Thermodynamic optimization of organic Rankine

Proceedings of the Institution of Mechanical Engineers, Part A: http://pia.sagepub.com/ Journal of Power and Energy

Thermodynamic Optimization of Organic Rankine Cycles at Several Condensing Temperatures: Case Study of Waste Heat Recovery in a Natural Gas Compressor Station I Saavedra, J C Bruno and A Coronas Proceedings of the Institution of Mechanical Engineers, Part A: Journal of Power and Energy 2010 224: 917 DOI: 10.1243/09576509JPE998 The online version of this article can be found at: http://pia.sagepub.com/content/224/7/917

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Thermodynamic optimization of organic Rankine cycles at several condensing temperatures: case study of waste heat recovery in a natural gas compressor station I Saavedra, J C Bruno∗ , and A Coronas Department of Mechanical Engineering, CREVER (Group of Applied Thermal Engineering), Universitat Rovira i Virgili, Tarragona, Spain The manuscript was received on 5 March 2010 and was accepted after revision for publication on 5 May 2010. DOI: 10.1243/09576509JPE998

Abstract: In this article, the authors performed the thermodynamic optimization of organic Rankine cycles (ORCs) using several working fluids and considered the effect of three heat rejection media in the condenser: cooling water, ambient air, and hot water at high temperature for co- or trigeneration applications. The ORC system was modelled and optimized using the Aspen Hysys process simulator. The objective function is the maximization of turbine power output. Most natural gas compression stations use no heat recovery system. In this study, they applied the optimization procedure to the recovery of waste heat from gas turbines used to drive natural gas compressors in natural gas pumping stations. They used operational data from an existing pumping station to illustrate the potential benefits of ORC systems for this application, taking into account also non-thermodynamic aspects such as toxicity, flammability, and investment cost return. The highest ORC turbine output powers were obtained with aromatic hydrocarbons, then with aromatic fluorocarbons, n-hydrocarbons, and siloxanes (in that order). For the case studied here – a gas turbine of 2.6 MW of mechanical power– the proposed ORC can produce around 1 MWe, with a thermal efficiency of up to 24 per cent depending on the working fluid and condensing temperature. Keywords: organic Rankine cycle, waste heat recovery, optimization, natural gas compressor station 1

INTRODUCTION AND OBJECTIVES

Electricity is often generated in power plants that are based on the Rankine cycle and use water as the working fluid. With low- and medium-temperature heat sources or small-to-medium plant sizes, however, heat can be converted into electricity more efficiently in organic Rankine cycle (ORC) power plants, which use organic fluids rather than water as the working fluid. This type of cycle recovers heat at lower temperatures than water and leads to higher cycle efficiencies [1]. In the ORC, the working fluid is heated in the evaporator and the generated vapour is expanded in a turbine or expander that converts mechanical work ∗ Corresponding

author: Department of Mechanical Engineering,

CREVER (Group of Applied Thermal Engineering), Universitat Rovira i Virgili, Avda. Paisos Catalans 26, Tarragona 43007, Spain. email: [email protected] JPE998

into electricity by means of a generator. The vapour at the turbine outlet is condensed and the fluid returns to the evaporator by means of a pump to close the cycle (see Fig. 1). For applications at relatively high temperatures, a heat transfer fluid (HTF) such as diathermic oil is often used in a closed loop to transfer heat from the source to the evaporator. Depending on the fluid and cycle operating conditions, a regenerator is sometimes used to preheat the working fluid inlet to the evaporator, thus increasing the efficiency of the cycle. Working fluids for ORCs have high molecular masses and, because of their molecular complexity (i.e. their heat capacity), their vapour saturation curves have positive slopes in temperature–entropy (T –S) diagrams. These fluids generally have much lower relative drops in enthalpy during the expansion of steam and the saturated vapour phase becomes superheated after isentropic expansion. An ORC therefore requires only a single- or double-stage expander, which is simpler than a multi-stage expander in steam Proc. IMechE Vol. 224 Part A: J. Power and Energy


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I Saavedra, J C Bruno, and A Coronas

Fig. 1

Schematic diagram of the simulated ORC

Rankine cycles [2, 3]. See reference [4] for an analytical expression for predicting the saturated vapour slope in a T –S diagram for several fluids. For molecules with hydrogen bond interactions, such as water, ammonia, and ethanol (usually referred to as wet fluids), the enthalpy of vaporization is comparatively large. Wet fluids are therefore not generally suitable for ORC systems, because they become saturated after a large drop in enthalpy in the turbine, which may get damaged by the condensing fluid. Dry and isentropic fluids can prevent this. In practice, a near-vertical saturated vapour line is desirable so that there is little moisture during expansion and so that it is unnecessary to condense superheated vapour. To select suitable organic fluids for a given application, the most important factors to consider include thermodynamic properties such as critical temperature, molecular complexity, and molecular weight, which affect the cycle efficiency, operating parameters, and environmental impact. Most organic fluids have relatively low critical pressures and are therefore usually operated at lower pressures. They must also be compatible with the construction materials and chemically stable at high temperatures (unlike water, most organic fluids suffer chemical decomposition and deterioration at high temperatures and pressures). Cost and availability are other factors to consider. 1.1

Review of the literature and state-of-the-art

It is beyond the scope of this article to make a detailed review of all the proposed cycles, working fluids, and ORC applications. Instead, the most recent research in this field and the most innovative technologies available commercially will be mentioned briefly. To illustrate previous research on the development of ORC engines and for a comprehensive historical review of Italian activities, see reference [5]. Many studies have compared ORC efficiencies using several working fluids. Hung et al. [6] and Maizza and Maizza [7] studied the use of several

