A study on optimal composition of zeotropic working fluid in an Organic Rankine Cycle (ORC) for low grade heat recovery

Accepted Manuscript A study on optimal composition of zeotropic working fluid in an Organic Rankine Cycle (ORC) for low grade heat recovery K. Satanphol, W. Pridasawas, B. Suphanit PII:

S0360-5442(17)30198-6

DOI:

10.1016/j.energy.2017.02.024

Reference:

EGY 10317

To appear in:

Energy

Received Date: 2 September 2016 Revised Date:

22 December 2016

Accepted Date: 5 February 2017

Please cite this article as: Satanphol K, Pridasawas W, Suphanit B, A study on optimal composition of zeotropic working fluid in an Organic Rankine Cycle (ORC) for low grade heat recovery, Energy (2017), doi: 10.1016/j.energy.2017.02.024. 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.

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A study on optimal composition of zeotropic working fluid in an Organic Rankine Cycle (ORC) for low

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K. Satanphol, W. Pridasawas, B. Suphanit*

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grade heat recovery

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Department of Chemical Engineering, Faculty of Engineering, King Mongkut’s University of Technology Thonburi, 126 Pracha Utit Rd., Tungkru, Bangkok 10140,

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Thailand.

*Corresponding author: B. Suphanit Tel. +66-2-470-9222 ext. 401

Fax. +66-2-428-3534

E-mail: [email protected]

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Abstract The Organic Rankine Cycle (ORC) is an interesting heat recovery alternative for low grade heat at the present time. In an ORC, the irreversibility of heat transfer in the cycle

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could be reduced by the application of zeotropic working fluid. In this work, the potential of zeotropic working fluid application in an ORC for low grade heat recovery was investigated. The types of fluid, the composition and the operating conditions that

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achieved the maximum net work output were determined through flowsheet modeling and optimization in Aspen Plus v.8.4 simulation software. The working fluids

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considered in this study were the components found in the 400-series refrigerant blends from REFPROP database. Among the group of pure working fluids in this study, the ORC using R-227ea provided the best performance in terms of net work output. In case of zeotropic working fluid, the optimum performance fluid was the blend of R(32.1/13.4/38.8/15.7).

Besides,

when

including

the

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218/227ea/C318/245fa

environmental factor into consideration, the optimum low-GWP blend consisting of R290/152a/600a/601a (35.1/38.1/22.4/4.4) was determined. The possible binary and

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ternary blends of the resulting constituents were also investigated. In addition, the

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thermo-economic analysis of each working fluid was carried out and discussed.

Keywords

Low grade heat, maximum net work output, zeotropic working fluid, Organic Rankine Cycle

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1. Introduction

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In various kinds of processes such as fuel combustion in power plants, chemical reactions in industrial, electronic or biological processes, a large amount of waste heat is commonly lost to environment, leading to low energy efficiency in the process. In

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addition, the waste heat release is also a contributing factor to the global warming which is threatening the world at the present time. Any measures to reduce or recover the

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waste heat are therefore considered with high priority in all kinds of processes. A massive amount of medium (230-650oC) and low grade (<230oC) waste heat [1] is emitted from certain types of equipment such as gas turbine, steam boiler, internal combustion engine (ICE), and furnace. The recovery of the medium and low grade

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waste heat could be carried out in numerous ways. It may be recovered as thermal energy to be used directly in the process. In some cases, the recovered energy usage may be rather limited if the waste heat temperature level is too low. One of the

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interesting recovery approaches being focused on recently was the application of an Organic Rankine Cycle (ORC) to recover waste heat and to transform it into power or

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electricity. The ORC is versatile to recover energy from various sources such as geothermal, solar or waste heat from industrial processes [2, 3]. In an ORC, low boiling compounds (mostly organics) are used as the working fluid to recover medium or low grade waste heat. In general, the critical temperature of a working fluid used in an ORC is much lower than that of water in a typical steam cycle at the same operating pressure [4]. This is a clear advantage of the ORC over the conventional Rankine especially in the low temperature region. Apart from the ORC, the other interesting power cycles for

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waste heat recovery are Maloney-Robertson cycle and Kalina cycle which use ammonia solution as the working fluid. However, when compared them with an ORC under similar size, it was found that the ORC could provide approximately the same or better

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performance at a much lower pressure [5, 6]. Consequently, much larger attention was focused on the ORC as shown by thousands of related journal articles published in recent years. The overview of the ORC in various aspects, e.g. cycle architectures, types

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of working fluids, thermodynamic and economic performances or equipment hardware, can also be found in some recent review articles [7-10]. In our work, however, not every

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aspect involving the ORC was considered. Only the influence of zeotropic working fluid on the ORC performance was focused on.

Depending on the slope of vapor saturation line on the temperature-entropy diagram, the

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working fluids used in an ORC can typically be divided into three types; i.e. wet fluid (negative slope), dry fluid (positive slope) and isentropic fluid (infinite slope) [11]. For dry or isentropic working fluid under subcritical region, the condition at the turbine inlet

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is only needed to be saturated or low degree of superheated vapor, respectively. Under these circumstances, there will be no operating difficulty involving the liquid droplet

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formation when these fluids expand across the turbine. On the other hand, the turbine inlet condition of most wet fluids in an ORC must be superheated to a certain degree otherwise the liquid droplet formation during expansion may occur. The damage on turbine blades could then follow as a consequence. The results from several research works also suggested that dry and isentropic fluids were the most favorable types of fluids for an ORC [12-18].

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The performance of an ORC can be affected by several factors. The choice of working fluid is one of the key factors among others such as the cycle operating region (subcritical or transcritical) and the cycle configuration. Numerous works on working

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fluid selection for an ORC have been studied to choose the right working fluid in various aspects [11-13, 19, 20]. Several screening criteria for choosing a suitable working fluid such as cost, cycle performance, environmental impact, thermal stability,

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flammability or toxicity may be applied. Basically, the cycle performance is the main objective used to identify an appropriate working fluid. In general, this could be

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accomplished by the optimization of either the net work output [14, 21] or the thermodynamic efficiencies [22-24]. Other economic indicators such as the ratio of net work output to total cost [21], the levelized cost of electricity (LCOE) [24], or the ratio of heat exchanger area to net work output [25] might also be set as objectives. Many

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researchers [14, 16, 18, 20, 26] tried to relate the cycle performance with the critical temperature of the working fluid in order to establish simple criteria for fluid selection. For instance, He et al. [16] investigated 22 pure working fluids operating in subcritical

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region; and found that the working fluids with the critical temperatures close to the heat source temperature tend to perform better in terms of net power output than those with

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higher or lower critical temperatures. However, their investigation was only limited to the subcritical ORCs. The deduction from their work might not be valid for other working fluids operating in the transcritical region. Ayachi et al. [18] showed that the working fluid with the best thermodynamic performance operated in the transcritical region. The optimum critical temperature could be identified. The fluids with the critical temperatures lower than the optimum critical value tended to operate optimally in the transcritical region but with lower cycle performance. On the other hand, the fluids with

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the critical temperatures higher than the optimum critical value tended to operate optimally in the subcritical region. The optimum critical temperature from Ayachi et al. [18] was around 0.7-0.8 times the heat source temperature. Others may suggest different

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criteria such as 0.5 times [20] or 0.8 times [27] the heat source temperature. These criteria may vary due to the differences in heat source temperature or the objective

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function used in various studies.

As discussed above, an ORC operating in the transcritical region was found to deliver

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better performance than the same ORC operating in the subcritical region due to better temperature driving force distribution in the evaporator. The temperature differences between the heat source and the working fluid along the heat transfer path were more uniform in case of the transcritical ORC than those in the subcritical ORC. The

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temperature profile of the working fluid in this situation may also be called temperature glide [28]. In general, the larger the temperature glide, the better the performance of an ORC. Although the efficiency could be improved by operating an ORC in the

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transcritical region, the heat transfer area requirement in an evaporator increases as a result of decreasing in temperature driving force. In addition, the effect of changes in

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some physical properties of fluid mixture, e.g. thermal conductivity and viscosity in the evaporator could also degrade the heat transfer characteristic, thereby leading to a larger area requirement as above. The evaporator cost was therefore inevitably higher in case of the transcritical ORC.

