it’s way through the fin section between the tubes that carry the refrigerant. So process 4–1 involves constant pressure evaporation of sub-cooled liquid into saturated vapor at section 1. In the process 2–3, the refrigerant is compressed from evaporator pressure to condensing pressure. This is followed by the constant pressure process 3–4 in which the refrigerant is desuperheated and condensed to the saturation liquid state. During this process, the latent heat of condensation is transferred to water circulating through the condenser. After the condenser, the refrigerant passes through a throttle valve that leads to flashing of the refrigerant, creating a mixture of liquid and vapor states. The thermal analysis of the DX-SAHP system can be performed using different configurations of solar collectors. The bare solar collectors have relatively lower initial cost, but their performance is adversely impacted by higher energy losses [9]. Singlecover solar collectors have relatively higher initial cost, but the associated heat losses are relatively small compared to that from bare collectors. This translates into higher solar collector efficiency and higher overall system efficiency. In this study we have used a single cover flat plate solar collector. 3. Long-term thermal analysis Thermal performance of the SAHP system is primarily indicated by parameters such as heat pump coefficient of performance (COPH), the system effective coefficient of performance of (COPeffective), and primary energy ratio (PER). All these parameters
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Fig. 1. Schematic of solar-assisted heat pump system.
4. The compressor model
T
2
The compressor work, Wcomp, or the compressor power _ comp ð¼ m _ ref: W comp Þ for a given pressure ratio P2/P1 is determined W from the expression,
W comp ¼
3
4
1
s Fig. 2. T–s diagram for the solar assisted heat pump system.
are evaluated for specified geographical location based on chosen condensing temperature (T3). The thermal performance of the proposed system is analyzed for the city of Norfolk VA, USA. In order to predict the long-term averaged thermal performance of the system, monthly averaged instantaneous values of radiation in the collector plane ðIcoll Þ are calculated for each hour from the long-term averaged daily solar data for the Norfolk location. From the radiation values ðIcoll Þ, the hourly performance of the system is calculated for an entire day in the middle of the chosen month. The hourly performance parameters are integrated to yield daily, monthly and yearly performance data. The system analysis makes use of the typical meteorological year (TMY) data which includes for many locations, the long-term climatic mean conditions such as the monthly average daily solar radiation on a horizontal surface ðHÞ, the monthly average clearness index ðK T Þ, and the monthly average maximum and minimum air temperature ðT max and T min Þ and the latitude of the location (a) [43]. The air temperature variation during the day is calculated by assuming a sinusoidal function to represent the air temperature variation during the collector operating hours. The solar radiation model for calculation radiation in the collector plane ðIcoll Þ is described in some details in Ref. [43]. The radiation module takes into account the available solar energy that can be captured by the collector based on meteorological data obtained for the specified geographical location. Output of radiation model is in the form of available solar energy (Icoll) at a specified slope angle of the collector. The system data used in the analysis are given in Table 1.
P1 v 1
gcomp
k k1
# " k1 P2 k 1 P1
ð1Þ
where k is the ratio of specific heats, and for R-134a it has a value of 1.106. The assumption of ideal gas behavior during the compression process turns out to be fairly reasonable since the compressor work calculated in this manner is somewhat overestimated compared to the compressor work calculated directly from the refrigerant property tables. The extent of overestimation is roughly 5–7% for the range of pressure ratios considered in this study. Since the coefficient of performance of the heat pump (COPH) is inversely related to the compressor work, this leads to conservative or underestimated values of COPH. The thermodynamic properties at state points 1 and 3 are calculated from curve fitting the thermodynamic table data, using polynomial functions. 5. The collector model The collector model is used to determine the collector temperature T1 for given values of ambient temperature Ta and Icoll , collector parameters (sa, F0 , Ac and UL), refrigerant properties h1 and h2 and the heat pump parameters (the displacement volume rate VD). The steady state energy balance on the collector, expressed by Eq. (2), states that the net energy absorbed by the refrigerant circulating through the collector equals incident solar radiation minus the heat loss from the collector.
VD
v1
ðh1 h3 Þ ¼ F 0 Ac ½Icoll ðsaÞ U L ðT 1 T a Þ
ð2Þ
From the above equation one can solve for T1 as
T1 ¼ Ta þ
Icoll ðsaÞ VD ðh1 h4 Þ UL v 1 UL F 0 Ac
ð3Þ
Since h1 and v1 are functions of collector temperature T1, the procedure for solving for T1 is iterative. A computer program, incorporating location dependent radiation data, collector data, economic data, heat pump and load data, was developed and implemented. A flow chart for the program is given in Fig. 3.
