- Email: [email protected]
Selection of refrigerants for a modified organic Rankine cycle
Accepted Manuscript SELECTION OF REFRIGERANTS FOR A MODIFIED ORGANIC RANKINE CYCLE Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski, Kamil Futyma, Zbigniew Rudowicz PII:
S0360-5442(18)32228-X
DOI:
https://doi.org/10.1016/j.energy.2018.11.024
Reference:
EGY 14122
To appear in:
Energy
Received Date: 21 November 2017 Revised Date:
5 November 2018
Accepted Date: 8 November 2018
Please cite this article as: Kajurek J, Rusowicz A, Grzebielec A, Bujalski W, Futyma K, Rudowicz Z, SELECTION OF REFRIGERANTS FOR A MODIFIED ORGANIC RANKINE CYCLE, Energy, https:// doi.org/10.1016/j.energy.2018.11.024. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ACCEPTED MANUSCRIPT The modified organic Rankine cycle is proposed An efficiency of ORC for different refrigerant is obtained The R245fa as a refrigerant is selected as the best solution for proposed modified Rankine cycle
AC C
EP
TE D
M AN U
SC
RI PT
An efficiency is compared with classic ORC system
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017 19 - 22 września 2017, Beskid Śląski
SELECTION OF REFRIGERANTS FOR A MODIFIED ORGANIC RANKINE CYCLE
Zakład Mechaniczny ‘Mestil’ sp. z o.o. ul. Walczaka 25, 66-407 Gorzów Wlkp. email: [email protected]
SC
2
RI PT
Jakub Kajurek1, Artur Rusowicz1, Andrzej Grzebielec 1, Wojciech Bujalski 1, Kamil Futyma 1, Zbigniew Rudowicz 2 1 Institute of Heat Engineering, Warsaw University of Technology, ul. Nowowiejska 21/25, 00-665 Warszawa email: [email protected]
M AN U
Keywords: Waste heat, Organic Rankine Cycle, working fluid, refrigerant Abstract
AC C
EP
TE D
The use of waste heat in many branches of industry is limited due to temperature in the range of 30 to 100°C. One of the methods of using waste heat are devices that implement the Organic Rankine Cycle (ORC). In currently used ORC systems, the heat source temperature is at least 80°C, while the low temperature heat source (usually atmospheric air) has a temperature of 30°C. The work analyzes the influence of the organic fluids properties on the performance of the novel, proposed installation with electric power output 1kW driven by the waste heat and working based on the ORC. In comparison to a commonly used ORC installation the examined installation is characterized by one essential difference: expansion process. The expansion device used in a typical ORC structure, is replaced by two connected tanks filled by secondary fluid, between which the hydraulic turbine is installed. The basic operation parameters in nominal conditions were determined for ten selected refrigerants: R134a, R152a, R227ea, R236fa, R245fa, R290, R600a, R717, R1234yf, R1234ze(E). The boiling point 80°C and the condensing temperature 30°C were used as nominal conditions. Analogical calculations were also done for conventional ORC system working with the same parameters and providing the same electric power output. As a result of the analysis, it turned out that for the proposed device, the best refrigerant is ammonia, for which the 5.68% efficiency was obtained. Comparing the results with the achievements for a standard ORC system, it is almost twice lower (for ammonia ORC systems, the efficiency is 9.93). However, in the experimental installation it was decided to use R245fa refrigerant, because unlike ammonia it is not flammable and is not toxic. According to the analysis, the R245fa was one of the best refrigerants in terms of efficiency and, what is the most important, it did not mix with the glycerol used in the tanks as a secondary fluid. Theoretical analysis for fluid R245fa has shown that the total energy efficiency coefficient for whole proposed system will be at level 4.72%, while as a result of the experiment, it managed to achieve an efficiency level of 2.51%.
1. Introduction The ORC is an unconventional and very promising technology for electricity generation that has been gaining increasing popularity [1–4]. ORC-based systems employ the steam Rankine 1
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
AC C
EP
TE D
M AN U
SC
RI PT
cycle except that instead of water, organic or inorganic fluids characterized by a low boiling point are used as the working fluid [3,5]. Despite numerous advantages, including high latent and specific heat, chemical stability in a wide range of temperatures, low viscosity, non-toxicity, non-flammability, and availability, water has a peculiar disadvantage, namely a high normal boiling point of 100°C, which renders water unusable in practice in low-temperature systems [6–9]. On the other hand, organic refrigerants are characterized by a low normal boiling point, often below 0°C [10]. There is a lot of research studies on selecting an optimum fluid for use in certain conditions [11–15]. In Ref. [16], the effect of working fluid properties on the irreversibility of a cycle supplied by a 10 MW waste heat source at 600 K is analyzed. Among the examined refrigerants the R113 and R123 were characterized by the highest cycle irreversibility which, at the same time, decreased at a lower rate with increasing evaporation pressure compared to the other fluids. Studies presented in Ref. [17] focus on the effect of the evaporation pressure of six chosen refrigerants on the power output of a Rankine cycle supplied from a heat source with temperature varying from 80°C to 200°C and the condensation temperature of 20°C. It was found that an increase in pressure leads to an increase in the system power output to the specified maximum value for all the fluids. The highest power output was achieved for R227ea between 80°C and 160°C, and for R245fa between 160°C and 200°C. In Ref. [6], Mikielewicz recommends some working fluids for use in residential ORC systems for generating 4 kW of heat and electricity. Among proposed fluids at the turbine inlet temperature of 170°C and the condensation temperature of 50°C, R123, R141b and ethanol achieved the highest efficiency. In the research presented in Ref. [18], heat transfer area minimization and heat recovery efficiency served as criteria for comparing the following fluids: R123, R134a, R141b, R142b, R152a, R227ea, R236ea, R236fa, R245ca, R245fa, R600, R600a and n-pentane. Also in this case the results were related to the heat source temperature. R123 performed as the best fluid between 100°C and 180°C, and R141b for heat sources at over 180°C. In Ref. [19], an attempt was made to select a working fluid which for a given supply temperature, varying in the study from 150°C to 250°C and at constant condenser cooling water temperature of 5°C will provide maximum power output. Some of fluids operated in a supercritical cycle. The highest power output between 150°C and 175°C was achieved for R134a, between 175°C and 185°C for R142b, between 185°C and 210°C for R245fa, between 210°C and 225°C for R245ca, between 225°C and 230°C for R365mfc, and between 230°C and 250°C for R123. In turn, Schilling et al. concludes that the choice of the refrigerant should be directly connected to the ORC installation design [20]. In this case, interventions resulting from subsequent design will be avoided. Many studies have examined the influence of the properties of the single ORC component working fluids as well. Table 1 summarizes refrigerant recommendations depending on the temperature of the heat source. In addition to single-component working fluids, mixtures can also be used in ORCs. Such fluids make it possible to match the fluid to the heat source more effectively and to develop a substance exhibiting desirable properties, taking into account environmental and safety requirements. The temperature of heat sources that supply systems often varies. A characteristic of single-component substances is a constant phase change temperature. This is not the case with multi-component, zeotropic fluids, which in the process of phase change are subject to a so-called temperature glide [20–22]. As the fluid temperature changes when heat is supplied, the temperature profile of the refrigerant can be better aligned to that of the heat source [23]. In this way, heat and exergy losses are lower, while the system efficiency is higher [24]. Such positive consequences are possible only for a certain composition of the zeotropic mixture and thus for a particular temperature glide [25] . Zeotropic refrigerants were also investigated in a number of studies. Liu et al. [26] analyze the effect of the composition of R600a/R601a and R227ea/R245fa mixtures on the efficiency of 2
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
AC C
EP
TE D
M AN U
SC
RI PT
an ORC supplied by variable-temperature geothermal energy, with cooling water at 20°C as the lower heat source. For the molar ratio of 90% of R600a in an R600a/R601a mixture, an increase of 8% in efficiency was achieved compared to the pure fluids. The maximum efficiency with an R227ea/R245fa mixture occurred for the molar ratio of 80% of R227ea in the mixture, and was higher by 15% compared to a system employing a single-component refrigerant. The increase in efficiency compared to single-component refrigerants was achieved only for temperatures below 120°C; above that value, R227ea and R600a provided higher efficiency. In Ref. [21], an ORC system with pentane/hexane and R245fa/R365mfc mixtures was studied. The increase in efficiency for high source temperature varying from 150°C to 250°C was from 6% to 16%, while electric power increased by 20%. It was also demonstrated that using three-component mixtures in a system has no significant impact on improving the efficiency. In Ref. [24], the effect of the temperature glide of R600/R601, R600/R601a, R600a/R601, and R600a/R601a mixtures on the operation of an ORC system was studied. The results has shown that the increase in the efficiency of a system using mixtures compared to a system employing single-component refrigerants is possible only for a temperature glide that is lower than a change in the temperature of the fluid supplied to the evaporator. Similar conclusions were presented in Ref. [25], which demonstrated that the temperature of the heat source determines the selection of the mixture. As a results from the research is fact that for a constant temperature range the efficiency and output power of a system employing single-component refrigerants are higher than those of a system working with mixtures. Another approach for choosing the correct mixture as a working medium was proposed by Oyewunmi and Marides in the publication [27] who as a criterion applied the cost of the construction of the installation and its subsequent operation. In their studies, it was shown that the installation driven by heat source of temperature 98°C from the thermodynamic point of view should be used mixtures of 50% n-pentane and 50% n-hexane or 60% R-245fa and 40% R-227ea. However, for clean components, the installation cost is 14% cheaper. The other results were obtained by Le, Kheiri et al. [28], whose economic analysis has shown that the use of pure refrigerant R245fa is the least viable. The theoretical analysis assumes that the system generates 1 kW of electric power. And on the basis of this input data, the other elements of the system were selected. Unamba C.K. et al. for a typical ORC installation with an electric output power of 1 kW, used the R245fa as refrigerant obtained a thermal efficiency of the system from 0 to 6% depending on the working conditions [29]. This work propose a new ORC design. In the proposed installation, the gas expansion turbine is replaced with an innovative solution of secondary fluid tanks and a liquid turbine. The main purpose of the research is the selection of the most appropriate working medium. Ten single component refrigerant was analyzed. Mixtures were not analyzed in this work due to the risk of the mixture separation in glycerol (secondary fluid). The influence of each working fluids properties on the performance of the proposed apparatus is discussed and the final guidelines for optimal selection are presented as well. Similar calculations were also performed for conventional ORC system that provides the same power output and works with the same refrigerants and under the same conditions as the proposed installation. The selected refrigerant has been tested experimentally.