In a transcritical ORC, a good temperature glide can be achieved in an evaporator. However, on the condenser side, a temperature glide does not occur if the working fluid

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is pure. To improve the temperature glide in both evaporator and condenser, the other possible alternative, i.e. the application of a zeotropic mixture, may be used instead. The number of research works on the application of zeotropic working fluid in an ORC

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increased remarkably during the last couple of years. Several researchers had already confirmed the benefit of gliding improvement from zeotropic fluid in both evaporator and condenser on the ORC performance [17, 20, 22, 23, 29-34]. Some of them found

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that the performance improvement was mainly a consequence of the good gliding match in the condenser [20, 22, 23, 30]. The zeotropic mixtures mostly considered in the past

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literatures were binary mixtures of some preselected fluids. Hence, the studies on optimal zeotropic composition in an ORC were primarily carried out through sensitivity analysis over the whole composition range [17, 22, 29-32]. The other optimization methods such as the genetic algorithm [20], and a Generalized Reduced Gradient

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(GRG) nonlinear multistart algorithm [23] were also used in the determination of optimal binary composition. The multi-objective optimization was also investigated by Clarke et al. [35] and Feng et al. [36] in order to simultaneously optimize the

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thermodynamic and economic performances. The studies on optimal composition of ternary or multicomponent working fluids were rare. Chys et al. [29] investigated the

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effect of the third additional component of several ternary working fluids on the ORC performance. They found that the addition of the third component into a binary fluid yielded insignificant improvement in terms of both power output and cycle efficiency when compared to the original binary fluid. Recently, Chaitanya et al. [37] used the sequential quadratic programming (SQP) method to optimize composition of several zeotropic mixtures which are mostly hydrocarbons in an ORC. The results showed that several good performance ternary and multicomponent fluids could be found. There was

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still no clear conclusion on the suitable number of components to be used in a zeotropic fluid. Therefore, the identification of high-performance multicomponent zeotropic fluid should be considered further since there may be some good combinations waiting to be

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discovered.

In this work, the potential of applying zeotropic working fluid in an ORC for low grade

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heat was explored. Our approach was different from those in some previous works. Rather than considering some preselected multicomponent mixtures then performing the

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optimization [29, 37], the working fluid selection and the optimization of fluid compositions were carried out simultaneously. The compositions of fluid species which were not suitable would be reduced to zero during the optimization. The number of fluid species in the zeoptropic fluids could then be reduced to optimum. Our detail

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optimization approach was described in section 3. The objective function considered in this work was to maximize the net work output. Both transcritical and subcritical conditions were considered in this work to evaluate the maximum possible performance

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of an ORC operating with each working fluid. All simulations and optimizations of zeotropic fluid composition used in an ORC were carried out by Aspen Plus V8.4

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simulation software.

2. Working fluids

In this study, the selected working fluids were the components found in the ASHRAE R400-series refrigerant blends and some additional fluorocarbon components from REFPROP database [38]. As shown in Table 1, a total of 24 working fluids were investigated. All of them were listed in order of their normal boiling points. The other

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related thermodynamic properties such as critical temperature and pressure, the molecular weight, the type of fluid expansion characteristic and the environmental, health and safety indices were also shown on the same table.

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Insert Table 1 In this work, the major criterion used to evaluate the zeotropic working fluid was the net

the first priority was also investigated.

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3. ORC modeling and optimization

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work output. Besides, the scenario in which the environmental issue was considered as

To optimize the performance of an ORC operating with a pure or zeotropic working fluid, a simulation model of a conventional ORC is necessary. Aspen Plus V8.4 simulation software was applied for this task. A conventional ORC model consists of 4

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main pieces of equipment; i.e. an evaporator, a turbine, a condenser and a pump (cf. Fig. 1). In the evaporator (EVAP), the working fluid is heated up from the subcooled liquid to the saturated or the superheated vapor condition by recovering heat from a waste heat

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or geothermal heat source. After leaving the evaporator, the hot working fluid vapor is expanded across the turbine (TURBINE) to generate work. The turbine exhaust vapor is

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then condensed into the saturated liquid by heat rejection through the condenser (CONDENSE) using cooling water as the cooling medium. Finally, the condensed liquid is pumped back to the evaporator to continue cycle operation. Insert Figure 1 Insert Table 2

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The thermodynamic properties of a working fluid at any particular point in an ORC were calculated by REFPROP model [38] in Aspen Plus software. All ORC optimizations were executed with regard to a set of assumptions, specifications and

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constraints as shown in Table 2. The objective function used in this work was the maximization of net work output. The optimization variables considered in case of pure and zeotropic working fluids were the mass flow rate of each constituent, the evaporator

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pressure (or pump discharge pressure), the condenser pressure (or turbine exit pressure), and the degree of superheating of the working fluid at the evaporator exit. Here, we

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have made an assumption on the lower bound of the degree of superheating of the working fluid at the evaporator exit. Instead of using 0K to represent the saturated vapor condition, a marginal value of 1K was specified in order to avoid any discontinuities which may arise at the vapor-liquid phase boundary during optimization. If an optimum

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solution appears to be at 1K of superheating, it is considered to be an optimum at the saturated vapor condition. For the optimization in the subcritical region of a pure working fluid, the critical pressure was taken as the upper bound value of the evaporator

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pressure. However, for some working fluids, the evaporator pressure may approach the critical region during the optimization. It is therefore possible that the optimum

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condition in the evaporator may lie above the critical point of that working fluid. The operating region considered in the optimization will be switched from the subcritical to the transcritical region. In such cases, the upper bound of the evaporator pressure will be set arbitrarily to a value higher than the critical value. The degree of superheating variable will also be replaced by the working fluid temperature at the evaporator exit.

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Due to the non-linear nature of this optimization problem, the sequential quadratic programming (SQP) method was selected to solve this problem. Here, the SQP was used as a black-box method by converging tear stream variables and design

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specifications separately in two nested loops. In this manner, the number of derivative evaluations in any iteration could be reduced since tear stream variables were converged separately. Moreover, this convergence approach is more robust than the default

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approach by Aspen Plus in which tear variables are converged simultaneously with other optimization variables. In this problem, the flowsheet convergence involves three

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separate loops. First, at any particular values of optimization variables, the inner loop is used to achieve tear variable convergence (pump inlet pressure) via the Wegstein method. Next, the cooling water (CW) flow is adjusted to achieve the specified CW return temperature via the Broyden method. Finally, the optimization block is

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converged by the SQP method in the outer loop. The optimization tolerance was set at the default value of 0.001.

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The optimal zeotropic composition was also determined via optimization in which the flow rate of each constituent was varied simultaneously with other operating variables.

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However, due to a large number of pure components considered in this study, all constituent flow values could not be varied simultaneously. The total number of components in this study exceeded the allowable limit in REFPROP model which was 20 components [38]. Moreover, the optimization of a large number of variables was time-consuming. The initial guess which could produce a converged flowsheet result was also very difficult to obtain. Hence, several optimization problems were carried out sequentially in this study. Starting with only a small group of randomly chosen

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components (around 6 components), the optimization was executed until reaching convergence. Some constituent flow values were reduced to zero at optimum. These constituents were then replaced with the other unconsidered components. The

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optimization was then repeated until all pure components were considered in the problem. Any solution obtained could not however be guaranteed as a global optimum due to the non-linearity of the system. The final solution largely depended on the initial

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guess. Hence, several initial guesses were tried to ensure the best possible solution from

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the optimization.

Yet, there is still another issue left to be resolved. Unlike the case of pure fluid in which the upper bound value of the evaporator pressure is generally set at the critical pressure of working fluid, the appropriate upper bound value in case of a zeotropic fluid is

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undetermined ahead of the optimization. The critical pressure of a zeotropic fluid depends on the fluid composition, thereby varying during optimization. Specifying the largest critical pressure value of the existing constituents in a zeotropic fluid as the

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upper bound sometimes leads to an optimum which might be poorer than those obtained from some pure fluids. To solve this problem, a series of optimization were then

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executed. Starting from choosing an arbitrarily low pressure value as the upper bound of the evaporator pressure, the optimization is performed repeatedly, each time with a small increase in the upper bound value. Once reaching the value in which the flash convergence in the evaporator can no longer be achieved at optimum, the maximum evaporator pressure of an ORC is then determined. This situation generally occurs as the condition approaches the critical region. For mixtures consisting of constituents with a wide range of volatilities as shown by a critical temperature ratio greater than 2, the

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property calculation at saturation generally fails [38]. The optimization region of the wide-boiling zeotropic fluids was mostly limited in the subcritical region due to this limitation. The overall optimization procedures for optimal zeotropic composition in an

Insert Figure 2

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4. Thermo-economic analysis

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ORC were summarized as shown in Figure 2.

