Comparative performance analysis of low-temperature Organic Rankine Cycle (ORC) using pure and zeotropic working fluids

Comparative performance analysis of low-temperature Organic Rankine Cycle (ORC) using pure and zeotropic working fluids

Applied Thermal Engineering 54 (2013) 35e42 Contents lists available at SciVerse ScienceDirect Applied Thermal Engineering journal homepage: www.els...

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Applied Thermal Engineering 54 (2013) 35e42

Contents lists available at SciVerse ScienceDirect

Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng

Comparative performance analysis of low-temperature Organic Rankine Cycle (ORC) using pure and zeotropic working fluids S. Aghahosseini*, I. Dincer Faculty of Engineering and Applied Science, University of Ontario Institute of Technology, 2000 Simcoe Street North, Oshawa, Ontario, Canada L1H 74K

h i g h l i g h t s < Combined energy and exergy analysis is conducted for Organic Rankine Cycle. < Comparative assessment is performed for different pure and zeotropic working fluids. < Exergy and energy efficiency, cycle irreversibility, and required external heat are analyzed. < Toxicity, flammability, ODP and GWP of considered working fluids are studied. < Environmental benefits of the renewable/waste heat-based ORC are investigated.

a r t i c l e i n f o

a b s t r a c t

Article history: Received 15 October 2012 Accepted 15 January 2013 Available online 4 February 2013

In this paper, a comprehensive thermodynamic analysis of the low-grade heat source Organic Rankine Cycle (ORC) is conducted and the cycle performance is analyzed and compared for different pure and zeotropic-mixture working fluids. The comparative performance evaluation of the cycle using a combined energy and exergy analysis is carried out by sensitivity assessment of the cycle certain operating parameters such as efficiency, flow rate, irreversibility, and heat input requirement at various temperatures and pressures. The environmental characteristics of the working fluids such as toxicity, flammability, ODP and GWP are studied and the cycle CO2 emission is compared with different fuel combustion systems. R123, R245fa, R600a, R134a, R407c, and R404a are considered as the potential working fluids. Results from this analysis provide valuable insight into selection of the most suitable working fluids for power generating application at different operating conditions with a minimal environmental impact. Ó 2013 Elsevier Ltd. All rights reserved.

Keywords: Organic Rankine Cycle Exergy Energy efficiency Zeotropic mixture Irreversibility CO2 emission

1. Introduction Concerns of energy industries have increased over utilization of fossil fuels towards global warming, air pollution and stratospheric ozone depletion. Also, waste heat energy being released from process industries and power plants causes serious thermal pollution [1]. In this context, utilization of the renewable and industrial waste heat for electricity generation has become a significant point of interest. In addition, due to the fact that the thermal efficiency of the conventional steam power generation becomes uneconomically low when the gaseous steam temperature drops below 370  C, using water as a working fluid become considerably less efficient and more costly [2].

* Corresponding author. Tel.: þ1 6472201020. E-mail address: [email protected] (S. Aghahosseini). 1359-4311/$ e see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.applthermaleng.2013.01.028

In recent years, Organic Rankine Cycle (ORC) has become a field of intense research and development as a promising technology for conversion of low-grade heat into useful work and hence electricity. The heat source can be of various origins such as solar radiation [3], biomass combustion [4], geothermal energy [5] or waste heat from process industries [6,7]. Some actual applications have been installed for recovering geothermal and waste heat for power generation in various locations [8,9]. Examples are the plants in Altheim, Austria, with a power production of 1 MW [10] and in Neustadt-Glewe, Germany, with a power production of 0.2 MW [11]. Unlike in the steam power cycle where vapour steam is the working fluid, ORCs employ organic fluids, namely refrigerants or hydrocarbons. Right selection of a working fluid is crucial to achieve higher energetic and exergetic efficiencies. Optimum utilization of the available heat source in different operating conditions involves various trade-offs. Moreover, the organic working fluid must be carefully selected based on safety and environmental properties