2. Description of the proposed experimental installation The paper aims to select an optimal working fluid for the proposed experimental installation schematically depicted in Fig. 1. Although the working fluid in the cycle of the system under consideration is subject to the same changes as in the ORC, i.e. isobaric evaporation, expansion, isobaric condensation, and liquid compression, essential differences can be noticed 3
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
AC C
EP
TE D
M AN U
SC
RI PT
[26,29]. An expansion device, which in classic ORC systems can be a turbine or a piston, or screw expander, is replaced with two cylindrical tanks filled by intermediate liquid - glycerol. The generator in the proposed installation is not driven at the expense of a drop in the enthalpy of the refrigerant in the turbine, but following a change in the height of liquid columns in tanks. This solution has been proposed due to the fact that water turbines are usually characterized by higher efficiency than gas turbines (especially those used in ORC systems). As with a conventional ORC system, the system under consideration can be driven by low-temperature heat source. Then the heat carrier, at a temperature higher than that of the organic refrigerant, flows through the evaporator (1) and makes the refrigerant to evaporate. The operation cycle of the device starts when the level of the glycerol is maximum in one of the tanks (5.2) and minimum in the second one (5.1). When a liquid level sensor indicates the extreme level of the liquid in the tanks, valves A open, while valves B close. The refrigerant under high pressure, equal to evaporation pressure, flows to tank 5.2. At the same time vapour, which filled the tank in the previous cycle, starts to flow out of tank 5.1 to a volume of low pressure, equal to the condensation pressure. While the pressure difference occurs, the secondary fluid flows from the high-pressure tank (5.2) to the low-pressure one (5.1), performing work in the turbine (T) located between them. When, as a consequence of pressure exerted by the organic refrigerant on the liquid stored in the high-pressure tank, the intermediate liquid in the low-pressure tank is at its maximum level, the inlet valves switch positions: the A valves close, while the B valves open. Then the glycerol flows from tank 5.1 to tank 5.2, permanently driving the turbine (T). The direction of flow of the intermediate fluid through the turbine will not change, since the turbine’s inlet and outlet connections are interconnected with a collecting manifold. Valve positions will switch each time the glycerol attains extreme levels, thus providing the flow through the pump and an on-going electricity generation in the generator (G). Further thermal process of the refrigerant are the same as in a classic ORC. The refrigerant flows out of the low-pressure tank, it reaches the condenser (4), where it undergoes a phase change, giving heat out of the system. From the condenser the refrigerant is again fed to the high-pressure part of the set-up via the pump (9) located in the liquid line. The proposed experimental installation is equipped with necessary accessories to control the operation. Tank 6 serves as a so-called supercharging device that boosts the liquid flow through the turbine if the generator load increases. Also in this case, the flow is forced by the organic refrigerant under high pressure. The component marked as 7 is a liquid refrigerant tank and at the same time it works as a secondary fluid separator. The correct level of fluids in the tanks is provided by the expansion vessel and the pump (10). Operating parameters, i.e. pressure, temperature, and flow, are controlled by pressure, temperature, and level transmitters, and flow meters.
3. Modelling approach Basic thermodynamic transformations for the proposed ORC system are presented in Fig. 2. In order to determine the basic parameters of the experimental system, it was additionally assumed that system is in a steady state. Pressure drops in the evaporator, condenser and other elements of the installation are negligible, also the heat losses to the environment are negligible. Refrigerant flows into the tank and exerts an even pressure on the surface of the secondary fluid. Secondary fluid does not mix and does not dissolve with the refrigerant. The efficiency of all items including turbines and pumps are known: • isentropic turbine efficiency: = 0.92; • mechanical efficiency of the turbine gear: = 0.95; • electrical efficiency of the generator: = 0.93; 4
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
• • • • •
power output system efficiency: = 0.98; refrigerant expansion efficiency in the tanks: = 0.92; isentropic pump efficiency: = 0.85; pump gear efficiency: = 0.95; electrical efficiency of the pump engine: = 0.98. Operation cycle starts from position 1 (Fig. 2.). The power Np required for increasing the pressure from condensation pC to evaporation pE level is equal to: ℎ(
!)
− ℎ$ %
where pump efficiency: = ∙ ∙
(1)
RI PT
=
(2)
ℎ( !) − ℎ$ %
M AN U
ℎ = ℎ$ +
SC
while the enthalpy at point 2, at the end of the pumping process, is equal to:
(3)
Next step in cycle is evaporation process – line 2-3 in Fig. 2. The heat flux supplied in this process () must firstly warms up the refrigerant to the phase change temperature and then evaporates it. Required heat flux is described as: () = (),, + (),- = (ℎ. − ℎ )
(4)
TE D
where (),, is heat flux needed from warimg up refrigerant and (),- is heat flux needed for evaporating: (5)
EP
(),, = (ℎ/ − ℎ ) (),- = (ℎ. − ℎ/ )
(6)
AC C
For the classic ORC system, the next step in cycle is the expansion in the turbine, in which the work is equal to: 012 = (ℎ. − ℎ3 )
(7)
the electric power at the terminals of the generator is: = 012
(8)
and total efficiency of the ORC system is equal to: 12 =
− () 5
(9)
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
In the experimental installation (Figure 1), the expansion device is replaced by two tanks filled by a secondary fluid. During operation of the system in steady state, the liquid flow rate through the turbine (T in Fig. 1) is constant, so the volumetric flow of the secondary , caused by the refrigerant at evaporation pressure is equal to refrigerant volumetric fluid 45 flow in point 3 (Fig. 2): = 4. 45
(10)
RI PT
where: 4. = 6.
(11)
SC
The displacement volume of the tanks 47 will therefore be equal to the volume of liquid flowing through the reservoirs at a time corresponding to half cycle of the device:
The time of valves change is equal:
M AN U
89: ;: =$/ 47 = = 45 4
=$/ =
89: ;: 445
(12)
(13)
TE D
The theoretical power PT,th of the turbine and the electric power of the system NEL,2 supplied with liquid flowing from the high-pressure tank to the low-pressure tank will be: (@ − @2 ) ?,!, = 45
(14)
, = A -B C ?!,
(15)
AC C
EP
and the efficiency of the system is expressed as: D =
, − ()
(16)
In both solutions, in the experimental installation and the classic ORC system, the expansion medium flows to the condenser, where heat is transferred out of the system: (2 = (ℎ3 − ℎ$ )
(17)
The end of the expansion of “dry” fluid (difference between dry, wet and isentropic fluids are shown in Fig. 3.), shown in Figure 2, is located in the area of superheated gas. Heat transferred in condenser outside of system can be divided into two heat flux: (2, , - chilling the gas to the condensation temperature and (2, 7 - real condensation process:
6
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
(18) (2, , = (ℎ3 − ℎ3/ ) (2,
7
= (ℎ3/ − ℎ$ )
(19)
M AN U
SC
RI PT
In the case of “wet” fluids operating in circuits without superheated gas, the end of the expansion is located in the area of wet gas. The heat transferred to environment is therefore equal to the heat of condensing the refrigerant from a given degree of dryness to a saturated liquid. For “isentropic” fluids, this heat is equal to the total heat of condensation for a given temperature. The position of the final expansion point in the wet gas area for “wet” refrigerant can contribute to the dropping of organic liquid droplets in the flow tanks. Depending on the type of secondary fluid, the wet compound may dissolve in it. The return of the working fluid to the evaporator in accordance with Figure 1 will be impossible, as there is no installation element responsible for separating the components of the resulting mixture. The oil separator (No. 8 in Figure 1) serves to separate the medium from the intermediate liquid, but the one that has been entrained by the pairs of the compound. The continuous increase in the amount of liquid refrigerant in the flow tanks may contribute to operational problems related to, e.g. with inoperative inlet and outlet valves (A and B in Figure 1). Therefore, in the case of wet substances, it seems necessary to overheat the compound pairs by the value for which the end of the expansion would be on the saturation line (Figure 2). Then the total amount of heat received from the upper source will be equal to the sum of the heat of heating the liquid to the saturation state (),, , heat of evaporation (),- and heat of overheating (), : (20)
TE D
() = (),, + (),- + (),
4. Results and analysis
AC C
EP
In the first step of the analysis, it was decided to analytically select the refrigerant most suited to the proposed installation. The results of the theoretical analysis were compared with the results for a typical ORC installation working under the same conditions. Calculations for the proposed experimental set-up were performed for the electric power output of 1 kW. In addition, it was decided that the dimensions of the tanks are in every case the same: • height: ; = 0.5 ; • diameter: 9 = 0.2 . The same simplifications was also enabled in the analysis for a conventional ORC system that provides the same power output and works with the same refrigerants and under the same conditions as for the proposed installation. For both the solutions basic set-ups were analyzed. In line with the description, the proposed installation operation is similar to the operation of the ORC with basic processes shown in Fig. 2. Nominal operating parameters for the calculations are as the following: • evaporation temperature: 80°C; • condensation temperature: 30°C. The selection of the refrigerant for a given temperature range is sometimes based on the vapour saturation curve in the T-s diagram [1,30–33]. In the group of refrigerants considered, 7
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
AC C
EP
TE D
M AN U
SC
RI PT
