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A comparative study of pure and zeotropic mixtures in low-temperature solar Rankine cycle
Applied Energy 87 (2010) 3366–3373
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A comparative study of pure and zeotropic mixtures in low-temperature solar Rankine cycle J.L. Wang, L. Zhao *, X.D. Wang Department of Thermal Energy and Refrigeration Engineering, School of Mechanical Engineering, Tianjin University, No. 92 Weijin Road, Tianjin 300072, PR China
a r t i c l e
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Article history: Received 22 March 2010 Received in revised form 10 May 2010 Accepted 12 May 2010 Available online 17 June 2010 Keywords: Zeotropic mixture Rankine cycle Solar Low temperature
a b s t r a c t The paper presents an on site experimental study of a low-temperature solar Rankine cycle system for power generation. The cycle performances of pure fluid M1 (R245fa) and zeotropic mixtures M2 (R245fa/R152a, 0.9/0.1) and M3 (R245fa/R152a, 0.7/0.3) are compared, respectively, based on the experimental prototype. The experiments have been conducted under constant volume flow rate of different fluids. The results show that, with the component of R152a increasing, the system pressure level increases and the power output varies accordingly, which provides an additional means of capacity adjustment. The collector efficiency and thermal efficiency of zeotropic mixtures are comparatively higher than pure fluid of R245fa in the experimental condition, which indicates that zeotropic mixtures have the potential for overall efficiency improvement. Due to the non-isothermal condensation of zeotropic mixture, the condensing heat could be partially recovered by adding an external heat exchanger. Thus, compared with pure fluid R245fa the system overall efficiency of zeotropic mixtures could be improved. Crown Copyright Ó 2010 Published by Elsevier Ltd. All rights reserved.
1. Introduction In the energy-to-power conversion industry, the thermal efficiency becomes uneconomically low when the gaseous steam temperature drops below 700 F (371 °C), the bulk of the energy loses and systems using water as working fluid become significantly less efficient and capital cost increase [1]. To overcome this disadvantage and still exploit the basic Rankine cycle technology developed over the years, systems based on working fluids such as hydrocarbons or refrigerants are being developed and researched in recent years. This is so called organic Rankine cycles (ORCs). The organic fluids is most promising for Rankine cycle power generation systems in that they can use low grade heat from a variety origins, such as geothermal energy [2,3], solar radiation [4], biomass combustion [5], waste heat from the industrial process [6–8]. It should be pointed out that, in the solar powered Rankine cycle research area, some important solar Rankine cycle systems have been developed and examined recent years, such as using freon R245fa [9] and natural fluid-CO2 [10–13] as working fluid. In the design of the cycle system, one important factor should be considered is the working fluid selection. Regarding the working fluids, several studies envisage the utilization of different working fluids that are more appropriate for low temperature heat source than tradition steam power cycles, such as, benzene * Corresponding author. Tel.: +86 22 81590264; fax: +86 22 27404188. E-mail address: [email protected] (L. Zhao).
[14], n-pentane [15], isobutane [16], and some cryogens like the freons R11, R113, R114, Rl23 and R134a [17–20]. It can be seen that the previous studies of ORCs regarded more about single component organic fluids. However, an important limitation of pure fluids is the constant temperature evaporation which is not suitable for sensible heat sources such as waste heat. The mixtures have variable temperature during the phase change process, which could be used to reduce the mismatch of temperature profiles between heat transfer fluid and the evaporating or condensing working fluid mixtures. Thus, the system irreversibilities can be minimized. Another advantage of the mixture is that it could be acquired the fluids that have the same thermodynamic properties as the pure ones through different component mass friction. This could greatly extend the range of candidate working fluid selection for low temperature Rankine cycle. Presently, it can be noticed that there is a significantly arising interest in multi-component mixture research. Radermacher demonstrated the mutual influence of working fluid mixtures properties on Rankine cycle performance, and counter-flow heat exchangers are suggested in the system for the mixtures [21]. Angelino and Paliano compared n-pentane and mixture of n-butane and n-hexane (50%/50%) by simulating liquid geothermal resource for electricity generation. The results show that mixture yields 6.8% more electricity than n-pentane and 25% less air are used with potential benefits in both cooler frontal area and fan power consumption [22]. Aleksandra studied Rankine cycle with heat source temperature of 80–115 °C by using different working
0306-2619/$ - see front matter Crown Copyright Ó 2010 Published by Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2010.05.016
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Nomenclature A 0 h1, h2 I _ m Q_ E _P W _T W
collector aperture area, m2 enthalpy of different state point in T–S diagram, kJ/kg global solar radiation, W/m2 mass flow rate of working fluids, kg/s heat absorbed in the collector, kW power consumed by working fluid pump, kW power generated by expander, kW
fluids, both natural and synthetic as well as mixtures [23]. Wang and Zhao made a theoretical analysis of zeotropic mixtures R245fa/R152a used in low-temperature solar Rankine cycles [24]. However, since the research in this field is in a very early stage, very few experimental studies on low temperature Rankine cycle for power generation using zeotropic mixtures are reported. In order to analyze the thermal performance of low-temperature solar Rankine cycle system using zeotropic mixtures, the paper concentrates on relative experimental researches. An experimental prototype is constructed and tested, such as the one that this paper described. The flat-plate collector, which does not need a complex tracking system and thus has low cost, has been considered, and the experimental evaluation and cycle performances could be a tool for their further optimization.
gC gR gT gsys
collector efficiency, % Rankine cycle efficiency, % turbine efficiency, % overall efficiency, %
vaporizes and finally superheats at the collector outlet. High pressure vapor is directed to the expander which is coupled with a generator. The exhaust vapor of the expander passes through the condenser. Condensation and a small degree of sub-cooling occur. Liquid is then sucked by the pump and the cycle recommences.
2. Working fluid selection In ORCs applications, the choice of working fluid is important since the fluid must not only possess thermophysical properties that match the application but also have adequate chemical stability at the desired working temperature. Here, R245fa and R152a are chosen as the components of the mixture because they both have zero ODP (Ozone Depression Potential) and lower GWP (Global Warming Potential), which have less environmental impact. The property details of the compositions are shown in Table 1. Here, pure fluid R245fa is noted as M1 and the mass fraction of zeotropic mixtures M2 and M3 are 0.9/0.1 and 0.7/0.3, respectively. Under this mass fraction, the fluids have comparatively low critical pressure, which is acceptable in the experimental condition. The temperature–entropy diagrams of these fluids are shown in Fig. 1. In Fig. 1, pressures of 2.0 MPa and 0.2 Mpa for different fluids are drawn, respectively. It is clear that as the mass fraction of R152a increases the pressure level rises and at the same time it is desirable to have the high pressure resistant equipment. The system economic efficiency needs to be further revalued.