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assessment. General criterion such as cycle thermodynamic performance, fluid stability limit, flammability, safety, and environmental impact could be considered to analyze using different working fluids. As an example, utilizing the non-flammable and non-toxic refrigerants is promoted previously as attractive working fluids. R113 and R114 have been also banned because of their ozone layer depletion potential. It should be mentioned that this regulation will include R123 in the near future [12]. Vijayaraghavan and Goswami [13], Badr et al. [14], Hettiarachchi et al. [15], Saleh et al. [16], and Tchanche et al. [17] are some of the researchers who have analyzed the characteristics of different working fluids in various ORC applications. A large number of previous studies regarded mostly to pure components as the working fluid for ORC performance assessment. However, using single working fluid component brings substantial deficiencies. In the most studied applications, the temperatures of a pure working fluid remain constant during evaporating and condensing processes, whereas the temperatures of the heating and cooling sources are changing during the heat transfer process. Consequently, pinch point imposes larger temperature differences leading to higher system irreversibility which considerably decreases the cycle exergy efficiency. In other words, an important limitation of using pure working fluids is the constant temperature of evaporation and condensation that is not suitable for sensible heat sources such as waste heat. In contrast, working fluid mixtures present variable temperature profile during the phase change process, which could considerably reduce the mismatch between heating or cooling sources and the evaporating or condensing working fluid mixtures respectively. So, the cycle overall efficiency could increase noticeably since system irreversibilities can be minimized. Wang and Zhao [18] presented a theoretical analysis of zeotropic mixtures R245fa/ R152a in the low-temperature solar Rankine cycle. Radermacher [19] analyzed the mutual influence of working fluid mixtures properties on the ORC overall performance and suggested simple counter-flow heat exchangers. In the present work, energy and exergy analyses of the lowtemperature heat source ORC are conducted for different pure and zeotropic-mixture working fluids and results are studied and compared. Performance comparison between pure and multicomponent mixtures as a working fluid, which is the key missing part in a majority of the previous works, is also included. The cycle energy and exergy efficiencies, total irreversibility, external heat requirements, and mass flow rate of the potential working fluid are calculated and compared for a 100 kW power generation system. In addition, the environmental characteristics of the working fluids such as toxicity, flammability, ODP (Ozone Depletion Potential) and GWP (Global Warming Potential) are studied and the cycle CO2 emission is compared with different fuel combustion systems.

Fig. 1. Schematic of a typical Organic Rankine Cycle (ORC).

classified into three categories. Those are dry, isentropic, and wet depending on the slope of the cycle Tes diagram to be positive, infinite, and negative respectively. Also ORC can be classified in two groups according to the level of expander inlet pressure, including supercritical ORCs and sub-critical ORCs which is the one investigated in the present study. Figs. 2 and 3 show Tes diagrams of two types of ORC processes with the negative slope of the saturated vapour curves. As it is shown in Fig. 2, the working fluid leaves the condenser as saturated liquid, state point 1. Then, it is compressed by the liquid pump to the sub-critical pressure, state point 2. The working fluid then is heated up in the evaporator until it becomes superheated vapour, state point 3. The superheated vapour flow is then expanded after to the condensing pressure, state point 4. At the condensing pressure, the working fluid lies in the two-phase region. The two-phase fluid passes through the condenser where heat is removed until it becomes a saturated liquid, state point 1. The processes in Fig. 3 are similar to those in Fig. 2 with the only difference being that the state point 4 after expansion lies in the superheated vapour region. Figs. 4 and 5 show Tes diagrams of the other two types of ORC processes with the positive slope of the saturated vapour curves. The state points 1 and 2 are in the same condition as the ORC system in Figs. 2 and 3. Starting from state 2, the working fluid is heated up in the evaporator at constant sub-critical pressure until it becomes saturated, state point 3 in Fig. 3, or it is superheated, state point 3 in Fig. 4. Then, it is expanded to state point 4, which is in the superheated vapour region. The key point for performance analysis of ORCs which have been also presented by Hung [22], Gurgenci [23], Yamamoto et al. [24],

