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Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid
Energy xxx (2015) 1e8
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Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid E. Galloni*, G. Fontana, S. Staccone DICeM, University of Cassino and Southern Latium, Italy
a r t i c l e i n f o
a b s t r a c t
Article history: Received 11 November 2014 Received in revised form 18 June 2015 Accepted 24 July 2015 Available online xxx
ORC (organic Rankine cycles) represent a sound solution for the exploitation of thermal energy available at low temperature. The prototype of a small ORC power plant has been realized at the Energy Systems Laboratory of Cassino University. In this paper, the plant design, the experimental methodology and the thermodynamic analysis of the work cycle have been illustrated. The aim of the work is to assess the feasibility of small-scale ORC plants. The basic idea is to analyze the performance of a small ORC plant able to exploit low-temperature heat sources. Thus, a simple organic Rankine cycle has been analyzed and R245fa as working fluid has been selected. Due to the small working fluid flow rates, a volumetric machine, in particular a scroll expander, has been chosen for mechanical power generation. The hot source temperature has been varied in the range 75e95 C and the cold sink temperature ranged between 20 C and 33 C. The R245fa vapor maximum pressure varied from 6 up to 10 bar. In this operating range, the best obtained results were: electric power equal to 1.2 kW, specific work about 20 kJ/kg and cycle efficiency slightly greater than 9 percent. © 2015 Elsevier Ltd. All rights reserved.
Keywords: Organic Rankine cycle Scroll expander ORC test rig Heat recovery
1. Introduction In recent years, the exploitation of low-temperature heat sources for power generation is, from a thermodynamic point of view, a more and more investigated issue. A lot of attention has been paid to the waste heat of such internal combustion power plants as reciprocating engines or gas turbines. However, in these examples, the heat recovery systems can be based on waste gases whose temperature is still relatively high with respect to the ambient temperature, ranging from about 250 C (micro gas turbines), 300e450 C (internal combustion engines) up to over 500 C (large gas turbines). If the exploitation of quite lower temperature thermal energy has to be considered - for instance, solar and geothermal sources or simply the waste heat of industrial processes - the development of suitable technologies for the energy conversion becomes a crucial point in the scenario of power generation.
* Corresponding author. Tel.: þ39 776 2994005; fax: þ39 776 2994365. E-mail address: [email protected] (E. Galloni).
This paper investigates the potential of a small-scale system (1e3 kW) able to exploit low-temperature thermal energy (approximately 100 C). When very small power plants are considered, as domestic plants delivering less than 10 kW, the necessity to propose sound solutions in the conversion of thermal to mechanical energy is particularly felt. Different technologies can be used for low-temperature heat conversion [1]. Stirling engines, ORC (organic Rankine cycles), thermo-electrical SeebeckePeltier systems and the NIFTE (NonInertive-Feedback Thermofluidic Engine) are compared in Fig. 1. In the external combustion engines (Stirling engines) the strong irreversibilities, due to the intense heat transfer processes, limit the actual conversion efficiency to values generally lower than 30%. Besides, these values can be reached if heat is provided at a temperature range between 600 and 800 C. Naturally, the conversion efficiency dramatically decreases if thermal energy is available at lower temperatures. Furthermore, small size engines unlikely reach efficiency values greater than 15 percent. Thermoelectric systems, based on the SeebeckePeltier effect in semiconductors, directly convert thermal energy into electrical energy. They need a temperature difference between the hot and
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Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
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Fig. 1. Technologies for low-temperature heat recovery.
the cold source lower than 100 C. At present, prototypes reach a few Watt power with efficiency levels lower than 5 percent. If the temperature difference between the heat source and the ambient is as low as 50 C, a new, interesting conversion technology is represented by the NIFTE (Non-Inertive-Feedback Thermofluidic Engine) which is a two-phase thermofluidic oscillator. Naturally, low-grade heat utilization means very low efficiency levels. However, this could be a negligible feature if the energy source cost is very low (or zero). Finally, the lack of mechanical moving parts is an undoubted advantage of this engine [2e4]. When thermal energy is available at temperature below 200 C or less, the most common working fluids used in thermal power plants become no more suitable for the energy conversion. Lowtemperature, high-pressure boiling fluids are more appropriate in exploiting low-grade heat sources. In this field, organic fluids can effectively be utilized giving life to the so-called ORC (organic Rankine cycles), i.e. vapor pressure plants using a cryogenic fluid instead of water [5]. Disregarding the supercritical cycles, vapor power cycles can be superheated or not, with or without internal heat exchangers (recuperated or un-recuperated cycles). Naturally, the cycle performances depend on the thermodynamic characteristics of the different working fluids (Fig. 2). Main properties of some organic fluids are reported in Table 1. They are classified according to the shape of the saturated vapor line in the temperature-entropy diagram. A dry fluid features a positive slope, while a wet fluid has a negative slope. Dry fluids are preferable because they do not condensate during the expansion phase. A lot of theoretical studies have been carried out in order to analyze the performances of organic Rankine cycles. Results show that, for a given fluid, both thermal efficiency and specific work rise with the maximum pressure cycle [5,6]. Studying the behavior of the organic fluids reported in Table 1, Mago et al. [7] have shown that these organic fluids do not need to be super-heated in order to increase the cycle thermal efficiency. In fact, at a given evaporation pressure, the cycle efficiency remains