Hydrochlorofluorocarbons (HCFC) and Hydrofluorocarbons (HFC) refrigerants in ORC cycles with similar performances. Hung [8] presented a simplified model to calculate and compare irreversibilities in ORC cycles using different fluids. Tchanche et al. [1] presented a fluid selection focused on small, lowtemperature solar ORC. Another option for ORC is to use multi-component working media, which allows non-isothermal phase change at both high and low temperatures. Angelino and Colonna di Paliano [9] studied the merits of siloxanes and hydrocarbon mixtures for ORC. The performance of supercritical cycles has also been explored, for example, in reference [10]. With regard to ORC optimization, the mixed-integer linear method has been used for a specific case – an ORC fired by biomass from the furniture industry to find the optimal economic solution [11]. The system performance analysis and optimization of an ORC system using HFC-245fa as working fluid driven by exhaust heat was presented in reference [12]. Hettiarachchi et al. [13] presented an optimization model using as the objective function the ratio of total heat exchanger area to net power output, using empirical correlations to obtain the required heat transfer coefficients. Dai et al. [14] performed the optimization using a genetic algorithm method. The dynamic modelling of ORC was carried out in reference [15]. Integrating ORCs into specific processes was proposed using the Pinch methodology in reference [16]. Other studies have developed computational tools to improve the quality of flow through the turbine stage of an ORC using computational fluid dynamics for dense gas dynamics. See, for example, reference [17], which studied the gas dynamic behaviour at high pressures and densities near the liquid–vapour critical point. Another area of research is the laboratory testing of complete ORC systems. See, for example, reference [18], which reports the laboratory results of a 1 kW ORC experimental unit using HCFC-123 as the working fluid, or reference [19]. Other studies have tested key components such as the expander [20]. ORC technology is commercially available mainly for the exploitation of sustainable energy sources and mainly in geothermal [13, 21] and biomass applications [11, 22, 23]. For geothermal applications, ORC is a mature technology with thousands of megawatts of several kinds of plants installed worldwide starting at a power range of a few hundred kW. Other less-developed applications use solar thermal energy plants, from the higher-capacity systems using parabolic trough collectors [24] to small-scale systems using evacuated tube collectors [25] and small hybrid solar/waste heat systems [26]. Other applications are still in the early stages of development. Small-scale distributed generation systems, such as fuel cells [10] and micro-gas turbines [27], can be integrated with ORC to build high-efficiency,

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Thermodynamic optimization of ORCs at several condensing temperatures

small-capacity Brayton–Rankine combined cycles for buildings or small-capacity industrial applications. United Technologies [28] anticipates achieving 40 per cent electric efficiency by coupling the new Capstone 200 kW micro-turbine with their new ORC system. Other applications are bottoming cycles combined with medium-size gas turbines [29] and, more rarely, the recovery of waste heat from gas turbines used to drive natural gas pipeline compressors [30]. ORCs could also be integrated with an ejector cycle to provide cooling as a trigeneration plant [31]. It has also been proposed for integration with reverse osmosis desalination systems to produce fresh water by using solar thermal energy [19, 32]. Compact ORC standard modules are currently available in container size with nominal capacities of 200–2200 kWe and using specially formulated silicone as the working fluid [33]. Customized plants for combined heat and power (CHP) and heat recovery applications also exist. Other manufacturers (e.g. United Technologies, Tri-o-gen, and Ormat, use other working fluids). In commercial applications, a suitable intermediate HTF (also called thermal oil circuit) can be used between the waste heat recovery boiler and the ORC turbo-generator. This provides several advantages, the main of which one is that, as the adopted working fluid inside the ORC turbo-generator is protected from excessive temperatures, high temperature spots are avoided in all operating conditions. The main disadvantage, however, is that this introduces an additional temperature difference between the combustion gases and the turbo-generator. Future developments of ORC technology will focus on further improving electric efficiency, for example, by using two-stage ORC cycles and combined hot air turbine–ORC cycles [23].

1.2

Justification and objectives

The pressure of natural gas in gas transmission pipeline networks must be maintained along the pipeline through the use of compressor stations driven by gas turbines, reciprocating engines, or electric motors. Virtually, all the heat contained in the exhaust gases of gas turbine engines in compressor stations is discarded into the atmosphere. This lost heat represents 65–75 per cent of fuel energy, depending on the turbine’s thermal efficiency [34]. In 2009, in Spain alone, the company that owns the gas transmission network (Enagas) operates 14 compression stations with a total installed capacity of 380 MW. In the same year, in Russia Gazprom has a total capacity exceeding 47 500 MW and operates 281 compressor stations. An efficient system to recover waste heat from most of these plants could provide an enormous primary energy saving (PES) with environmental and economic benefits for most parts of the world. JPE998

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This waste heat from small- and medium-size gas turbines in compression stations is not recovered by using combined steam cycles, mainly because of the high cost of combined cycles in small-size plants, and the large and expensive air-cooled steam condensers and their demanding operation (blow-down, licensed operators, etc.) [34]. Popov [34] proposed an air bottoming cycle as a low-cost alternative to the steam bottoming cycle. Another alternative is to use ORC as the bottoming cycle. Only a few ORC plants are currently operating in natural gas compression stations. These include the 7 MW plant of the TransCanada Gold Creek compressor station near Grand Prairie (Alberta, Canada) [35]; the 5 MWe Kerrobert compressor station in Saskatchewan (Canada), which is the first of a series of four systems of this type; and the gas compressor stations under construction in Almendralejo and Huelva (Spain), which, like the example analysed in this article, convert recovered heat into electrical energy without the need for additional fuel. In all these applications, the fluid used is n-pentane. The previous review of the literature shows that many articles have analysed several types of organic fluids, but have not used a systematic optimization procedure to select the optimal operation conditions for each fluid in order to compare them properly. Moreover, this fluid screening and optimization have not focused on the specific power range of gas turbines used to drive natural gas compressors. In this study, the thermodynamic optimization of ORCs was performed using several working fluids, in order to compare them properly and assess how three different heat rejection media in the condenser (cooling water, ambient air, and hot water at high temperature for co- or trigeneration applications) affect this comparison. The ORC system is modelled and optimized using the Aspen Hysys process simulator [36]. The objective function is the maximization of the turbine power output by varying the mass flow of the working and HTFs and the turbine inlet temperature. The problem is restricted to minimum approach temperatures in some heat exchangers. The optimization procedure is applied to the recovery of waste heat from a gas turbine used to drive natural gas compressors in natural gas pumping stations. The optimum cycle performance is evaluated with several working fluids, including n-hydrocarbons, aromatic hydrocarbons, aromatic fluorocarbons, cyclic hydrocarbons, and siloxanes. The operational data of a real pumping station are used to illustrate the potential benefits of ORC systems for this application. 2

ORC MODELLING AND OPTIMIZATION

The process flow diagram of the ORC studied is shown in Fig. 1. The authors optimized the performance of this ORC using several organic fluids and the Proc. IMechE Vol. 224 Part A: J. Power and Energy


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Aspen Hysys process simulator [36], which contains a multi-variable steady-state optimizer. Once they built the flowsheet and obtained a converged solution, they used the optimizer to find which operating conditions maximize or minimize the objective function. In their case, this function is the maximization of turbine power output. The optimizer has its own spreadsheet from which all the cycle variables can be accessed and the objective function and optimization constraints can be defined. The variables imported from the flowsheet whose values are manipulated to maximize the objective function are known as primary variables. These variables are: (a) the mass flowrate of the working fluid (mWF ); (b) the mass flowrate of the HTF (mHTF ); (c) the turbine inlet temperature (T1 ).