Fig. 1. T–S diagram of different working fluids.
3. System description 3.1. Experimental apparatuses Fig. 2 illustrates the basic solar Rankine cycle system. The thermodynamic process can be briefly described as follows: the liquid organic working fluid is compressed with a working fluid pump which forces the fluid through the flat-plate collector. Heat transfers from solar radiation to the liquid, and the fluid is preheated,
Fig. 2. Basic solar Rankine cycle for electricity generation.
Table 1 Property details of working fluids.
a
Fluid
Molecular weight (g/mol)
Normal boiling pointa (°C)
Critical pressure (MPa)
Critical temperature (°C)
ODP
GWP
Safety group
R245fa R152a
134.05 66.05
15.14 24.02
3.65 4.52
154.01 113.26
0 0
950 120
B1 A2
The normal boiling point is boiling point at a pressure of 1 atmosphere.
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3.2. Data acquisition system The system is instrumented comprehensively. Each set of measurements is taken in a time step of 1 min and the logged data are recorded as a function of time. The testing parameters include temperature, pressure, flow rate and solar radiation. Temperature T-type thermocouples with accuracy of ±0.5 °C are mounted in the system to provide temperature measurement. The testing points are shown in Fig. 3 and the ambient temperature is also recorded. Pressure transducers provide pressure values for collector outlet and throttling outlet. The accuracy of the transducer of points 1 and 2, as shown in Fig. 3, is ±3 kPa and ±1 kPa, respectively. Solar radiation to measure the global solar radiation, a pyranometer with accuracy less than ±0.05 W is installed on the 45° slop surface which has the same angle of collector. Fig. 3. Experimental layout of the low-temperature solar Rankine cycle.
The experimental prototype mainly consists of two flat plate solar collector, a throttling valve, an air-cooled condenser, a storage tank, a fluid pump, a radiometer and date acquisition system. Fig. 3 shows the layout of the experimental system. Because the power output of the experimental prototype is too small, there is no real expander available for this system. Therefore, a throttling valve is used in the system to expand the vapor instead of a real expander. By adjusting the valve opening state, the pressure difference of the realistic expander can be simulated accordingly. Obviously, the fluid temperature and its thermodynamic and transport properties and states at the valve outlet are different from those of the true expander condition. However, the cycle performance for the power production can be derived based on thermodynamic analysis. As the fluid passing through the valve is a throttling process, the power generated from the expander can be calculated by setting the expander efficiency as 0.85 [25]. During the experiment, only one flat-plate collector, with its internal tube tank diameter of 12 mm, is used. The collector aperture area is 0.6 m2. It possesses one glass cover sheet and a high efficiency solar absorber plate whose absorptivity is 0.95 and emissivity is 0.17. Fig. 4 presents a photo of the experimental prototype.
The output signals of these instruments are connected to a PC through an Agilent 34980A series data logger and data are recorded as functions of time simultaneously. A diaphragm metering pump, which could be regulated in the range of 0–100% of rated flow, is used in the prototype to feed the fluids and control volume flow rate. 4. Assumption and calculation method The paper presented is the results of calculations regarding the Rankine cycle, which is shown in Fig. 3. Thermodynamic parameters of working fluids have been taken from the computer program of REFPROP8.0 [26], which is developed by the National Institute of Standards and Technology. Some additional parameters must be set to perform the calculation in this case: 1. Each component is considered as a steady-state-flow system. 2. The kinetic and potential energies as well as friction losses are neglected. 3. The expander isentropic efficiency gT is set at 0.85. 4. Furthermore, composition shift is neglected in the experimental system. 5. No pressures drop in the heat exchangers and pipelines have been considered. For the conditions, the cycle T–S diagram of pure fluid R245fa and zeotropic mixtures are illustrated in Figs. 5 and 6, respectively.
Fig. 4. A photo of the experimental prototype.
Fig. 5. Schematic diagram for R245fa Rankine cycle.
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valve open state was also kept in the same open state without being adjusted. 5. Results and discussion
Fig. 6. Schematic diagram for zeotropic mixtures Rankine cycle.
Based on the above measured parameters, it is possible to get a set of calculated values for pure fluid R245fa and zeotropic mixtures, and thus obtain a global understanding of the system behavior and performance. These values include:
_ T ¼g m _ ðh1 h20 Þ Power generated by the expander W T
ð1Þ
_P Power consumed by the working fluid pump W _ ðh5 h40 Þ ¼m
ð2Þ
_ ðh1 h5 Þ Heat absorbed in the collector Q_ E ¼ m Collector efficiency gC ¼
Q_ E IA
Rankine cycle efficiency gR ¼
ð3Þ ð4Þ
_P _ T W W _ QE
Overall efficiency gsys ¼ gC gR ¼
_ T W _P W IA
ð5Þ
ð6Þ
where h1, h20 , h40 and h5 are the enthalpies of corresponding state _ is the fluid mass flow rate of the system, points in T–S diagram. m which could be converted from the experimental volume flow rate. And, I and A represent the global solar radiation on the collectors’ inclination surface and collector aperture area, respectively. It should be pointed out that the overall efficiency gsys in Eq. (6) is the ratio of the net power output to the solar radiation on the collector’s surface. In order to evaluate the cycle performances of different working fluid, the volume flow rate was set as constant of 1.3 L/h and the
Fig. 7. Solar radiation during the experiment.