2. Methodology The components of an ORC are essentially similar to the conventional Rankine Cycle which consists of a pump, evaporator, expander and condenser. The working fluid is saturated liquid when passing out the condenser and is then pumped to the evaporator to gain heat from a heating source. Resulting hot pressurized working fluid that could be saturated or superheated expands in the expander and generates useful work. The layout of a typical ORC is shown in Fig. 1. The expander is considered similar to the scroll expander investigated by Zamfirescu and Dincer [20] and Quoilin et al. [21]. The appropriate selection of working fluids for different operation conditions, as it was mentioned earlier, is the most important criteria to system performance. Thermodynamic properties of selected working fluid will affect the system efficiencies and the cycle environmental impact. Technically, the working fluid can be

Fig. 2. ORC with a negative slope of the saturated vapour curve and wet vapour at the expander outlet.

S. Aghahosseini, I. Dincer / Applied Thermal Engineering 54 (2013) 35e42

Fig. 3. ORC with a negative slope of the saturated vapour curve and superheated vapour at the expander inlet.

and Somayaji et al. [25] is that the organic working fluid must be operated at saturated vapour condition before getting into the expander to reduce the total irreversibility of the cycle. The other important issue in ORCs application is the selection of appropriate working fluid to have thermo physical properties that match the application and possess adequate chemical stability at the desired working temperature. Moreover, they should have adequate ODP and GWP values to minimize cycle environmental impacts. GWP, Global Warming Potential, is a measurement, usually over a 100-year period, of how much effect a refrigerant will have on global warming in relation to carbon dioxide, which has a GWP of 1. The lower the value of GWP, the better the refrigerant is for the environment. ODP, Ozone Depletion Potential, is a measure of the relative capability of a particular chemical to destroy ozone. ODP is measured against chlorofluorocarbon R11, which has an assigned potential of 1. The less the value of ODP, the better the refrigerant is for the ozone layer and the environment [26]. The selected organic working fluids for this study are R123, R245fa, R600a, R134a, R407c, and R404a with the boiling points ranging from 47  C to 28  C. It is found that in the temperature range that the low-temperature heat ORC works, usually below 200  C, few pure organic working fluids are isentropic, but the most of them are dry or wet. Wet working fluids include R407c, R404a,

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Fig. 5. ORC with a non-negative slope of the saturated vapour curve and superheated vapour at the expander inlet.

R134a, and R143a while dry working fluids include R245fa, R123, and R600a. Table 1 provides physical properties and Table 2 presents safety and environmental data of the considered working fluids respectively. It should be mentioned that all refrigerants are classified by ASHRAE into safety groups A and B with a lower and higher level of toxicity respectively. They are also categorized according to the level of flammability into three groups from 3 as a highly flammable to 1 with no flame propagation [27]. 3. System modelling For the cycle performance modelling, various operating conditions are analyzed and compared for the mentioned working fluids in order to determine the operating condition that presents the best thermal efficiency with a minimum irreversibility. This evaluation is performed using a combined first and second law of thermodynamics analysis by varying certain system operating parameters at various temperatures and pressures. The process model is developed with the Engineering Equation Solver (EES) software [28]. It is assumed that the cycle reaches the steady-state condition and the pressure drop and heat losses to the environment in the evaporator, condenser, expander and pump are neglected. Because of the thermodynamic irreversibility occurring in each of the components, such as non-isentropic expansion, non-isentropic compression and heat transfer over a finite temperature difference, the exergy analysis methodology is employed to evaluate the practical cycle performance. Considering P0 ¼ 100 kPa and T0 ¼ 25  C to be the ambient pressure and temperature as the reference state. The isentropic efficiencies of the expander and pump are assumed 85% and 80% respectively. The condenser temperature was kept constant at 25  C, while the maximum pressure used for fluids is kept 3000 kPa, close to the working fluids’ critical Table 1 Physical properties of the selected working fluids. Working fluids

Fig. 4. ORC with a non-negative slope of the saturated vapour curve and saturated vapour at the expander inlet.