approximately constant (wet fluids) or slightly decreases (dry fluids) with the increment of the turbine inlet temperature. Exploiting low-temperature thermal sources, Saleh et al. [6] have confirmed that dry fluids show a decrease of thermal efficiency with superheating, while, for some wet fluids, combining internal heat exchange and superheating could be advantageous. The internal heat exchange allows an efficiency increase of 1e3 percent with respect to the un-recuperated cycle, while the specific work does not significantly change [1]. The characteristics of working fluid greatly affect the cycle performances. The most common working fluids used for small scale low-temperature plants are R134a and R245fa. In subcritical cycles, R134a can work under 100 C. It allows achieving a thermal efficiency less than 10% delivering a specific work of about 15 kJ/kg. R245fa fluid is able to work up to 150 C. Assuming an evaporation temperature equal to 120 C, it is possible to reach a thermal efficiency of about 12% obtaining an output specific work of about 30 kJ/kg [1,6]. At present, these thermal power plants cover a wide power range, from 10 up to about 3000 kW [8]. Using a temperature drop often smaller than 100 , the useful work per mass unit of working fluid is generally small in ORC plants. For this reason, in order to not increase too much the mass flow rates thus the plant sizes, a few dozen kW power seems a reasonable size for ORC plants. This feature makes them attractive in the scenario of distributed power generation when low-temperature thermal energy can be exploited. In particular, the development of mini ORC technologies, suitable for residential applications (a few kW), has received more and more attention over the last years [9e13], but the achievement of satisfactory efficiencies is still a challenge [14e17]. In fact, the organic Rankine cycle efficiency strongly depends on the used expander. Actually, positive displacement expanders are suitable for small size plants; unfortunately, these small-size expanders are often characterized by poor levels of expansion efficiency (less than 75%) that penalize the overall energy conversion efficiency [8]. 2. Thermodynamic analysis and mini-ORC plant design The general purpose of this paper is to give a contribution in the optimization of small-scale organic Rankine cycle plants for lowgrade heat recovery. An experimental methodology has been adopted in order to evaluate the performance of such a plant exploiting thermal energy available at a temperature level somewhat less than 100 C. In particular, the objective of the paper is to illustrate the development of a mini-ORC plant prototype. The demanded power target is about 1 kW, while the thermal energy is available at about 100 C or less. As mentioned above, recuperated organic Rankine cycles show a little efficiency gain compared to un-recuperated cycles. As a consequence, for an output size of a few kW, it seems inconvenient to complicate the plant by adding an internal heat exchanger, so the simple lay-out, shown in Fig. 3, has been considered. The choice of working fluid is a crucial point in an ORC plant design. The working fluid should be environment friendly, not aggressive, not flammable and not too onerous for the power plant economy. Naturally, the choice is conditioned by the temperature of the available heat source. Considering an evaporation temperature equal to 100 C and a condensation temperature equal to 30 C, thermodynamic analyses reported in Ref. [6] show that few fluids allow obtaining overall efficiency higher than 10 percent. Among these, flammable fluids (hexane as example) have been discarded for safety criteria. Some other fluids have been disregarded in order to avoid pressure levels at the condenser beneath the atmospheric pressure (as an example,
Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
E. Galloni et al. / Energy xxx (2015) 1e8
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Fig. 2. Rankine and superheated vapor cycles for water (non organic wet fluid, left) and R245fa (organic dry fluid, right).
Table 1 Comparison of water and organic fluid properties. Fluid
Type
Molecular weight (kg/kmol)
Boiling point ( C) at atmospheric pressure
Critical pressure (bar)
Critical temperature ( C)
Water R134a R245fa R245ca Propane Isobutane
wet wet dry dry wet dry
18.015 102.03 134.05 134.05 44.10 58.12
99.97 26.07 14.90 25.13 42.09 11.61
220.64 40.59 42.5 39.25 42.48 36.4
373.95 101.06 154.05 174.42 96.70 134.7
the R601 fluid) so preventing air entrance into the process. At the end, R245fa has been chosen because it is safe, wide available, and allows reaching cycle efficiencies higher than 10% when not very high levels (about 10 bar) for cycle maximum pressures are adopted. Fig. 4 shows the predicted performances of a Rankine cycle operating with R245fa. Thermodynamic data of the working fluid have been calculated by means of the CoolProp libraries [18]. A standard thermodynamic analysis has been carried out. No pressure drops have been considered both in the evaporator and in
the condenser. Energy losses through the pump have been neglected, while numerous values for both the vapor quality at the expander inlet and the expander isentropic efficiency have been considered. The system efficiency has been calculated as:
hth ¼
_p _ eW W _ Q
(1)
h
while the specific work output is:
w¼
Fig. 3. Mini ORC lay-out.
_p _ eW W m_
(2)
As expected, the system efficiency increases with both the reduction of the condenser pressure and the increment of the evaporation pressure. As an example, assuming equal to 1.0 the expander efficiency (i.e. analyzing the ideal cycle behavior) and keeping the turbine inlet at dry saturated conditions (Q3 ¼ 1), when the evaporation pressure is equal to 10 bar (i.e. T3 ¼ 89.7 C) and the condenser pressure is equal to 1.8 bar (i.e. T1 ¼ 30.3 C), the ideal efficiency is equal to 13.6%, and the ideal specific work is 30.9 kJ/kg. The vapor quality at the turbine inlet slightly affects the system efficiency. For the above mentioned case, if Q3 ¼ 0.9 the efficiency decreases by 0.8%, while if Q3 ¼ 0.8 the efficiency decreases by 2.4%. On the contrary, the vapor quality greatly affects the ideal specific work: if Q3 ¼ 0.9 the specific work decreases by 7.1%, while if Q3 ¼ 0.8 it decreases by 14.8%.
Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
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E. Galloni et al. / Energy xxx (2015) 1e8
Fig. 4. ORC thermal efficiency (left) and specific work (right) versus evaporation pressure at different condenser pressures (p1), vapor qualities at the turbine inlet (Q3) and expander efficiencies (hh,exp).