Apart from the optimization of net work output, the optimum ORC using each working

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fluid was also analyzed in terms of both thermodynamic and economic performances. The thermodynamic analysis was carried out by evaluating the exergy efficiency of each piece of equipment and the overall cycle. All ORCs were also evaluated and compared on the economic point of view. The exergy and economic analysis of an ORC are

4.1 Exergy analysis

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briefly described here.

To evaluate the efficiency of energy conversion in a thermodynamic cycle, the analysis

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was generally considered in terms of the converted quantity according to the first law of thermodynamics. However, the second law analysis was also applied as an additional

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tool to analyze the quality of energy or the potential of thermal energy to produce work. According to the second law, the exergy efficiency was a good performance indicator for each piece of equipment in a cycle. By considering this value, inefficient pieces of equipment in the cycle can be identified and then investigated for any possible alternatives in order to improve the cycle performance. Here, the concise fundamentals of exergy analysis and exergy efficiency calculation were described. Exergy of a process stream at a given temperature and pressure (T and P) is defined as the maximum

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amount of work output which can be obtained as the stream changes reversibly from that given state to the state of equilibrium with the environment at T0 and P0, hence given by:
 =  −  

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(1)

where H and S are the stream enthalpy and entropy with reference to the ambient

conditions T0 and P0. The typical values of reference temperature and pressure are 25oC

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and 1 atm, respectively.

As a matter of fact, exergy of a stream is a direct function of its enthalpy and entropy,

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which are both functions of state, the exergy itself is therefore a state function. Thus, the exergy value of a given stream can be calculated from the stream properties (composition, temperature and pressure). If a stream changes from state 1 to state 2, the exergy change of the stream is given by:

(2)

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∆
  =
 −
 =  −   −   −  

According to the second law of thermodynamics, all natural processes are irreversible and therefore lead to the degradation of energy. Whenever energy is transformed or

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transferred, its potential to produce useful work or exergy is reduced forever. For that reason, to accomplish a certain exergy output, a real process always requires a higher

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exergy input. The difference between them is the exergy lost due to irreversibility.
 −
 =
 =


(3)

The exergy loss value provides a measure of thermodynamic efficiency of a process. The maximum thermodynamic efficiency may be achieved by a reversible process in which there is no exergy loss. Reversible processes, however, have no driving forces and can never be achieved in practice. The exergy efficiency of a process can be defined as follows:

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= 1 −

!""

(4)



The exergy efficiency used in this work was the rational [39] or functional [2] exergy efficiency. This definition of exergy efficiency was suitable for the system or equipment

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with good understanding of purpose or behavior. In an ORC, the exergy input could be either the heat or work flow into a unit, or the exergy provided by a stream flowing through that unit. Similarly, the exergy output could be either the exergy taken up by a

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stream flowing through a unit, or the heat or work flow out of the unit. Applying eq. (3) and (4) in an ORC, the exergy loss and the exergy efficiency of each piece of equipment

Pump:

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can be determined as follows:


,$%$ = &'$%$ & − ∆
(),$%$ ,$%$ = 1 −

!"",*+* &(*+* &

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Evaporator:


,,-.$ = |∆
0( | − ∆
(),,-.$

Turbine:

!"",123* |∆45 |

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,,-.$ = 1 −

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,6 = &∆
(),6 & − '6 

,6 = 1 − &∆!"",78 & 59,78

(5) (6)

(7) (8)

(9) (10)

Condenser:


,:; = &∆
(),:; & − ∆
<( 

,:; = 1 − &∆!"",=> & 59,=>

(11) (12)

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For an ORC, the overall exergy loss is the combination of all exergy losses from all pieces of equipment. The same definition of exergy efficiency as in eq. (4) can be applied to the overall system as shown in eq. (14).

∆@5 A(78 45 |A&(*+* &

,-,. = |∆



= 1 − |∆ !"",2173!! |A&( 45

(14)

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4.2 Economic analysis

*+* &

(13)

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,-,. =
,$%$ +
,,-.$ +
,6 +
,:;

In addition to the thermodynamic performance of an ORC, the economic performance is

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also another essential factor to be considered. The working fluid capable of yielding not only good energy performance but also low cost of electricity production would be very attractive for being used in an ORC. However, this might not be the case in general. For example, in transcritical ORCs, the systems could deliver good energy performance

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since they operated with large pressure change across the turbine. On the other hand, the capital investment was rather high in such cases due to the high pressure condition in the evaporator and the turbine. This led to poor economy which was in contrast to the

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good energy performance. Basically, the cost of electricity generated from an ORC mainly depends on the capital investment of the ORC itself. The operating costs such as

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cooling water cost, labor cost, or maintenance cost are typically estimated in proportion to the capital investment. In this study, however, these estimated operating costs were not considered in the analysis. The economic performance of an ORC operating with each working fluid was therefore evaluated on the basis of the estimated specific purchased equipment cost (SPEC) in $/kWe. The detail of equipment cost estimation was described in the appendix. The electricity production capacity was calculated from

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the net electricity produced from an ORC. The electricity conversion efficiencies were assumed to be 100% and 96% for pump and turbine, respectively.

5.1 Optimization of ORC using pure fluid

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5. Results and discussion

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Although the optimizations of ORCs using numerous pure fluids had been done extensively in the past, the assumptions used in previous studies were different from

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those used in this work. The comparison of performance between pure and zeotropic fluids would therefore be inconsistent if the results from some previous works were used. The optimizations of ORCs using pure fluids were then carried out under the assumptions used in this study. The results were evaluated and employed as bases for

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comparison with the performance of ORCs using optimal zeotropic working fluids. For each considered working fluid, all variables affecting the net work output as described in section 3 were optimized simultaneously. The subcritical region was considered at

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the beginning. However, for some cases in which the optimization results approached the critical region, the optimization problems were then relocated to the transcritical

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region instead. From Table 3, it was found that the ORCs using pure working fluids with the critical temperatures approximately below 100oC tended to operate optimally in the transcritical condition. The optimum critical temperature was found to be around 0.7 times the heat source temperature. This coincided with the results from some previous works [18, 27]. There were 10 working fluids of this kind which ranked top on the list as shown in Table 3. The remaining fluids which operated optimally in the subcritical region then followed in subsequent order except R-C318 which had an outstanding

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performance above others in the subcritical group. From Table 3, R-227ea was the best performance fluid in terms of the net work output while R-123 was the worst performance fluid under the assumptions in this study. The best pure fluid obtained (R-

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227ea) was similar to the result at the same heat source temperature from ChagnonLessard et al. [40]. The ranking of the top 5 pure fluids was also in close agreement with the results obtained by Walraven [41] for simple ORCs. The first law or thermal

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efficiencies of all pure working fluids were in range of 8.37-10.30%. The results showed that there was no distinct linkage between the net work output and the thermal

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efficiency obtained by these working fluids. On the other hand, the net work output was largely affected by the amount of heat recovery in the evaporator. The ORC operating with high heat recovery level in the evaporator tended to produce a large net work

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output.

The amount of heat recovery in the evaporator depended on the temperature profiles of both hot and cold streams. Effective heat recovery could be achieved in the evaporator

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having appropriate temperature glide or rather uniform temperature difference along the heat transfer path such as those found in the evaporators operating at the supercritical

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condition. On the contrary, the heat recovery in the evaporator operating at the subcritical condition was less effective due to the pinch limitation generated by the phase-change location. From Fig. 3, the ORCs operating in the transcritical region, e.g. R-227ea, R-115, R-143a, R-125, or R-218, exhibited relatively uniform temperature differences in the evaporators. The temperature differences in these evaporators were also very close to the given minimum temperature approach (10oC). Hence, the heat recovery of the evaporator operating at the supercritical condition was more effective

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than those operating at the subcritical condition. For working fluids in the subcritical group, dry and isentropic fluids, e.g. R-236fa, R-124, R-12, R-600a, operated optimally near the saturated vapor condition while wet fluids such as R-E170 operated optimally

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at a significant degree of superheating at the evaporator outlet as shown in Fig. 3. Considered in terms of the optimum operating pressure in the evaporator or the turbine inlet pressure (PT,in), it was found that the transcritical ORCs operated at a much higher

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pressure than the subcritical ORCs did. The potential to generate work through the turbine was therefore high in case of the transcritical ORCs. In condenser, the

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temperature profiles of the heat-exchanged streams were rather similar in both transcritical and subcritical ORCs. The phase change occurred at constant temperature hence there was no temperature gliding on the working fluid side. For some working fluids such as R-227ea or R-143a, the internal pinch which limited the condensing

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temperature could be found. For others such as R-290 or R-134a, the pinch may occur at the hot end of the condenser. For all pure working fluids, the condensing temperature in the condenser were found to be in range of 40-45oC. Since the condensing temperature

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and pressure were directly related, the condensing pressure was limited by this temperature range. Hence, the potential to generate more work through the turbine was

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reduced.