R123 R245fa R600a (Isobutane) R134a R407c (R134a, R125, R32) R404a (R134a, R125, R143a)

Physical properties Tbp ( C)

Tc ( C)

Pc (MPa)

27.8 25.13 11.7 26.1 43.6 46.39

183.7 174.42 135 101 86.8 72

3.66 3.93 3.64 4.06 4.59 3.71

Tbp: normal boiling temperature; Tc: critical temperature; Pc: critical pressure.

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Table 2 Safety and environmental data of the selected working fluids. Working fluids

Safety data

Environmental data

ASHRAE safety group

ODP

GWP

R123 R245fa R600a (Isobutane) R134a R407c (R134a, R125, R32) R404a (R134a, R125, R143a)

B1 A1 A3 A1 A1 A1

0.02 0 0 0 0 0.04

0.02 1030 4 1610 1800 3300

ODP: Ozone Depletion Potential; GWP: Global Warming Potential.

pressures. The temperature differential of the evaporator and condenser with the cycle are assumed to be constant at 10  C. Mass, energy and exergy balance equations are employed for each component to determine the heat input requirement, the rate of irreversibility, and energy and exergy efficiencies. The steadystate and steady-flow conditions are assumed for all processes. A general mass balance can be expressed in rate form as

X

_ in ¼ m

i

X

_ out m

(1)

i

_ is the mass flow rate, and the subscript “in” and “out” where m stand for inlet and outlet respectively and “i” corresponds to the each process stream. The energy and exergy balance equations can be written as follow,

X

E_ in þ Q_ ¼

i

X i

X

_ E_ out þ W

(2)

i

_ þ Ex in

! X T0 _ _ þ Ex _ _ out þ W 1 Qj ¼ Ex des Tj i

Fig. 7. The variation of ORC energy efficiency with the expander inlet pressure, P3, for zeotropic-mixture working fluids.

The irreversibility rates for the cycle components are assessed using a rearranged form of exergy balance equation and is equal to the component exergy destruction as given below,

_ I_ ¼ Ex des ¼

X i

_  Ex in

X

_ þ _ out  W Ex

i

1

! T0 _ Qj Tj

(5)

The cycle total irreversibility is equal to the sum of its component irreversibility.

I_cycle ¼ I_pump þ I_eva þ I_exp þ I_cond

(6)

The cycle energy efficiency is calculated as follow,

(3)

hcycle ¼

_ net W Q_

(7)

in

where the exergy flow rate of the streams are calculated as below,

_ ¼ m½ðh _ Ex i i  h0 Þ  T0 ðsi  s0 Þ

(4)

_ net can be calculated as below, where the net power output W

_ exp h _ _ net ¼ W W gen  W pump

(8)

where the “h” and “s” are the enthalpy and entropy of the state point “i” respectively and the subscript zero indicates properties at the reference state ambient condition.

Here, hgen denotes generator efficiency. The cycle exergy efficiency is calculated as where the “Teva” is the evaporator temperature.

Fig. 6. The variation of ORC energy efficiency with the expander inlet pressure, P3, for pure working fluids.

Fig. 8. The variation of ORC exergy efficiency with the expander inlet pressure, P3, for pure working fluids.

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Fig. 9. The variation of ORC exergy efficiency with the expander inlet pressure, P3, for zeotropic-mixture working fluids.

_ net W

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Fig. 11. The variation of ORC external heat required with the expander inlet pressure, P3, for pure working fluids.

In this part, parametric sensitivity analysis of the ORC performance indicators such as exergy and energy efficiencies, cycle irreversibility, and external heat requirement is conducted for different pure and zeotropic-mixture working fluids and results are compared. Figs. 6 and 7 present the variation of ORC energy efficiency with the expander inlet pressure for the selected pure and zeotropic-mixture working fluids while keeping the expander inlet temperature at saturated condition. It is demonstrated that the cycle energy efficiency increases with the increment of the expander inlet pressure. The results are consistent for all working fluids and reveals that higher expander inlet pressure increases both the net output power and the evaporator heat requirement which finally leads to improve in the cycle overall thermal efficiency. It means that the percentage of increase of the net output power is higher than the percentage of increase of the evaporator heat requirement. Therefore, the ratio of the net output power to