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Table 1 Standard system parameters for thermal analysis.
1 2 3 4 5 6 7 8 9 10 11
Parameter
Value
Condensing temperature Collector area Collector slope Collector efficiency factor Transmittance-absorptance product Collector heat loss coefficient Compressor efficiency Compressor displacement Number of persons in house Ground reflectance Average hot water consumption per person
60 °C (variable) 3 m2 (variable) 36.9° 0.85 0.9 6 W/°C m2 0.8 0.0003518 m3/s (variable) 4 0.2 0.07571 m3/day (20 Gal/ day)
For a given location, all the data described above are provided as inputs. First the value of the collector temperature at a given time of the day is assumed and values of enthalpies at state points 1 and 4 are calculated from the polynomial fit for the refrigerant properties. From the collector model (Eq. (2)), a new value of T1 is calculated and compared with the previously assumed value of T1. This iterative procedure is continued until a converged value of T1 is obtained. Using the compressor, collector and thermal modules, monthly averaged instantaneous values of compressor work, solar energy absorbed and auxiliary energy are calculated. The values of these parameters are integrated over a given month, and eventually over all months to yield the annual values. The thermal energy delivered by the heat pump Q_ H , and the solar energy absorbed by the refrigerant, Q_ solar are calculated from the following equations
_ coll ðh2 h3 Þ Q_ H ¼ m
ð4Þ
_ coll ðh1 h4 Þ Q_ solar ¼ m
ð5Þ
The coefficient of performance of the heat pump (COPH) and the collector efficiency are calculated from Eqs. (6)–(8) listed below.
COPH ¼
QH _ comp W
ð6Þ
Fig. 3. Flow chart of thermal analysis computational algorithm.
where
_ coll ¼ m
gcoll ¼
6. Economic analysis
VD
ð7Þ
v1 Q_ solar
ð8Þ
Icoll Ac
Since the system requires auxiliary energy due to intermittency of available solar energy, the effective coefficient of performance is determined as
COPeffectiv e ¼
Q_ load _ _ Aux Þ ðW comp þ W
ð9Þ
_ Aux is the auxiliary energy required to meet the load in the where W absence of solar radiation. From the integrated monthly energy fluxes, one can calculate auxiliary energy ðQ_ Aux Þ and solar energy ðQ_ solar Þ as a fraction of thermal load ðQ_ load Þ
fsolar ¼
Q_ solar Q_ load
and f Aux ¼
Q_ Aux Q_ load
In the present study, economic analysis is performed using the life cycle cost (LCC) method. The life cycle cost is the sum of all costs associated with the SAHP system over its’ entire lifetime, in present cost, that takes into account the time value of money. Several factors such as initial cost, operating cost, and maintenance cost need to be taken into account. The method for calculating LCC is well established and readers are referred to Ref. [44]. Taking into account the utility inflation rate, and the monetary inflation rate and the present worth (PW) factors, the LCC can be expressed as
LCC ¼ PWSystem þ PWInstillation þ PWmaintenance þ PWEnergy usage PWtax
ð11Þ
Using LCC one can compare the economic sustainability of the proposed SAHP system with other alternatives such as the solar electric water heater. The system cost and other related economic parameters are shown in Table 3.
ð10Þ
Table 2 provides the monthly thermal performance data of the system for all months of the year for the City of Norfolk, Virginia.