Finally, the expander efficiency (hh,exp) greatly affects both the thermal efficiency and the specific work. For the reference case, if hh,exp ¼ 0.8 both the efficiency and the specific work decrease by 20%, while if hh,exp ¼ 0.6 they decrease by 40%. The ORC plant (Fig. 5) has been designed basing on the results provided by the preliminary analysis mentioned above. Two liquid to liquid heat exchangers have been used for both heating the liquid up to evaporation and cooling the vapor leaving the expander in order to condense it. The condensate is collected in a tank arranged below the condenser. A rotative vane pump moves the fluid through the pipeline; at the discharge side, it mounts an internal safety relief valve that retains the fluid pressure below 10 bar. The pump is supplied by the tank mentioned above. In order to avoid the liquid evaporation, so cavitation phenomena, the pump has been positioned just below the tank. Copper lines have been used to link the main components: flow sections have been calculated in order to have very low friction losses. Due to the small mass flow rates involved in this plant, for the expansion phase a volumetric machine has been chosen. In particular, the E15H22N4.25L Air Squared scroll expander [19,20] able to operate with several types of refrigerant gases, has been selected. Its volumetric expansion ratio is equal to 3.5 and the displacement is 12 cc/rev. This scroll expander is rated up to 1 kW, based on a maximum inlet pressure of 13.5 bar; running at 3000 rpm, the Air Squared Manufacturing states an expansion efficiency slightly less than 0.8when it operates with lubricated fluid. Currently, the expander directly drives a two pole magnetogenerator. Obviously, this is not the optimal generator for this kind of applications (it was just the cheapest available) and its efficiency is particularly low. Consequently, it has been waived to measure the actual electrical power delivered during the test.
Fig. 5. The mini ORC prototype. 1: evaporator; 2: scroll expander; 3: condenser.
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The proposed prototype system requires high initial investment costs. The total cost is about 5000 Euros, composed mainly by 3000 EUR for the expander and 1000 EUR for the heat exchangers. The high cost is due to the purchase of the elements with spot orders from retailers. The amount could be significantly reduced in the case of a mass production. It is worth noting that a refrigerator pump, more or less of the same size of the proposed expander, is priced at 300e400 Euros. The pilot plant features just 1 kW. It makes sense to think of producing plants at least for the sizes: 3e5e10 kW. Therefore it is reasonable to expect that the cost could become about 1500 EUR/ kW. In the following, we will consider the actual work that the fluid delivers to the expander scroll in order to calculate the expander efficiency and to evaluate the cycle thermal efficiency. In other terms, these efficiencies do not account for energy losses occurring between the expander rotor and the alternator output shaft. 3. Experimental setup In order to test the mini-ORC plant, a test rig has been set up. The heat carrier is water coming from a purpose-built boiler. This boiler is able to provide up to 11 kW to the water flow; the heat output can be adjusted to control the temperature of the water entering the hot side of the evaporator in the range 70e100 C. The condenser coolant is cold water coming from a heat rejection device. Fig. 6 shows a schematic sketch of the test rig. At inlet/outlet of main components the state point of R245fa is determined by means of both temperature and pressure measurements. At the pump inlet, fluid in the saturated liquid phase has been assumed, so, in this case, only pressure values have been measured. This hypothesis seems to be reasonable because the pump sucks the fluid from a tank which is arranged just below the condenser and collects the condensate coming out from this heat exchanger. At the expander inlet, fluid could be in the wet region. Actually, during some tests the measured temperature T3 has been found equal to the saturation temperature corresponding to the measured pressure P3. In this case, the vapor quality has been calculated by means of the following relationship:
Q3 ¼
h3 h3l Dhvap
5
(3)
where, referring to the measured temperature T3, h3l is the enthalpy of the saturated liquid, Dhvap is the latent heat of vaporization, and:
h3 ¼
Q_ h m_ $cw $ðTwhi Twho Þ þ h1 ¼ wh þ h1 _ m m_
(4)
Temperatures have been measured by means of PT100 class B thermometers, while four pressure transducers have been flushmounted on the refrigerant lines. Low pressure measurements (P1 and P4 transducers) have an inaccuracy equal to 0.012 bar while high pressure measurements (P2 and P3 transducers) have an inaccuracy equal to 0.12 bar. The water mass flow rates through the heat exchangers have been measured by means of Adafruit 833 liquid flow meters (inaccuracy less than 0.015 kg/s). R245fa mass flow rate is not directly measured. It is determined via the energy balance at the condenser. Under steady-state conditions:
m_ ¼
Q_ c m_ wc $cw $ðTwco Twci Þ ¼ h4 h1 h4 h1
(5)
where cw is the water specific heat. In theory, the system is quite simple because it consists of a few elements. However, the complexity lies in sizing and properly placing each part in order to obtain the maximum efficiency. About mounting, the most complex part concerns the correct positioning of the sensors for correctly detecting the physical quantities of interest. 4. Results Several tests have been conducted at different hot source and cold sink temperatures. The hot source temperature has been varied from 75 C to 95 C, while, according to the environment conditions, the cold sink temperature ranged between 20 C and 33 C.
Fig. 6. Test bench sketch.
Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
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During the tests, both the hot and the cold water flow rates have been kept approximately constant. The water drop temperature was about 10 C in both the heat exchangers. Due to its simplicity, the system is easy to handle during both the start-up phase and the steady operation. The start-up phase takes only a few minutes. The temperature of the hot fluid determines the evaporation temperature of the refrigerant, and then the back pressure acting on the pump. Measurements have shown that: - Wet vapor enters the expander. Depending on the heat exchanger efficiency, R245fa inlet temperatures range from 90 C to 71 C. Correspondingly, the vapor pressure varies from 10 bar to 6.1 bar, while the vapor quality is in the range 0.871e0.939. - Superheated vapor leaves the expander. Its temperature ranges from 31.3 C to 43.0 C. - The exhaust vapor condenses through the heat exchanger. The condensate is blown into the tank where its pressure ranges from 1.5 bar to 2.4 bar. The corresponding saturation temperature ranges from about 25.9 C to 39.0 C. - Sub-cooled liquid fluid leaves the pump at the maximum pressure described above; its temperature ranges from 27.6 C to 41.1 C. Fig. 7 shows both the thermal efficiency and the specific work calculated as:
hth
h h4 ðh2 h1 Þ ¼ 3 h3 h2
w ¼ h3 h4 ðh2 h1 Þ
(6) (7)
while Figs. 8 and 9 show plant input and output power calculated as:
_ ¼ m$Dh _ W
(8)
Fig. 8. Mini ORC work rates at different test conditions.
about 10%, while the specific work is within the range 9e20 kJ/kg. The expander output power ranges from about 0.4 to about 1.2 kW. The maximum output power is obtained when the heat rate entering the system is slightly less than 11 kW. Naturally, also the cold source temperature influences the cycle performances causing the data dispersion highlighted both in Figs. 7 and 8. With regard to this variable, Fig. 10 shows the benefit obtained by reducing the coolant's temperature.