the evaporator required heat increases with the turbine inlet pressure. Fig. 6 also illustrates that R123 which has the highest boiling point temperature has the best performance among the other pure organic fluids, while R134a which has the lowest boiling point temperature represents the worst cycle energy efficiency. Fig. 7 shows the same results for the zeotropic-mixture fluids. It could be concluded that the higher the boiling point temperature of the working fluid, the better the cycle energy efficiency. Figs. 8 and 9 depict the variation of ORC exergy efficiency with the expander inlet pressure for the selected pure and zeotropicmixture working fluids respectively. Fig. 8 shows that the exergy efficiency of the cycle decreases when the expander inlet pressure increases. Decrease in the cycle exergy efficiency represents an increment in the cycle irreversibility rate which means the higher exergy destruction. The results from this figure are in contrast with the cycle energy efficiency. Consequently, the optimum operating point could be selected by trade-offs between energy and exergy efficiencies. It is also revealed that the higher the boiling point temperature of the working fluid, the lower the cycle exergy efficiency. The cycle exergy efficiency for zeotropic-mixtures fluids shows optimum point at the specific expander inlet pressure. The fact is

Fig. 10. ORC total irreversibility versus expander inlet pressure, P3.

Fig. 12. The variation of ORC external heat required with the expander inlet pressure, P3, for zeotropic-mixture working fluids.

jcycle ¼

Q_ in



T 1  0 Teva



(9)

4. Results and discussion

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Fig. 15. The variation of ORC energy and exergy efficiency with the increase of expander inlet temperature from saturation point for R245fa. Fig. 13. The variation of ORC working fluid mass flow rate with the expander inlet pressure, P3, for pure working fluids.

these results should be considered with the other performance indicators at different operating pressures and temperatures. Fig. 10 shows the effects of increase in the expander inlet pressure on the cycle total irreversibility. It is demonstrated that the ORC total irreversibility increases with an increment in the expander inlet pressure. The lowest irreversibility rate is obtained for R123. It means that the higher the normal boiling temperature of a working fluid results in the lower ORC irreversibility. The cycle external heat requirement in the evaporator to generate 100 kW of an electric power is evaluated with the change in the expander inlet pressure. Results for the pure and zeotropicmixture working fluids are presented in Figs. 11 and 12 respectively. It is assumed that the generator converts expander shaft work to electricity without any loss. It is shown that the required external heat decreases when the expander inlet pressure rises. The reason is that with an increment in the expander inlet pressure, the required working fluid mass flow rate is reduced. Figs. 13 and 14 show the calculated working fluid mass flow rate for the cases analyzed in Figs. 11 and 12. It is depicted that the required mass flow rate decreases with the increment of the expander inlet pressure due to an increase in the net power output of the cycle with an increment in the expander inlet pressure. The

Fig. 14. The variation of ORC working fluid mass flow rate with the expander inlet pressure, P3, for zeotropic-mixture working fluids.

results are in line with what was presented in Figs. 6 and 7 since an increment of the net power output represents an increase in the cycle energy efficiency. It can be also understood that fluids with the lower practical operating pressure and the higher enthalpy of vaporization require lower mass flow rate and therefore lower external heat requirement. The variation of ORC energy and exergy efficiency with the expander inlet temperature is analyzed for all working fluids and is depicted just for R245fa, R407c and R404a in Figs. 15 and 16 because of similar behaviour. It should be also mentioned that evaporation pressure is kept constant at 2 MPa. These figures reveal the effects of using superheated working fluid on the energetic and exergetic efficiency of the cycle. The temperature range is from saturation temperature of each working fluid at P ¼ 2 MPa to 30  C above that point. It is illustrated that the energy efficiency of the cycle remains almost constant or negligibly decreases but the exergy efficiency decreases with an increment of the expander inlet temperature. This reflects that not only the organic fluids do not need to be superheated like water to increase the cycle energy efficiency, but also it is exergetically beneficial to work at the saturated condition. However, it should be considered that the organic fluids are restricted to a small range of applicability depending on their thermodynamic conditions. Fig. 17 shows the variation of ORC external required heat with an isentropic efficiency of the expander. It is illustrated that the

Fig. 16. The variation of ORC energy and exergy efficiency with the increase of expander inlet temperature from saturation point for R407c and R404a.