T8 − T6 5

(1)

T8 T10 − 5

(2)

The method the authors selected to solve the problem was the mixed method, which is valid for inequality constraints only. This method takes advantage of the global convergence characteristics of the Box method and the efficiency of the Sequential Quadratic Programming (SQP) method. It starts the minimization with the BOX method using a very loose convergence tolerance (50 times the desired tolerance). After convergence, the SQP method is used to locate the final solution using the desired tolerance [36]. To simulate and optimize the ORC cycles for each working fluid, the authors considered three condensing temperatures: dissipating heat with water using a cooling water tower (35 ◦ C), dissipating heat using an air condenser (50 ◦ C), and dissipating heat producing hot water at 90 ◦ C, which can be used for heating or cooling using an absorption chiller. The following definitions of equipment efficiencies and parameters were used. Regenerator effectiveness (considering constant specific heat capacity) T3 − T2 T2 − T 5

(3)

Evaporator effectiveness (considering constant specific heat capacity) εevap =

T7 − T8 T7 − T 6

ηth =

Wnet Qevap

(4)

(5)

where the net power output of the cycle is Wnet = Wt − Wp

(6)

and the electrical power output can be estimated as We = ηeg Wnet

(7)

The cooling capacity of a thermally driven chiller when the heat from the condenser is recovered as hot water at 90 ◦ C is calculated as Qcold = COP Qcond

These primary variables are interdependent. However, the value of these variables can be set freely and all the other variables will adapt to these values producing a problem solution that has to be optimized by adjusting these primary variables. The constraint functions, defined using the nomenclature in Fig. 1, are

εreg =

ORC thermal efficiency is defined as

(8)

where Qcond is the heat dissipated in the condenser and the COP corresponds to that of the thermally driven chiller. In this study, the authors considered a COP of 0.7, which is typical for a single-effect water/LiBr absorption chiller. In this model they assume: (a) saturated vapour at the turbine inlet; (b) saturated liquid at the condenser exit; (c) no pressure changes in the units other than the turbine and pumps; (d) negligible heat losses; (e) steady-state operation; (f ) isentropic efficiencies to characterize the performance of the turbine and circulation pump. The accuracy of the results depends strongly on a suitable prediction of the working fluid’s thermodynamic properties. In this study, they calculated the thermodynamic properties of the various fluids using the Peng–Robinson Stryek–Vera (PRSV) equation of state as implemented in the process simulator. This thermodynamic properties model was selected according to the recommendations in the literature [37, 38]. To validate the results of the PRSV equation of state as implemented in Aspen Hysys, the authors compared them to correlations of experimental data as implemented in the Engineering Equation Solver (EES) software [39]. The organic hydrocarbons selected for the validation test were n-heptane (linear hydrocarbon), cyclohexane (cyclic hydrocarbon), and toluene (aromatic hydrocarbon) modelled using the high-accuracy thermodynamic data from references [40] (n-heptane), [41] (cyclohexane), and [42] (toluene). They compared the saturation pressure at different temperatures for the three types of hydrocarbons. The deviation between the equation of state used and the experimental correlated data was 1– 2 per cent for n-heptane and cyclohexane and 1–6 per cent for toluene. The PRSV equation of state is the one usually used in the literature to predict the


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thermodynamic properties of siloxanes [9, 17]. A specific multi-parameter equation of state for selected siloxanes was proposed [43] and compared with the one used in the present study. However, it was concluded that the improvement when using the multiparametric equation of state rather than the PRSV equation of state is limited in many aspects of interest, including vapour and liquid P-v-T data [43]. 3

RESULTS AND DISCUSSION

The operational data used as input for the model were obtained from an operating natural gas compressor plant in Tivissa (Spain). This plant consists of three turbo-compressors with a total capacity to pump 1 640 000 N m3 /h of natural gas with a pressure increase of over 20 bar. The proposed ORC for this plant aims to recover waste heat from the exhaust gas turbine used to drive a natural gas compressor station and convert it into electrical power. The ORC working Table 1

Annual average operational data for the gas turbine-driven natural gas compressors

Natural gas composition (minor components included in the methane percentage) Methane (%mol) Ethane (%mol) Propane (%mol) Nitrogen (%mol)

89.49 7.71 2.49 0.61

Gas turbine data Air inlet temperature (◦ C) Natural gas inlet temperature (◦ C) Natural gas inlet pressure (bar a) Combustion air pressure (bar a) Natural Gas consumption (N m3 /h) Natural Gas consumption (kW) Net power output (kW) Net efficiency (%) Exhaust gas turbine (◦ C) Oxygen in exhaust gas (%)

23.3 24.4 13.5 7.3 850 10 221 2635 25.8 405 16.8

Bar a, absolute pressure in contrast with relative or gauge pressure.