in nominal operating conditions of the set-up, R134a, R152a, R290, and R717 are wet fluids, while R227ea, R2236fa, R245fa, and R600a are dry fluids (Tab. 2). R1234yf and R1234ze(E) can be considered as isentropic fluids, even though their final expansion point is not exactly on the vapour saturation line. In the group of fluids under consideration, all the wet refrigerants are characterized by higher working pressures than the dry ones. Ammonia, propane, and R134a (Tab. 2) are characterized by the highest evaporation and condensation pressures in nominal operating conditions. On the other hand, R245fa, R236fa, and R600a are refrigerants with the lowest evaporation and condensation pressures. Although R245fa features the lowest values of working pressures and pressures ratio is the highest; for R290 the pressures ratio is the lowest. It should be noted that in none of the cases the condensation pressure is lower than the atmospheric pressure, which will protect the system from the possibility of moisture condensation due to leakages. Ammonia has the highest heat of evaporation and condensation (Tab. 2); these values are almost three times higher than for the other fluids. The lowest heat of phase change in the temperature range considered is observed for R227ea. One quantity that is essential for the proposed installation is the specific volume of vapours of the refrigerant flowing into the flow tanks. It is the value that determines the volumetric flow rate of the glycerol flowing through the expansion device. Among the organic substances considered, ammonia R717 features the highest specific volume of saturated vapour under the evaporation pressure. Slightly lower values of over 20 dm3/kg are also observed for R600a and R245fa. R134a, R227ea, and R1234yf are fluids with the smallest specific volume of vapour, slightly over 5 dm3/kg. The volumetric flow rate can also be expressed as a ratio of the system power output (1 kWe in the considered case) to the difference between evaporation and condensation pressures. The larger the pressure difference for a constant power output of the system, the lower the volumetric flow rate of the intermediate liquid is required. In the group of considered fluids, a system employing ammonia requires the lowest flow rate of the liquid, since the pressure difference for ammonia is the largest. As with working pressures, R290 and R134a (Tab. 2) are next and require a higher flow rate of slightly over 0.5 l/s. On the other hand, if refrigerants with the smallest difference between evaporation and condensation pressures (R245fa, R236fa, and R600a) are used in a system, the highest volumetric flow rate is required (Tab. 2). According to the pressure values, a relation can be observed such that the volumetric flow rate of the wet refrigerants is lower than that of the dry ones. The time after which valves are switched is inversely proportional to the volumetric flow rate of the refrigerant vapour under evaporation pressure, which is why for fluids featuring a high volumetric flow rate the switching time will be short, while for those with a low volumetric flow rate the switching time will be longer (Tab. 2). The values of the specific volume of saturated vapour under evaporation pressure, the evaporation and condensation pressures, and the system power output determine the mass flow rate of the refrigerant. This parameter is directly proportional to the power output and inversely proportional to the product of the pressure difference and the specific volume of vapour. In the group of considered fluids, the lowest flow rate is required in a system based on ammonia. Next is a system using isobutane (R600a), with one of the lowest differences between working pressures and, on the other hand, one of the highest specific volumes of saturated vapour. Third is propane (R290) which has quite a large pressure difference and an average specific volume of saturated vapour. The highest mass flow rate is required in a system using R227ea, which is a consequence of its low specific volume of saturated vapour. Other fluids with a relatively low mass flow rate include R1234yf and R236fa. Despite the smallest difference between working pressures among all the analysed fluids, and due to a relatively high specific 8
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
AC C
EP
TE D
M AN U
SC
RI PT
volume of saturated vapour, R245fa is a refrigerant requiring an average mass flow rate compared to the other fluids. In each case the mass flow rate of a refrigerant in the proposed experimental installation is higher than the mass flow rate of the refrigerant in a conventional ORC system that provides the same electric power output as the system considered. For all fluids, this value is almost twice as high as in the classic ORC system. Minimum and maximum flow rates in both the solutions are achieved for the same working fluids: R717 (the lowest one) and R227ea (the highest one). The efficiency of both the proposed set-up and the classic ORC system is equal to the ratio of the electric power output less the power to drive the refrigerant pump to the heat supplied. The system efficiency increases when the pump power and the amount of heat supplied to the system are smaller. Among the selected organic fluids, the least power of the fluid pump is required by R245fa (the smallest pressure difference and a relatively low volumetric flow of the liquid refrigerant through the pump) and R717 (the lowest rate of volumetric flow of the liquid through the pump and the largest pressure difference). Such a low value of the rate of volumetric flow of the fluid through the pump and the small pressure difference in the case of R245fa are crucial for the pump driving power. The highest power of the refrigerant pump is required in a system using R290 and R1234yf, i.e. refrigerants with a high volumetric flow rate of the liquid. These are also fluids with the highest temperature at the end of the pumping process (Tab. 2). The high value of energy needed to drive the pump may result from the fact that it is partly used for heating the liquid. The least amount of heat drawn from the heat source is a characteristic of ammonia, which has the highest heat of evaporation compared to the other organic fluids. The small amount of heat is a consequence of a very low mass flow rate of R717. It is followed by R600a, R290, and R152a, i.e. refrigerants with a relatively high heat of phase change. On the other hand, R227ea, characterized by the highest mass flow rate and the lowest heat of evaporation, requires the largest amount of heat to be supplied to the system (Tab. 2). As with the mass flow rate, both the pump driving power and the heat drawn from the upper heat source are higher in the proposed installation than in the ORC system. The maximum and minimum pump driving power in the proposed set-up and in the ORC system is achieved again for the same refrigerants: R290 (the maximum power) and R245fa (the minimum power). The minimum and maximum values of heat supplied from the upper heat source also occur for the same fluids. Due to a low required power for the liquid refrigerant pump and a small amount of heat provided from the heat source, among all the considered fluids it is ammonia that guarantees the highest efficiency of both the proposed set-up and the conventional ORC system (Fig. 4). It is followed by isobutane in terms of efficiency of the proposed installation Despite a relatively small pressure difference and due to a high specific volume of saturated vapour and thus a small mass flow rate and a small amount of heat drawn, R600a provides the efficiency of 4.87%. In the case of the conventional ORC system, second best efficiency, equal to 9.50%, is achieved with R245fa. This is caused by a lower power of the refrigerant pump and the same amount of heat drawn from the heat source compared to isobutane. R152a provides a slightly lower efficiency than that of the proposed installation using isobutane. R245fa takes the fourth place. R1234yf is characterized by the lowest efficiency. For this fluid, the required power of the liquid refrigerant pump is the highest. Almost the same efficiency is achieved with R227ea, which requires the largest amount of heat to be supplied from the heat source. Also in the case of the conventional ORC system, R227ea and R1234yf provide minimum efficiencies (Fig. 4). The differences between corresponding values of efficiency for both the solutions are small and amount to up to 2 percentage points. For each refrigerant, the efficiency of the 9
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
RI PT
conventional ORC system is higher than that of the proposed set-up. The largest difference in terms of the efficiency of the proposed set-up occurs for R1234yf (nearly 130%) and R134a (113%), while the smaller one for ammonia (70%) and isobutane (90%). In terms of energy efficiency, the best refrigerant was ammonia. Isobutane was on the second place and R152a on the third. Due to their flammability, these refrigerants were eliminated from the experimental part of research. Therefore, the refrigerant R245fa was used for the experiment. The installation presented in figure 1 was filled with refrigerant R245fa. During operation, the device maintained in the evaporator boiling pressure corresponding to 80°C, and in the condenser the pressure corresponding to 30°C of saturation. At the turbinegenerator assembly, a flywheel was mounted, which enabled operation at a constant speed even when the tanks were switched. For these conditions thermal efficiency of 2.51% was obtained. Achieved electric power output was at level 800 W.
SC
5. CONCLUSIONS
AC C
EP
TE D
M AN U
In this study the selection of refrigerants for a modified ORC as well as the influence of their properties on the performance of the proposed installation are described. Ten different working fluid were examined in terms of the thermal efficiency for proposed apparatus. The achieved results were also compared with an operation of a typical ORC system that provides the same power output and works with the same refrigerants and under the same conditions as the proposed installation. ORC systems with liquid turbines turned out to have lower efficiency than systems with gas turbines. Among the refrigerants analyzed, in nominal operating conditions of the proposed installation (Te=80°C, Tc=30°C), the maximum efficiency is provided for ammonia (R717), which is characterized by both the highest specific volume of saturated vapour flowing into the tanks and the largest difference between working pressures. For all the organic fluids examined, these two quantities are critical to the efficiency of the system, since they determine the refrigerant mass flow rate, the feed pump power, and the amount of heat drawn from the lower heat source. Selecting a suitable organic refrigerant from the group of the proposed fluids can lead to a 57% increase in efficiency compared to a refrigerant for which the system was the least efficient. Nevertheless, irrespective of the properties of an organic fluid, the efficiency of the system is very low, from 3.62% for the worst refrigerant (R1234yf) to 5.68% for the best one (R717). In the case of the other substances, the efficiencies are between 4% and 5%. The values are nearly twice as high for the same fluids used in the same conditions in the classic ORC system, where the efficiency for the best fluid (again ammonia) amounts to 9.93% and for the worst one (again R1234yf) 8.38%. The R245fa refrigerant was used in the experimental part because refrigerants characterized by better energy efficiency R717, R600a and R152a are flammable. For refrigerant R245fa thermal efficiency of 2.51% was obtained. Despite the fact that installation was designed to generate 1000W, it was experimentally achieved a maximum electric output of 800W.
Acknowledgements The presented results were financed from the European Regional Development Fund under the Intelligent Development 2014-2020 Program. Project entitled "An innovative thermal power plant with a liquid turbine and a hydraulic-gas generator that utilizes a waste heat flow" implemented as part of the National Center for Research and Development contest: Szybka Ścieżka. 10
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
References Chen H, Goswami DY, Stefanakos EK. A review of thermodynamic cycles and working fluids for the conversion of low-grade heat. Renew Sustain Energy Rev 2010;14:3059–67. doi:10.1016/j.rser.2010.07.006.
[2]
Hung TC. 98/01752 A review of organic Rankine cycles (ORCs) for the recovery of low-grade waste heat: Hung, T. C. et al. Energy, 1997, 22, (7), 661–667. Fuel Energy Abstr 1998;39:151-. doi:http://dx.doi.org/10.1016/S0140-6701(98)97894-8.
[3]
Kajurek J, Rusowicz A, Grzebielec A, Bujalski W, Futyma K, Szczęśniak A, et al. Influence of working fluid properties on performance of organic Rankine cycle. Rynek Energii 2017:68–79.
[4]
Chacartegui R, Sánchez D, Muñoz JM, Sánchez T. Alternative ORC bottoming cycles FOR combined cycle power plants. Appl Energy 2009;86:2162–70. doi:10.1016/j.apenergy.2009.02.016.
[5]
Kajurek J, Rusowicz A, Grzebielec A. The Influence of Stack Position and Acoustic Frequency on the Performance of Thermoacoustic Refrigerator with the Standing Wave. Arch Thermodyn 2017;38. doi:10.1515/aoter-2017-0026.
[6]
Mikielewicz D, Mikielewicz J. A thermodynamic criterion for selection of working fluid for subcritical and supercritical domestic micro CHP. Appl Therm Eng 2010;30:2357–62. doi:10.1016/j.applthermaleng.2010.05.035.
[7]
Quoilin S, Broek M Van Den, Declaye S, Dewallef P, Lemort V. Techno-economic survey of organic rankine cycle (ORC) systems. Renew Sustain Energy Rev 2013;22:168–86. doi:10.1016/j.rser.2013.01.028.
[8]
Saleh B, Koglbauer G, Wendland M, Fischer J. Working fluids for low-temperature organic Rankine cycles. Energy 2007;32:1210–21. doi:10.1016/j.energy.2006.07.001.
[9]
Tchanche BF, Lambrinos G, Frangoudakis A, Papadakis G. Low-grade heat conversion into power using organic Rankine cycles - A review of various applications. Renew Sustain Energy Rev 2011;15:3963–79. doi:10.1016/j.rser.2011.07.024.