Insert Table 3 Insert Figure 3

5.2 Optimization of ORC using zeotropic working fluid

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After evaluating the ORC performance of each pure fluid, the ORC using zeotropic working fluid was further explored to identify any possible improvement on the net work output. By applying the optimization procedure outlined in section 3, it was found

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that from the total of 24 fluid species considered, the optimum blend of zeotropic working fluid consisted of R-218, R-227ea, R-C318, and R-245fa with the composition of 32.1, 13.4, 38.8, and 15.7 wt%, respectively. The number of fluid species was

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reduced from 24 to only 4 at optimum by using the proposed methodology. Both fluid selection and optimal compositions were obtained from the optimization. The boiling

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range of this optimum blend was quite wide. The critical temperature ratio of the heaviest (R-245fa) to the lightest constituents (R-218) in the blend was around 2.14 which was higher than the limitation of property calculation around critical region in REFPROP model [38]. Hence, the optimum zeotropic fluid was found to operate under

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the subcritical region. The evaporator pressure of this blend was 2.73 MPa which was significantly lower than that of the best pure fluid in the blend (3.37 MPa for R-227ea). This blend also operated near vapor saturation at the evaporator outlet. When comparing

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the performance of this blend with R-227ea, the maximum net work output of the ORC using this blend was 461.61 kW (cf. Table 4) which was around 8.2% higher than that

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of the ORC using R-227ea. The thermal efficiency was however lowered by around 0.22% points. The improvement in the net work output resulting in this case was largely due to high heat recovery in the evaporator. The amount of heat recovery in the evaporator in this case was 5,188 kW which was around 10.9% higher than that in case of R-227ea. Insert Table 4

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When this optimum zeotropic blend was considered in terms of health and safety issues, it was found that three constituents namely R-218, R-227ea, and R-C318 were classified in A1 group while R-245fa was in B1 group according to ASHRAE standard 34-2013

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[42]. All constituents in this blend were in the non-flame propagation group (group 1). Therefore, this blend was safe in terms of flammability. However, the toxicity of this blend must be evaluated before assigning the safety classification to the blend since it

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contains R-245fa which is toxic (group B).

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On the environmental issue, this blend has zero Ozone Depletion Potential (ODP), but the weight-average Global Warming Potential (GWP) is very high at 7,475. This blend does not correspond to the GWP criteria of the fourth generation refrigerants which delimited the acceptability of the working fluid with GWP below 150 [43]. If the

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environmental concern is on the top priority when selecting a working fluid for an ORC, only low-GWP components should be allowed in the consideration. In this study, there were only 9 low-GWP components, i.e. R-1270, R-290, R-E170, R-152a, R-600a, R-

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600, R-123, R-601a, and R-601. The optimization of the ORC using zeotropic working fluids in this low-GWP group could be carried out in the same manner as earlier. It was

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found that the optimum blend of the low-GWP working fluids consisted of R-290, R152a, R-600a, and R-601a with the composition of 35.1, 38.1, 22.4, and 4.4 wt%, respectively. When compared to the performance of the best pure fluid in this blend (R290), the maximum net work output of the ORC using this low-GWP blend was 403.88 kW (cf. Table 4) which was around 6.4% higher than that of the ORC using R-290. The thermal efficiency was also higher by 0.44% points. The amount of heat recovery in the evaporator in this case was 4,212 kW which was only 1.5% higher than that in case of

22

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R-290. The improvement in the net work output resulting in this case was to a lesser extent affected by a marginal increase in heat recovery in the evaporator but rather by the improvement in the cycle efficiency. This optimum blend operated at the subcritical

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condition and a low degree of superheat at the evaporator outlet. The maximum and the minimum pressures (the turbine inlet and outlet pressures) in the ORC using this blend were also lower than those in case of R-290. This low-GWP blend has zero ODP. The

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weight-average GWP is 63 which is below the environmental criteria. Also, this lowGWP blend has low toxicity (group A). The only concern for this low-GWP blend is on

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the flammability issue since most of the constituents are highly flammable (A3 group) [42].

5.3 Optimization of ORC using binary and ternary zeotropic fluids

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From the results in the previous section, the optimal number of constituents in both optimal blends was four. However, the optimal zeotropic fluids obtained may not be the global optimal results due to the non-linearity of the system. It is therefore worth

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considering the possible binary and ternary fluids containing the constituents presented in the optimal fluids. Some equivalent or even better performance fluids may be found.

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The optimization results of all possible binary and ternary fluids for both optimal zeotropic fluids were shown in Table 5. The results of some possible ternary or binary fluids, e.g. R-218/227ea/C318 or R-C318/245fa, may be missing since they were reduced to binary or pure fluids at optimal. Insert Table 5 For the group of R-218, R-227ea, R-C318 and R-245fa, it was found that the ternary zeotropic R-218/C318/245fa and R-218/227ea/245fa could deliver the equivalent or

23

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slightly better performance in terms of the net power output than the optimal quaternary zeotropic R-218/227ea/C318/245fa. For optimal ternary zeotropic fluids, R-218 and R-

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245fa are important constituents together with either R-C318 or R-227ea.

For the group of R-290, R-152a, R-600a, and R-601a, it was found that the ternary zeotropic R-290/152a/600a could perform slightly better in terms of net power output

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than the optimal R-290/152a/600a/601a. It could be seen that the presence of R-601a in the optimal R-290/152a/600a/601a was only 4.4%. When R-601a was removed from the

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blend, it did not show any detrimental effect on the ORC performance. The other ternary and binary fluids did not perform better than the optimal R-290/152a/600a/601a and R-290/152a/600a.

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5.4 Thermo-economic analysis

All optimized ORCs using pure fluids and zeotropic working fluids were further analyzed in terms of exergy efficiency and the ORC investment cost as described in

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section 4. The exergy efficiencies of all pieces of equipment and the ORC system using each working fluid were determined. Their results were shown in Table 6.

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Insert Table 6

In case of the optimum ORCs using pure working fluids, it can be seen that the exergy efficiencies of pump, turbine, and condenser in all cases were 85.37, 86.25, and 26.09%, respectively on average with only a small variation. On the other hand, a much larger variation in the exergy efficiencies of the evaporator can be observed. Therefore, it was obvious that the evaporator was the efficiency dominating equipment in the ORC system using pure fluid. The exergy efficiencies of the evaporator in the transcritical

24

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ORCs (82-87%) were in a higher range than those in the subcritical ORCs (72-76% excluding R-C318). The overall exergy efficiencies of the ORCs using pure working fluids also followed relatively the same trend as the net work outputs. In most of the

subcritical ones (cf. Table 6). Insert Figure 4

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transcritical ORCs, the overall exergy efficiencies were higher than those of the

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In case of the optimal ORCs using zeotropic working fluids, it can be seen that the exergy efficiencies of pump and turbine of the ORCs using both optimum zeotropic

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fluids were very close to those using pure fluids. For the ORC using the optimum R128/227ea/C318/245fa blend, the exergy efficiency of the condenser was improved by 7% points when compared to that in case of R-227ea. This was the result of a significant gliding improvement in the condenser as shown in Fig. 4a). The temperature difference

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along the temperature profiles in the condenser was rather uniform at the minimum allowable value (10K). On the contrary, the exergy efficiency of the evaporator was deteriorated by 3.8% points due to the subcritical operation in the zeotropic fluid case

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when compared to the transcritical operation in case of R-227ea. However, the temperature gliding in the evaporator was better than those using pure fluids operating

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under the subcritical region. A large proportion of uniform temperature difference could be observed along the temperature profiles in the evaporator. The suitable gliding matches found in both evaporator and condenser were as a result of the optimal fluid composition. In this case, the gliding improvement in the condenser was the dominating side since the best possible gliding match had been achieved on this side with almost uniform temperature difference along the profiles. These findings were consistent with

25

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those found in some previous literatures [20, 22, 23, 30]. The overall exergy efficiency of the ORC using this optimum blend was also higher than that in case of R-227ea.

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For the ORC using the optimum low-GWP blend (R-290/152a/600a/601a), a similar observation could be found in both condenser and evaporator as shown in Fig. 4b). The exergy efficiency of the condenser was improved by 6.7% points while the exergy

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efficiency of the evaporator was deteriorated by 4% points when compared to that in case of R-290 (cf. Table 6). The overall exergy efficiency of the ORC using this

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optimum low-GWP blend was a little higher than that in case of R-290.