7. Primary energy resource utilization For analysis of primary energy utilization in competing systems for domestic hot water (DHW) applications, one can use the
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Table 2 Monthly thermal performance of SAHP system for the City of Norfolk, Virginia. Month
QLoad (kJ)
QH (kJ)
Wc (kJ)
QAux (kJ)
COPH
COPeffective
PERoverall
January February March April May June July August September October November December
1834249.00 1656741.10 1746604.40 1563606.90 1462840.20 1225478.40 1201085.20 1179075.90 1119741.80 1244711.10 1542687.60 1702978.90
980841.44 997399.10 1444560.50 1563606.90 1462840.20 1225478.40 1201085.20 1179075.90 1119741.80 1244711.10 1123494.90 988435.75
224066.16 207337.25 284766.20 278685.12 289305.78 280105.75 290158.25 289580.75 279336.97 233339.90 223628.38 224564.36
853407.56 659342.00 302043.88 0.00 0.00 0.00 0.00 0.00 0.00 0.00 419192.75 714543.10
4.38 4.81 5.07 5.61 5.06 4.38 4.14 4.07 4.01 5.33 5.02 4.40
1.70 1.91 2.98 5.61 5.06 4.38 4.14 4.07 4.01 5.33 2.40 1.81
1.05 1.14 1.41 1.85 1.67 1.44 1.37 1.34 1.32 1.76 1.28 1.08
primary energy ratio (PER), a parameter that characterizes the overall efficiency of a number of interconnected energy transformation processes. For instance, in the DHW applications, in order to determine the PER one must trace the energy flow all the way back to the primary energy source (coal, natural gas, oil, etc.) to determine how many kilo Joules of thermal energy are delivered for end-use for every kJ of primary energy utilized. The ratio of the delivered thermal energy and the primary energy is known as the primary energy ratio (PER). Fig. 4 shows the energy flow streams for two competing water heaters namely, an electric water heater, and a solar-assisted heat pump with electric resistor as a back-up system. In the case of an electric water heater, the efficiency of the thermal power plant (nth) and the electrical transmission efficiency (ntrans) must be considered to account for losses in those systems. As a result, the PER of the overall system, a combination of the power plant, transmission and electric water heater, is nthntrans if the heat losses from the water tank are neglected. For a coal fired power plant, with typical values of nth = 0.35 and ntrans = 0.90 yields a PER value of 0.315. This represents an effective yield of 31.5% with the remainder 68.5% of the primary energy being lost as waste heat in the power plant and the utility transmission lines. It should also be noted that even though electric water heater are marketed as environmentally clean devices, in reality there are emissions of global warming gases such as CO2 and other pollutants at the power plant that must be attributed to electric heater for each kW h of electrical energy used in it. For the solar- assisted heat pump system with and electric resistance back-up system, the chart showing energy flow from the primary energy source to the end-use (delivered) thermal energy is illustrated in Fig. 4. Using the first law of thermodynamics for energy balances across each sub-system, one can show that the PERSAHP, the primary energy ratio of the SAHP water heater system can be expressed as follows.
Table 3 System input parameters for economic analysis.
1 2 3 4 5 6 7 8 9 10 11 12
Parameter
Value
Life span-SAHP system Initial system cost -SAHP system Installation cost -SAHP system Maintenance cost per year-SAHP system Federal tax incentive-SAHP system (30% system cost) Life span-electric heater Initial system cost – electric heater Installation cost – electric heater Maintenance cost per year-electric heater Electricity cost per kW h Utility inflation rate Monetary inflation rate
20 years $3750 $2000 $40 $1725 20 years $500 $200 $20 $0.1 10% 7%
PERSAHP 1 i ¼h ðntrans nth Þ fAux þ ð1fAux Þ
ð12Þ
COPH
In terms of fsolar (the fraction of load met by solar energy), the above equation can be rewritten as,
PERSAHP 1 ¼ ðntrans nth Þ ð1 fsolar Þ
ð13Þ
where
fsolar ¼ ð1 fAux Þ 1 The term
1 COPH
1 1 COP H
ð14Þ
can be regarded as the limiting value for
fsolar since any value of fsolar above it will result in a negative value for fAux. This limiting value of fsolar increases with COPH. For instance, for COPH = 5.0, the limiting value of fsolar is 0.8 and for COPH = 3.0, it is 0.67. Since the PER of the electric hot water equals ntransnth, PERSAHP can be expressed as
PERSAHP 1 ¼ 1 fsolar PERelec
ð15Þ
The primary energy ratio (PER) is defined as the ratio of the thermal energy delivered to the primary energy consumed. The variation of PERSAHP/PERelec as a function of fsolar is shown in Fig. 5. It is noted that as fsolar approaches 0.0, fAux equals 1.0 and the SAHP system is 100% reliant on the auxiliary electric heat, and the PERSAHP is accordingly equal to PERelec. At the other extreme, as fsolar approaches 1, the SAHP system operates in the solar only system mode, and the ratio approaches infinity asymptotically. The latter limit is not technically feasible as it would require COPH to approach infinity. This is also not economically feasible as it would require a very large solar collector field to achieve 100% solar only limit. In practical systems, fsolar is generally limited to a value in the range of 0.50–0.70. Since PERSAHP/PERelec is always greater than 1.0, from the perspective of efficient primary resource utilization, the SAHP water heater is always a better alternative than the electric water heater. However, from an economic viewpoint, this may not always be the case because of the tradeoffs between higher capital cost (for the SAHP system) and higher operating cost (for the electric water heater). 8. Discussion of results-annualized system performance In this section, the effects of changes in the collector area, compressor displacement volume (VD) and the load temperature on
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Fig. 4. Energy flow diagram for the SAHP system.