AC C
EP
TE D
M AN U
SC
RI PT
[1]
[10] Grzebielec A. Termodynamiczne podstawy przenoszenia ciepła w termoakustycznych urządzeniach chłodniczych. Chłodnictwo 2009;44:12–6. [11] Roy JP, Mishra MK, Misra A. Parametric optimization and performance analysis of a waste heat recovery system using Organic Rankine Cycle. Energy 2010;35:5049–62. doi:10.1016/j.energy.2010.08.013. [12] Shokati N, Ranjbar F, Yari M. Comparative and parametric study of double flash and single flash/ORC combined cycles based on exergoeconomic criteria. Appl Therm Eng 2015;91:479–95. doi:10.1016/j.applthermaleng.2015.08.031. [13] Zhai H, An Q, Shi L, Lemort V, Quoilin S. Categorization and analysis of heat sources for organic Rankine cycle systems. Renew Sustain Energy Rev 2016;64:790–805. doi:10.1016/j.rser.2016.06.076. 11
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
[14] Papadopoulos AI, Stijepovic M, Linke P, Seferlis P, Voutetakis S. Multi-level Design and Selection of Optimum Working Fluids and ORC Systems for Power and Heat Cogeneration from Low Enthalpy Renewable Sources. Comput Aided Chem Eng 2012;30:66–70. doi:10.1016/B978-0-444-59519-5.50014-9. [15] Bao J, Zhao L. A review of working fluid and expander selections for organic Rankine cycle. Renew Sustain Energy Rev 2013;24:325–42. doi:10.1016/j.rser.2013.03.040.
RI PT
[16] Hung T-C. Waste heat recovery of organic Rankine cycle using dry fluids. Energy Convers Manag 2001;42:539–53. doi:10.1016/S0196-8904(00)00081-9. [17] Lakew AA, Bolland O. Working fluids for low-temperature heat source. Appl Therm Eng 2010;30:1262–8. doi:10.1016/j.applthermaleng.2010.02.009.
SC
[18] Wang ZQ, Zhou NJ, Guo J, Wang XY. Fluid selection and parametric optimization of organic Rankine cycle using low temperature waste heat. Energy 2012;40:107–15. doi:10.1016/j.energy.2012.02.022.
M AN U
[19] Maizza V, Maizza A. Unconventional working fluids in organic Rankine-cycles for waste energy recovery systems. Appl Therm Eng 2001;21:381–90. doi:10.1016/S13594311(00)00044-2. [20] Schilling J, Gross J, Bardow A. Integrated design of ORC process and working fluid using process flowsheeting software and PC-SAFT. Energy Procedia, vol. 129, 2017, p. 129–36. doi:10.1016/j.egypro.2017.09.184. [21] Chys M, van den Broek M, Vanslambrouck B, De Paepe M. Potential of zeotropic mixtures as working fluids in organic Rankine cycles. Energy 2012;44:623–32. doi:10.1016/j.energy.2012.05.030.
TE D
[22] Rusowicz A, Grzebielec A. Legal and technical aspects of replacement refrigerants in refrigeration and air conditioning. J Civ Eng Environ Archit 2015;62:359–68. [23] Li Y-R, Du M-T, Wu C-M, Wu S-Y, Liu C. Potential of organic Rankine cycle using zeotropic mixtures as working fluids for waste heat recovery. Energy 2014;77:509–19. doi:10.1016/j.energy.2014.09.035.
EP
[24] Liu Q, Duan Y, Yang Z. Effect of condensation temperature glide on the performance of organic Rankine cycles with zeotropic mixture working fluids. Appl Energy 2014;115:394–404. doi:10.1016/j.apenergy.2013.11.036.
AC C
[25] Zhao L, Bao J. Thermodynamic analysis of organic Rankine cycle using zeotropic mixtures. Appl Energy 2014;130:748–56. doi:10.1016/j.apenergy.2014.03.067. [26] Heberle F, Preißinger M, Brüggemann D. Zeotropic mixtures as working fluids in Organic Rankine Cycles for low-enthalpy geothermal resources. Renew Energy 2012;37:364–70. doi:10.1016/j.renene.2011.06.044. [27] Oyewunmi OA, Markides CN. Thermo-Economic and Heat Transfer Optimization of Working-Fluid Mixtures in a Low-Temperature Organic Rankine Cycle System. Energies 2016;9:448. doi:10.3390/en9060448. [28] Le VL, Kheiri A, Feidt M, Pelloux-Prayer S. Thermodynamic and economic optimizations of a waste heat to power plant driven by a subcritical ORC (Organic Rankine Cycle) using pure or zeotropic working fluid. Energy 2014. doi:10.1016/j.energy.2014.10.051. [29] Unamba CK, White M, Sapin P, Freeman J, Lecompte S, Oyewunmi OA, et al. Experimental Investigation of the Operating Point of a 1-kW ORC System. Energy 12
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
Procedia, vol. 129, 2017, p. 875–82. doi:10.1016/j.egypro.2017.09.211. [30] Feng Y, Hung TC, Greg K, Zhang Y, Li B, Yang J. Thermoeconomic comparison between pure and mixture working fluids of organic Rankine cycles (ORCs) for low temperature waste heat recovery. Energy Convers Manag 2015;106:859–72. doi:10.1016/j.enconman.2015.09.042.
RI PT
[31] Hærvig J, Sørensen K, Condra TJ. Guidelines for optimal selection of working fluid for an organic Rankine cycle in relation to waste heat recovery. Energy 2016;96:592–602. doi:10.1016/j.energy.2015.12.098. [32] He C, Liu C, Gao H, Xie H, Li Y, Wu S, et al. The optimal evaporation temperature and working fluids for subcritical organic Rankine cycle. Energy 2012;38:136–43. doi:10.1016/j.energy.2011.12.022.
AC C
EP
TE D
M AN U
SC
[33] He C, Liu C, Zhou M, Xie H, Xu X, Wu S, et al. A new selection principle of working fluids for subcritical organic Rankine cycle coupling with different heat sources. Energy 2014;68:283–91. doi:10.1016/j.energy.2014.02.050.
13
ACCEPTED MANUSCRIPT Table 1. Recommended refrigerants Recommended refrigerant R32, R125, R134a, R143a, R124, R227ea, R290, R1234yf, R1270 R114, R141b, R123, R124, R245fa, R601a, R1243ze R123, R141b, R1234ze, RC218, R236fa, R236ea, R600, R601 benzene, paraxylene, toulene, hexane
AC C
EP
TE D
M AN U
SC
RI PT
Source temperature [°C] <100 100-120 120-160 160-200 >200
1
Wet
A2
133
0
Dry
A1
3580
0
Heat of condensation qk, kJ/kg 1.32
1.35
For temperatures between the condensation point and the evaporation point.
0.76
0.67
Volumetric flow rate of refrigerant V3, l/s 0.94
41.6
1.35
42.7
30.5
8.18 11.06 19,24 17.24 0.205 4.13 23.5
Heat of evaporation Qp, kW
Heat drawn from the upper source QGZC, kW
Heat delivered to the lower source QDZC, kW
Driving power of the refrigerant pump PPC, kW
Efficiency of the experimental Installation ηD , %
Valve switching time τ, s
10.68
20.7
4.79
16.7
3.65
0.146 0.202
15.89 19.71
17.84 21.88
11.77 11.20
6.07
11.7
4.30
0.105
18.63
20.82
12.69
8.13
0.1039 0.0618 0.1652 0.1206
Heat for heating up liquid Qdp, kW
Mass flow rate of refrigerant mR, kg/s
30
30
Refrigerant post-expansion temperature T4, °C
30.9
1
31.1
1
31.3
0.941
Refrigerant post-compression temperature T2, °C
0.91
0.959
0.91
Vapour dryness fraction at the end of expansion
1.27
0.95
1.40
30
31.9
0.948
0.0118
2.05
3.12
326.71
189.8
1079
3139.9
Wet
A3
3
0
R290
7.7
4.72
0.053
17.83
20.04
13.76
6.28
0.0902
2.05
873.96
1167.2
4142
Wet
B2L
0
0
R717
25.8
4.08
0.275
15.94
17.76
9.76
8.00
11.7
4.87
0.104
16.39
18.38
12.08
6.30
37.4
5.68
0.090
14.17
16.03
12.44
3.59
21.8
3.62
0.265
18.28
20.30
10.18
RI PT
0.72
30
31.4
0.998
0.0055
1.06
1.61
10.12
0.42
30
31.0
0.87
0.0295
3.19
4.78
141.24
78.36
783.5
2519.4
Isentropic
A2L
4
0
R1234yf
0.1297
0.0479
1.34
40.3
30.6
1
0.028
1.87
2.66
323.33 1144.41
252.58
404.7
1343.8
Dry
A3
4
0
R600a
0.0142
0.0515
0.61
SC
40.5
30.3
1
0.0227
M AN U
1.99
0.0064 0.0123 0.0057 0.0112
1.03
1.55
Specific volume of saturated vapour v3, m3/kg
Average specific heat of vapour cp for ps, kJ/(kg·K)
2
187.33
173.1 273.16 108.71 142.95
177.8 152.51
321
789.3
Dry
B1
1050
0
R245fa
105.29
106.42 190.28 67.81
Heat of evaporation qp, kJ/kg
689.8 528.4
770.2
Condensation pressure pk, kPa
Average specific heat of liquid 1 cw for pp, kJ/(kg·K)
Dry
A1
9820
0
2633.2 2342.4 1858.3 1249.1
TE D
Evaporation pressure pp, kPa
Type of refrigerant
EP
A1
Safety group Wet
1430
GWP
AC C 0
R134a R152a R227ea R236fa
ODP
Quantity
17.9
4.26
0.166
17.53
19.58
11.56
8.02
0.1057
0.88
32.1
32.6
1
0.0083
0.99
1.51
163.05
110.19
578.3
2007.7
Isentropic
A2L
6
0
R1234ze(E)
ACCEPTED MANUSCRIPT Table 2. Basic cycle parameters
AC C
EP
TE D
M AN U
SC
RI PT
ACCEPTED MANUSCRIPT
Fig. 1. Schematic illustration of the proposed experimental installation.