The economic performance of the optimum ORC using each working fluid was represented in terms of the estimated SPEC as shown in Table 6. For the ORCs using

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pure working fluids, the estimated SPECs of the transcritical ORCs ($5,825-8,911/kWe) were in a higher range than those of the subcritical ORCs ($4,499-5,531/kWe). Therefore, the high-capacity ORCs operating in the transcritical region typically led to

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poor economy. However, that was not the case when using the optimum zeotropic working fluid, the economic performance was also better when compared to the best

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pure fluid in the blend. The estimated SPEC of the ORC using the optimum R128/227ea/C318/245fa blend was $5,493/kWe which was lower than that in case of R227ea. Similarly, the estimated SPEC of the ORC using the optimum low-GWP (R290/152a/600a/601a) blend was lower than that in case of R-290. It was clearly seen that the optimum zeotropic fluid was better than its best pure constituent in terms of both thermodynamic and economic performance.

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6. Concluding remarks In our contribution, the potential of applying zeotropic working fluid in an ORC for low grade waste heat was thoroughly explored via optimization. The proposed optimization

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strategy has shown to be sufficiently effective in finding suitable constituents of a zeotropic fluid, their optimal compositions and operating conditions for an ORC. By maximizing the net work output, the optimum zeotropic fluid found in this work was a

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quaternary blend consisting of R-218/227ea/C318/245fa (32.1/13.4/38.8/15.7). Both thermodynamic and economic performances of this optimum zeotropic blend were

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better than those of the best pure fluid in the blend (R-227ea). This was mainly as a result of thermodynamic improvement in the ORC condenser and higher heat recovery in the evaporator when compared to that of the best pure fluid. The major advantages of the optimum zeotropic blend over the best pure fluid were lower operating pressure

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range and lower capital investment. In addition, when the other criterion such as environmental issue was considered in high priority, only the fluids passing the criterion were considered. The optimum low-GWP blend, i.e. R-290/152a/600a/601a (35.1/38.1/

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22.4/4.4), could also be obtained. Similarly, both thermodynamic and economic performances of this optimum low-GWP blend were better than those of the best pure

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fluid in this blend (R-290). However, the improvement in this case was primarily due to better thermodynamic efficiency in the ORC condenser since the increase in heat recovery in the evaporator was only marginal when compared to that of R-290.

The optimum zeotropic fluids obtained from the proposed optimization strategy could not be guaranteed as the global optimum due to the highly non-linear nature of the system. There could be other zeotropic fluids which may perform better. When

27

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considering the other possible ternary and binary fluids consisting of the similar constituents as in the optimal fluids, it was found that some ternary fluids, i.e. R218/C318/245fa, R-218/227ea/245fa, or R-290/152a/600a, could deliver equivalent or

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slightly better performance than the optimal fluids in the same group of components. It can be concluded that the optimal number of constituents in the zeotropic fluid was

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either 3 or 4 under the conditions and components used in this study.

For future work, other optimization approaches such as stochastic methods or genetic

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algorithm may be considered. The other limitation in this work was the property calculation problem of wide-boiling zeotropic fluids in REFPROP model when approaching the critical region. Future improvement in the property calculation of zeotropic fluids around the critical region could help unveiling new high-performance

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zeotropic ORCs.

Appendix: Capital cost estimation.

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The cost of each piece of major equipment in an ORC was estimated by the cost models as outlined by Turton et al. [44]. The purchased costs of equipment were based on the

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cost survey in 2001 (CEPCI = 394.3). The costs calculated from the models were then adjusted to the present costs when accounting for inflation. The CEPCI value in 2012 (584.6) was used in this study [45]. The purchased cost of the equipment, at ambient operating pressure and using carbon steel in construction, could be estimated by the following equation: log 10 CP0 2001 = K 1 + K 2 log 10 A + K 3 [log 10 A]2 Where CP0 2001 is the purchased cost of the equipment in 2001.

(A-1)

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ACCEPTED MANUSCRIPT A is the capacity or size parameter for the equipment.

K1, K2, and K3 are parameters for various kinds of equipment and capacity. The values used in this study were given in Table A1.

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Insert Table A1

The purchased equipment cost must be corrected with the bare module factor (FBM)

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which accounted for the material (FM) and pressure (FP) factors in construction. The corrected value, now called the bare module cost (CBM), is therefore used as the

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estimated capital cost in the economic analysis. NOP = N$ QOP

(A-2)

For heat exchanger and pump, the bare module factor (FBM) can be calculated as follows:

(A-3)

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FBM = B1 + U FP FM

Where B1 and B2 are constant parameters which depend on type of equipment as shown in Table A2.

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Insert Table A2

In the heat exchanger area estimation, the overall heat transfer coefficients were

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assumed constant as shown in Table A3: [46] Insert Table A3

For turbine, the bare module factor (FBM) was fixed at 3.5. The pressure factor (FP) in equation (3) can be determined as follows: log 10 FP 2001 = C1 + C2 log 10 P + C3 [log 10 P]2 Where P is pressure in barg; Fp (2001) is pressure factor in 2001; and

(A-4)

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C1, C2, and C3 are constant parameters which depend on operating pressure as shown in Table A4.

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Insert Table A4

Atmospheric

CW

Cooling water

Ex

Exergy

Flow

Mass flow rate

GWP

Global warming potential

H

Enthalpy

MW

Molecular weight

ODP

Ozone depletion potential

P

Pressure

Q

Heat load

S

Entropy

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Specific purchased equipment cost

Temperature

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T

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SPEC

W

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Atm

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Nomenclature

Work

WF

Working fluid



Efficiency

∆

Subscripts:

Change, or difference

30

Critical state

CW

Cooling water

cond

Condenser

evap

Evaporator

Ex

Exergy

HW

Hot water

in

Input

irr

Irreversible

net

Net output

out

Output

turb

Turbine

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C

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Boiling point

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EP

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B

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ACCEPTED MANUSCRIPT References

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[35] Clarke J, McLeskey JT. Multi-objective particle swarm optimization of binary geothermal power plants. Applied Energy. 2015;138:302-14. [36] Feng Y, Hung T, 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 Conversion and Management. 2015;106:85972. [37] Chaitanya Prasad GS, Suresh Kumar C, Srinivasa Murthy S, Venkatarathnam G. Performance of an organic Rankine cycle with multicomponent mixtures. Energy. 2015;88:690-6. [38] NIST. NIST Standard Reference Database 23. Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 80. Gaithersburg2007. [39] Kotas TJ. The exergy method of thermal plant analysis. 1st ed. Great Britain: Butterworths, 1985. [40] Chagnon-Lessard N, Mathieu-Potvin F, Gosselin L. Geothermal power plants with maximized specific power output: Optimal working fluid and operating conditions of subcritical and transcritical Organic Rankine Cycles. Geothermics. 2016;64:111-24. [41] Walraven D, Laenen B, D’haeseleer W. Comparison of thermodynamic cycles for power production from low-temperature geothermal heat sources. Energy Conversion and Management. 2013;66:220-33. [42] ANSI/ASHRAE. ANSI/ASHRAE Standard 34-2013, Designation and Safety Classification of Refrigerants. Safety Group Classifications. USA2013. [43] Calm JM, Hourahan GC. Physical, Safety, and Environmental Data for Current and Alternative Refrigerants. The 23rd International Congress of Refrigeration. Prague, Czech Republic2011. [44] Turton R, Bailie RC, Whiting WB, Shaeiwitz JA. Analysis, synthesis, and design of chemical processes. 3rd ed. Upper Saddle River, N.J.: Prentice Hall, 2009. [45] Lozowski D. Economic Indicators. Chemical Engineering. USA: Access Intelligence; 2014. p. 72. [46] Hewitt GF, Shires GL, Bott TR. Process Heat Transfer. Boca Raton, FL: CRC Press, 1994. [47] Drescher U, Brüggemann D. Fluid selection for the Organic Rankine Cycle (ORC) in biomass power and heat plants. Applied Thermal Engineering. 2007;27(1):223-8.

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Captions for the Figure(s)

Fig. 1. Flow diagram of a conventional ORC.

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Fig. 2. Optimization procedure for optimal zeotropic composition in ORCs. Fig. 3. Temperature profiles in the evaporator and condenser of optimal

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ORCs operating with various pure fluids. Solid lines represent the temperature profiles of hot water and cooling water. Dash lines represent

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the temperature profiles of working fluids.