The experimental procedure was conducted in spring (April 2009) in the city of Tianjin, China. Fig. 7 shows the solar radiation for the 1st, 6th and 16th of April 2009, respectively. It can be seen that the solar radiation of 1st and 6th are almost identical, and the maximum value are both 1055 W/m2 which appeared at 11:48 a.m. and 12:02 a.m., respectively. Affected by the air quality, the overall solar radiation of 16th is slightly lower and the maximum value is appeared at 12:02 a.m. with value of 1001 W/m2. Compared with 1st and 6th, the solar radiation of 16th does not result to an extreme change of system operation and this radiation deviation is almost insignificant, which could be neglected. Fig. 8 shows the collector inlet and outlet temperature of different working fluids during the experiment. Since the process of phase change occurs in the collector, the collector outlet temperature is greatly influenced by solar radiation and ambient temperature. Thus, the outlet temperatures fluctuate during the experiment. However, under the daily constant flow rate, the overall trend of the outlet temperatures of different fluids follows that of the solar radiation, and the maximum outlet temperatures are 105.9 °C (M1, 12:56 a.m.), 99.86 °C (M2, 12:55 a.m.) and 101.56 °C (M3, 13:03 p.m.), respectively. During the experiment, an interesting observation concerning the behavior of the system is that compared with the time of maximum value of solar radiation, a hysteresis of the fluids maximum outlet temperature is observed. It could be noticed that the time lag is about 1 h. The identified time lag of hysteresis is directly related to the system thermal inertia of solar collector. The same phenomenon could be found in the following operation in the afternoon, that is, when the solar radiation descends from the maximum, collector outlet temperature exhibit no trend of decreasing but maintain the same states of high temperature. It is because that the collector internal tube tank and thermal insulation material act as a thermal storage during the experiment and heat is stored when the solar radiation is sufficient. Though there is collector heat loss, the working fluids could still get the sufficient heat for fluids evaporating and superheating and the collector outlet temperatures could achieve high values. When the stored heat and solar radiation are lower than collector heat loss, the outlet temperatures begin to drop, so as the degree of superheating. Therefore, the time of outlet temperature beginning to decrease is also behind that of solar radiation by 1 h. According to Manolakos [4], the thermal inertia of the system makes the operation more stable and smoothly reacted to the variation of solar radiation.
Fig. 8. Collector inlet and outlet temperature.
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In Fig. 8, as the collector inlet temperature of M1 is comparatively higher during the experiment, the outlet temperature of M1 is higher than M2 under the approximately similar solar radiation. Though the solar radiation of 16th April has small deviation compared with 1st and 6th, the outlet temperature of M1 still reach high value. Fig. 9 shows the pressure of different working fluid in the experimental monitoring point. As mentioned above, the pressure level of the mixtures increases with the adding of R152a. Obviously, M3 has the highest pressure level and M1 has the lowest. Influenced by the collector outlet temperature fluctuation, the collector outlet pressures of three fluids show the similar variation. It should be noted that, for zeotropic mixtures M2 and M3, the composition shift, which probably leads to composition change during the phase change process, results in collector outlet pressure fluctuating acutely during the operation. The mechanism of this phenomenon needs further study. Since the air-cooled condenser is used in the prototype, the pressures after the valve have the relationship to ambient temperature and thus behave smooth trend. The ambient temperature, together with the condenser inlet and outlet temperature, is illustrated in Fig. 10. Condenser inlet temperature, namely the valve outlet temperature, has the great relationship to the collector outlet temperature as the fluid passing the valve is a process of throttling. At the same time, affected by the ambient temperature the condenser outlet temperatures increase with the ambient temperature ascending. The condenser outlet temperatures of three working fluids are all higher than ambient temperature about 2 or 3 °C on average. The ambient temperature of Fig. 10a is higher than that of Fig. 10b and c, which leads to a higher collector inlet temperature of M1 in Fig. 8. Fig. 11 shows different degree of superheating at collector outlet. It can be seen that the collector outlet states of M1 are superheating in the whole experimental process. However, the outlet states of M2 and M3 could be divided into three stages: liquid–vapor phase at the beginning, superheating in the steady state, liquid–vapor phase at the end of the experiment. This fact could be explained that, compared with M2 and M3, M1 has the higher collector inlet temperature and lower latent heat during evaporation, which brings on the large degree of superheating. As the working fluid phase change process plays an important role in the whole cycle, when incident radiation is lower, heat absorption in the collector will decrease and therefore boiling is incomplete. This will have a negative effect on the system operation, which could be seen in the following analysis. In order to illustrate the superheating of different fluid under same constant flow rate, Table 2 presents the degree of superheating of different fluids during the experiment. Pure working fluid M1
Fig. 10. Ambient temperature and condenser inlet and outlet temperature.
Fig. 9. Collector outlet pressure and after valve pressure.
has the maximum superheating of 54.89 °C which is higher than M2 and M3 of 7.06 °C and 13.87 °C, respectively. Accordingly, M1 has the maximum average degree of superheating. This result in collector heat loss increases and collector efficiency decreases. Because a high degree of superheating in the expander inlet would not improve Rankine cycle efficiency [8], the Rankine cycle efficiency of M1 does not show any special improvement. It deserves to mention that an optimized regulation of fluid flow rate is required in order to minimize the superheating heat loss, which in turn increases the collector efficiency. The heat absorbed by the working fluids in the collector could be calculated from Eq. (3). Enthalpy difference, together with heat absorption in the flat-plate collector, is illustrated in Fig. 12. During the period of 8:34–9:24 (6th April) and 8:34–9:41 (16th April), as the fluid do not finish the phase change process in the collector, fluid M2 and M3 have the lower heat absorption, so as the enthalpy difference. From Fig. 12a the calculation results show that, in the
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Fig. 12. Specific enthalpy difference and heat absorbed by the fluid in the collector.
be seen that fluids heat absorption are almost identical during the superheating period. This can be explained that, with the component of R152a increasing, the mixtures have lower critical density and the mass flow rates are different correspondingly. According to the calculation, under the volume flow rate of 1.3 L/ h the mass flow rate of M1, M2 and M3 is 1.73 kg/s, 1.69 kg/s and 1.51 kg/s, respectively. Therefore, if the fluids were in the same mass flow rate, zeotropic mixtures will show a potential to improve the collector efficiency. Fig. 13 shows the collector efficiency during the experiment. As mentioned above, in the periods of 8:34–9:24 (6th April) and 8:34– 9:41 (16th April), M2 and M3 are in the liquid–vapor states and the collector efficiencies are lower as 4.19% and 12.27%, respectively. From 9:50 to 15:19, the fluids are all in the state of superheating Fig. 11. Superheating at the collector outlet.