ACCEPTED MANUSCRIPT R227ea 2
20,0
2'
3
RI PT
10,0 9,0 8,0 7,0 6,0
1
4
SC
5,0
3,0
160
180
200
220
240
M AN U
4,0
260
280
300
320
340
360
AC C
EP
TE D
Enthalpy, kJ/(kg K) Fig. 2. Pressure-enthalpy diagram for a theoretical R227ea cycle
Fig. 3. ORC diagram for dry, wet and isentropic fluids
380
400
ACCEPTED MANUSCRIPT Experimental
ORC
8.38
8.93
9.93
9.27
8.57
9.50
9.00
8.40
8.80
10
4.26
3.62
5.68
4.87
4.08
4.72
RI PT
4
4.30
3.65
6
4.79
8 4.13
Efficiency, %
9.28
12
2
SC
0
AC C
EP
TE D
M AN U
Fig. 4. Theoretical efficiency of the proposed installation vs. typical ORC installation
S0360-5442(18)32228-X
DOI:
https://doi.org/10.1016/j.energy.2018.11.024
Reference:
EGY 14122
To appear in:
Energy
Received Date: 21 November 2017 Revised Date:
5 November 2018
Accepted Date: 8 November 2018
Please cite this article as: Kajurek J, Rusowicz A, Grzebielec A, Bujalski W, Futyma K, Rudowicz Z, SELECTION OF REFRIGERANTS FOR A MODIFIED ORGANIC RANKINE CYCLE, Energy, https:// doi.org/10.1016/j.energy.2018.11.024. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ACCEPTED MANUSCRIPT The modified organic Rankine cycle is proposed An efficiency of ORC for different refrigerant is obtained The R245fa as a refrigerant is selected as the best solution for proposed modified Rankine cycle
AC C
EP
TE D
M AN U
SC
RI PT
An efficiency is compared with classic ORC system
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017 19 - 22 września 2017, Beskid Śląski
SELECTION OF REFRIGERANTS FOR A MODIFIED ORGANIC RANKINE CYCLE
Zakład Mechaniczny ‘Mestil’ sp. z o.o. ul. Walczaka 25, 66-407 Gorzów Wlkp. email: [email protected]
SC
2
RI PT
Jakub Kajurek1, Artur Rusowicz1, Andrzej Grzebielec 1, Wojciech Bujalski 1, Kamil Futyma 1, Zbigniew Rudowicz 2 1 Institute of Heat Engineering, Warsaw University of Technology, ul. Nowowiejska 21/25, 00-665 Warszawa email: [email protected]
M AN U
Keywords: Waste heat, Organic Rankine Cycle, working fluid, refrigerant Abstract
AC C
EP
TE D
The use of waste heat in many branches of industry is limited due to temperature in the range of 30 to 100°C. One of the methods of using waste heat are devices that implement the Organic Rankine Cycle (ORC). In currently used ORC systems, the heat source temperature is at least 80°C, while the low temperature heat source (usually atmospheric air) has a temperature of 30°C. The work analyzes the influence of the organic fluids properties on the performance of the novel, proposed installation with electric power output 1kW driven by the waste heat and working based on the ORC. In comparison to a commonly used ORC installation the examined installation is characterized by one essential difference: expansion process. The expansion device used in a typical ORC structure, is replaced by two connected tanks filled by secondary fluid, between which the hydraulic turbine is installed. The basic operation parameters in nominal conditions were determined for ten selected refrigerants: R134a, R152a, R227ea, R236fa, R245fa, R290, R600a, R717, R1234yf, R1234ze(E). The boiling point 80°C and the condensing temperature 30°C were used as nominal conditions. Analogical calculations were also done for conventional ORC system working with the same parameters and providing the same electric power output. As a result of the analysis, it turned out that for the proposed device, the best refrigerant is ammonia, for which the 5.68% efficiency was obtained. Comparing the results with the achievements for a standard ORC system, it is almost twice lower (for ammonia ORC systems, the efficiency is 9.93). However, in the experimental installation it was decided to use R245fa refrigerant, because unlike ammonia it is not flammable and is not toxic. According to the analysis, the R245fa was one of the best refrigerants in terms of efficiency and, what is the most important, it did not mix with the glycerol used in the tanks as a secondary fluid. Theoretical analysis for fluid R245fa has shown that the total energy efficiency coefficient for whole proposed system will be at level 4.72%, while as a result of the experiment, it managed to achieve an efficiency level of 2.51%.
1. Introduction The ORC is an unconventional and very promising technology for electricity generation that has been gaining increasing popularity [1–4]. ORC-based systems employ the steam Rankine 1
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
AC C
EP
TE D
M AN U
SC
RI PT
cycle except that instead of water, organic or inorganic fluids characterized by a low boiling point are used as the working fluid [3,5]. Despite numerous advantages, including high latent and specific heat, chemical stability in a wide range of temperatures, low viscosity, non-toxicity, non-flammability, and availability, water has a peculiar disadvantage, namely a high normal boiling point of 100°C, which renders water unusable in practice in low-temperature systems [6–9]. On the other hand, organic refrigerants are characterized by a low normal boiling point, often below 0°C [10]. There is a lot of research studies on selecting an optimum fluid for use in certain conditions [11–15]. In Ref. [16], the effect of working fluid properties on the irreversibility of a cycle supplied by a 10 MW waste heat source at 600 K is analyzed. Among the examined refrigerants the R113 and R123 were characterized by the highest cycle irreversibility which, at the same time, decreased at a lower rate with increasing evaporation pressure compared to the other fluids. Studies presented in Ref. [17] focus on the effect of the evaporation pressure of six chosen refrigerants on the power output of a Rankine cycle supplied from a heat source with temperature varying from 80°C to 200°C and the condensation temperature of 20°C. It was found that an increase in pressure leads to an increase in the system power output to the specified maximum value for all the fluids. The highest power output was achieved for R227ea between 80°C and 160°C, and for R245fa between 160°C and 200°C. In Ref. [6], Mikielewicz recommends some working fluids for use in residential ORC systems for generating 4 kW of heat and electricity. Among proposed fluids at the turbine inlet temperature of 170°C and the condensation temperature of 50°C, R123, R141b and ethanol achieved the highest efficiency. In the research presented in Ref. [18], heat transfer area minimization and heat recovery efficiency served as criteria for comparing the following fluids: R123, R134a, R141b, R142b, R152a, R227ea, R236ea, R236fa, R245ca, R245fa, R600, R600a and n-pentane. Also in this case the results were related to the heat source temperature. R123 performed as the best fluid between 100°C and 180°C, and R141b for heat sources at over 180°C. In Ref. [19], an attempt was made to select a working fluid which for a given supply temperature, varying in the study from 150°C to 250°C and at constant condenser cooling water temperature of 5°C will provide maximum power output. Some of fluids operated in a supercritical cycle. The highest power output between 150°C and 175°C was achieved for R134a, between 175°C and 185°C for R142b, between 185°C and 210°C for R245fa, between 210°C and 225°C for R245ca, between 225°C and 230°C for R365mfc, and between 230°C and 250°C for R123. In turn, Schilling et al. concludes that the choice of the refrigerant should be directly connected to the ORC installation design [20]. In this case, interventions resulting from subsequent design will be avoided. Many studies have examined the influence of the properties of the single ORC component working fluids as well. Table 1 summarizes refrigerant recommendations depending on the temperature of the heat source. In addition to single-component working fluids, mixtures can also be used in ORCs. Such fluids make it possible to match the fluid to the heat source more effectively and to develop a substance exhibiting desirable properties, taking into account environmental and safety requirements. The temperature of heat sources that supply systems often varies. A characteristic of single-component substances is a constant phase change temperature. This is not the case with multi-component, zeotropic fluids, which in the process of phase change are subject to a so-called temperature glide [20–22]. As the fluid temperature changes when heat is supplied, the temperature profile of the refrigerant can be better aligned to that of the heat source [23]. In this way, heat and exergy losses are lower, while the system efficiency is higher [24]. Such positive consequences are possible only for a certain composition of the zeotropic mixture and thus for a particular temperature glide [25] . Zeotropic refrigerants were also investigated in a number of studies. Liu et al. [26] analyze the effect of the composition of R600a/R601a and R227ea/R245fa mixtures on the efficiency of 2
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
AC C
EP
TE D
M AN U
SC
RI PT
an ORC supplied by variable-temperature geothermal energy, with cooling water at 20°C as the lower heat source. For the molar ratio of 90% of R600a in an R600a/R601a mixture, an increase of 8% in efficiency was achieved compared to the pure fluids. The maximum efficiency with an R227ea/R245fa mixture occurred for the molar ratio of 80% of R227ea in the mixture, and was higher by 15% compared to a system employing a single-component refrigerant. The increase in efficiency compared to single-component refrigerants was achieved only for temperatures below 120°C; above that value, R227ea and R600a provided higher efficiency. In Ref. [21], an ORC system with pentane/hexane and R245fa/R365mfc mixtures was studied. The increase in efficiency for high source temperature varying from 150°C to 250°C was from 6% to 16%, while electric power increased by 20%. It was also demonstrated that using three-component mixtures in a system has no significant impact on improving the efficiency. In Ref. [24], the effect of the temperature glide of R600/R601, R600/R601a, R600a/R601, and R600a/R601a mixtures on the operation of an ORC system was studied. The results has shown that the increase in the efficiency of a system using mixtures compared to a system employing single-component refrigerants is possible only for a temperature glide that is lower than a change in the temperature of the fluid supplied to the evaporator. Similar conclusions were presented in Ref. [25], which demonstrated that the temperature of the heat source determines the selection of the mixture. As a results from the research is fact that for a constant temperature range the efficiency and output power of a system employing single-component refrigerants are higher than those of a system working with mixtures. Another approach for choosing the correct mixture as a working medium was proposed by Oyewunmi and Marides in the publication [27] who as a criterion applied the cost of the construction of the installation and its subsequent operation. In their studies, it was shown that the installation driven by heat source of temperature 98°C from the thermodynamic point of view should be used mixtures of 50% n-pentane and 50% n-hexane or 60% R-245fa and 40% R-227ea. However, for clean components, the installation cost is 14% cheaper. The other results were obtained by Le, Kheiri et al. [28], whose economic analysis has shown that the use of pure refrigerant R245fa is the least viable. The theoretical analysis assumes that the system generates 1 kW of electric power. And on the basis of this input data, the other elements of the system were selected. Unamba C.K. et al. for a typical ORC installation with an electric output power of 1 kW, used the R245fa as refrigerant obtained a thermal efficiency of the system from 0 to 6% depending on the working conditions [29]. This work propose a new ORC design. In the proposed installation, the gas expansion turbine is replaced with an innovative solution of secondary fluid tanks and a liquid turbine. The main purpose of the research is the selection of the most appropriate working medium. Ten single component refrigerant was analyzed. Mixtures were not analyzed in this work due to the risk of the mixture separation in glycerol (secondary fluid). The influence of each working fluids properties on the performance of the proposed apparatus is discussed and the final guidelines for optimal selection are presented as well. Similar calculations were also performed for conventional ORC system that provides the same power output and works with the same refrigerants and under the same conditions as the proposed installation. The selected refrigerant has been tested experimentally.