Fig. 4. Temperature profiles in the evaporator and condenser of the ORCs operating with (a) the optimum zeotropic fluid and (b) the optimum lowGWP zeotropic fluid. Solid lines represent the temperature profiles of hot

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working fluids.

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water and cooling water. Dash lines represent the temperature profiles of

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Table 1 Physical properties of selected pure working fluids. [43] TB (°C)

Type

MW

CH2F2 CHF2CF3 CH3CH=CH2 CH3CF3 CH3CH2CH3 CHClF2 CClF2CF3 CF3CF2CF3 CCl2F2 CH2FCF3 CH3OCH3 CH3CHF2 CF3CHFCF3 CHClFCF3 CH(CH3)2CH3 -(CF2)4CF3CH2CF3 CH3CH2CH2CH3 CHF2CH2CF3 CH2FCF2CHF2 CHCl2CF3 (CH3)2CHCH2CH3 CH3CClF2 CH3CH2CH2CH2CH3

5.78 3.62 4.6 3.76 4.25 4.97 3.16 2.68 4.13 4.06 5.37 4.52 2.91 3.66 3.63 2.78 3.22 3.80 3.64 3.93 3.66 3.38 4.04 3.37

78.105 66.015 91.7 72.73 96.68 96.15 80 71.9 111.8 101.03 126.95 113.29 101.68 122.5 134.65 115.22 124.92 151.97 154.05 174.42 183.79 187.25 137.14 196.55

-51.70 -48.11 -47.7 -47.34 -42.08 -40.83 -39.11 -36.70 -29.79 -26.07 -24.84 -24.02 -16.36 -12.10 -11.87 -5.98 -1.45 -0.55 15.30 25.25 27.83 27.84 32.00 36.06

Wet Wet Wet Dry Wet Wet Dry Isentropic Isentropic Wet Wet Wet Dry Isentropic Isentropic Isentropic Isentropic Dry Dry Dry Dry Dry Isentropic Dry

52.02 120.02 42.08 84.04 44.10 86.47 154.47 188.02 120.91 102.03 46.07 66.05 170.03 136.48 58.12 200.03 152.04 58.12 134.05 134.05 152.93 72.15 100.50 72.15

Table 2

ASHRAE 34 Safety Group A2L A1 A3 A2L A3 A1 A1 A1 A1 A1 A3 A2 A1 A1 A3 A1 A1 A3 B1 B1 A3 A2 A3

ODP

GWP 100 yrs

Atm Life (years)

0 0 0 0 0 0.04 0.57 0 0.82 0 0d 0 0 0.02 0 0 0 0 0 0 0.01 0 0.06 0

716 3420 <20 4180 ~20 1790 7230 8830 10900 1370 133 3580 619 ~20 10300 9820 ~20 1050 726 77 ~20 2220 ~20

5.2 28.2 0.001 47.1 0.041 11.9 1020 2600 100 13.4 0.015 1.5 38.9 5.9 0.016 3200 242 0.018 7.7 6.5 1.3 0.009 17.2 0.009

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TC (°C)

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PC (MPa)

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R-32 R-125 R-1270 R-143a R-290 R-22 R-115 R-218 R-12 R-134a R-E170 R-152a R-227ea R-124 R-600a R-C318 R-236fa R-600 R-245fa R-245ca R-123 R-601a R-142b R-601

Chemical formula

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Working fluid

Assumptions, specifications, and constraints in a conventional ORC model.

5

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2 3 4

Assumptions, specifications, and constraints The conditions of the hot water heat source were assumed to be at 140oC,and 6 bar. The hot water flow rate was 50,000 kg/h. The isentropic efficiencies of pump and turbine were 85%. The minimum temperature approaches in evaporator and condenser were specified at 10K. The condition of working fluid at the turbine exit must be either saturated or superheated vapor. This can be achieved by setting the vapor fraction of working fluid at the turbine exit to one. The minimum allowable pressure at turbine exit is 5 kPa. [47] The operating temperature range of the cooling water was set between 25 and 35oC. The heat loss in all heat exchangers and pipelines was neglected. There was no pressure drop in the evaporator and condenser. There was no chemical loss from the system.

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No. 1

6 7 8 9 10

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36

Table 3

QEvap (kW) 4,677 4,506 4,297 4,584 4,092 4,750 4,078 4,074 4,148 3,925 3,541 3,703 3,610 3,680 3,505 3,568 3,357 3,220 3,344 3,263 3,425 3,370 3,300 3,281

MWF (kg/h) 143,221 152,383 83,729 120,753 52,993 160,795 116,947 45,570 45,792 81,015 67,081 83,069 84,433 95,747 34,706 46,329 55,660 23,929 29,330 53,617 53,561 30,187 27,158 60,322

THW,out (°C) 60.2 63.2 66.8 61.8 70.3 59.0 70.5 70.6 69.4 73.2 79.8 77.0 78.6 77.4 80.4 79.3 82.9 85.3 83.2 84.6 81.3 82.7 83.9 84.2

TEvap,in (°C) 46.5 48.8 49.0 48.7 49.2 47.3 44.6 49.3 48.3 47.6 48.1 44.9 45.6 46.7 45.1 46.5 44.4 44.6 44.7 45.3 44.4 44.2 44.0 44.3

TT,in (°C) 111.3 124.6 127.8 125.4 128.8 125.6 104.7 114.3 108.4 111.8 125.1 94.8 94.5 96.3 93.4 98.1 96.5 126.9 94.6 98.3 90.4 89.9 95.9 97.1

DSPH (°C) 1.0 1.0 1.0 4.7 1.2 8.0 5.7 37.3 3.7 5.9 2.2 1.0 7.3 9.4

SC

Eff (%) 9.12 9.17 9.59 8.78 9.73 8.37 9.37 9.37 9.15 9.66 10.30 9.05 9.06 8.74 9.08 8.79 9.16 9.54 9.16 9.38 8.85 8.97 8.99 8.98

M AN U

Wnet (kW) 426.56 413.35 411.96 402.26 398.19 397.40 382.16 381.64 379.46 379.24 364.82 335.23 327.17 321.79 318.38 313.66 307.57 307.22 306.39 306.04 303.15 302.20 296.68 294.61

TE D

Refrigerant R-227ea R-115 R-143a R-125 R-32 R-218 R-C318 R-1270 R-290 R-134a R-22 R-236fa R-124 R-12 R-600a R-152a R-245fa R-E170 R-600 R-142b R-245ca R-601a R-601 R-123

TT,out (°C) 51.6 60.2 66.5 72.3 57.3 74.5 64.2 45.1 45.0 45.0 53.6 57.0 51.0 45.0 56.6 45.0 62.4 80.3 60.5 54.8 61.2 63.3 69.1 64.6

PT,in (MPa) 3.37 5.94 6.38 6.75 7.35 5.93 2.21 5.53 4.56 4.29 5.30 1.70 2.09 2.86 1.71 2.88 1.02 2.70 1.27 1.79 0.70 0.56 0.46 0.59

PT,out (MPa) 0.79 1.42 1.96 2.11 2.73 1.32 0.54 1.84 1.53 1.16 1.71 0.49 0.67 1.08 0.59 1.04 0.29 0.97 0.42 0.59 0.20 0.17 0.13 0.18

Cycle type TransTransTransTransTransTransSubTransTransTransTransSubSubSubSubSubSubSubSubSubSubSubSubSub-

EP

Rank 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24

RI PT

Optimization results of ORCs operating with pure working fluids.