Table 2 Degree of superheating during the experiment. Working fluid
M1
M2
M3
Time of superheating
8:39– 15:22 54.89
9:50– 15:40 47.83
9:40– 15:19 41.02
35.19
27.86
26.37
Maximum degree of superheating (°C) Average degree of superheating (°C)
superheating period of 9:50–15:19, the averaged specific enthalpy difference of M3 is 281.17 kJ/kg, which is higher than M2’s 258.31 kJ/kg and M1’s 248.97 kJ/kg. Therefore, under the similar solar radiation the zeotropic mixtures have large specific heat absorption. Fig. 12b shows the actual heat absorption of different working fluid under the same volume flow rate of 1.3 L/h. It can
Fig. 13. Collector efficiency.
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and the collector efficiency curves exhibit the shape of concave. The minimum point of the concave curve is the time of maximum solar radiation. It’s because that, when the solar radiation reaches its maximum value at noon, the heat loss of collector also reaches its maximal, and thus the collector efficiency shows the minimal value. When the solar radiation begin to fall, just as the system thermal inertia mentioned above, the attenuation of the collector outlet temperature lags behind that of the solar radiation and the collector efficiencies behave the trend of ascending. Integrating Figs. 7 and 12b, Fig. 13 shows that the collector efficiencies of three fluids are almost the same in the period of superheating. M3 shows slightly higher collector efficiency than M1 and M2 by 7.91% and 6.45%, respectively, as is presented in Table 3. In order to evaluate the cycle performance of different working fluid, it deserves to concentrate on the states of superheating from 9:50 to 15:19, which are the steady states of three working fluids. The following analyses are based on this period. The power produced from the expander is shown in the following Fig. 14. As zeotropic mixture M3 has the lower superheating degree at collector outlet, a comparatively higher power output is obtained at the maximum of 9.06 W. Affected by the fluctuation of collector outlet temperature and pressure, the power output of M2 varies from 4.69 W to 7.69 W. The average values could be found in Table 3. The average power output of M3 is higher than that of M1 and M2 by 29.10% and 28.03%, respectively. It can be seen that zeotropic mixtures have the great ability for power output, and the system capacity adjustment could be easily realized under different composition. The Rankine cycle efficiencies and the overall efficiencies can be calculated from Eqs. (5) and (6). Table 3 summarizes a set of characteristic value of cycle performance for different working fluids,
Table 3 Comparison of different fluid cycle performance. Working fluid
M1 (April 1st)
M2 (April 6th)
M3 (April 16th)
Period of superheating
9:50– 15:19 118.87
9:50– 15:19 120.17
9:50– 15:19 118.86
1.04
0.88
1.07
5.98 21.25 4.16 0.88
6.03 21.54 4.29 0.92
7.72 22.93 5.59 1.28
Average heat absorbed in the collector, W Average power consumed by working fluid pump, W Average power output, W Average collector efficiency, % Average Rankine cycle efficiency, % Average overall efficiency, %
which is useful to get an idea of the overall performance of the system. The averaged Rankine cycle efficiency of M3 is the highest of 5.59%. On the contrary, M1 has the lowest of 4.16%. The system overall efficiency has a great relationship to collector efficiency and Rankine cycle efficiency, so the average overall efficiency of M1, M2 and M3 is 0.88%, 0.92% and 1.28%, respectively. These low values, in some extent, are in close relation to the lack of regulation of flow rate of working fluids. As the zeotropic mixtures have the large amount of condensation heat, measures could be taken to recover part of condensation heat. For example, an external heat exchanger could be added between the expander and condenser to recover low temperature heat which could be used for domestic hot water supply. Due to the temperature glide of zeotropic mixtures during condensation, they would get the hot water whose outlet temperatures are higher than that of pure fluids. Therefore, the overall efficiency of zeotropic mixtures can be improved.
6. Conclusion An investigation on low-temperature solar Rankine cycle performance has been performed based on an experimental prototype by using zeotropic mixtures and pure fluid R245fa. Based on the constant flow rate of the fluid, the current work mentioned above is largely dedicated to make a system comparison between fluids M1, M2 and M3 to identify the cycle performance. The main results of the experimental measurements can be extracted as follow: The thermal inertia of the collector leads to the maximum outlet temperature lagged behind that of the solar radiation about 1 h. Under sufficient solar radiation to evaporate the working fluids in the collector, the flow rate is an important factor affecting the cycle performance. In the following procedure, a further optimization of flow rate regulation is needed. Due to the larger latent heat of evaporation, fluid M3 has large specific enthalpy difference in the collector, which has a great potential to improve the collector efficiency comparing with the pure R245fa. In the experimental superheating period, the average power output of M3 is higher than that of M1 and M2 by 29.10% and 28.03%, respectively. It can be seen that the power output varies accordingly and the system capacity adjustment could be easily realized under different composition of zeotropic mixtures. The average overall efficiency of M1, M2 and M3 is 0.88%, 0.92% and 1.28%, respectively. For zeotropic mixtures, there is a great potential to improve the overall efficiency by introducing an external heat exchanger to recover the partial condensation heat. Acknowledgement The authors would like to acknowledge the financial support provided by the Program for New Century Excellent Talents in University.
Appendix A. Uncertainty analysis
Fig. 14. Power output from the expander.
In every measured parameter there is an error between the measured and real value. Therefore, in order to evaluate the experiment data thoroughly and more reliable, an uncertainty analyses is necessary. In the following Table A there are the measured instruments and their uncertainty. Table B shows the total experimental uncertainty for every calculated value derived from measured data.