2. Description of the proposed experimental installation The paper aims to select an optimal working fluid for the proposed experimental installation schematically depicted in Fig. 1. Although the working fluid in the cycle of the system under consideration is subject to the same changes as in the ORC, i.e. isobaric evaporation, expansion, isobaric condensation, and liquid compression, essential differences can be noticed 3
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
AC C
EP
TE D
M AN U
SC
RI PT
[26,29]. An expansion device, which in classic ORC systems can be a turbine or a piston, or screw expander, is replaced with two cylindrical tanks filled by intermediate liquid - glycerol. The generator in the proposed installation is not driven at the expense of a drop in the enthalpy of the refrigerant in the turbine, but following a change in the height of liquid columns in tanks. This solution has been proposed due to the fact that water turbines are usually characterized by higher efficiency than gas turbines (especially those used in ORC systems). As with a conventional ORC system, the system under consideration can be driven by low-temperature heat source. Then the heat carrier, at a temperature higher than that of the organic refrigerant, flows through the evaporator (1) and makes the refrigerant to evaporate. The operation cycle of the device starts when the level of the glycerol is maximum in one of the tanks (5.2) and minimum in the second one (5.1). When a liquid level sensor indicates the extreme level of the liquid in the tanks, valves A open, while valves B close. The refrigerant under high pressure, equal to evaporation pressure, flows to tank 5.2. At the same time vapour, which filled the tank in the previous cycle, starts to flow out of tank 5.1 to a volume of low pressure, equal to the condensation pressure. While the pressure difference occurs, the secondary fluid flows from the high-pressure tank (5.2) to the low-pressure one (5.1), performing work in the turbine (T) located between them. When, as a consequence of pressure exerted by the organic refrigerant on the liquid stored in the high-pressure tank, the intermediate liquid in the low-pressure tank is at its maximum level, the inlet valves switch positions: the A valves close, while the B valves open. Then the glycerol flows from tank 5.1 to tank 5.2, permanently driving the turbine (T). The direction of flow of the intermediate fluid through the turbine will not change, since the turbine’s inlet and outlet connections are interconnected with a collecting manifold. Valve positions will switch each time the glycerol attains extreme levels, thus providing the flow through the pump and an on-going electricity generation in the generator (G). Further thermal process of the refrigerant are the same as in a classic ORC. The refrigerant flows out of the low-pressure tank, it reaches the condenser (4), where it undergoes a phase change, giving heat out of the system. From the condenser the refrigerant is again fed to the high-pressure part of the set-up via the pump (9) located in the liquid line. The proposed experimental installation is equipped with necessary accessories to control the operation. Tank 6 serves as a so-called supercharging device that boosts the liquid flow through the turbine if the generator load increases. Also in this case, the flow is forced by the organic refrigerant under high pressure. The component marked as 7 is a liquid refrigerant tank and at the same time it works as a secondary fluid separator. The correct level of fluids in the tanks is provided by the expansion vessel and the pump (10). Operating parameters, i.e. pressure, temperature, and flow, are controlled by pressure, temperature, and level transmitters, and flow meters.
3. Modelling approach Basic thermodynamic transformations for the proposed ORC system are presented in Fig. 2. In order to determine the basic parameters of the experimental system, it was additionally assumed that system is in a steady state. Pressure drops in the evaporator, condenser and other elements of the installation are negligible, also the heat losses to the environment are negligible. Refrigerant flows into the tank and exerts an even pressure on the surface of the secondary fluid. Secondary fluid does not mix and does not dissolve with the refrigerant. The efficiency of all items including turbines and pumps are known: • isentropic turbine efficiency: = 0.92; • mechanical efficiency of the turbine gear: = 0.95; • electrical efficiency of the generator: = 0.93; 4
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
• • • • •
power output system efficiency: = 0.98; refrigerant expansion efficiency in the tanks: = 0.92; isentropic pump efficiency: = 0.85; pump gear efficiency: = 0.95; electrical efficiency of the pump engine: = 0.98. Operation cycle starts from position 1 (Fig. 2.). The power Np required for increasing the pressure from condensation pC to evaporation pE level is equal to: ℎ(
!)
− ℎ$ %
where pump efficiency: = ∙ ∙
(1)
RI PT
=
(2)
ℎ( !) − ℎ$ %
M AN U
ℎ = ℎ$ +
SC
while the enthalpy at point 2, at the end of the pumping process, is equal to:
(3)
Next step in cycle is evaporation process – line 2-3 in Fig. 2. The heat flux supplied in this process () must firstly warms up the refrigerant to the phase change temperature and then evaporates it. Required heat flux is described as: () = (),, + (),- = (ℎ. − ℎ )
(4)
TE D
where (),, is heat flux needed from warimg up refrigerant and (),- is heat flux needed for evaporating: (5)
EP
(),, = (ℎ/ − ℎ ) (),- = (ℎ. − ℎ/ )
(6)
AC C
For the classic ORC system, the next step in cycle is the expansion in the turbine, in which the work is equal to: 012 = (ℎ. − ℎ3 )
(7)
the electric power at the terminals of the generator is: = 012
(8)
and total efficiency of the ORC system is equal to: 12 =
− () 5
(9)
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
In the experimental installation (Figure 1), the expansion device is replaced by two tanks filled by a secondary fluid. During operation of the system in steady state, the liquid flow rate through the turbine (T in Fig. 1) is constant, so the volumetric flow of the secondary , caused by the refrigerant at evaporation pressure is equal to refrigerant volumetric fluid 45 flow in point 3 (Fig. 2): = 4. 45
(10)
RI PT
where: 4. = 6.
(11)
SC
The displacement volume of the tanks 47 will therefore be equal to the volume of liquid flowing through the reservoirs at a time corresponding to half cycle of the device:
The time of valves change is equal:
M AN U
89: ;: =$/ 47 = = 45 4
=$/ =
89: ;: 445
(12)
(13)
TE D
The theoretical power PT,th of the turbine and the electric power of the system NEL,2 supplied with liquid flowing from the high-pressure tank to the low-pressure tank will be: (@ − @2 ) ?,!, = 45
(14)
, = A -B C ?!,
(15)
AC C
EP
and the efficiency of the system is expressed as: D =
, − ()
(16)
In both solutions, in the experimental installation and the classic ORC system, the expansion medium flows to the condenser, where heat is transferred out of the system: (2 = (ℎ3 − ℎ$ )
(17)
The end of the expansion of “dry” fluid (difference between dry, wet and isentropic fluids are shown in Fig. 3.), shown in Figure 2, is located in the area of superheated gas. Heat transferred in condenser outside of system can be divided into two heat flux: (2, , - chilling the gas to the condensation temperature and (2, 7 - real condensation process:
6
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
(18) (2, , = (ℎ3 − ℎ3/ ) (2,
7
= (ℎ3/ − ℎ$ )
(19)
M AN U
SC
RI PT
In the case of “wet” fluids operating in circuits without superheated gas, the end of the expansion is located in the area of wet gas. The heat transferred to environment is therefore equal to the heat of condensing the refrigerant from a given degree of dryness to a saturated liquid. For “isentropic” fluids, this heat is equal to the total heat of condensation for a given temperature. The position of the final expansion point in the wet gas area for “wet” refrigerant can contribute to the dropping of organic liquid droplets in the flow tanks. Depending on the type of secondary fluid, the wet compound may dissolve in it. The return of the working fluid to the evaporator in accordance with Figure 1 will be impossible, as there is no installation element responsible for separating the components of the resulting mixture. The oil separator (No. 8 in Figure 1) serves to separate the medium from the intermediate liquid, but the one that has been entrained by the pairs of the compound. The continuous increase in the amount of liquid refrigerant in the flow tanks may contribute to operational problems related to, e.g. with inoperative inlet and outlet valves (A and B in Figure 1). Therefore, in the case of wet substances, it seems necessary to overheat the compound pairs by the value for which the end of the expansion would be on the saturation line (Figure 2). Then the total amount of heat received from the upper source will be equal to the sum of the heat of heating the liquid to the saturation state (),, , heat of evaporation (),- and heat of overheating (), : (20)
TE D
() = (),, + (),- + (),
4. Results and analysis
AC C
EP
In the first step of the analysis, it was decided to analytically select the refrigerant most suited to the proposed installation. The results of the theoretical analysis were compared with the results for a typical ORC installation working under the same conditions. Calculations for the proposed experimental set-up were performed for the electric power output of 1 kW. In addition, it was decided that the dimensions of the tanks are in every case the same: • height: ; = 0.5 ; • diameter: 9 = 0.2 . The same simplifications was also enabled in the analysis for a conventional ORC system that provides the same power output and works with the same refrigerants and under the same conditions as for the proposed installation. For both the solutions basic set-ups were analyzed. In line with the description, the proposed installation operation is similar to the operation of the ORC with basic processes shown in Fig. 2. Nominal operating parameters for the calculations are as the following: • evaporation temperature: 80°C; • condensation temperature: 30°C. The selection of the refrigerant for a given temperature range is sometimes based on the vapour saturation curve in the T-s diagram [1,30–33]. In the group of refrigerants considered, 7
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
AC C
EP
TE D
M AN U
SC
RI PT