40 35
PER solar /PER elec.
30 25 20 15 10 5 0
0
0.2
0.4
0.6
0.8
1
fSolar Fig. 5. Variation of PERSAHP/PERelec as a function of fsolar.
the long-term system performance are discussed. Fig. 6 shows the variation of the system life cycle cost (LCC) with the collector area for a specified load temperature and compressor displacement volume. The presented load temperature is 60 °C and the volume
displacement of the compressor is 0.0003518 m3/s. From Fig. 6, one observes that a minimum value of LCC occurs at a collector area of 3 m2. Further increase in collector area beyond the minimum point yields a higher LCC. This is the optimal configuration from the long-term economic perspective due to two counteracting effects, namely the capital cost and the operating cost of the system. The capital cost of the system for a fixed load temperature and compressor displacement volume increases with increasing collector area while the system operational cost decreases. Increased solar energy collection raises the collector/evaporator temperature, leading to a lower temperature lift in the heat pumping process. This translates into reduced compressor power input, and consequently lower operational cost. The variation of the LCC with the compressor displacement volume is shown in Fig. 7. For a fixed 3 m2 solar collector area, it can be seen that a minimum value of LCC is obtained at VD = 0.0003518 m3/s. The reason for this occurrence has to do with two opposing trends _ comp Þ, and the auxinvolving the variation of compressor power ðW _ Aux Þ, as the compressor volume is changed. iliary electrical powerðW Since the capital cost is fixed for a given collector area, the varia_ Aux and W _ comp . As higher tion in LCC is solely due to changes in W values of compressor displacement are chosen the refrigerant mass flow rate through the system also increases. This results in greater _ comp . Since W _ Aux is the differheat pump capacity ðQ_ H Þ and higher W ence between the fixed load the heat pump capacity ðQ_ H Þ, the net _ Aux . These two effects transpire to yield an result is a decrease in W optimal value of VD as shown in Fig. 7.
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14000
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o
Tc=50 o Tc=55 o Tc=60 o Tc=65 o Tc=70 2
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2
Collector area (m ) Fig. 6. Variation of life cycle cost with collector area (VD = 35.18 105 m3/s, load temperature = 60 °C).
Fig. 8. Variation of life cycle cost (LCC) with collector area and load temperature.
3
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Ac=3 Ac=4 Ac=5 Electical heater
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Fig. 7. Variation of life cycle cost (LCC) with compressor displacement (VD) (Ac = 3.0 m2/s, load temperature = 60 °C).
Fig. 9. Variation of overall annual per with collector area and load temperature.
The variation of LCC as a function of the collector area and the load temperature is shown in Fig. 8. For comparison the LCC of the electric water heater is also displayed. One notes that as the load temperature increases the LCC increases for a given collector area (fixed capital cost) because the operating costs are higher due to the higher temperature lift during the heat pumping process. Also, for a given load temperature, the LCC value increases with area since the capital cost is greater than the reduction in the operating cost due to enhanced heat pump efficiency. Towards the higher end of the load temperature range, as seen in Fig. 8, the heat pump LCC becomes comparable to the LCC of the electric hot water system. Unlike the electric hot water heater whose thermal and economic performance is independent of the load temperature, the performance of the SAHP system depends strongly on the load temperature. Fig. 9 shows the variation of the annual PER of the SAHP system as a function of the collector area and the load temperature. For comparison, the PERelec of the electric hot water heater is also included. For all collector areas and load temperatures the PERSAHP values are higher than unity and also well above the PERelec value of 0.33. This indicates that the SAHP system utilizes primary energy far more efficiently compared to the electric hot water heater over the load temperature range considered in this study. The PERSAHP value decreases with the increasing load temperature