J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373
Table A. Measured parameters and uncertainty. Instruments
Parameters
Range
Uncertainty (%)
Pyranometer
Solar radiation
0–2000 (W/m2)
5.0
Working fluid pump
Volume flow rate
0–6.5 (L/ h)
5.0
Thermocouples
Collector inlet temperature Collector outlet temperature Condenser inlet temperature Condenser outlet temperature Pump inlet temperature Pump outlet temperature
10– 100 °C 10– 100 °C 10– 100 °C 10– 100 °C 10– 100 °C 10– 100 °C
1.5
Collector outlet pressure Throttling outlet pressure
0– 1.5 MPa 0– 0.5 MPa
1.0
Pressure transducer
2.0 2.0 1.5 1.5 1.5
0.7
Table B. Uncertainty of every variable. Calculated variable
Total experimental uncertainty (%)
Power generated by the expander Power consumed by the working fluid pump Heat absorbed in the collector Collector efficiency Rankine cycle efficiency Overall efficiency
5.72 5.56 5.68 7.57 6.59 8.27
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
A comparative study of pure and zeotropic mixtures in low-temperature solar Rankine cycle J.L. Wang, L. Zhao *, X.D. Wang Department of Thermal Energy and Refrigeration Engineering, School of Mechanical Engineering, Tianjin University, No. 92 Weijin Road, Tianjin 300072, PR China
a r t i c l e
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Article history: Received 22 March 2010 Received in revised form 10 May 2010 Accepted 12 May 2010 Available online 17 June 2010 Keywords: Zeotropic mixture Rankine cycle Solar Low temperature
a b s t r a c t The paper presents an on site experimental study of a low-temperature solar Rankine cycle system for power generation. The cycle performances of pure fluid M1 (R245fa) and zeotropic mixtures M2 (R245fa/R152a, 0.9/0.1) and M3 (R245fa/R152a, 0.7/0.3) are compared, respectively, based on the experimental prototype. The experiments have been conducted under constant volume flow rate of different fluids. The results show that, with the component of R152a increasing, the system pressure level increases and the power output varies accordingly, which provides an additional means of capacity adjustment. The collector efficiency and thermal efficiency of zeotropic mixtures are comparatively higher than pure fluid of R245fa in the experimental condition, which indicates that zeotropic mixtures have the potential for overall efficiency improvement. Due to the non-isothermal condensation of zeotropic mixture, the condensing heat could be partially recovered by adding an external heat exchanger. Thus, compared with pure fluid R245fa the system overall efficiency of zeotropic mixtures could be improved. Crown Copyright Ó 2010 Published by Elsevier Ltd. All rights reserved.
1. Introduction In the energy-to-power conversion industry, the thermal efficiency becomes uneconomically low when the gaseous steam temperature drops below 700 F (371 °C), the bulk of the energy loses and systems using water as working fluid become significantly less efficient and capital cost increase [1]. To overcome this disadvantage and still exploit the basic Rankine cycle technology developed over the years, systems based on working fluids such as hydrocarbons or refrigerants are being developed and researched in recent years. This is so called organic Rankine cycles (ORCs). The organic fluids is most promising for Rankine cycle power generation systems in that they can use low grade heat from a variety origins, such as geothermal energy [2,3], solar radiation [4], biomass combustion [5], waste heat from the industrial process [6–8]. It should be pointed out that, in the solar powered Rankine cycle research area, some important solar Rankine cycle systems have been developed and examined recent years, such as using freon R245fa [9] and natural fluid-CO2 [10–13] as working fluid. In the design of the cycle system, one important factor should be considered is the working fluid selection. Regarding the working fluids, several studies envisage the utilization of different working fluids that are more appropriate for low temperature heat source than tradition steam power cycles, such as, benzene * Corresponding author. Tel.: +86 22 81590264; fax: +86 22 27404188. E-mail address: [email protected] (L. Zhao).
[14], n-pentane [15], isobutane [16], and some cryogens like the freons R11, R113, R114, Rl23 and R134a [17–20]. It can be seen that the previous studies of ORCs regarded more about single component organic fluids. However, an important limitation of pure fluids is the constant temperature evaporation which is not suitable for sensible heat sources such as waste heat. The mixtures have variable temperature during the phase change process, which could be used to reduce the mismatch of temperature profiles between heat transfer fluid and the evaporating or condensing working fluid mixtures. Thus, the system irreversibilities can be minimized. Another advantage of the mixture is that it could be acquired the fluids that have the same thermodynamic properties as the pure ones through different component mass friction. This could greatly extend the range of candidate working fluid selection for low temperature Rankine cycle. Presently, it can be noticed that there is a significantly arising interest in multi-component mixture research. Radermacher demonstrated the mutual influence of working fluid mixtures properties on Rankine cycle performance, and counter-flow heat exchangers are suggested in the system for the mixtures [21]. Angelino and Paliano compared n-pentane and mixture of n-butane and n-hexane (50%/50%) by simulating liquid geothermal resource for electricity generation. The results show that mixture yields 6.8% more electricity than n-pentane and 25% less air are used with potential benefits in both cooler frontal area and fan power consumption [22]. Aleksandra studied Rankine cycle with heat source temperature of 80–115 °C by using different working
0306-2619/$ - see front matter Crown Copyright Ó 2010 Published by Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2010.05.016
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Nomenclature A 0 h1, h2 I _ m Q_ E _P W _T W
collector aperture area, m2 enthalpy of different state point in T–S diagram, kJ/kg global solar radiation, W/m2 mass flow rate of working fluids, kg/s heat absorbed in the collector, kW power consumed by working fluid pump, kW power generated by expander, kW
fluids, both natural and synthetic as well as mixtures [23]. Wang and Zhao made a theoretical analysis of zeotropic mixtures R245fa/R152a used in low-temperature solar Rankine cycles [24]. However, since the research in this field is in a very early stage, very few experimental studies on low temperature Rankine cycle for power generation using zeotropic mixtures are reported. In order to analyze the thermal performance of low-temperature solar Rankine cycle system using zeotropic mixtures, the paper concentrates on relative experimental researches. An experimental prototype is constructed and tested, such as the one that this paper described. The flat-plate collector, which does not need a complex tracking system and thus has low cost, has been considered, and the experimental evaluation and cycle performances could be a tool for their further optimization.
gC gR gT gsys
collector efficiency, % Rankine cycle efficiency, % turbine efficiency, % overall efficiency, %
vaporizes and finally superheats at the collector outlet. High pressure vapor is directed to the expander which is coupled with a generator. The exhaust vapor of the expander passes through the condenser. Condensation and a small degree of sub-cooling occur. Liquid is then sucked by the pump and the cycle recommences.
2. Working fluid selection In ORCs applications, the choice of working fluid is important since the fluid must not only possess thermophysical properties that match the application but also have adequate chemical stability at the desired working temperature. Here, R245fa and R152a are chosen as the components of the mixture because they both have zero ODP (Ozone Depression Potential) and lower GWP (Global Warming Potential), which have less environmental impact. The property details of the compositions are shown in Table 1. Here, pure fluid R245fa is noted as M1 and the mass fraction of zeotropic mixtures M2 and M3 are 0.9/0.1 and 0.7/0.3, respectively. Under this mass fraction, the fluids have comparatively low critical pressure, which is acceptable in the experimental condition. The temperature–entropy diagrams of these fluids are shown in Fig. 1. In Fig. 1, pressures of 2.0 MPa and 0.2 Mpa for different fluids are drawn, respectively. It is clear that as the mass fraction of R152a increases the pressure level rises and at the same time it is desirable to have the high pressure resistant equipment. The system economic efficiency needs to be further revalued.