It is worth noting that cycle performances increase almost linearly with the maximum cycle pressure. At the test conditions mentioned above, the thermal efficiency ranges from about 5% to
Fig. 7. Mini ORC thermal efficiencies and specific works at different test conditions.
Fig. 9. Mini ORC heat rates at different test conditions.
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Fig. 10. Influence of the condenser pressure on the Mini ORC performances.
Table 2 Test results, operating point characterized by the best thermal efficiency. Direct measured values. Water temperature at evaporator inlet
Twhi
[ C]
95.7
Water temperature at evaporator outlet Water temperature at condensator inlet Water temperature at condensator outlet Hot water mass flow rate Cold water mass flow rate R245fa pressure at scroll inlet R245fa temperature at scroll inlet R245fa pressure at scroll outlet R245fa temperature at scroll outlet R245fa pressure at pump outlet R245fa temperature at pump outlet R245fa pressure at pump inlet
Twho Twci Twco m_ wh m_ wc P3 T3 P4 T4 P2 T2 P1
[ C] [ C] [ C] [kg/s] [kg/s] [bar] [ C] [Bar] [ C] [bar] [ C] [Bar]
86.8 26.8 36.3 0.288 0.248 9.95 89.7 2.17 38.4 9.99 35.8 2.02
Table 2 and 3 describe the operating point characterized by the best thermal efficiency achieved during the tests, while Fig. 11 shows the corresponding cycle state points. The best value achieved for thermal efficiency was equal to 9.3 percent. In order to have a reference value, a Carnot cycle, evolving between the given hot source temperature (Twhi ¼ 95.7 C) and the cold sink temperature (Twci ¼ 26.8 C), can be assumed. In this case, the efficiency of the Carnot cycle would be equal to 18.7 percent. Table 3 Test results, operating point characterized by the best thermal efficiency. Calculated data. R245fa temperature at pump inlet
T1
[ C]
33.8
R245fa vapor quality at scroll inlet R245fa flow mass rate Specific work Evaporator heat rate Condensator heat rate Expander work rate Pump work rate Thermal efficiency Expander efficiency Pump efficiency Carnot cycle efficiency
Q3 m_ w Q_ h Q_ c _e W _p W
[e] [kg/s] [kJ/kg] [kW] [kW] [kW] [kW] [%] [%] [%] [%]
0.915 0.052 19.35 10.88 9.87 1.17 0.16 9.28 84.9 79.7 18.7
hth hh,exp hh,pum hCar
Fig. 11. R245fa state points relative to the test results shown in Table 2.
Saturated RF245a vapor leaves the evaporator. Its quality is equal to 0.915. As mentioned above, (Fig. 4) this parameter has a small influence on the cycle efficiency, but it affects the specific work provided by the power plant. Finally, the results show that both the scroll expander and the pump have a very good hydraulic efficiency, calculated comparing the actual work (i.e. the mechanical energy transferred to or from the rotor) to the isentropic ideal work [21]. 5. Conclusions In this paper, a simple and cheap small-scale ORC plant has been designed, realized and tested at numerous operative conditions. In the proposed study, an un-recuperated Rankine cycle has been adopted and a volumetric scroll expander, for the mechanical power generation, has been chosen. This expander featured an acceptable expansion efficiency and a very good volumetric efficiency. Furthermore, it showed silent and reliable operation. When thermal energy is introduced at a temperature less than 100 C, the mini-ORC plant is able to reach a thermal efficiency equal to about 9%, a specific work equal to about 20 kJ/kg and provides an output power somewhat greater than 1 kW. Authors want to precise that, at present, the prototype is still under development and the reported results come from partial investigations relative to the evolution of the working fluid. However, these results prove that mini-ORC plant is a promising technology in the few kW power range when low-temperature thermal energy is available as a primary energy source. In particular, it can be attractive both in the field of the waste heat recovery and in the scenario of distributed power generation for domestic usage. References [1] Bianchi M, De Pascale A. Bottoming cycles for electric energy generation: parametric investigation of available and innovative solutions for the exploitation of low and medium temperature heat sources. Appl Energy 2011;88:1500e9.