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fluid should take heat from the flue gas by using a diathermic oil to avoid local overheating and use a regenerator to improve cycle thermal efficiency (see Fig. 1). All the gas turbine operational data are shown in Table 1. These data were obtained in the compressor station from one of the three turbo-compressor units. The low temperature of the exhaust gas may be explained by the high excess of combustion air (calculated from available data as 360 per cent). The authors assume that the gas turbine will nearly always work at its nominal power conditions and so its exhaust gas temperature will be almost constant throughout the year. This simplification is in close agreement with the actual operating conditions of the natural gas compressor station. The compression unit plant produces exhaust gas at an average temperature of 405 ◦ C with a mass flowrate of 52 200 kg/h. In the case of an ORC for trigeneration, the chilled water produced may be used for local air-conditioning or for cooling the natural gas at the compressor outlet [34]. In this study, both steam and ORCs are compared at their optimal operation conditions for each condensing temperature. The potential organic working fluids considered cover a wide range of critical temperatures and molecular complexity and were selected, because they belong to the kind of fluids usually considered for ORC applications. These fluids are divided into the following groups: n-hydrocarbons, aromatic hydrocarbons, aromatic fluorocarbons, cyclic hydrocarbons, and siloxanes. Their thermo-physical properties are shown in Table 2. The specifications of the cycle in this analysis are shown in Table 3. An increase in the turbine inlet temperature (or an increase in pressure in saturated cycles) produces an increase in turbine output and cycle efficiency. However, when the turbine inlet temperature approaches the working fluid critical temperature, the turbine output starts to decline and an optimum value is reached. The optimum operation point is also constrained by the other problem

Table 2 Thermo-physical properties of the fluids considered in this study Group

Fluid

Molecular mass

Boiling point (◦ C)

ρ (kg/m3 )

Tc (◦ C)

Pc (kPa)

n-hydrocarbons

n-pentane n-heptane n-octane

72.2 100.2 114.2

36.1 98.4 125.7

629.7 686.8 705.7

196.5 267.0 295.4

3375 2737 2497

Aromatic hydrocarbons

Iso-octane Benzene Toluene Ethylbenzene Propylbenzene

114.2 78.1 92.1 106.2 120.2

99.2 80.1 110.6 132.2 159.2

695.0 882.2 870.0 870.0 864.8

270.8 288.9 318.6 343.9 365.2

2568 4924 4100 3607 3200

Cyclic hydrocarbons

Cyclohexane

84.2

80.7

781.8

280.1

4053

Aromatic fluorocarbons

Hexafluorobenzene Pentafluorobenzene

186.1 168.1

80.3 85.8

1626.0 1508.0

243.6 257.9

3200 3520

Lineal siloxanes

Hexamethyldisiloxane (MM) Decamethyltetrasiloxane (MD2 M) Tetradecamethylhexasiloxane (MD4 M) Octamethylcyclotetrasiloxane (D4 )

162.4 310.7 459.0 296.6

100.5 194.4 259.8 175.0

769.8 857.9 894.3 949.4

245.6 326.3 380.1 313.4

1914 1227 804 1332

Cyclic siloxanes

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Table 3

Specifications for the ORC and main parameters for the case study

Organic fluid Turbine

Upper limit: Tc∗ Lower limit: 30% approximately below Tc 80% 50 ◦ C, heat dissipation with air 35 ◦ C, heat dissipation with water 90 ◦ C, production of hot water Upper limit: 20 000 kg/h Lower limit: 40 000 kg/h 75% 82% 87%

Inlet temperature, T1 Isentropic efficiency, ηis Condensing temperature, T4

Pump

Mass flow, mWF

Regenerator Evaporator

Isentropic efficiency, ηis Regenerator efficiency, εreg Evaporator efficiency, εevap

Heat transfer fluid Evaporator

320 ◦ C Upper limit: 30 000 kg/h Lower limit: 40 000 kg/h 200 kPa

Inlet temperature, T7 Mass flow, mHTF Pressure, PHTF

Heat source Gas turbine exhaust gas

Except for those fluids whose Tc is higher than T9 .

Net power, W net (kW)

variables: the minimum approach temperatures in the cycle heat exchanger units, the temperatures in the waste heat stream, and the maximum temperature for the HTF and working fluid flow-rates calculated with the waste heat recovered. Figure 2 shows an example for n-heptane and the condensing temperatures considered. Table 4 shows an example of results for each stream and plant component for one of the fluids (n-heptane) at this optimum operation point. In the next sections (3.1–3.8), the authors will analyse the performance of an ORC cycle with the various fluids and the proposed optimization procedure.

3.1 Turbine output power Figure 3 shows the turbine power output (Wt ) of the working fluids at their optimal operating conditions for a condensing temperature of 90 ◦ C. In this analysis, authors assumed the same turbine efficiency for all the working fluids as a basis for comparison, though in practice this is not possible. The Wt obtained is in the 590–867 kW range. Benzene produced the highest power capacity (867 kW), probably because it shows the highest critical pressure. Siloxanes, water, and n-pentane obtained the lowest turbine output for their

1600

30

1400

25

1200

20

1000

15

800

10

600

5

400 170

190

210

230

250

Thermal Efficiency (%)

a

405 ◦ C 100 kPa 52 200 kg/h O2 : 16.812%mol N2 : 7.533%mol H2 O: 3.6982%mol CO2 : 1.9576%mol

Temperature, T9 Gas pressure Flowrate (mg ) Composition

0 270

Tinlet turbine (ºC) Wnet, Tcond 35ºC Eff, Tcond 35ºC

Fig. 2

Wnet, Tcond 50ºC Eff, Tcond 50ºC

Wnet, Tcond 90ºC Eff, Tcond 90ºC

Optimum performance of the ORC plant as a function of the turbine inlet temperature for several condensing temperatures (T4 )


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Table 4

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Example of results for each stream and plant component for one of the fluids (n-heptane)

Stream

Fluid

Pressure (kPa)

1 2 3 4 5 6 7 8 9 10

n-heptane n-heptane n-heptane n-heptane n-heptane n-heptane HTF HTF Exhaust gas Exhaust gas

2594 9.54 9.54 9.54 2594 2594 200 200 100 100

Unit

Parameter

Value

Turbine ORC pump Evaporator Regenerator Condenser Gas heat recovery

Work (kW) Work (kW) Duty (kW) Duty (kW) Duty (kW) Duty (kW)

1121 38 4035 1475 2952 4035

Temperature (◦ C)

Vapour quality

Mass flowrate (kg/h)

263.0 157.8 58.0 35.0 36.1 117.9 320.0 144.2 405.0 149.2

1.0 Superheated Superheated 0.0 Subcooled Subcooled – – – –

26 284 26 284 26 284 26 284 26 284 26 284 37 391 37 391 52 200 52 200

Benzene

900

Cyclohexane 850

Pentafluorobenzene N-Octane

800

Toluene

Wt (kW)

750

n-Heptane Ethylbenzene

700

Hexafluorobenzene 650

Propylbenzene Iso-Octane

600

MM

550 500 180

D4 MD2M 200

220

240

260

280

300

320

MD4M n-Pentane

Turbine inlet temperature (ºC)