Fig. 1. T–S diagram of different working fluids.
3. System description 3.1. Experimental apparatuses Fig. 2 illustrates the basic solar Rankine cycle system. The thermodynamic process can be briefly described as follows: the liquid organic working fluid is compressed with a working fluid pump which forces the fluid through the flat-plate collector. Heat transfers from solar radiation to the liquid, and the fluid is preheated,
Fig. 2. Basic solar Rankine cycle for electricity generation.
Table 1 Property details of working fluids.
a
Fluid
Molecular weight (g/mol)
Normal boiling pointa (°C)
Critical pressure (MPa)
Critical temperature (°C)
ODP
GWP
Safety group
R245fa R152a
134.05 66.05
15.14 24.02
3.65 4.52
154.01 113.26
0 0
950 120
B1 A2
The normal boiling point is boiling point at a pressure of 1 atmosphere.
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3.2. Data acquisition system The system is instrumented comprehensively. Each set of measurements is taken in a time step of 1 min and the logged data are recorded as a function of time. The testing parameters include temperature, pressure, flow rate and solar radiation. Temperature T-type thermocouples with accuracy of ±0.5 °C are mounted in the system to provide temperature measurement. The testing points are shown in Fig. 3 and the ambient temperature is also recorded. Pressure transducers provide pressure values for collector outlet and throttling outlet. The accuracy of the transducer of points 1 and 2, as shown in Fig. 3, is ±3 kPa and ±1 kPa, respectively. Solar radiation to measure the global solar radiation, a pyranometer with accuracy less than ±0.05 W is installed on the 45° slop surface which has the same angle of collector. Fig. 3. Experimental layout of the low-temperature solar Rankine cycle.
The experimental prototype mainly consists of two flat plate solar collector, a throttling valve, an air-cooled condenser, a storage tank, a fluid pump, a radiometer and date acquisition system. Fig. 3 shows the layout of the experimental system. Because the power output of the experimental prototype is too small, there is no real expander available for this system. Therefore, a throttling valve is used in the system to expand the vapor instead of a real expander. By adjusting the valve opening state, the pressure difference of the realistic expander can be simulated accordingly. Obviously, the fluid temperature and its thermodynamic and transport properties and states at the valve outlet are different from those of the true expander condition. However, the cycle performance for the power production can be derived based on thermodynamic analysis. As the fluid passing through the valve is a throttling process, the power generated from the expander can be calculated by setting the expander efficiency as 0.85 [25]. During the experiment, only one flat-plate collector, with its internal tube tank diameter of 12 mm, is used. The collector aperture area is 0.6 m2. It possesses one glass cover sheet and a high efficiency solar absorber plate whose absorptivity is 0.95 and emissivity is 0.17. Fig. 4 presents a photo of the experimental prototype.
The output signals of these instruments are connected to a PC through an Agilent 34980A series data logger and data are recorded as functions of time simultaneously. A diaphragm metering pump, which could be regulated in the range of 0–100% of rated flow, is used in the prototype to feed the fluids and control volume flow rate. 4. Assumption and calculation method The paper presented is the results of calculations regarding the Rankine cycle, which is shown in Fig. 3. Thermodynamic parameters of working fluids have been taken from the computer program of REFPROP8.0 [26], which is developed by the National Institute of Standards and Technology. Some additional parameters must be set to perform the calculation in this case: 1. Each component is considered as a steady-state-flow system. 2. The kinetic and potential energies as well as friction losses are neglected. 3. The expander isentropic efficiency gT is set at 0.85. 4. Furthermore, composition shift is neglected in the experimental system. 5. No pressures drop in the heat exchangers and pipelines have been considered. For the conditions, the cycle T–S diagram of pure fluid R245fa and zeotropic mixtures are illustrated in Figs. 5 and 6, respectively.
Fig. 4. A photo of the experimental prototype.
Fig. 5. Schematic diagram for R245fa Rankine cycle.
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valve open state was also kept in the same open state without being adjusted. 5. Results and discussion
Fig. 6. Schematic diagram for zeotropic mixtures Rankine cycle.
Based on the above measured parameters, it is possible to get a set of calculated values for pure fluid R245fa and zeotropic mixtures, and thus obtain a global understanding of the system behavior and performance. These values include:
_ T ¼g m _ ðh1 h20 Þ Power generated by the expander W T
ð1Þ
_P Power consumed by the working fluid pump W _ ðh5 h40 Þ ¼m
ð2Þ
_ ðh1 h5 Þ Heat absorbed in the collector Q_ E ¼ m Collector efficiency gC ¼
Q_ E IA
Rankine cycle efficiency gR ¼
ð3Þ ð4Þ
_P _ T W W _ QE
Overall efficiency gsys ¼ gC gR ¼
_ T W _P W IA
ð5Þ
ð6Þ
where h1, h20 , h40 and h5 are the enthalpies of corresponding state _ is the fluid mass flow rate of the system, points in T–S diagram. m which could be converted from the experimental volume flow rate. And, I and A represent the global solar radiation on the collectors’ inclination surface and collector aperture area, respectively. It should be pointed out that the overall efficiency gsys in Eq. (6) is the ratio of the net power output to the solar radiation on the collector’s surface. In order to evaluate the cycle performances of different working fluid, the volume flow rate was set as constant of 1.3 L/h and the
Fig. 7. Solar radiation during the experiment.