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[2] Smith TCB. Power dense thermofluidic oscillators for high load applications. In: IECEC 2004: proceedings of the 2nd international energy conversion engineering conference; 2004. p. 2004e5758. Providence (RI), USA. [3] Solanki R, Mathie R, Galindo A, Markides CN. Modelling of a two-phase thermofluidic oscillator for low-grade heat utilisation: accounting for irreversible thermal losses. Appl Energy 2013;106:337e54. [4] Solanki R, Galindo A, Markides CN. Dynamic modelling of a two-phase thermofluidic oscillator for efficient low grade heat utilization: effect of fluid inertia. Appl Energy 2012;89:156e63. [5] Delgado-Torres A, Garcia-Rodriguez L. Analysis and optimization of the lowtemperature solar Organic Rankine Cycle (ORC). Energy Convers Manag 2010;51:2846e56. [6] Saleh B, Koglbauer G, Wendland M, Fischer J. Working fluids for lowtemperature Organic Rankine Cycles. Energy 2007;32:1210e21. [7] Mago PJ, Chamra LM, Somayali C. Performance analysis of different working fluids for use in Organic Rankine Cycles. IMECHE J Power Energy 2007;221: 255e64. [8] Quoilin S, VanDenBroek M, Declaye S, Dewallef P, Lemort V. Techno-economic survey of Organic Rankine Cycle (ORC) systems. Renew Sustain Energy Rev 2013;22:168e86. [9] Dong L, Liu H, Riffat S. Development of small-scale and micro-scale biomassfuelled CHP systems e a literature review. Appl Therm Eng 2009;29. [10] Georges E, Declaye S, Dumont O, Quoilin S, Lemort V. Design of a small-scale Organic Rankine Cycle engine used in a solar power plant. Int J Low-Carbon Technol 2013;8:i34e41. [11] Bracco R, Clemente S, Micheli D, Reini M. Experimental tests and modelization of a domestic-scale ORC (Organic Rankine Cycle). Energy 2013;58: 107e16.
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Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid E. Galloni*, G. Fontana, S. Staccone DICeM, University of Cassino and Southern Latium, Italy
a r t i c l e i n f o
a b s t r a c t
Article history: Received 11 November 2014 Received in revised form 18 June 2015 Accepted 24 July 2015 Available online xxx
ORC (organic Rankine cycles) represent a sound solution for the exploitation of thermal energy available at low temperature. The prototype of a small ORC power plant has been realized at the Energy Systems Laboratory of Cassino University. In this paper, the plant design, the experimental methodology and the thermodynamic analysis of the work cycle have been illustrated. The aim of the work is to assess the feasibility of small-scale ORC plants. The basic idea is to analyze the performance of a small ORC plant able to exploit low-temperature heat sources. Thus, a simple organic Rankine cycle has been analyzed and R245fa as working fluid has been selected. Due to the small working fluid flow rates, a volumetric machine, in particular a scroll expander, has been chosen for mechanical power generation. The hot source temperature has been varied in the range 75e95 C and the cold sink temperature ranged between 20 C and 33 C. The R245fa vapor maximum pressure varied from 6 up to 10 bar. In this operating range, the best obtained results were: electric power equal to 1.2 kW, specific work about 20 kJ/kg and cycle efficiency slightly greater than 9 percent. © 2015 Elsevier Ltd. All rights reserved.
Keywords: Organic Rankine cycle Scroll expander ORC test rig Heat recovery
1. Introduction In recent years, the exploitation of low-temperature heat sources for power generation is, from a thermodynamic point of view, a more and more investigated issue. A lot of attention has been paid to the waste heat of such internal combustion power plants as reciprocating engines or gas turbines. However, in these examples, the heat recovery systems can be based on waste gases whose temperature is still relatively high with respect to the ambient temperature, ranging from about 250 C (micro gas turbines), 300e450 C (internal combustion engines) up to over 500 C (large gas turbines). If the exploitation of quite lower temperature thermal energy has to be considered - for instance, solar and geothermal sources or simply the waste heat of industrial processes - the development of suitable technologies for the energy conversion becomes a crucial point in the scenario of power generation.
* Corresponding author. Tel.: þ39 776 2994005; fax: þ39 776 2994365. E-mail address: [email protected] (E. Galloni).
This paper investigates the potential of a small-scale system (1e3 kW) able to exploit low-temperature thermal energy (approximately 100 C). When very small power plants are considered, as domestic plants delivering less than 10 kW, the necessity to propose sound solutions in the conversion of thermal to mechanical energy is particularly felt. Different technologies can be used for low-temperature heat conversion [1]. Stirling engines, ORC (organic Rankine cycles), thermo-electrical SeebeckePeltier systems and the NIFTE (NonInertive-Feedback Thermofluidic Engine) are compared in Fig. 1. In the external combustion engines (Stirling engines) the strong irreversibilities, due to the intense heat transfer processes, limit the actual conversion efficiency to values generally lower than 30%. Besides, these values can be reached if heat is provided at a temperature range between 600 and 800 C. Naturally, the conversion efficiency dramatically decreases if thermal energy is available at lower temperatures. Furthermore, small size engines unlikely reach efficiency values greater than 15 percent. Thermoelectric systems, based on the SeebeckePeltier effect in semiconductors, directly convert thermal energy into electrical energy. They need a temperature difference between the hot and
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Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
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Fig. 1. Technologies for low-temperature heat recovery.