AC C

Table 4

Optimization results of ORCs operating with zeotropic working fluids. Refrigerant blend R-218/227ea/C318/245fa (32.1/13.4/38.8/15.7) R-290/152a/600a/601a (35.1/38.1/22.4/4.4)

Wnet (kW)

Eff (%)

461.61

8.90

403.88

9.59

QEvap (kW)

MWF (kg/h)

THW,out (°C)

TEvap,in (°C)

TT,in (°C)

DSPH (K)

TT,out (°C)

PT,in (MPa)

PT,out (MPa)

Cycle type

5,188

145,775

51.4

36.5

102.3

1.0

53.9

2.73

0.68

Sub-

4,212

47,987

68.2

39.7

105.2

6.3

46.5

3.65

1.14

Sub-

ACCEPTED MANUSCRIPT

37

Table 5

Eff (%)

QEvap (kW)

MWF (kg/h)

THW,out (°C)

TEvap,in (°C)

TT,in (°C)

DSPH (K)

TT,out (°C)

PT,in (MPa)

PT,out (MPa)

Cycle type

463.01

8.87

5,221

149,933

50.9

36.5

103.1

1.0

54.5

2.74

0.67

Sub-

461.31

9.13

5,055

134,345

53.7

36.9

103.4

1.0

51.3

2.87

0.70

Sub-

397.03

8.27

4,800

145,960

58.1

45.5

95.7

53.2

2.53

0.79

Sub-

426.25

9.13

4,669

140,851

60.4

44.1

SC

1.2

105.3

1.0

57.1

2.63

0.65

Sub-

446.41

8.74

5,107

156,903

52.8

39.5

103.5

1.0

57.1

2.63

0.66

Sub-

426.80

8.10

5,271

146,112

50.0

36.6

100.9

1.0

54.8

2.90

0.84

Sub-

419.99

9.15

4,588

127,853

61.8

44.1

103.2

1.0

52.8

2.74

0.71

Sub-

404.06

9.73

4,154

47,067

69.2

40.6

106.5

5.3

45.3

3.69

1.11

Sub-

379.17

8.69

4,364

43,869

65.6

38.4

98.5

4.8

50.7

3.22

1.19

Sub-

385.33

9.57

4,025

39,247

71.5

38.3

99.5

1.0

45.9

3.08

0.98

Sub-

378.52

9.56

3,961

46,045

72.6

38.9

98.9

1.0

46.7

2.67

0.82

Sub-

378.72

8.77

4,317

56,324

66.4

43.3

102.0

10.7

46.0

3.83

1.38

Sub-

373.80

9.78

3,822

45,260

74.9

41.0

102.1

1.0

45.0

2.95

0.87

Sub-

357.91

9.27

3,863

34,187

74.2

36.1

94.2

1.0

60.7

1.19

0.37

Sub-

389.67

9.58

398.14

9.44

361.19

9.34

EP

TE D

M AN U

Wnet (kW)

AC C

Refrigerant blend R-218/C318/245fa (35.9/50.5/13.6) R-218/227ea/245fa (24.3/55.1/20.6) R-218/227ea (0.4/99.6) R-227ea/C318 (33.5/66.5) R-218/C318 (22.7/77.3) R-218/245fa (68.2/31.8) R-227ea/245fa (89.8/10.2) R-290/152a/600a (34.03/30.35/35.62) R-290/152a/601a (74.1/12/13.9) R-290/600a/601a (62.4/32.8/4.8) R-152a/600a/601a (47.1/46.6/6.3) R-290/152a (24.3/75.7) R-152a/600a (40.9/59.1) R-600a/601a (66.1/33.9) R-290/600a (76.2/23.8) R-290/601a (91.1/8.9) R-152a/601a (73.8/26.2)

RI PT

Optimization results of ORCs operating with binary and ternary zeotropic working fluids.

4,067

41,735

70.7

42.8

105.0

4.3

45.0

3.76

1.20

Sub-

4,219

41,829

68.1

41.0

103.7

6.2

47.2

3.75

1.25

Sub-

3,866

43,497

74.2

37.4

98.6

2.0

50.4

2.16

0.69

Sub-

38

ACCEPTED MANUSCRIPT

Table 6 Results from thermo-economic analysis of ORCs operating with various working fluids. Refrigerant

Wnet

Exergy efficiency Pump

Evaporator

Turbine

Condenser

Pure R-227ea R-115 R-143a R-125 R-32 R-218 R-C318 R-1270 R-290 R-134a R-22 R-236fa R-124 R-12 R-600a R-152a R-245fa R-E170 R-600 R-142b R-245ca R-601a R-601 R-123

426.56 413.35 411.96 402.26 398.19 397.40 382.16 381.64 379.46 379.24 364.82 335.23 327.17 321.79 318.38 313.66 307.57 307.22 306.39 306.04 303.15 302.20 296.68 294.61

85.40% 84.40% 84.27% 83.77% 85.12% 83.58% 85.54% 85.23% 85.36% 85.44% 85.47% 85.76% 85.70% 85.74% 85.74% 85.72% 85.95% 85.71% 85.83% 85.77% 85.77% 85.83% 85.93% 85.91%

84.65% 86.99% 86.63% 86.35% 83.59% 87.28% 80.24% 83.27% 81.94% 82.22% 82.36% 75.17% 75.06% 74.88% 74.12% 74.01% 73.22% 75.78% 73.28% 74.08% 71.95% 72.23% 72.14% 71.77%

86.02% 86.32% 86.53% 86.74% 86.20% 86.82% 86.46% 85.81% 85.81% 85.81% 86.05% 86.20% 85.98% 85.81% 86.18% 85.81% 86.38% 86.96% 86.32% 86.10% 86.34% 86.41% 86.61% 86.45%

26.66% 26.54% 25.93% 24.91% 26.63% 24.09% 26.11% 26.13% 26.10% 26.11% 26.64% 26.55% 26.55% 26.12% 26.51% 26.10% 26.28% 24.88% 26.35% 26.55% 26.29% 26.26% 25.74% 26.20%

57.67% 60.95% 60.18% 58.91% 57.91% 58.83% 54.06% 57.44% 55.96% 56.15% 57.37% 50.11% 50.14% 49.87% 49.45% 48.96% 48.36% 50.23% 48.60% 49.60% 46.99% 47.30% 46.99% 46.97%

5,979 8,207 7,342 8,183 6,303 8,911 5,092 6,826 6,410 5,825 5,903 4,898 5,078 5,531 4,964 5,296 4,628 5,001 4,767 4,907 4,547 4,531 4,499 4,548

Zeotropic Blend R-218/227ea/C318/245fa R-218/C318/245fa R-218/227ea/245fa R-218/227ea R-227ea/C318 R-218/C318 R-218/245fa R-227ea/245fa

461.61 463.01 461.31 397.03 426.25 446.41 426.80 419.99

85.15% 85.13% 85.17% 85.55% 85.47% 85.24% 85.11% 85.50%

80.87% 81.26% 81.01% 79.76% 82.87% 82.02% 79.99% 81.93%

86.09% 86.12% 86.00% 86.07% 86.21% 86.22% 86.13% 86.06%

33.68% 33.32% 33.96% 26.75% 27.39% 29.93% 29.70% 27.91%

58.31% 58.42% 58.72% 53.04% 56.55% 57.11% 54.95% 56.23%

5,493 5,525 5,438 5,647 5,395 5,529 5,794 5,385

Low-GWP Blend R-290/152a/600a/601a R-290/152a/600a R-290/152a/601a R-290/600a/601a R-152a/600a/601a R-290/152a R-152a/600a R-600a/601a R-290/600a R-290/601a R-152a/601a

403.88 404.06 379.17 385.33 378.52 378.72 373.80 357.91 389.67 398.14 361.19

85.23% 85.24% 85.28% 85.30% 85.35% 85.35% 85.41% 85.44% 85.35% 85.29% 85.42%

77.98% 79.26% 74.19% 75.81% 75.30% 76.98% 77.20% 72.17% 79.41% 77.88% 73.37%

85.85% 85.81% 86.01% 85.84% 85.85% 85.81% 85.80% 86.32% 85.81% 85.87% 85.99%

32.80% 31.71% 32.02% 33.39% 32.85% 29.16% 30.91% 32.35% 29.77% 31.97% 32.14%

56.67% 57.20% 52.74% 55.17% 54.35% 53.38% 55.13% 51.15% 56.29% 56.17% 52.15%

5,499 5,533 5,497 5,298 5,114 5,707 5,239 4,627 5,643 5,592 4,852

SC

M AN U

TE D

EP

RI PT

(kW)

AC C

Overall

Estimated SPEC ($/kWe)

39

ACCEPTED MANUSCRIPT

Table A1 Constant parameters used in cost model eq. (A-1). K1

K2

K3

A

Min. Size

Max. Size

4.8306

-0.8509

0.3187

Area (m2)

10

1,000

2.7051

1.4398

-0.1776

Power (kW)

100

4,000

3.8696

0.3161

0.1220

Power (kW)

0.1

200

SC

Table A2

RI PT

Equipment Evaporator and condenser (Floating Head type) Turbine (Axial type) Pump (Reciprocating type)

Constant parameters used in bare module factor eq. (A-3). B1 1.63 1.89

Table A3

B2 1.66 1.35

FM 1 1.5

M AN U

Equipment Floating Head exchanger Reciprocating pump

Overall heat transfer coefficients used in the heat transfer area estimation.

