Thermodynamic analysis of organic Rankine cycle used for flue gases from biogas combustion

Energy Conversion and Management 153 (2017) 627–640

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Thermodynamic analysis of organic Rankine cycle used for flue gases from biogas combustion

MARK

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Roshaan Mudasar, Faraz Aziz, Man-Hoe Kim

School of Mechanical Engineering, Kyungpook National University, Daegu 41566, South Korea

A R T I C L E I N F O

A B S T R A C T

Keywords: Organic Rankine cycle Biogas CHP application Toluene Design constraints

This study elaborates an investigation on organic Rankine cycle (ORC) for integration with a sewage plant for generation of power and district heating purposes. Flue gases from combustion of biogas is considered as heat source. Initially, design constraints of the cycle i.e. pump exit temperature, source exit temperature, cooling fluid exit temperature, and flue gases exit temperature are applied to check the ORC applicability. Furthermore, fifteen sets of working conditions with combinations of evaporation pressure, source temperature, and pinch point condition are passed through a step by step energetic and exergetic analysis. Seven best set of working conditions are shortlisted to examine the thermal efficiency of the cycle for combined heat and power (CHP) generation. In addition, the research also highlights the importance of biogas technology with respect to environment. The results show that the maximum work output of 156.4 kW is produced at source temperature of 345 °C, evaporation pressure of 36 bar, and pinch point of 5 °C at evaporator and 10 °C at condenser. This research demonstrates that biogas-fired ORC systems depicts an efficient solution for domestic scale power generation applications in rural areas.

1. Introduction The renewable energy has immense importance in energy sector as fossil fuels are vanishing from earth due to rapid industrial revolution. The researchers nowadays are focused towards the development of alternative energy resources which can reduce their dependence on nonrenewable resources. Burning primitive fuels produce certain gases in the atmosphere, which are acknowledged as the greatest environmental threat. The waste management is yet another trouble of urban development, which is being mishandled and eventually pollute our environment. Wastewater management, involving sewage waste at large, is considered to be an integral part of ecosystem-based management system that operates across social, economic, and environmental dimensions of sustainable development [1]. Energy outage of urban and remote areas can be resolved if biomass, especially animal and human waste, are employed as a remedy which are available abundantly in nature [2]. Sewage waste and cow manure are potential biomass which can produce biogas (CH4-55%, CO2-35% and small traces of other gases) from anaerobic digestion (AD) to use as a source of heat and power [3]. This biogas can be burned in combustor and the resulting flue gases can provide the heat for the operation of power cycles to generate power. Such domestic renewable energy resources can provide a viable cost-effective solution on one hand and simultaneously is a

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fight against climate change on the other hand. The organic Rankine cycle (ORC) has been regarded as the best option in industry to recover heat from low temperature sources, i.e. biomass, solar, geothermal, and waste heat from power plants [4,5]. The main advantages of ORC are the simplicity of the system and the availability of components [6,7]. The organic fluid is the heat carrier within the cycle, better adjusted than water to low temperatures [8]. Generally, the organic working fluid has higher molecular mass and lower boiling point compared to water. Detailed investigation regarding the selection of organic fluid can be found in the study of Quoilin et al. [9]. Search for the new working fluid is a continuous process, several researchers have explained the selection procedure for different temperature range organic Rankine cycles. Rahbar et al. [10] suggested a thermo-physical methodical to shortlist the working fluids corresponding to the application specific temperature. Further study was conducted by Zhou et al. [11] on mixtures as candidates for working fluids in the organic Rankine cycle due to their temperature glide feature. They also studied the influence of composition shift on system net power output and heat transfer process. However, selection of working fluid takes heed of environmental and safety factors. Various researches have been carried out to investigate the EHS (environment, health and safety) criteria of the organic working fluids for ORC [12–14]. Typically exhaust flue gases from gas turbine have

Corresponding author. E-mail addresses: [email protected] (R. Mudasar), [email protected] (F. Aziz), [email protected] (M.-H. Kim).

http://dx.doi.org/10.1016/j.enconman.2017.10.034 Received 20 June 2017; Received in revised form 3 September 2017; Accepted 14 October 2017 0196-8904/ © 2017 Elsevier Ltd. All rights reserved.