in nominal operating conditions of the set-up, R134a, R152a, R290, and R717 are wet fluids, while R227ea, R2236fa, R245fa, and R600a are dry fluids (Tab. 2). R1234yf and R1234ze(E) can be considered as isentropic fluids, even though their final expansion point is not exactly on the vapour saturation line. In the group of fluids under consideration, all the wet refrigerants are characterized by higher working pressures than the dry ones. Ammonia, propane, and R134a (Tab. 2) are characterized by the highest evaporation and condensation pressures in nominal operating conditions. On the other hand, R245fa, R236fa, and R600a are refrigerants with the lowest evaporation and condensation pressures. Although R245fa features the lowest values of working pressures and pressures ratio is the highest; for R290 the pressures ratio is the lowest. It should be noted that in none of the cases the condensation pressure is lower than the atmospheric pressure, which will protect the system from the possibility of moisture condensation due to leakages. Ammonia has the highest heat of evaporation and condensation (Tab. 2); these values are almost three times higher than for the other fluids. The lowest heat of phase change in the temperature range considered is observed for R227ea. One quantity that is essential for the proposed installation is the specific volume of vapours of the refrigerant flowing into the flow tanks. It is the value that determines the volumetric flow rate of the glycerol flowing through the expansion device. Among the organic substances considered, ammonia R717 features the highest specific volume of saturated vapour under the evaporation pressure. Slightly lower values of over 20 dm3/kg are also observed for R600a and R245fa. R134a, R227ea, and R1234yf are fluids with the smallest specific volume of vapour, slightly over 5 dm3/kg. The volumetric flow rate can also be expressed as a ratio of the system power output (1 kWe in the considered case) to the difference between evaporation and condensation pressures. The larger the pressure difference for a constant power output of the system, the lower the volumetric flow rate of the intermediate liquid is required. In the group of considered fluids, a system employing ammonia requires the lowest flow rate of the liquid, since the pressure difference for ammonia is the largest. As with working pressures, R290 and R134a (Tab. 2) are next and require a higher flow rate of slightly over 0.5 l/s. On the other hand, if refrigerants with the smallest difference between evaporation and condensation pressures (R245fa, R236fa, and R600a) are used in a system, the highest volumetric flow rate is required (Tab. 2). According to the pressure values, a relation can be observed such that the volumetric flow rate of the wet refrigerants is lower than that of the dry ones. The time after which valves are switched is inversely proportional to the volumetric flow rate of the refrigerant vapour under evaporation pressure, which is why for fluids featuring a high volumetric flow rate the switching time will be short, while for those with a low volumetric flow rate the switching time will be longer (Tab. 2). The values of the specific volume of saturated vapour under evaporation pressure, the evaporation and condensation pressures, and the system power output determine the mass flow rate of the refrigerant. This parameter is directly proportional to the power output and inversely proportional to the product of the pressure difference and the specific volume of vapour. In the group of considered fluids, the lowest flow rate is required in a system based on ammonia. Next is a system using isobutane (R600a), with one of the lowest differences between working pressures and, on the other hand, one of the highest specific volumes of saturated vapour. Third is propane (R290) which has quite a large pressure difference and an average specific volume of saturated vapour. The highest mass flow rate is required in a system using R227ea, which is a consequence of its low specific volume of saturated vapour. Other fluids with a relatively low mass flow rate include R1234yf and R236fa. Despite the smallest difference between working pressures among all the analysed fluids, and due to a relatively high specific 8
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
AC C
EP
TE D
M AN U
SC
RI PT
volume of saturated vapour, R245fa is a refrigerant requiring an average mass flow rate compared to the other fluids. In each case the mass flow rate of a refrigerant in the proposed experimental installation is higher than the mass flow rate of the refrigerant in a conventional ORC system that provides the same electric power output as the system considered. For all fluids, this value is almost twice as high as in the classic ORC system. Minimum and maximum flow rates in both the solutions are achieved for the same working fluids: R717 (the lowest one) and R227ea (the highest one). The efficiency of both the proposed set-up and the classic ORC system is equal to the ratio of the electric power output less the power to drive the refrigerant pump to the heat supplied. The system efficiency increases when the pump power and the amount of heat supplied to the system are smaller. Among the selected organic fluids, the least power of the fluid pump is required by R245fa (the smallest pressure difference and a relatively low volumetric flow of the liquid refrigerant through the pump) and R717 (the lowest rate of volumetric flow of the liquid through the pump and the largest pressure difference). Such a low value of the rate of volumetric flow of the fluid through the pump and the small pressure difference in the case of R245fa are crucial for the pump driving power. The highest power of the refrigerant pump is required in a system using R290 and R1234yf, i.e. refrigerants with a high volumetric flow rate of the liquid. These are also fluids with the highest temperature at the end of the pumping process (Tab. 2). The high value of energy needed to drive the pump may result from the fact that it is partly used for heating the liquid. The least amount of heat drawn from the heat source is a characteristic of ammonia, which has the highest heat of evaporation compared to the other organic fluids. The small amount of heat is a consequence of a very low mass flow rate of R717. It is followed by R600a, R290, and R152a, i.e. refrigerants with a relatively high heat of phase change. On the other hand, R227ea, characterized by the highest mass flow rate and the lowest heat of evaporation, requires the largest amount of heat to be supplied to the system (Tab. 2). As with the mass flow rate, both the pump driving power and the heat drawn from the upper heat source are higher in the proposed installation than in the ORC system. The maximum and minimum pump driving power in the proposed set-up and in the ORC system is achieved again for the same refrigerants: R290 (the maximum power) and R245fa (the minimum power). The minimum and maximum values of heat supplied from the upper heat source also occur for the same fluids. Due to a low required power for the liquid refrigerant pump and a small amount of heat provided from the heat source, among all the considered fluids it is ammonia that guarantees the highest efficiency of both the proposed set-up and the conventional ORC system (Fig. 4). It is followed by isobutane in terms of efficiency of the proposed installation Despite a relatively small pressure difference and due to a high specific volume of saturated vapour and thus a small mass flow rate and a small amount of heat drawn, R600a provides the efficiency of 4.87%. In the case of the conventional ORC system, second best efficiency, equal to 9.50%, is achieved with R245fa. This is caused by a lower power of the refrigerant pump and the same amount of heat drawn from the heat source compared to isobutane. R152a provides a slightly lower efficiency than that of the proposed installation using isobutane. R245fa takes the fourth place. R1234yf is characterized by the lowest efficiency. For this fluid, the required power of the liquid refrigerant pump is the highest. Almost the same efficiency is achieved with R227ea, which requires the largest amount of heat to be supplied from the heat source. Also in the case of the conventional ORC system, R227ea and R1234yf provide minimum efficiencies (Fig. 4). The differences between corresponding values of efficiency for both the solutions are small and amount to up to 2 percentage points. For each refrigerant, the efficiency of the 9
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
RI PT
conventional ORC system is higher than that of the proposed set-up. The largest difference in terms of the efficiency of the proposed set-up occurs for R1234yf (nearly 130%) and R134a (113%), while the smaller one for ammonia (70%) and isobutane (90%). In terms of energy efficiency, the best refrigerant was ammonia. Isobutane was on the second place and R152a on the third. Due to their flammability, these refrigerants were eliminated from the experimental part of research. Therefore, the refrigerant R245fa was used for the experiment. The installation presented in figure 1 was filled with refrigerant R245fa. During operation, the device maintained in the evaporator boiling pressure corresponding to 80°C, and in the condenser the pressure corresponding to 30°C of saturation. At the turbinegenerator assembly, a flywheel was mounted, which enabled operation at a constant speed even when the tanks were switched. For these conditions thermal efficiency of 2.51% was obtained. Achieved electric power output was at level 800 W.
SC
5. CONCLUSIONS
AC C
EP
TE D
M AN U
In this study the selection of refrigerants for a modified ORC as well as the influence of their properties on the performance of the proposed installation are described. Ten different working fluid were examined in terms of the thermal efficiency for proposed apparatus. The achieved results were also compared with an operation of a typical ORC system that provides the same power output and works with the same refrigerants and under the same conditions as the proposed installation. ORC systems with liquid turbines turned out to have lower efficiency than systems with gas turbines. Among the refrigerants analyzed, in nominal operating conditions of the proposed installation (Te=80°C, Tc=30°C), the maximum efficiency is provided for ammonia (R717), which is characterized by both the highest specific volume of saturated vapour flowing into the tanks and the largest difference between working pressures. For all the organic fluids examined, these two quantities are critical to the efficiency of the system, since they determine the refrigerant mass flow rate, the feed pump power, and the amount of heat drawn from the lower heat source. Selecting a suitable organic refrigerant from the group of the proposed fluids can lead to a 57% increase in efficiency compared to a refrigerant for which the system was the least efficient. Nevertheless, irrespective of the properties of an organic fluid, the efficiency of the system is very low, from 3.62% for the worst refrigerant (R1234yf) to 5.68% for the best one (R717). In the case of the other substances, the efficiencies are between 4% and 5%. The values are nearly twice as high for the same fluids used in the same conditions in the classic ORC system, where the efficiency for the best fluid (again ammonia) amounts to 9.93% and for the worst one (again R1234yf) 8.38%. The R245fa refrigerant was used in the experimental part because refrigerants characterized by better energy efficiency R717, R600a and R152a are flammable. For refrigerant R245fa thermal efficiency of 2.51% was obtained. Despite the fact that installation was designed to generate 1000W, it was experimentally achieved a maximum electric output of 800W.
Acknowledgements The presented results were financed from the European Regional Development Fund under the Intelligent Development 2014-2020 Program. Project entitled "An innovative thermal power plant with a liquid turbine and a hydraulic-gas generator that utilizes a waste heat flow" implemented as part of the National Center for Research and Development contest: Szybka Ścieżka. 10
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
References Chen H, Goswami DY, Stefanakos EK. A review of thermodynamic cycles and working fluids for the conversion of low-grade heat. Renew Sustain Energy Rev 2010;14:3059–67. doi:10.1016/j.rser.2010.07.006.
[2]
Hung TC. 98/01752 A review of organic Rankine cycles (ORCs) for the recovery of low-grade waste heat: Hung, T. C. et al. Energy, 1997, 22, (7), 661–667. Fuel Energy Abstr 1998;39:151-. doi:http://dx.doi.org/10.1016/S0140-6701(98)97894-8.
[3]
Kajurek J, Rusowicz A, Grzebielec A, Bujalski W, Futyma K, Szczęśniak A, et al. Influence of working fluid properties on performance of organic Rankine cycle. Rynek Energii 2017:68–79.
[4]
Chacartegui R, Sánchez D, Muñoz JM, Sánchez T. Alternative ORC bottoming cycles FOR combined cycle power plants. Appl Energy 2009;86:2162–70. doi:10.1016/j.apenergy.2009.02.016.
[5]
Kajurek J, Rusowicz A, Grzebielec A. The Influence of Stack Position and Acoustic Frequency on the Performance of Thermoacoustic Refrigerator with the Standing Wave. Arch Thermodyn 2017;38. doi:10.1515/aoter-2017-0026.
[6]
Mikielewicz D, Mikielewicz J. A thermodynamic criterion for selection of working fluid for subcritical and supercritical domestic micro CHP. Appl Therm Eng 2010;30:2357–62. doi:10.1016/j.applthermaleng.2010.05.035.
[7]
Quoilin S, Broek M Van Den, Declaye S, Dewallef P, Lemort V. Techno-economic survey of organic rankine cycle (ORC) systems. Renew Sustain Energy Rev 2013;22:168–86. doi:10.1016/j.rser.2013.01.028.
[8]
Saleh B, Koglbauer G, Wendland M, Fischer J. Working fluids for low-temperature organic Rankine cycles. Energy 2007;32:1210–21. doi:10.1016/j.energy.2006.07.001.
[9]
Tchanche BF, Lambrinos G, Frangoudakis A, Papadakis G. Low-grade heat conversion into power using organic Rankine cycles - A review of various applications. Renew Sustain Energy Rev 2011;15:3963–79. doi:10.1016/j.rser.2011.07.024.