Water

Fig. 3 Turbine output power at the optimal operating conditions for each fluid at a condensing temperature of 90 ◦ C

low critical temperature or pressure depending on the fluid. At a condensing temperature of 50 ◦ C, the turbine output power ranged from 830 to 1175 kW, while at a condensing temperature of 35 ◦ C, it ranged from 920 to 1300 kW. Figure 4 shows the specific turbine power output at the optimal operating conditions when the condensing temperature was 90 ◦ C. Here, water showed the maximum specific power, because its mass flowrate was around six times lower than the average for the other fluids. Benzene and octane had the best results of the organic fluids. The ranking of the fluids JPE998

at the other two condensing temperatures was very similar. 3.2

Cycle efficiency

Figure 5 shows the thermal efficiency ηth at the optimal operating conditions when the condensation temperature was set at 90 ◦ C. These values ranged from 12 to 22 per cent. Here, decamethyltetrasiloxane (MD2 M) obtained the highest efficiency (22.4 per cent) and n-pentane the lowest. At condensing temperatures of 50 and 35 ◦ C, the values ranged from 18 to 27 per cent Proc. IMechE Vol. 224 Part A: J. Power and Energy


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450 Hexafluorobenzene Pentafluorobenzene MM n-Pentane MD2M D4 MD4M Iso-Octane Propylbenzene Ethylbenzene n-Heptane Toluene Cyclohexane n-Octane Benzene Water

400 350

Wesp (kJ/kg)

300 250 200 150 100 50 0 180

200

220

240

260

280

300

320


Fig. 4

Specific turbine output power at the optimal operating conditions for each fluid at a condensing temperature of 90 ◦ C

24 MD2M D4

22

n-Octane MD4M Benzene

20

Cyclohexane

th

η (%)

Iso-Octane n-Heptane

18

,

Propylbenzene Ethylbenzene Toluene

16

Pentafluorobenzene MM

14

Hexafluorobenzene n-Pentane Water

12 180

200

220

240

260

280

300

320


Fig. 5

Cycle efficiency at the optimal operating conditions for each fluid at a condensing temperature of 90 ◦ C

and from 20 to 29 per cent, respectively. A working fluid that performed well at a given condensing temperature also performed well at another temperature with respect to the other fluids. However, their relative positions when the fluids were sorted according to cycle efficiency were not exactly the same. The good cycle efficiency shown by the three siloxanes with the highest critical temperature could explain their commercialization for ORC systems at relatively high driving temperatures, as in the case of biomass combustion. All fluids except n-pentane had a better cycle efficiency than water.

3.3 Working fluid mass flowrate Figure 6 shows the mass flowrate at the optimal working conditions for each fluid when the condensation temperature was set at 90 ◦ C. The two fluorocarbons considered required the highest mass flowrate at the optimum point. The hydrocarbons (including the alkenes), followed by the siloxanes, required the lowest mass flowrate, though all fluids had a higher mass flowrate than water. The mass flow-rates at the three condensing temperatures were similar for all fluids.


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Benzene

64500

Cyclohexane Pentafluorobenzene 54500

Toluene Ethylbenzene n-Octane

Mf (kg/h)

44500

Propylbenzene n-Heptane 34500

Hexafluorobenzene Iso-Octane Water

24500

D4 MM 14500

MD2M MD4M

4500 180

n-Pentane 200

220

240

260

280

300

320


Fig. 6

3.4

Mass flowrate of the different organic fluids at the optimal operating conditions and a condensing temperature of 90 ◦ C

Density at the turbine outlet

3.5 Toxicity and flammability

The density of the working fluid at the turbine outlet affects the design and size of the turbine. A low density at the turbine outlet means that a higher volume is required in the turbine or expander and also in the condenser. A higher volume flowrate implies a need for more construction material and therefore a higher cost. As Fig. 7 shows, the fluorocarbons, followed by the hydrocarbons, benefit most from a high density of vapour. Except for the siloxanes, all other fluids performed better than water in this respect.

In addition to thermodynamic properties and energy performance, toxicity and flammability are the main factors to consider when selecting the working fluid. Some fluids (e.g. toluene, propylbenzene, and benzene), have a restrictive toxicology and a limited level of occupational exposure. Benzene, for example, has the lowest permissible level at which a worker can be exposed without causing adverse health effects. The fact that it is the most toxic of all the fluids analysed detracts from its good energy performance. On

3

Density at the turbine outlet (kg/m )

12 MD4M MD2M Propylbenzene D4 Water Ethylbenzene n-Octane Toluene n-Heptane Iso-Octane Cyclohexane Benzene MM Pentafluorobenzene Hexafluorobenzene n-Pentane

10

8

6

4

2

0 180

200

220

240

260

280

300

320

Turbine Inlet temperature (ºC)

Fig. 7

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Density at the turbine output at the optimal operating conditions for each fluid and a condensing temperature of 90 ◦ C Proc. IMechE Vol. 224 Part A: J. Power and Energy Downloaded from pia.sagepub.com at CARLETON UNIV on June 28, 2014

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the other hand, n-pentane and cyclohexane have less restrictive toxicity limits. With regard to flammability, benzene, toluene, and propylbenzene do not exceed their autoignition temperatures at the maximum operating temperature of the cycle, but cyclohexane and n-octane do. n-Hydrocarbons generally have low toxicity but high flammability. For those fluids that can also be used as refrigerants, the ASHRAE 34 standard provides a good safety classification for fluids.