the cold source lower than 100 C. At present, prototypes reach a few Watt power with efficiency levels lower than 5 percent. If the temperature difference between the heat source and the ambient is as low as 50 C, a new, interesting conversion technology is represented by the NIFTE (Non-Inertive-Feedback Thermofluidic Engine) which is a two-phase thermofluidic oscillator. Naturally, low-grade heat utilization means very low efficiency levels. However, this could be a negligible feature if the energy source cost is very low (or zero). Finally, the lack of mechanical moving parts is an undoubted advantage of this engine [2e4]. When thermal energy is available at temperature below 200 C or less, the most common working fluids used in thermal power plants become no more suitable for the energy conversion. Lowtemperature, high-pressure boiling fluids are more appropriate in exploiting low-grade heat sources. In this field, organic fluids can effectively be utilized giving life to the so-called ORC (organic Rankine cycles), i.e. vapor pressure plants using a cryogenic fluid instead of water [5]. Disregarding the supercritical cycles, vapor power cycles can be superheated or not, with or without internal heat exchangers (recuperated or un-recuperated cycles). Naturally, the cycle performances depend on the thermodynamic characteristics of the different working fluids (Fig. 2). Main properties of some organic fluids are reported in Table 1. They are classified according to the shape of the saturated vapor line in the temperature-entropy diagram. A dry fluid features a positive slope, while a wet fluid has a negative slope. Dry fluids are preferable because they do not condensate during the expansion phase. A lot of theoretical studies have been carried out in order to analyze the performances of organic Rankine cycles. Results show that, for a given fluid, both thermal efficiency and specific work rise with the maximum pressure cycle [5,6]. Studying the behavior of the organic fluids reported in Table 1, Mago et al. [7] have shown that these organic fluids do not need to be super-heated in order to increase the cycle thermal efficiency. In fact, at a given evaporation pressure, the cycle efficiency remains
approximately constant (wet fluids) or slightly decreases (dry fluids) with the increment of the turbine inlet temperature. Exploiting low-temperature thermal sources, Saleh et al. [6] have confirmed that dry fluids show a decrease of thermal efficiency with superheating, while, for some wet fluids, combining internal heat exchange and superheating could be advantageous. The internal heat exchange allows an efficiency increase of 1e3 percent with respect to the un-recuperated cycle, while the specific work does not significantly change [1]. The characteristics of working fluid greatly affect the cycle performances. The most common working fluids used for small scale low-temperature plants are R134a and R245fa. In subcritical cycles, R134a can work under 100 C. It allows achieving a thermal efficiency less than 10% delivering a specific work of about 15 kJ/kg. R245fa fluid is able to work up to 150 C. Assuming an evaporation temperature equal to 120 C, it is possible to reach a thermal efficiency of about 12% obtaining an output specific work of about 30 kJ/kg [1,6]. At present, these thermal power plants cover a wide power range, from 10 up to about 3000 kW [8]. Using a temperature drop often smaller than 100 , the useful work per mass unit of working fluid is generally small in ORC plants. For this reason, in order to not increase too much the mass flow rates thus the plant sizes, a few dozen kW power seems a reasonable size for ORC plants. This feature makes them attractive in the scenario of distributed power generation when low-temperature thermal energy can be exploited. In particular, the development of mini ORC technologies, suitable for residential applications (a few kW), has received more and more attention over the last years [9e13], but the achievement of satisfactory efficiencies is still a challenge [14e17]. In fact, the organic Rankine cycle efficiency strongly depends on the used expander. Actually, positive displacement expanders are suitable for small size plants; unfortunately, these small-size expanders are often characterized by poor levels of expansion efficiency (less than 75%) that penalize the overall energy conversion efficiency [8]. 2. Thermodynamic analysis and mini-ORC plant design The general purpose of this paper is to give a contribution in the optimization of small-scale organic Rankine cycle plants for lowgrade heat recovery. An experimental methodology has been adopted in order to evaluate the performance of such a plant exploiting thermal energy available at a temperature level somewhat less than 100 C. In particular, the objective of the paper is to illustrate the development of a mini-ORC plant prototype. The demanded power target is about 1 kW, while the thermal energy is available at about 100 C or less. As mentioned above, recuperated organic Rankine cycles show a little efficiency gain compared to un-recuperated cycles. As a consequence, for an output size of a few kW, it seems inconvenient to complicate the plant by adding an internal heat exchanger, so the simple lay-out, shown in Fig. 3, has been considered. The choice of working fluid is a crucial point in an ORC plant design. The working fluid should be environment friendly, not aggressive, not flammable and not too onerous for the power plant economy. Naturally, the choice is conditioned by the temperature of the available heat source. Considering an evaporation temperature equal to 100 C and a condensation temperature equal to 30 C, thermodynamic analyses reported in Ref. [6] show that few fluids allow obtaining overall efficiency higher than 10 percent. Among these, flammable fluids (hexane as example) have been discarded for safety criteria. Some other fluids have been disregarded in order to avoid pressure levels at the condenser beneath the atmospheric pressure (as an example,
Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
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Fig. 2. Rankine and superheated vapor cycles for water (non organic wet fluid, left) and R245fa (organic dry fluid, right).
Table 1 Comparison of water and organic fluid properties. Fluid
Type
Molecular weight (kg/kmol)
Boiling point ( C) at atmospheric pressure
Critical pressure (bar)
Critical temperature ( C)
Water R134a R245fa R245ca Propane Isobutane
wet wet dry dry wet dry
18.015 102.03 134.05 134.05 44.10 58.12
99.97 26.07 14.90 25.13 42.09 11.61
220.64 40.59 42.5 39.25 42.48 36.4
373.95 101.06 154.05 174.42 96.70 134.7
the R601 fluid) so preventing air entrance into the process. At the end, R245fa has been chosen because it is safe, wide available, and allows reaching cycle efficiencies higher than 10% when not very high levels (about 10 bar) for cycle maximum pressures are adopted. Fig. 4 shows the predicted performances of a Rankine cycle operating with R245fa. Thermodynamic data of the working fluid have been calculated by means of the CoolProp libraries [18]. A standard thermodynamic analysis has been carried out. No pressure drops have been considered both in the evaporator and in
the condenser. Energy losses through the pump have been neglected, while numerous values for both the vapor quality at the expander inlet and the expander isentropic efficiency have been considered. The system efficiency has been calculated as:
hth ¼
_p _ eW W _ Q
(1)
h
while the specific work output is:
w¼
Fig. 3. Mini ORC lay-out.
_p _ eW W m_
(2)
As expected, the system efficiency increases with both the reduction of the condenser pressure and the increment of the evaporation pressure. As an example, assuming equal to 1.0 the expander efficiency (i.e. analyzing the ideal cycle behavior) and keeping the turbine inlet at dry saturated conditions (Q3 ¼ 1), when the evaporation pressure is equal to 10 bar (i.e. T3 ¼ 89.7 C) and the condenser pressure is equal to 1.8 bar (i.e. T1 ¼ 30.3 C), the ideal efficiency is equal to 13.6%, and the ideal specific work is 30.9 kJ/kg. The vapor quality at the turbine inlet slightly affects the system efficiency. For the above mentioned case, if Q3 ¼ 0.9 the efficiency decreases by 0.8%, while if Q3 ¼ 0.8 the efficiency decreases by 2.4%. On the contrary, the vapor quality greatly affects the ideal specific work: if Q3 ¼ 0.9 the specific work decreases by 7.1%, while if Q3 ¼ 0.8 it decreases by 14.8%.
Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
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Fig. 4. ORC thermal efficiency (left) and specific work (right) versus evaporation pressure at different condenser pressures (p1), vapor qualities at the turbine inlet (Q3) and expander efficiencies (hh,exp).
Finally, the expander efficiency (hh,exp) greatly affects both the thermal efficiency and the specific work. For the reference case, if hh,exp ¼ 0.8 both the efficiency and the specific work decrease by 20%, while if hh,exp ¼ 0.6 they decrease by 40%. The ORC plant (Fig. 5) has been designed basing on the results provided by the preliminary analysis mentioned above. Two liquid to liquid heat exchangers have been used for both heating the liquid up to evaporation and cooling the vapor leaving the expander in order to condense it. The condensate is collected in a tank arranged below the condenser. A rotative vane pump moves the fluid through the pipeline; at the discharge side, it mounts an internal safety relief valve that retains the fluid pressure below 10 bar. The pump is supplied by the tank mentioned above. In order to avoid the liquid evaporation, so cavitation phenomena, the pump has been positioned just below the tank. Copper lines have been used to link the main components: flow sections have been calculated in order to have very low friction losses. Due to the small mass flow rates involved in this plant, for the expansion phase a volumetric machine has been chosen. In particular, the E15H22N4.25L Air Squared scroll expander [19,20] able to operate with several types of refrigerant gases, has been selected. Its volumetric expansion ratio is equal to 3.5 and the displacement is 12 cc/rev. This scroll expander is rated up to 1 kW, based on a maximum inlet pressure of 13.5 bar; running at 3000 rpm, the Air Squared Manufacturing states an expansion efficiency slightly less than 0.8when it operates with lubricated fluid. Currently, the expander directly drives a two pole magnetogenerator. Obviously, this is not the optimal generator for this kind of applications (it was just the cheapest available) and its efficiency is particularly low. Consequently, it has been waived to measure the actual electrical power delivered during the test.
Fig. 5. The mini ORC prototype. 1: evaporator; 2: scroll expander; 3: condenser.
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The proposed prototype system requires high initial investment costs. The total cost is about 5000 Euros, composed mainly by 3000 EUR for the expander and 1000 EUR for the heat exchangers. The high cost is due to the purchase of the elements with spot orders from retailers. The amount could be significantly reduced in the case of a mass production. It is worth noting that a refrigerator pump, more or less of the same size of the proposed expander, is priced at 300e400 Euros. The pilot plant features just 1 kW. It makes sense to think of producing plants at least for the sizes: 3e5e10 kW. Therefore it is reasonable to expect that the cost could become about 1500 EUR/ kW. In the following, we will consider the actual work that the fluid delivers to the expander scroll in order to calculate the expander efficiency and to evaluate the cycle thermal efficiency. In other terms, these efficiencies do not account for energy losses occurring between the expander rotor and the alternator output shaft. 3. Experimental setup In order to test the mini-ORC plant, a test rig has been set up. The heat carrier is water coming from a purpose-built boiler. This boiler is able to provide up to 11 kW to the water flow; the heat output can be adjusted to control the temperature of the water entering the hot side of the evaporator in the range 70e100 C. The condenser coolant is cold water coming from a heat rejection device. Fig. 6 shows a schematic sketch of the test rig. At inlet/outlet of main components the state point of R245fa is determined by means of both temperature and pressure measurements. At the pump inlet, fluid in the saturated liquid phase has been assumed, so, in this case, only pressure values have been measured. This hypothesis seems to be reasonable because the pump sucks the fluid from a tank which is arranged just below the condenser and collects the condensate coming out from this heat exchanger. At the expander inlet, fluid could be in the wet region. Actually, during some tests the measured temperature T3 has been found equal to the saturation temperature corresponding to the measured pressure P3. In this case, the vapor quality has been calculated by means of the following relationship:
Q3 ¼
h3 h3l Dhvap
5
(3)
where, referring to the measured temperature T3, h3l is the enthalpy of the saturated liquid, Dhvap is the latent heat of vaporization, and:
h3 ¼
Q_ h m_ $cw $ðTwhi Twho Þ þ h1 ¼ wh þ h1 _ m m_
(4)
Temperatures have been measured by means of PT100 class B thermometers, while four pressure transducers have been flushmounted on the refrigerant lines. Low pressure measurements (P1 and P4 transducers) have an inaccuracy equal to 0.012 bar while high pressure measurements (P2 and P3 transducers) have an inaccuracy equal to 0.12 bar. The water mass flow rates through the heat exchangers have been measured by means of Adafruit 833 liquid flow meters (inaccuracy less than 0.015 kg/s). R245fa mass flow rate is not directly measured. It is determined via the energy balance at the condenser. Under steady-state conditions:
m_ ¼
Q_ c m_ wc $cw $ðTwco Twci Þ ¼ h4 h1 h4 h1
(5)
where cw is the water specific heat. In theory, the system is quite simple because it consists of a few elements. However, the complexity lies in sizing and properly placing each part in order to obtain the maximum efficiency. About mounting, the most complex part concerns the correct positioning of the sensors for correctly detecting the physical quantities of interest. 4. Results Several tests have been conducted at different hot source and cold sink temperatures. The hot source temperature has been varied from 75 C to 95 C, while, according to the environment conditions, the cold sink temperature ranged between 20 C and 33 C.
Fig. 6. Test bench sketch.
Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104
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During the tests, both the hot and the cold water flow rates have been kept approximately constant. The water drop temperature was about 10 C in both the heat exchangers. Due to its simplicity, the system is easy to handle during both the start-up phase and the steady operation. The start-up phase takes only a few minutes. The temperature of the hot fluid determines the evaporation temperature of the refrigerant, and then the back pressure acting on the pump. Measurements have shown that: - Wet vapor enters the expander. Depending on the heat exchanger efficiency, R245fa inlet temperatures range from 90 C to 71 C. Correspondingly, the vapor pressure varies from 10 bar to 6.1 bar, while the vapor quality is in the range 0.871e0.939. - Superheated vapor leaves the expander. Its temperature ranges from 31.3 C to 43.0 C. - The exhaust vapor condenses through the heat exchanger. The condensate is blown into the tank where its pressure ranges from 1.5 bar to 2.4 bar. The corresponding saturation temperature ranges from about 25.9 C to 39.0 C. - Sub-cooled liquid fluid leaves the pump at the maximum pressure described above; its temperature ranges from 27.6 C to 41.1 C. Fig. 7 shows both the thermal efficiency and the specific work calculated as:
hth
h h4 ðh2 h1 Þ ¼ 3 h3 h2
w ¼ h3 h4 ðh2 h1 Þ
(6) (7)
while Figs. 8 and 9 show plant input and output power calculated as:
_ ¼ m$Dh _ W
(8)
Fig. 8. Mini ORC work rates at different test conditions.
about 10%, while the specific work is within the range 9e20 kJ/kg. The expander output power ranges from about 0.4 to about 1.2 kW. The maximum output power is obtained when the heat rate entering the system is slightly less than 11 kW. Naturally, also the cold source temperature influences the cycle performances causing the data dispersion highlighted both in Figs. 7 and 8. With regard to this variable, Fig. 10 shows the benefit obtained by reducing the coolant's temperature.
It is worth noting that cycle performances increase almost linearly with the maximum cycle pressure. At the test conditions mentioned above, the thermal efficiency ranges from about 5% to
Fig. 7. Mini ORC thermal efficiencies and specific works at different test conditions.
Fig. 9. Mini ORC heat rates at different test conditions.
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Fig. 10. Influence of the condenser pressure on the Mini ORC performances.
Table 2 Test results, operating point characterized by the best thermal efficiency. Direct measured values. Water temperature at evaporator inlet
Twhi
[ C]
95.7
Water temperature at evaporator outlet Water temperature at condensator inlet Water temperature at condensator outlet Hot water mass flow rate Cold water mass flow rate R245fa pressure at scroll inlet R245fa temperature at scroll inlet R245fa pressure at scroll outlet R245fa temperature at scroll outlet R245fa pressure at pump outlet R245fa temperature at pump outlet R245fa pressure at pump inlet
Twho Twci Twco m_ wh m_ wc P3 T3 P4 T4 P2 T2 P1
[ C] [ C] [ C] [kg/s] [kg/s] [bar] [ C] [Bar] [ C] [bar] [ C] [Bar]
86.8 26.8 36.3 0.288 0.248 9.95 89.7 2.17 38.4 9.99 35.8 2.02
Table 2 and 3 describe the operating point characterized by the best thermal efficiency achieved during the tests, while Fig. 11 shows the corresponding cycle state points. The best value achieved for thermal efficiency was equal to 9.3 percent. In order to have a reference value, a Carnot cycle, evolving between the given hot source temperature (Twhi ¼ 95.7 C) and the cold sink temperature (Twci ¼ 26.8 C), can be assumed. In this case, the efficiency of the Carnot cycle would be equal to 18.7 percent. Table 3 Test results, operating point characterized by the best thermal efficiency. Calculated data. R245fa temperature at pump inlet
T1
[ C]
33.8
R245fa vapor quality at scroll inlet R245fa flow mass rate Specific work Evaporator heat rate Condensator heat rate Expander work rate Pump work rate Thermal efficiency Expander efficiency Pump efficiency Carnot cycle efficiency
Q3 m_ w Q_ h Q_ c _e W _p W
[e] [kg/s] [kJ/kg] [kW] [kW] [kW] [kW] [%] [%] [%] [%]
0.915 0.052 19.35 10.88 9.87 1.17 0.16 9.28 84.9 79.7 18.7
hth hh,exp hh,pum hCar
Fig. 11. R245fa state points relative to the test results shown in Table 2.
Saturated RF245a vapor leaves the evaporator. Its quality is equal to 0.915. As mentioned above, (Fig. 4) this parameter has a small influence on the cycle efficiency, but it affects the specific work provided by the power plant. Finally, the results show that both the scroll expander and the pump have a very good hydraulic efficiency, calculated comparing the actual work (i.e. the mechanical energy transferred to or from the rotor) to the isentropic ideal work [21]. 5. Conclusions In this paper, a simple and cheap small-scale ORC plant has been designed, realized and tested at numerous operative conditions. In the proposed study, an un-recuperated Rankine cycle has been adopted and a volumetric scroll expander, for the mechanical power generation, has been chosen. This expander featured an acceptable expansion efficiency and a very good volumetric efficiency. Furthermore, it showed silent and reliable operation. When thermal energy is introduced at a temperature less than 100 C, the mini-ORC plant is able to reach a thermal efficiency equal to about 9%, a specific work equal to about 20 kJ/kg and provides an output power somewhat greater than 1 kW. Authors want to precise that, at present, the prototype is still under development and the reported results come from partial investigations relative to the evolution of the working fluid. However, these results prove that mini-ORC plant is a promising technology in the few kW power range when low-temperature thermal energy is available as a primary energy source. In particular, it can be attractive both in the field of the waste heat recovery and in the scenario of distributed power generation for domestic usage. References [1] Bianchi M, De Pascale A. Bottoming cycles for electric energy generation: parametric investigation of available and innovative solutions for the exploitation of low and medium temperature heat sources. Appl Energy 2011;88:1500e9.
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Please cite this article in press as: Galloni E, et al., Design and experimental analysis of a mini ORC (organic Rankine cycle) power plant based on R245fa working fluid, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.07.104