AC C
EP
TE D
M AN U
SC
RI PT
[1]
[10] Grzebielec A. Termodynamiczne podstawy przenoszenia ciepła w termoakustycznych urządzeniach chłodniczych. Chłodnictwo 2009;44:12–6. [11] Roy JP, Mishra MK, Misra A. Parametric optimization and performance analysis of a waste heat recovery system using Organic Rankine Cycle. Energy 2010;35:5049–62. doi:10.1016/j.energy.2010.08.013. [12] Shokati N, Ranjbar F, Yari M. Comparative and parametric study of double flash and single flash/ORC combined cycles based on exergoeconomic criteria. Appl Therm Eng 2015;91:479–95. doi:10.1016/j.applthermaleng.2015.08.031. [13] Zhai H, An Q, Shi L, Lemort V, Quoilin S. Categorization and analysis of heat sources for organic Rankine cycle systems. Renew Sustain Energy Rev 2016;64:790–805. doi:10.1016/j.rser.2016.06.076. 11
ACCEPTED MANUSCRIPT
Jakub Kajurek, Artur Rusowicz, Andrzej Grzebielec, Wojciech Bujalski
[14] Papadopoulos AI, Stijepovic M, Linke P, Seferlis P, Voutetakis S. Multi-level Design and Selection of Optimum Working Fluids and ORC Systems for Power and Heat Cogeneration from Low Enthalpy Renewable Sources. Comput Aided Chem Eng 2012;30:66–70. doi:10.1016/B978-0-444-59519-5.50014-9. [15] Bao J, Zhao L. A review of working fluid and expander selections for organic Rankine cycle. Renew Sustain Energy Rev 2013;24:325–42. doi:10.1016/j.rser.2013.03.040.
RI PT
[16] Hung T-C. Waste heat recovery of organic Rankine cycle using dry fluids. Energy Convers Manag 2001;42:539–53. doi:10.1016/S0196-8904(00)00081-9. [17] Lakew AA, Bolland O. Working fluids for low-temperature heat source. Appl Therm Eng 2010;30:1262–8. doi:10.1016/j.applthermaleng.2010.02.009.
SC
[18] Wang ZQ, Zhou NJ, Guo J, Wang XY. Fluid selection and parametric optimization of organic Rankine cycle using low temperature waste heat. Energy 2012;40:107–15. doi:10.1016/j.energy.2012.02.022.
M AN U
[19] Maizza V, Maizza A. Unconventional working fluids in organic Rankine-cycles for waste energy recovery systems. Appl Therm Eng 2001;21:381–90. doi:10.1016/S13594311(00)00044-2. [20] Schilling J, Gross J, Bardow A. Integrated design of ORC process and working fluid using process flowsheeting software and PC-SAFT. Energy Procedia, vol. 129, 2017, p. 129–36. doi:10.1016/j.egypro.2017.09.184. [21] Chys M, van den Broek M, Vanslambrouck B, De Paepe M. Potential of zeotropic mixtures as working fluids in organic Rankine cycles. Energy 2012;44:623–32. doi:10.1016/j.energy.2012.05.030.
TE D
[22] Rusowicz A, Grzebielec A. Legal and technical aspects of replacement refrigerants in refrigeration and air conditioning. J Civ Eng Environ Archit 2015;62:359–68. [23] Li Y-R, Du M-T, Wu C-M, Wu S-Y, Liu C. Potential of organic Rankine cycle using zeotropic mixtures as working fluids for waste heat recovery. Energy 2014;77:509–19. doi:10.1016/j.energy.2014.09.035.
EP
[24] Liu Q, Duan Y, Yang Z. Effect of condensation temperature glide on the performance of organic Rankine cycles with zeotropic mixture working fluids. Appl Energy 2014;115:394–404. doi:10.1016/j.apenergy.2013.11.036.
AC C
[25] Zhao L, Bao J. Thermodynamic analysis of organic Rankine cycle using zeotropic mixtures. Appl Energy 2014;130:748–56. doi:10.1016/j.apenergy.2014.03.067. [26] Heberle F, Preißinger M, Brüggemann D. Zeotropic mixtures as working fluids in Organic Rankine Cycles for low-enthalpy geothermal resources. Renew Energy 2012;37:364–70. doi:10.1016/j.renene.2011.06.044. [27] Oyewunmi OA, Markides CN. Thermo-Economic and Heat Transfer Optimization of Working-Fluid Mixtures in a Low-Temperature Organic Rankine Cycle System. Energies 2016;9:448. doi:10.3390/en9060448. [28] Le VL, Kheiri A, Feidt M, Pelloux-Prayer S. Thermodynamic and economic optimizations of a waste heat to power plant driven by a subcritical ORC (Organic Rankine Cycle) using pure or zeotropic working fluid. Energy 2014. doi:10.1016/j.energy.2014.10.051. [29] Unamba CK, White M, Sapin P, Freeman J, Lecompte S, Oyewunmi OA, et al. Experimental Investigation of the Operating Point of a 1-kW ORC System. Energy 12
ACCEPTED MANUSCRIPT
XXIII Zjazd Termodynamików 2017, 19 - 22 września 2017, Beskid Śląski
Procedia, vol. 129, 2017, p. 875–82. doi:10.1016/j.egypro.2017.09.211. [30] Feng Y, Hung TC, Greg K, Zhang Y, Li B, Yang J. Thermoeconomic comparison between pure and mixture working fluids of organic Rankine cycles (ORCs) for low temperature waste heat recovery. Energy Convers Manag 2015;106:859–72. doi:10.1016/j.enconman.2015.09.042.
RI PT
[31] Hærvig J, Sørensen K, Condra TJ. Guidelines for optimal selection of working fluid for an organic Rankine cycle in relation to waste heat recovery. Energy 2016;96:592–602. doi:10.1016/j.energy.2015.12.098. [32] He C, Liu C, Gao H, Xie H, Li Y, Wu S, et al. The optimal evaporation temperature and working fluids for subcritical organic Rankine cycle. Energy 2012;38:136–43. doi:10.1016/j.energy.2011.12.022.
AC C
EP
TE D
M AN U
SC
[33] He C, Liu C, Zhou M, Xie H, Xu X, Wu S, et al. A new selection principle of working fluids for subcritical organic Rankine cycle coupling with different heat sources. Energy 2014;68:283–91. doi:10.1016/j.energy.2014.02.050.
13
ACCEPTED MANUSCRIPT Table 1. Recommended refrigerants Recommended refrigerant R32, R125, R134a, R143a, R124, R227ea, R290, R1234yf, R1270 R114, R141b, R123, R124, R245fa, R601a, R1243ze R123, R141b, R1234ze, RC218, R236fa, R236ea, R600, R601 benzene, paraxylene, toulene, hexane
AC C
EP
TE D
M AN U
SC
RI PT
Source temperature [°C] <100 100-120 120-160 160-200 >200
1
Wet
A2
133
0
Dry
A1
3580
0
Heat of condensation qk, kJ/kg 1.32
1.35
For temperatures between the condensation point and the evaporation point.
0.76
0.67
Volumetric flow rate of refrigerant V3, l/s 0.94
41.6
1.35
42.7
30.5
8.18 11.06 19,24 17.24 0.205 4.13 23.5
Heat of evaporation Qp, kW
Heat drawn from the upper source QGZC, kW
Heat delivered to the lower source QDZC, kW
Driving power of the refrigerant pump PPC, kW
Efficiency of the experimental Installation ηD , %
Valve switching time τ, s
10.68
20.7
4.79
16.7
3.65
0.146 0.202
15.89 19.71
17.84 21.88
11.77 11.20
6.07
11.7
4.30
0.105
18.63
20.82
12.69
8.13
0.1039 0.0618 0.1652 0.1206
Heat for heating up liquid Qdp, kW
Mass flow rate of refrigerant mR, kg/s
30
30
Refrigerant post-expansion temperature T4, °C
30.9
1
31.1
1
31.3
0.941
Refrigerant post-compression temperature T2, °C
0.91
0.959
0.91
Vapour dryness fraction at the end of expansion
1.27
0.95
1.40
30
31.9
0.948
0.0118
2.05
3.12
326.71
189.8
1079
3139.9
Wet
A3
3
0
R290
7.7
4.72
0.053
17.83
20.04
13.76
6.28
0.0902
2.05
873.96
1167.2
4142
Wet
B2L
0
0
R717
25.8
4.08
0.275
15.94
17.76
9.76
8.00
11.7
4.87
0.104
16.39
18.38
12.08
6.30
37.4
5.68
0.090
14.17
16.03
12.44
3.59
21.8
3.62
0.265
18.28
20.30
10.18
RI PT
0.72
30
31.4
0.998
0.0055
1.06
1.61
10.12
0.42
30
31.0
0.87
0.0295
3.19
4.78
141.24
78.36
783.5
2519.4
Isentropic
A2L
4
0
R1234yf
0.1297
0.0479
1.34
40.3
30.6
1
0.028
1.87
2.66
323.33 1144.41
252.58
404.7
1343.8
Dry
A3
4
0
R600a
0.0142
0.0515
0.61
SC
40.5
30.3
1
0.0227
M AN U
1.99
0.0064 0.0123 0.0057 0.0112
1.03
1.55
Specific volume of saturated vapour v3, m3/kg
Average specific heat of vapour cp for ps, kJ/(kg·K)
2
187.33
173.1 273.16 108.71 142.95
177.8 152.51
321
789.3
Dry
B1
1050
0
R245fa
105.29
106.42 190.28 67.81
Heat of evaporation qp, kJ/kg
689.8 528.4
770.2
Condensation pressure pk, kPa
Average specific heat of liquid 1 cw for pp, kJ/(kg·K)
Dry
A1
9820
0
2633.2 2342.4 1858.3 1249.1
TE D
Evaporation pressure pp, kPa
Type of refrigerant
EP
A1
Safety group Wet
1430
GWP
AC C 0
R134a R152a R227ea R236fa
ODP
Quantity
17.9
4.26
0.166
17.53
19.58
11.56
8.02
0.1057
0.88
32.1
32.6
1
0.0083
0.99
1.51
163.05
110.19
578.3
2007.7
Isentropic
A2L
6
0
R1234ze(E)
ACCEPTED MANUSCRIPT Table 2. Basic cycle parameters
AC C
EP
TE D
M AN U
SC
RI PT
ACCEPTED MANUSCRIPT
Fig. 1. Schematic illustration of the proposed experimental installation.
ACCEPTED MANUSCRIPT R227ea 2
20,0
2'
3
RI PT
10,0 9,0 8,0 7,0 6,0
1
4
SC
5,0
3,0
160
180
200
220
240
M AN U
4,0
260
280
300
320
340
360
AC C
EP
TE D
Enthalpy, kJ/(kg K) Fig. 2. Pressure-enthalpy diagram for a theoretical R227ea cycle
Fig. 3. ORC diagram for dry, wet and isentropic fluids
380
400
ACCEPTED MANUSCRIPT Experimental
ORC
8.38
8.93
9.93
9.27
8.57
9.50
9.00
8.40
8.80
10
4.26
3.62
5.68
4.87
4.08
4.72
RI PT
4
4.30
3.65
6
4.79
8 4.13
Efficiency, %
9.28
12
2
SC
0
AC C
EP
TE D
M AN U
Fig. 4. Theoretical efficiency of the proposed installation vs. typical ORC installation