and this implies that the SAHP system is utilizing a primary energy less efficiently. For a fixed load temperature gain in PERSAHP are achieved. This is due to elevated collector temperature that results _ comp in a smaller temperature lift. The resultant reduced value of W implies a decrement in the primary energy resource demand for a given thermal load requirement. Fig. 9 also suggests that the SAHP system is best suited from the efficient energy resource utilization perspective, for water heating or preheating applications in the 50– 70 °C load temperature range. In this temperature range both the LCC and PER values are decidedly superior to the electric only system. 9. Conclusions The present study employed modeling and long-term averaged monthly solar data to simulate the long-term averaged thermal performance of a direct-expansion solar assisted heat pump. Results were obtained for a range of load temperature and for a number of collector areas and compressor displacement volumes. The thermo-economic analysis results of the SAHP system indicates that the system is an appropriate match for the temperature water heating applications. Since a large amount of primary energy is consumed in the USA in meeting the demands for low temperature
S.K. Chaturvedi et al. / Energy Conversion and Management 77 (2014) 550–557
(50–70 °C) domestic water heating applications, the SAHP water heaters are the ideal systems for such applications. These systems, as the present study reveals are both economical as well as energy conserving solutions compared to the electric-only hot water heaters. Reduced consumption of electricity and the capture of renewable solar energy through collectors should have a favorable environmental impact since reduced carbon emission would result if these systems are used as an alternative to the electric-only systems. The simulation of the system also indicates that its long-term economic performance (LCC) can be optimized with respect to the collector area and the compressor displacement volume for a specified load temperature. At load temperatures higher than 70 °C, the performance of the single stage heat pump considered in this study degrades to the extent that its cost and efficiency advantages over the electric only system are lost. In the higher load temperature range (>70 °C) a two stage heat pump system may be more appropriate [45] and its long-term thermo economic cost-efficiency benefit analysis should be considered in a future study.
References [1] World consumption of primary energy by energy type and selected country groups. Energy Information Adminstration, U.S. Department of Energy; 2006. [2] http://en.wikipedia.org/wiki/Global_warming. [3] Annual energy outlook. U.S. Department of Energy; 2006. [4] www.eere.energy.gov/consumer/your_home/water_heating/index.cfm/ mytopic=12760. [5] www.solarhome.org/newssolarwaterheaters.html. [6] Kush BA. Performance of heat pump at elevated evaporating temperatures with application to solar input. Sol Energy 1980;102:203–10. [7] Franklin J, Saaski F, Yamagiwa A. A high efficiency direct expansion solar panel. In: Flat plate solar collector conference, Orlando, Florida, USA; 1977. p. 187–95. [8] Sporn P, Ambrosc E. The heat pump and solar energy. In: The world symposium on applied solar energy, Phoenix, Arizona, USA; 1955. p. 1–5. [9] Chaturvedi SK, Chiang YF, Roberts AS. Analysis of two phase flow collectors with applications to heat pumps. Sol Energy Eng 1982;104:358–64. [10] Abou-Ziyan HZ, Ahmed MF, Metwally MN, Abd El-Hameed HM. Solar-assisted R22 and R134a heat pump systems for low-temperature applications. Appl Therm Eng 1997;17:455–69. [11] Chaturvedi SK, Chen DT, Kheireddine A. Thermal performance of a variable capacity direct expansion solar-assisted heat pump. Energy Convers Manage 1998;39:181–91. [12] Chaturvedi SK, Abdel-Salam TM, Sreedharan S. Two-stage direct expansion solar-assisted heat pump for high temperature applications. Appl Therm Eng 2009;29:2093–9. [13] Aziz W, Chaturvedi SK, Kheireddine A. Thermodynamic analysis of twocomponent, two-phase flow in solar collectors with application to a directexpansion solar-assisted heat pump. Energy 1999;24:247–59. [14] Li YW, Wang RZ, Wu JY, Xu YX. Experimental performance analysis on a directexpansion solar-assisted heat pump water heater. Appl Therm Eng 2007;27:2858–68. [15] Chaturvedi SK, Abazeri M. Transient simulation of a capacity modulated, direct-expansion. Sol Energy 1987;39:421–8. [16] Chaturvedi SK, Roberts AS, Mei V. Solar collector as heat pump evaporator. In: 13th Intersociety energy conversion conference, Boston, USA; 1979. p. 4–9. [17] Kuang YH, Sumathy K, Wang RZ. Study on a direct-expansion solar-assisted heat pump water heating system. Int J Energy Res 2003;27:531–48.