The experimental procedure was conducted in spring (April 2009) in the city of Tianjin, China. Fig. 7 shows the solar radiation for the 1st, 6th and 16th of April 2009, respectively. It can be seen that the solar radiation of 1st and 6th are almost identical, and the maximum value are both 1055 W/m2 which appeared at 11:48 a.m. and 12:02 a.m., respectively. Affected by the air quality, the overall solar radiation of 16th is slightly lower and the maximum value is appeared at 12:02 a.m. with value of 1001 W/m2. Compared with 1st and 6th, the solar radiation of 16th does not result to an extreme change of system operation and this radiation deviation is almost insignificant, which could be neglected. Fig. 8 shows the collector inlet and outlet temperature of different working fluids during the experiment. Since the process of phase change occurs in the collector, the collector outlet temperature is greatly influenced by solar radiation and ambient temperature. Thus, the outlet temperatures fluctuate during the experiment. However, under the daily constant flow rate, the overall trend of the outlet temperatures of different fluids follows that of the solar radiation, and the maximum outlet temperatures are 105.9 °C (M1, 12:56 a.m.), 99.86 °C (M2, 12:55 a.m.) and 101.56 °C (M3, 13:03 p.m.), respectively. During the experiment, an interesting observation concerning the behavior of the system is that compared with the time of maximum value of solar radiation, a hysteresis of the fluids maximum outlet temperature is observed. It could be noticed that the time lag is about 1 h. The identified time lag of hysteresis is directly related to the system thermal inertia of solar collector. The same phenomenon could be found in the following operation in the afternoon, that is, when the solar radiation descends from the maximum, collector outlet temperature exhibit no trend of decreasing but maintain the same states of high temperature. It is because that the collector internal tube tank and thermal insulation material act as a thermal storage during the experiment and heat is stored when the solar radiation is sufficient. Though there is collector heat loss, the working fluids could still get the sufficient heat for fluids evaporating and superheating and the collector outlet temperatures could achieve high values. When the stored heat and solar radiation are lower than collector heat loss, the outlet temperatures begin to drop, so as the degree of superheating. Therefore, the time of outlet temperature beginning to decrease is also behind that of solar radiation by 1 h. According to Manolakos [4], the thermal inertia of the system makes the operation more stable and smoothly reacted to the variation of solar radiation.
Fig. 8. Collector inlet and outlet temperature.
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In Fig. 8, as the collector inlet temperature of M1 is comparatively higher during the experiment, the outlet temperature of M1 is higher than M2 under the approximately similar solar radiation. Though the solar radiation of 16th April has small deviation compared with 1st and 6th, the outlet temperature of M1 still reach high value. Fig. 9 shows the pressure of different working fluid in the experimental monitoring point. As mentioned above, the pressure level of the mixtures increases with the adding of R152a. Obviously, M3 has the highest pressure level and M1 has the lowest. Influenced by the collector outlet temperature fluctuation, the collector outlet pressures of three fluids show the similar variation. It should be noted that, for zeotropic mixtures M2 and M3, the composition shift, which probably leads to composition change during the phase change process, results in collector outlet pressure fluctuating acutely during the operation. The mechanism of this phenomenon needs further study. Since the air-cooled condenser is used in the prototype, the pressures after the valve have the relationship to ambient temperature and thus behave smooth trend. The ambient temperature, together with the condenser inlet and outlet temperature, is illustrated in Fig. 10. Condenser inlet temperature, namely the valve outlet temperature, has the great relationship to the collector outlet temperature as the fluid passing the valve is a process of throttling. At the same time, affected by the ambient temperature the condenser outlet temperatures increase with the ambient temperature ascending. The condenser outlet temperatures of three working fluids are all higher than ambient temperature about 2 or 3 °C on average. The ambient temperature of Fig. 10a is higher than that of Fig. 10b and c, which leads to a higher collector inlet temperature of M1 in Fig. 8. Fig. 11 shows different degree of superheating at collector outlet. It can be seen that the collector outlet states of M1 are superheating in the whole experimental process. However, the outlet states of M2 and M3 could be divided into three stages: liquid–vapor phase at the beginning, superheating in the steady state, liquid–vapor phase at the end of the experiment. This fact could be explained that, compared with M2 and M3, M1 has the higher collector inlet temperature and lower latent heat during evaporation, which brings on the large degree of superheating. As the working fluid phase change process plays an important role in the whole cycle, when incident radiation is lower, heat absorption in the collector will decrease and therefore boiling is incomplete. This will have a negative effect on the system operation, which could be seen in the following analysis. In order to illustrate the superheating of different fluid under same constant flow rate, Table 2 presents the degree of superheating of different fluids during the experiment. Pure working fluid M1
Fig. 10. Ambient temperature and condenser inlet and outlet temperature.
Fig. 9. Collector outlet pressure and after valve pressure.
has the maximum superheating of 54.89 °C which is higher than M2 and M3 of 7.06 °C and 13.87 °C, respectively. Accordingly, M1 has the maximum average degree of superheating. This result in collector heat loss increases and collector efficiency decreases. Because a high degree of superheating in the expander inlet would not improve Rankine cycle efficiency [8], the Rankine cycle efficiency of M1 does not show any special improvement. It deserves to mention that an optimized regulation of fluid flow rate is required in order to minimize the superheating heat loss, which in turn increases the collector efficiency. The heat absorbed by the working fluids in the collector could be calculated from Eq. (3). Enthalpy difference, together with heat absorption in the flat-plate collector, is illustrated in Fig. 12. During the period of 8:34–9:24 (6th April) and 8:34–9:41 (16th April), as the fluid do not finish the phase change process in the collector, fluid M2 and M3 have the lower heat absorption, so as the enthalpy difference. From Fig. 12a the calculation results show that, in the
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Fig. 12. Specific enthalpy difference and heat absorbed by the fluid in the collector.
be seen that fluids heat absorption are almost identical during the superheating period. This can be explained that, with the component of R152a increasing, the mixtures have lower critical density and the mass flow rates are different correspondingly. According to the calculation, under the volume flow rate of 1.3 L/ h the mass flow rate of M1, M2 and M3 is 1.73 kg/s, 1.69 kg/s and 1.51 kg/s, respectively. Therefore, if the fluids were in the same mass flow rate, zeotropic mixtures will show a potential to improve the collector efficiency. Fig. 13 shows the collector efficiency during the experiment. As mentioned above, in the periods of 8:34–9:24 (6th April) and 8:34– 9:41 (16th April), M2 and M3 are in the liquid–vapor states and the collector efficiencies are lower as 4.19% and 12.27%, respectively. From 9:50 to 15:19, the fluids are all in the state of superheating Fig. 11. Superheating at the collector outlet.