Uliquid-Boiling Uliquid-Condensing Uliquid-Vapor

AC C

Table A4

EP

Uliquid-Supercritical Uliquid-Liquid

TE D

Overall heat transfer coefficient (W/m2.K)

Evaporator

Condenser

850 600

764

600

-

-

764

428

484

Constant parameters used in pressure factor eq. (A-4). Equipment Evaporator - Floating Head (Tube only) Condenser -Floating Head (Both shell & tube) Turbine -Axial gas Pump -Reciprocating

C1

C2

C3

Pressure (barg)

-0.00164

-0.00627

0.0123

5
0

0

0

P<5

0

0

0

-

-0.245382

0.259016

-0.01363

10
40

ACCEPTED MANUSCRIPT

HW-OUT

HW-IN

EVAP-IN

RI PT

TUR-IN

EVAP

TURBINE PUMP

SC

CW-OUT

CW-IN

TUR-OUT

M AN U

PUMP-IN

CONDENSE

AC C

EP

TE D

Figure 1

Figure 2.

41

ACCEPTED MANUSCRIPT 160

160

R-115

R-227ea 140

140

120

120

Hot water

Hot water 100

R-227ea (Evaporating)

80 60

R-227ea (Condensing) 40 Cooling water

20

R-115 (Evaporating)

80 60

R-115 (Condensing) 40 Cooling water

20

0

0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

0

500

1000

1500

Heat duty (kW) 160

R-125

R-143a 140

140

120

Hot water

R-143a (Condensing)

Cooling water

3500

4000

4500

5000

3500

4000

4500

5000

3500

4000

4500

5000

3500

4000

4500

5000

R-125 (Evaporating)

60

R-125 (Condensing)

40

Cooling water

20

0

0

500

1000

1500

2000

2500

3000

3500

4000

Heat duty (kW) 160

4500

0

5000

500

1000

1500

2000

2500

3000

Heat duty (kW)

160

R-32

R-218

140

140

120

120

Hot water

100

Hot water

100

R-32 (Evaporating)

80 60

R-32 (Condensing)

TE D

Temperature (°C)

80

M AN U

40

0

40

Cooling water

20 0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

80

160

R-C318

R-218 (Condensing) 40 Cooling water

20 0

5000

EP

140 120

0

500

1000

1500

R-1270 140

Temperature (°C)

AC C

R-C318 (Condensing)

40

Cooling water

Hot water

500

1000

1500

2000

2500

R-1270 (Evaporating) 80 60 R-1270 (Condensing) 40 Cooling water

20 0

0

0

3000

100

60

20

2500

160

120

R-C318 (Evaporating)

80

2000

Heat duty (kW)

Hot water

100

R-218 (Evaporating)

60

Heat duty (kW)

Temperature (°C)

3000

SC

Temperature (°C)

60

Temperature (°C)

Temperature (°C)

R-143a (Evaporating)

20

2500

Hot water

100

100 80

2000

Heat duty (kW)

160

120

RI PT

Temperature (°C)

Temperature (°C)

100

3000

3500

4000

4500

5000

0

500

1000

1500

2000

2500

3000

Heat duty (kW)

Heat duty (kW)

Figure 3

42

ACCEPTED MANUSCRIPT 160

160

R-290

R-134a

140

140 Hot water

120

120 Hot water 100

R-290 (Evaporating) 80 60 R-290 (Condensing) 40 Cooling water

20

R-134a (Evaporating)

80 60

R-134a (Condensing) 40 Cooling water

20

0

0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

0

500

1000

1500

Heat duty (kW) 160

R-236fa

R-22 140

140 Hot water

120

60 R-22 (Condensing)

3500

4000

4500

5000

Cooling water

3500

4000

4500

5000

3500

4000

4500

5000

3500

4000

4500

5000

R-236fa (Evaporating)

80 60

R-236fa (Condensing)

40

Cooling water

M AN U

40

20 0

0

500

1000

1500

2000

2500

3000

3500

4000

Heat duty (kW)

4500

5000

0

500

1000

1500

2000

2500

3000

Heat duty (kW)

160

160

R-12

R-124

140

140 Hot water

120

Hot water

120 100

100 R-124 (Evaporating)

80 60

R-124 (Condensing) 40 Cooling water

20 0 0

500

1000

1500

2000

TE D

Temperature (°C)

3000

SC

Temperature (°C)

R-22 (Evaporating) 80

0

2500

3000

3500

4000

4500

R-12 (Evaporating)

80 60

R-12 (Condensing) 40 Cooling water

20 0 0

5000

500

1000

1500

2000

160

R-600a

EP

140 Hot water

100

AC C

R-600a (Condensing)

40

Cooling water

500

1000

1500

2000

2500

R-152a (Evaporating)

80 60

R-152a (Condensing) 40 Cooling water

20

0

0

Hot water

100

60

20

R-152a 140 120

R-600a (Evaporating)

80

3000

160

Temperature (°C)

120

2500

Heat duty (kW)

Heat duty (kW)

Temperature (°C)

2500

Hot water

100

Temperature (°C)

Temperature (°C)

100

20

2000

Heat duty (kW)

160

120

RI PT

Temperature (°C)

Temperature (°C)

100

0 3000

3500

4000

4500

5000

0

500

Heat duty (kW)

1000

1500

2000

2500

3000

Heat duty (kW)

Figure 3 (cont.)

43

ACCEPTED MANUSCRIPT 160

160

R-245fa

R-E170

140

140

120

120

Hot water

100

R-245fa (Evaporating)

80 60

R-245fa (Condensing) 40 Cooling water

20

R-E170 (Evaporating)

80 60

R-E170 (Condensing) 40 Cooling water

20

0

0 0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

0

500

1000

1500

Heat duty (kW) 160

R-600

R-142b

140

140 Hot water

120

60 R-600 (Condensing)

3000

3500

4000

4500

5000

Cooling water

3500

4000

4500

5000

3500

4000

4500

5000

3500

4000

4500

5000

R-142b (Evaporating)

80 60

R-142b (Condensing)

40

Cooling water

M AN U

40

20 0

0

500

1000

1500

2000

2500

3000

3500

4000

Heat duty (kW) 160

4500

5000

0

500

1000

1500

2000

2500

3000

Heat duty (kW)

160

R-245ca

R-601a

140

140

120

120

Hot water

100

Hot water

100

R-245ca (Evaporating)

80 60

Cooling water

0 0

500

1000

1500

2000

TE D

R-245ca (Condensing) 40 20

2500

3000

3500

4000

4500

R-601a (Evaporating)

80 60

R-601a (Condensing) 40 Cooling water

20 0

5000

0

500

1000

1500

2000

Heat duty (kW) 160

R-601

EP

140 120

R-601 (Condensing)

40

Cooling water

500

1000

1500

2000

2500

R-123 140 Hot water

R-123 (Evaporating)

80 60

R-123 (Condensing) 40 Cooling water

20

0

0

160

Temperature (°C)

AC C

60

20

3000

100

R-601 (Evaporating)

80

2500

Heat duty (kW)

120

Hot water

100

Temperature (°C)

2500

SC

Temperature (°C)

R-600 (Evaporating) 80

0

Temperature (°C)

Hot water

100

Temperature (°C)

Temperature (°C)

100

20

2000

Heat duty (kW)

160

120

RI PT

Temperature (°C)

Temperature (°C)

100

Hot water

0 3000

3500

4000

4500

5000

0

500

Heat duty (kW)

1000

1500

2000

2500

3000

Heat duty (kW)

Figure 3 (cont.)

44

M AN U

SC

RI PT

ACCEPTED MANUSCRIPT

a)

160

TE D

R-290/152a/600a/601a (35.1/38.1/22.4/4.4)

140 120

Hot water

EP

80

R-290/152a/600a/601a (Evaporating)

60

AC C

Temperature (°C)

100

R-290/152a/600a/601a (Condensing)

40

Cooling water

20 0

0

500

1000

1500

2000

2500

3000

Heat duty (kW)

b) Figure 4

3500

4000

4500

5000

5500

ACCEPTED MANUSCRIPT

A study on optimal composition of zeotropic working fluid in an Organic Rankine Cycle (ORC) for low

RI PT

grade heat recovery

•

SC

Research highlights

In low grade heat recovery, the improvement on net work output of an ORC by

M AN U

the application of zeotropic working fluid was mainly due to the consequence of the good gliding match in the condenser and high degree of heat recovery in the evaporator. •

The optimization procedure used in this work can simultaneously select

•

TE D

appropriate constituents and their compositions in a zeotropic working fluid. The optimal number of constituents in the zeotropic fluid was either 3 or 4 under

AC C

EP

the assumptions used in this study.