557
[18] Huang BJ, Chyng JP. Performance characteristic of integral type solar-assisted heat pump. Sol Energy 2001;71:403–14. [19] Chaturvedi SK, Mei V. Thermal performance of SAHP system with combined collector/evaporator. In: AIAA terrestrial energy systems conference, Orlando, Florida, USA; 1979. [20] Kuang YH, Wang RZ, Yu LQ. Experimental study on solar assisted heat pump system for heat supply. Energy Convers Manage 2003;44:1089–98. [21] Kuang YH, Wang RZ. Performance of a multi-functional direct expansion solar assisted heat pump system. Sol Energy 2006;80:795–803. [22] Huang BJ, Chyng JP. Integral-type solar-assisted heat pump water heater. Renew Energy 1999;16:731–4. [23] Huang BJ, Lee JP, Chyng JP. Heat-pipe enhanced solar-assisted heat pump water heater. Sol Energy 2005;78:375–81. [24] Yumrutas R, Unsal M. Analysis of solar aided heat pump system with seasonal thermal energy storage in surface tanks. Energy 2000;25:1231–43. [25] Krakow KL, Lin SA. Solar source heat pump with refrigerant cooled solar collectors for cold climates. Trans ASHRAE 1982;88:417–39. [26] O’Dell MP, Mitchell JW, Beckman WA. Design method and performance of heat pump with refrigerant-filled solar collectors. Sol Energy Eng 1984;106:159–64. [27] Matsuki K, Shinobu Y, Yoshikawa M. A prototype direct-expansion solar heat pump system. Trans ASHRAE 1987;93. [28] Ito S, Miura N, Wang K. Performance of a heat pump using direct expansion solar collectors. Sol Energy 1999;65:189–96. [29] O’Dell MP, Mitchell JW, Beckman WA. Beckman, solar heat pump systems with refrigerant-filled collectors. Trans ASHRAE 1983;89:519–25. [30] Ito S, Miura N, Wang K. Performance of a heat pump using direct expansion solar collectors. Sol Energy 1998;65:189–96. [31] Hawlader MN, Chou SK, Ullah MZ. The performance of a solar assisted heat pump water heating system. Appl Therm Eng 2001;21:1049–65. [32] Ito S, Miura N, Takano Y. Studies of heat pumps using direct expansion type solar collectors. Sol Energy Eng 2005;127:60–4. [33] Chyng JP, Lee CP, Huang BJ. Performance analysis of a solar assisted heat pump water heater. Sol Energy 2003;74:33–44. [34] Gorozabel Chata FB, Chaturvedi SK, Almogbel A. Analysis of a direct expansion solar assisted heat pump using different refrigerants. Energy Convers Manage 2005;46:2614–24. [35] Fernández-Seara J, Piñeiro C, Dopazo JA, Fernandes F, Sousa PXB. Experimental analysis of a direct expansion solar assisted heat pump with integral storage tank for domestic water heating under zero solar radiation conditions. Energy Convers Manage 2012;59:1–8. [36] Kara O, Ulgen K, Hepbasli A. Exergetic assessment of direct-expansion solarassisted heat pump systems: review and modeling. Renew Sustain Energy Rev 2008;12:1383–401. [37] Kong XQ, Zhang D, Li Y, Yang QM. Thermal performance analysis of a directexpansion solar-assisted heat pump water heater. Energy 2011;36:6830–8. [38] Kuang YH, Wang RZ. Performance of a multi-functional direct-expansion solar assisted heat pump system. Sol Energy 2006;80:795–803. [39] Ozgener O, Hepbasli A. A review on the energy and exergy analysis of solar assisted heat pump systems. Renew Sustain Energy Rev 2007;11:482–96. [40] Tagliafico L, Scarpa F, Tagliafico G, Valsuani F. An approach to energy saving assessment of solar assisted heat pumps for swimming pool water heating. Energy Build 2012;55:833–40. [41] Chaturvedi SK, Mohieldin TO, Chen DT. Second-law analysis of solar-assisted heat pumps. Energy 1991;16:941–9. [42] Cervantes GJ, Torres-Reyes E. Experiments on a solar-assisted heat pump and an exergy analysis of the system. Appl Therm Eng 2002;22:1289–97. [43] Duffie JA, Beckman WA. Solar engineering of thermal processes. John Wiley & Sons Inc.; 1991. [44] Dhillon BS. Life cycle costing for engineers. Boca Raton, Florida, USA: CRC Press, Taylor & Francis Group; 2010. [45] Gagrani VD. Simulation and thermo-economic analysis of a solar-assisted heat pump system for domestic water heating application. Mechanical engineering. Norfolk, Virginia, USA: Old Dominion University; 2007.