Table 2 Degree of superheating during the experiment. Working fluid
M1
M2
M3
Time of superheating
8:39– 15:22 54.89
9:50– 15:40 47.83
9:40– 15:19 41.02
35.19
27.86
26.37
Maximum degree of superheating (°C) Average degree of superheating (°C)
superheating period of 9:50–15:19, the averaged specific enthalpy difference of M3 is 281.17 kJ/kg, which is higher than M2’s 258.31 kJ/kg and M1’s 248.97 kJ/kg. Therefore, under the similar solar radiation the zeotropic mixtures have large specific heat absorption. Fig. 12b shows the actual heat absorption of different working fluid under the same volume flow rate of 1.3 L/h. It can
Fig. 13. Collector efficiency.
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and the collector efficiency curves exhibit the shape of concave. The minimum point of the concave curve is the time of maximum solar radiation. It’s because that, when the solar radiation reaches its maximum value at noon, the heat loss of collector also reaches its maximal, and thus the collector efficiency shows the minimal value. When the solar radiation begin to fall, just as the system thermal inertia mentioned above, the attenuation of the collector outlet temperature lags behind that of the solar radiation and the collector efficiencies behave the trend of ascending. Integrating Figs. 7 and 12b, Fig. 13 shows that the collector efficiencies of three fluids are almost the same in the period of superheating. M3 shows slightly higher collector efficiency than M1 and M2 by 7.91% and 6.45%, respectively, as is presented in Table 3. In order to evaluate the cycle performance of different working fluid, it deserves to concentrate on the states of superheating from 9:50 to 15:19, which are the steady states of three working fluids. The following analyses are based on this period. The power produced from the expander is shown in the following Fig. 14. As zeotropic mixture M3 has the lower superheating degree at collector outlet, a comparatively higher power output is obtained at the maximum of 9.06 W. Affected by the fluctuation of collector outlet temperature and pressure, the power output of M2 varies from 4.69 W to 7.69 W. The average values could be found in Table 3. The average power output of M3 is higher than that of M1 and M2 by 29.10% and 28.03%, respectively. It can be seen that zeotropic mixtures have the great ability for power output, and the system capacity adjustment could be easily realized under different composition. The Rankine cycle efficiencies and the overall efficiencies can be calculated from Eqs. (5) and (6). Table 3 summarizes a set of characteristic value of cycle performance for different working fluids,
Table 3 Comparison of different fluid cycle performance. Working fluid
M1 (April 1st)
M2 (April 6th)
M3 (April 16th)
Period of superheating
9:50– 15:19 118.87
9:50– 15:19 120.17
9:50– 15:19 118.86
1.04
0.88
1.07
5.98 21.25 4.16 0.88
6.03 21.54 4.29 0.92
7.72 22.93 5.59 1.28
Average heat absorbed in the collector, W Average power consumed by working fluid pump, W Average power output, W Average collector efficiency, % Average Rankine cycle efficiency, % Average overall efficiency, %
which is useful to get an idea of the overall performance of the system. The averaged Rankine cycle efficiency of M3 is the highest of 5.59%. On the contrary, M1 has the lowest of 4.16%. The system overall efficiency has a great relationship to collector efficiency and Rankine cycle efficiency, so the average overall efficiency of M1, M2 and M3 is 0.88%, 0.92% and 1.28%, respectively. These low values, in some extent, are in close relation to the lack of regulation of flow rate of working fluids. As the zeotropic mixtures have the large amount of condensation heat, measures could be taken to recover part of condensation heat. For example, an external heat exchanger could be added between the expander and condenser to recover low temperature heat which could be used for domestic hot water supply. Due to the temperature glide of zeotropic mixtures during condensation, they would get the hot water whose outlet temperatures are higher than that of pure fluids. Therefore, the overall efficiency of zeotropic mixtures can be improved.
6. Conclusion An investigation on low-temperature solar Rankine cycle performance has been performed based on an experimental prototype by using zeotropic mixtures and pure fluid R245fa. Based on the constant flow rate of the fluid, the current work mentioned above is largely dedicated to make a system comparison between fluids M1, M2 and M3 to identify the cycle performance. The main results of the experimental measurements can be extracted as follow: The thermal inertia of the collector leads to the maximum outlet temperature lagged behind that of the solar radiation about 1 h. Under sufficient solar radiation to evaporate the working fluids in the collector, the flow rate is an important factor affecting the cycle performance. In the following procedure, a further optimization of flow rate regulation is needed. Due to the larger latent heat of evaporation, fluid M3 has large specific enthalpy difference in the collector, which has a great potential to improve the collector efficiency comparing with the pure R245fa. In the experimental superheating period, the average power output of M3 is higher than that of M1 and M2 by 29.10% and 28.03%, respectively. It can be seen that the power output varies accordingly and the system capacity adjustment could be easily realized under different composition of zeotropic mixtures. The average overall efficiency of M1, M2 and M3 is 0.88%, 0.92% and 1.28%, respectively. For zeotropic mixtures, there is a great potential to improve the overall efficiency by introducing an external heat exchanger to recover the partial condensation heat. Acknowledgement The authors would like to acknowledge the financial support provided by the Program for New Century Excellent Talents in University.
Appendix A. Uncertainty analysis
Fig. 14. Power output from the expander.
In every measured parameter there is an error between the measured and real value. Therefore, in order to evaluate the experiment data thoroughly and more reliable, an uncertainty analyses is necessary. In the following Table A there are the measured instruments and their uncertainty. Table B shows the total experimental uncertainty for every calculated value derived from measured data.
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Table A. Measured parameters and uncertainty. Instruments
Parameters
Range
Uncertainty (%)
Pyranometer
Solar radiation
0–2000 (W/m2)
5.0
Working fluid pump
Volume flow rate
0–6.5 (L/ h)
5.0
Thermocouples
Collector inlet temperature Collector outlet temperature Condenser inlet temperature Condenser outlet temperature Pump inlet temperature Pump outlet temperature
10– 100 °C 10– 100 °C 10– 100 °C 10– 100 °C 10– 100 °C 10– 100 °C
1.5
Collector outlet pressure Throttling outlet pressure
0– 1.5 MPa 0– 0.5 MPa
1.0
Pressure transducer
2.0 2.0 1.5 1.5 1.5
0.7
Table B. Uncertainty of every variable. Calculated variable
Total experimental uncertainty (%)
Power generated by the expander Power consumed by the working fluid pump Heat absorbed in the collector Collector efficiency Rankine cycle efficiency Overall efficiency
5.72 5.56 5.68 7.57 6.59 8.27
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