3.6

Selection of working fluids

Taking into account thermodynamics, flammability, and the toxicity of various fluids, the authors selected from each family of fluids those shown in Table 5. This table also shows the results at the optimum conditions for each organic fluid selected. Toluene produced the highest turbine power capacity at condensing temperatures of 35 ◦ C, whereas pentafluorobenzene produced the highest power at 50 and 90 ◦ C. However, the influence of other parameters, such as the optimal mass flowrate or the amount of waste heat that can be recovered from the turbine exhaust gas, means that fluids with maximum Wt do not necessarily provide the maximum thermal efficiency. The performance of the optimal steam Rankine cycle (Table 5) is worst than that of the selected organic fluids at a similar level as hexamethyldisiloxane (MM). Table 5

Also, for these kinds of small-scale power plants (about 1 MW or less), traditional steam cycles are not cost effective. Organic working fluids have, in comparison with water, a relatively low enthalpy difference between high pressure and expanded vapour. This leads to higher mass flows compared with water (see Table 5). The application of larger turbines due to the higher mass flow reduces the gap losses compared to water–steam turbine with the same power [44]. Also, in most cases, a single-stage turbine may be used instead of a multi-stage turbine, as usually required for steam. At a condensation temperature of 90 ◦ C, the dissipation of heat produces hot water that can be used in industrial or district heating networks or to drive absorption chillers. Table 6 shows the results obtained when the heat dissipated in the condenser is used to drive a typical single-effect water–LiBr absorption

Table 6

Cooling and heating output using the fluids selected

Fluid

Wnet (kW)

Qcond (kW)

Qcold (kW)

Wnet / Qcond

n-heptane Toluene Ethylbenzene Pentafluorobenzene Hexamethyldisiloxane

694 724 718 733 623

2869 3087 3026 3252 2904

2008 2161 2118 2276 2033

0.242 0.235 0.237 0.225 0.215

Energy performance of the selected organic fluids working at their optimal operation point and at different condensing temperatures (comparison with the steam Rankine cycle)

Group Tcond = 35 ◦ C nhydrocarbons Aromatic hydrocarbons Aromatic fluorocarbons Siloxanes Tcond = 50 ◦ C nhydrocarbons Aromatic hydrocarbons Aromatic fluorocarbons Siloxanes Tcond = 90 ◦ C nhydrocarbons Aromatic hydrocarbons Aromatic fluorocarbons Siloxanes

Fluid

Wt (kW)

ηth (%)

Qreg (kW)

Qevap (kW)

Wnet (kW)

TIT (◦ C)

P1 (kPa)

mf (kg/h)

mHTF (kg/h)

n-heptane

1121

26.84

1475

4035

1083

263.0

2594

26 284

37 391

Toluene Ethylbenzene Pentafluorobenzene

1222 1186 1212

27.87 27.66 25.57

761 915 645

4319 4251 4571

1204 1176 1169

249.0 239.0 248.6

1673 877 3083

26 228 26 351 56 210

36 781 36 909 36 399

1020 1032

25.13 24.30

2080 –

3908 4233

982 1029

242.6 240.0

1831 1291

41 723 5508

37 735 –

n-heptane

1005

24.77

1449

3898

966

263.3

2602

26 743

37 763


1065 1051 1084

24.93 25.51 23.58

723 911 630

4206 4082 4404

1049 1041 1038

235.0 236.0 249.5

1357 834 3123

27 156 26 810 56 822

36 999 36 807 36 642

914 935

23.02 23.51

2011 –

3799 3962

875 931

242.7 240.0

1835 1823

42 784 5300

38 081 –

n-heptane

737

19.49

1312

3563

694

263.9

2625

28 260

38 976


739 727 782

19.00 19.17 18.39

653 799 542

3812 3743 3985

724 718 733

228.0 227.0 250.4

1217 718 3166

28 135 27 998 58 874

38 036 38 269 37 518

667 646

17.66 16.74

1748 –

3527 3842

623 643

243.3 240.0

1850 1437

46 215 5445

39 138 –

MM Water

MM Water

MM Water

TIT, turbine inlet temperature. Proc. IMechE Vol. 224 Part A: J. Power and Energy Downloaded from pia.sagepub.com at CARLETON UNIV on June 28, 2014

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chiller. Pentafluorobenzene, followed by toluene, are the organic fluids that provide the highest cooling capacity while hexametyldisiloxane provides the highest waste heat recovery/power capacity ratio. 3.7

PES and environmental benefit

Depending on the condensation temperature, the plant produces electrical power (PESh = 0) or electrical power and heat simultaneously. The PES due to the ORC plant operation can be calculated as Wnet ηeg Qcond PES = PESe + PESh = Nop + (9) ηgrid ηb The efficiency for the electricity generator (ηeg ), the power from the electrical grid (ηgrid ), and the efficiency for a conventional boiler (ηb ) are 0.98, 0.4, and 0.9, respectively. As the number of running hours of the compressor station is very high, a utilization factor of 90 per cent (Nop = 7884 h per year) was considered for the ORC system. The CO2 emissions saved can be estimated as ECO2 = Wnet ηeg Nop fCO2,e + PESh fCO2,h

(10)

where the emission factor for the power generated ( fCO2,e ) is taken as 372 g CO2 per kWh of electrical power, which is reasonable for combined-cycle power plants [45]. The corresponding factor value for the whole electricity grid is more variable and difficult to determine and should be higher than that of a combined cycle power plant. In this case, the value for the combined cycle has been considered as a lower bond for the CO2 emissions saved, due to the electricity produced. The CO2 emission factor for the primary energy consumed in a boiler to produce an amount of heat equivalent to that produced by the ORC plant ( fCO2,h ) is assumed to be that corresponding to natural gas (i.e. 200 g of CO2 per kWh of primary energy [45]). If the plant were used to recover waste heat (at a condensing temperature of 90 ◦ C), depending on the organic fluid used, 14 500–16 000 tCO2 /year and 37– 42 GWh/year of natural gas could be saved. Without heat recovery in the condenser, the PES ranges from 17 to 23 GWh/year of natural gas and the CO2 emissions saved range from 6300 to 8500 tCO2 /year. Although these are rough estimates, they provide a good picture of the potential of this application. 3.8

Economic cost estimation

The capital costs of an ORC are still high and are mainly correlated to the capacity of the plant. It is difficult to evaluate the economic cost of the proposed ORC, because the main components of the system have not yet been designed. However, as a first approximation, the investment cost will be estimated using similar JPE998

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data from the literature [23, 30, 35, 46]. Recent studies report that the investment costs for a commercial ORC unit of 500 kW are about 1970 ¤/kW [47]. In this cost estimation, the authors assume a higher average cost of 3000 ¤/kWe installed, which may be reasonable for a demonstration technology that is not yet fully developed, though some studies have considered even higher costs [44]. Payback for the investment can be estimated from equation (11) PB =

ci Wnet ηeg ce Wnet ηeg N − cm Wnet ηeg N + (Qcond Nop /ηb )cNG (11)