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Fig. 19. CO2 emission for different fuel combustion systems versus ORC.

Fig. 17. The variation of ORC external heat requirement with the expander isentropic efficiency.

increases in an isentropic efficiency of the expander results in significant decrease in the amount of external heat required and consequently increase in the cycle energy and exergy efficiencies. It could be concluded that choosing appropriate type of expander substantially influence ORC overall performance. Fig. 18 shows the effects of ambient temperature on ORC exergy efficiency and total irreversibility for R134a. Technically ambient temperature greatly affects the condenser. As the ambient temperature gets close to the condenser temperature, the condenser irreversibility and consequently the cycle total irreversibility are reduced. It means that by increasing the ambient temperature the exergy efficiency of the cycle increases and in contrast, the cycle irreversibility decreases. Growing need for mitigation of greenhouse gas emissions would make renewable/waste heat based ORC technology one of the most promising power generation options in the near future. Generating electricity by utilizing renewable or waste heat at the temperatures as low as 150  C makes ORC as the standalone system very attractive for many residential applications. The other opportunity is the integration of ORC with the other industrial processes. The generated electricity could be either used internally or be sold to the power grid. Consequently, reduction in the grid power consumption results in significant decrease in CO2 emission. The specific carbon dioxide emission of different fuel combustion systems is

depicted in Fig. 19 and compared with the renewable/waste heat based ORC [29,30]. It is revealed that the application of ORC with an appropriate working fluid for power generation can substantially contribute to the reducing of CO2 emission per unit of useful energy produced. 5. Conclusions Comprehensive combined energy and exergy analysis of the low-grade heat source ORC is conducted and comparative assessment of using pure and zeotropic-mixture working fluid is carried out. Parametric sensitivity analysis of the cycle performance indicators such as energy and exergy efficiencies, cycle irreversibility rate, external heat requirement, and working fluid mass flow rate is developed at the various temperatures and pressures. Environmental characteristics of the selected working fluids such as toxicity, flammability, ODP and GWP are also studied. It is demonstrated that by increasing the expander inlet pressure, the ORC energy efficiency increases and it is also higher for the working fluids with the higher boiling point temperature. In contrast, sensitivity analysis shows that the exergy efficiency of the cycle decreases when the expander inlet pressure increases. Decrease in the cycle exergy efficiency represents an increment in the cycle irreversibility rate which means the higher exergy destruction. It is also revealed that the higher the normal boiling temperature of a working fluid results in the lower ORC irreversibility. The effects of using superheated working fluid on the energetic and exergetic efficiency of the cycle are also investigated. It is illustrated that the energy efficiency of the cycle remains almost constant or negligibly decreases but the exergy efficiency decreases with an increment of the inlet temperature from the saturation point. This reflects that not only the organic fluids do not need to be superheated, but also it is exergetically beneficial to work at the saturated condition. Moreover, the effects of expander isentropic efficiency on the cycle required external heat and ambient temperature on the cycle exergy efficiency are analyzed. It can be concluded that power generation with ORC using renewable heat sources like solar or geothermal systems and low-temperature industrial waste heat would significantly reduce CO2 emissions and offsets grid consumption. References

Fig. 18. ORC exergy efficiency and total irreversibility versus ambient temperature for R134a.

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Glossary _ mass flow rate, kg s1 m: _ energy flow rate, kW E: _ exergy flow rate, kW Ex: Q_ : heat rate, kW _ power, kW W: _ irreversibility rate, kW I: T: temperature, K T0: ambient temperature, K P0: ambient pressure, kPa h: specific enthalpy, kJ kg1 s: specific entropy, kJ kg1 h: energy efficiency J: exergy efficiency Subscripts in: inlet out: outlet i: process stream j: heat source (evaporation) or sink (condensation) des: destruction eva: evaporator exp: expander cond: condenser gen: generator