To calculate the running costs, only maintenance costs were considered as no additional fuel is needed. The electricity needed by the air cooling unit is considered small enough not to be considered in a first raw analysis. The maintenance costs are assumed to be low: 1 c¤/KWhe [23, 30, 48]. In this case, it was assumed that the electricity produced could be sold in the regulated market at a price of 65.81 ¤/MWh [49], with a corresponding additional feed-in tariff of 1.934 c¤/kWh [50]. At a condensation temperature of 90 ◦ C, the economic benefits would also include the saving associated with a lower consumption of natural gas than in a conventional boiler. The gas cost considered is 1.808 c¤/kWh [51] with a boiler efficiency of 92 per cent. As an example, let one consider an ORC that uses n-heptane as the working fluid. In this case, payback is about 5 years, but if the condensing temperature is 90 ◦ C, this falls to 2.4 years because of the economic benefit of using dissipated heat in the condenser for heating applications. Another factor with a potentially important economic impact on the unitary investment cost of the ORC plant is the choice of working fluid. Hexamethyldisiloxane, for example, is eight times more expensive than n-heptane. 4

CONCLUSIONS

In this study, the authors determined the optimal operation conditions for several working fluids in Rankine cycles in order to compare them properly. They analysed the system using an objective function to maximize the turbine power output and obtained the optimal operating conditions for several organic fluids for a specific case study: the recovery of waste heat in a natural gas compression station. Aromatic hydrocarbons provide the maximum turbine power output. These are followed by aromatic fluorocarbons, then n-hydrocarbons, and finally siloxanes. These results are the same for all three condensation temperatures studied. However, in addition to energy performance, one must also take into account toxicity and flammability. In these cases, the fluids Proc. IMechE Vol. 224 Part A: J. Power and Energy


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with the best overall performance are n-heptane, toluene, ethylbenzene, pentafluorobenzene, and MM for each type of fluid studied. At high condensing temperatures, the hot water produced in the condenser could be used to supply heat to an absorption chiller to then provide heating and/or cooling. Pentafluorobenzene and toluene produce the highest cooling capacity, while MM provides the highest waste heat recovery/power capacity ratio. For the natural gas compressor station with a maximum unitary pumping capacity of 500 000 N m3 /h of natural gas with a gas turbine of 2635 kW, an ORC could produce around 1 MWe depending on the working fluid and condenser temperature used. Also for this natural gas compressor station, it could save 37–42 GWh/year of natural gas and 14 500–16 000 tCO2 /year depending on the organic fluid used when the condenser heat is recovered for heating.When heat is not recovered, these figures are approximately 50 per cent lower. ACKNOWLEDGEMENTS The authors thank Mr Eduard Martin and ENAGAS, SA for their contribution to this study. © Authors 2010 REFERENCES 1 Tchanche, B. F., Papadakis, G., Lambrinos, G., and Frangoudakis, A. Fluid selection for a low-temperature solar organic Rankine cycle. Appl. Therm. Eng., 2009, 29, 2468–2476. 2 Andersen, W. and Bruno, T. Rapid screening of fluids for chemical stability in organic Rankine cycle applications. Ind. Engng Chem. Res., 2005, 44, 5560–5566. 3 Hung,T. C., Shai,T. Y., and Wang, S. K. A review of organic Rankine cycle (ORC) for the recovery of low grade waste heat. Energy, 1997, 22, 661–667. 4 Liu, B.-T., Chien, K.-H., and Wang, C.-C. Effect of working fluids on organic Rankine cycle for waste heat recovery. Energy, 2004, 29, 1207–1217. 5 Angelino, G., Gaia, M., and Macchi, E. A review of Italian activity in the field of organic Rankine cycles. In Proceedings of the International VD – Seminar, Zurich, 1984, pp. 465–482. 6 Hung,T. C., Shai,T. Y., and Wang, S. K. A review of organic Rankine cycles (ORC) for the recovery of low-grade waste heat. Energy, 1997, 23(7), 661–667. 7 Maizza, V. and Maizza, A. Unconventional working fluids in organic Rankine cycles for waste energy recovery systems. Appl. Therm. Eng., 2001, 21, 381–390. 8 Hung, T. C. Waste heat recovery of organic Rankine cycle using dry fluids. Energy Convers. Manage., 2001, 42, 539– 553. 9 Angelino, G. and Colonna di Paliano, P. Multicomponent working fluids for organic Rankine cycles (ORC). Energy, 1998, 23(6), 449–463.

10 Angelino, G. and Colonna di Paliano, P. Organic Rankine cycles (ORC) for energy recovery for molten carbonate fuel cells. In Proceedings of the 35th Intersociety Energy Conversion Engineering Conference (IECEC), Las Vegas, Nevada, 24–28 July 2000, vol. 2, pp. 1400–1409. 11 Chinese, D., Meneghetti, A., and Nardin, G. Diffused introduction of organic Rankine cycle for biomass-based power generation in an industrial district: a systems analysis. Int. J. Energy Res., 2004, 28, 1003–1021. 12 Wei, D., Lu, X., Lu, Z., and Gu, J. Performance analysis and optimization of organic Rankine cycle (ORC) for waste heat recovery. Energy Convers. Manage., 2007, 48, 1113–1119. 13 Hettiarachchi, H. D., Golubovic, M., and Worek, W. M. Optimum design criteria for an organic Rankine cycle using low-temperature geothermal heat sources. Energy, 2007, 32, 1698–1706. 14 Dai,Y.,Wang, J., and Gao, L. Parametric optimization and comparative study of organic Rankine cycle (ORC) for low grade waste heat recovery. Energy Convers. Manage., 2009, 50, 576–582. 15 Wei, D., Lu, X., Lu, Z., and Gu, J. Dynamic modelling and simulation of an organic Rankine cycle (ORC) system for waste heat recovery. Appl. Therm. Eng., 2008, 28, 1216– 1224. 16 Desai, N. B. and Bandyopadhyay, S. Process integration of organic Rankine cycle. Energy, 2009, 34, 1674–1686. 17 Colonna di Paliano, P. and Silva, P. Dense gas thermodynamic properties of single and multi-component fluids for fluid dynamics simulations. J. Fluids Engng Trans. ASME, 2003, 125(3), 414–427. 18 Yamamoto,T., Furuhata,T., Arai, N., and Mori, K. Design and testing of the organic Rankine cycle. Energy, 2001, 26, 239–251. 19 Manolakos, D., Kosmadakisa, G., Kyritsisa, S., and Papadakis, G. On site experimental evaluation of a lowtemperature solar organic Rankine cycle system for RO desalination. Sol. Energy, 2009, 83, 646–656. 20 Quoilin, S., Orosz, M., and Lemort, V. Modeling and experimental investigation of an organic Rankine cycle using scroll expander for small solar applications. In Proceedings of the Eurosun Conference, Lisbon, Portugal, 2008. 21 Broniki, L. Y. Implementing new power plant technologies – technical and economical aspects. In Proceedings of the International Conference on Geothermal energy and territory, Pomarance, Italy, 2004. 22 Drescher, U. and Brüggemann, D. Fluid selection for the organic Rankine cycle (ORC) in biomass power and heat plants. Appl. Therm. Eng., 2007, 27, 223–228. 23 Obernberger, I., Carlsen, H., and Biedermann, F. Stateof-art and future developments regarding small-scale biomass CHP systems with special focus on an ORC and Stirling engine technologies. In Proceedings of the International Nordic Bioenergy Conference, Jyväskylä, Finland, 2003. 24 Canada, S., Brosseau, D. A., and Price, H. Design and construction of the APS 1 MWe parabolic trough power plant. In Proceedings of the ISEC 2006, ASME International Solar Energy Conference, Denver, USA, 2006. 25 Kane, M., Larrain, D., and Allani, Y. Small hybrid solar systems. Energy, 2003, 28, 1427–1443.


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26 Mills, D. Advances in solar thermal electricity technology. Sol. Energy, 2004, 76, 19–31. 27 Invernizzi, C. Bottoming micro-Rankine cycles for micro-gas turbines. Appl. Therm. Eng., 2006, 27, 100–110. 28 Rosfjord, T., Wagner, T., and Knight, B. UTC microturbine CHP product development and launch. In Proceedings of the Fourth Annual Microturbine Applications Workshop, Canada, 2004. 29 Najjar, Y. S. H. Efficient use of energy by utilizing gas turbine combined systems. Appl. Therm. Eng., 2001, 21, 407–438. 30 Leibowitz, H. M. Generating electric power from waste heat using ORC technology. PTAC 2002 Climate Change and GHG Technology, Calgary, Alberta, Canada, 2002. 31 Oliveira, A. C., Afonso, C., Matos, J., Riffat, S., Nguyen, M., and Doherty, P. A combined heat and power system for buildings driven by solar energy and gas. In Proceedings of the International Heat Power Cycles Conference (HPC-2001), CNAM, Paris, France, 2001. 32 Bruno, J. C., López-Villada, J., Letelier, E., Romera, S., and Coronas, A. Modelling and optimisation of solar organic Rankine cycle engines for reverse osmosis desalination. Appl. Therm. Engng, 2008, 28, 2212–2226. 33 Turboden srl. Via Cernaia, 10, 25124 Brescia, Italy, http : / /www.turboden.eu/en/products/products-chp . php. 34 Popov, D. Application of a gas turbine with air bottoming cycle in natural gas compressor stations. In Proceedings of the Second International Gas Turbine Technology Conference, CAME-GT, Bled, Slovenia, 2004. 35 Leibowitz, H. and Schochet, D. Generating electric power from compressor station residual heat. Pipeline Gas J., 2001, 11, 24–26. 36 Aspen Hysys. Aspen Technology, Inc., Wheeler Road, Burlington, Massachusetts, USA, http://www.aspentech. com. 37 Carlson, E. C. Don’t gamble with physical properties for simulations. Chem. Engng Prog., 1996, October, 35–46. 38 Simulation basis, selecting property methods, Aspentech driving process profitability, 2004. 39 Engineering Equation Solver (EES), F-Chart Software, Professional Version, 2003. 40 Span, R. and Wagner, W. Equations of state for technical applications: II results for non-polar fluids. Int. J. Thermophys., 2003, 24, 41–109. 41 Penoncello, S. G., Jacobsen, R. T., and Goodwin, A. R. H. A thermodynamic property formulation for cyclohexane. Int. J. Thermophys., 1995, 16, 519–529. 42 Tillner-Roth, R. Fundamental equations of state, 1998 (Shaker, Verlag, Aachan). 43 Colonna di Paliano, P., Nannan, N. R., Guardone, A., and Lemmon, E. W. Multiparameter equations of state for selected siloxanes. Fluid Phase Equilib., 2006, 244, 193– 211. 44 Schuster, A., Karellas, S., Kakaras, E., and Spliethoff, H. Energetic and economic investigation of organic Rankine cycle applications. Appl. Therm. Eng., 2009, 29, 1808–1817. 45 Ministerio de industria y comercio. IDAE. Plan de energías renovables en España 2005–2010 (Spanish JPE998

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APPENDIX Notation ce ci cm cNG COP ECO2 fCO2,e fCO2,h HTF mg mHTF mWF Nop PB Pc PES PESe PESh PHTF Qcold

unitary price of electricity (¤/kWhe) unitary investment cost of the ORC plant (¤/kWe) unitary cost of maintenance (¤/kWhe) unitary cost of natural gas (¤/kWht) coefficient of performance avoided CO2 emissions CO2 emission factor for the power generated CO2 emission factor for heating heat transfer fluid exhaust gas mass flowrate (kg/h) HTF mass flowrate (kg/h) organic working fluid mass flowrate (kg/h) annual operation hours (h/year) payback time of the capital investment working fluid critical pressure (kPa) primary energy savings primary energy saving for electricity primary energy saving for heating HTF pressure (kPa) cooling capacity of an absorption chiller driven by the condenser recovered heat (kW) Proc. IMechE Vol. 224 Part A: J. Power and Energy


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Qcond Qevap Qgas Qreg T Tc We Wnet Wp


condenser duty (kW) evaporator duty (kW) exhaust gas heat exchanger duty (kW) regenerator duty (kW) temperature (◦ C) working fluid critical temperature (◦ C) ORC electrical power output (kW) ORC net mechanical power output (kW) pump mechanical power (kW)

Wt

turbine power output (kW)

εevap εreg ηb ηeg ηgrid ηis ηth ρ

evaporator effectiveness regenerator effectiveness boiler efficiency electrical generator efficiency electricity grid efficiency isentropic efficiency ORC thermal efficiency mass density (kg/m3 )


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