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Nomenclature AD DORC EHS CHP Tcritical Pcritical CH4 CO2 W m Ƞ h Q Cp e s T

I E

anaerobic digestion dual organic Rankine cycle environment, health and safety combined heat and power system critical temperature of working fluid critical pressure of working fluid methane carbon dioxide work [kW] mass flow rate [kg/sec] ORC cycle efficiency enthalpy [kJ/kg] heat [kW] heat capacity [kJ/kg·K] specific exergy [kW/kg] entropy [kJ/kg·K] temperature

exergy destruction [kW] exergy flow [kW]

Subscripts t, tur p, pum d wf in out w s net o cond evap ex

temperature of about 450 °C, therefore, working fluid chemical deteriorating temperature should be considered while selection. Some researchers opted to lower down the flue gas temperature from turbines by recuperator, but noticed a decrease in the exhaust gas temperature [15]. Others selected thermal oil as heat transfer medium between hot flue gases and ORC [16]. Biogas as a renewable energy fuel has plenty of applications as far as power generation industry is concerned [17]. Having major percentage of methane (CH4), it can either be burned directly for heating or directly passed to combustion engine to generate power [18]. The flame temperature of biogas varies in accordance with the percentage of methane present in the gas. The methane production in turn depends upon the biomass material dumped to decompose. Abdel-Hadi [19] reported that 460 °C flame temperature was achieved when methane content was 54.5% in the biogas produced from cattle dung and chicken manure. Mudasar and Kim [20] investigated on biomass collected from a sewage plant to produce biogas and power. When feces to water ratio was kept at 8:2 in 35 kg of waste sample, 0.35 m3/kg of biogas was attained which generated 26.8 kWh power. Small domestic scale power generation systems are developed

turbine pump destruction working fluid inflow outflow cooling water source/thermal oil total dead state condenser evaporator exergy

through ORC technology. Zhang et al. [21] proposed various combined power cycles incorporating organic Rankine cycle as bottoming cycle to utilise the waste heat from steam turbine, gas turbine, and thermoelectric generator. The results suggested that dual organic Rankine cycle (DORC) possessed high energy utilization capacity compared to other proposed cycles. The DORC ensued with 32.63 kW power with 26.55% thermal efficiency when R123 is employed as refrigerant. Camporeale et al. [22] worked on externally fired gas turbine with direct combustion of biomass coupled with ORC. They varied the evaporation pressure and superheating temperature of ORC to examine the influence upon different bottoming ORC cycles. Results revealed that in case of subcritical cycle, superheating caused a decrease in power output. Furthermore, thermal efficiency dropped for supercritical cycle with increment in the turbine inlet temperature. Benato and Macor [23] and Koc et al. [24] utilised exhaust of biogas fuelled engine as heat source for organic Rankine cycle analysis. A recap of the literature shows that several researchers investigated on ORC taking source as solar energy, geothermal energy, biomass direct burning, and waste heat from gas turbines. However, hot flue gases emanating from direct combustion of biogas is rarely acknowledged as

Fig. 1. Schematic diagram of ORC system based on biogas for CHP application.

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2.2. Energy analysis

source. Apart from this, very few work reflects the selection of optimum working conditions with reference to design constraints; source exit temperature and cooling fluid exit temperature. Moreover, little attention has been paid towards the overall environmental effect of system under study. In the present work, organic Rankine cycle has been analysed which uses flue gases from the direct combustion of biogas as a heat source. Source temperature has been varied to account for the methane present in biogas which controls the heating value. Thermodynamic cycle energy and exergy analysis have been carried out based on performance parameters such as work output, mass flow rate, energetic efficiency, exergetic efficiency, and component exergy destruction. Results and discussions include the evaluation of optimum conditions for ORC with toluene as working fluid, which are applied to estimate thermal efficiency for the CHP application. The work also highlights the environmental aspects of biogas based ORC technology and befitting working fluids.

The thermodynamic processes of a simple ORC system can be illustrated on a T-s diagram as shown in Fig. 2. For the thermodynamic analysis, some assumptions are made which are as follows: (1) all the components of the system are considered as steady-state flow devices (2) the pressure drops in the condenser, evaporator, pipes and tubes are neglected (3) the heat losses in the components are also neglected (4) isentropic efficiencies of turbine and pump are given (5) the working fluid enters the pump as a saturated liquid The power generated by the turbine can be given as:

Wt = mwf (h1−h2s)ηt = mwf (h1−h2) The heat rejected to the cooling water by the working fluid in the condenser is given as:

2. Methodology 2.1. Cycle processes

Qout = mwf (h2−h3) = mw (h8−h7) = mw c p w (T8−T7)

An organic Rankine cycle is similar to a conventional steam Rankine cycle except that it uses a low boiling point organic working fluid instead of steam. A simple organic Rankine cycle consists of four main components; a pump, an evaporator, a turbine and a condenser as shown in Fig. 1. The pump raises the pressure of a low boiling point working fluid and passes it to the evaporator. In the evaporator, the heat from the combustion of biogas via thermal oil is used to first preheat the working fluid to reach the saturated liquid state and then vaporize it by further heat addition at constant temperature. The high pressure vapour flows into the turbine from the evaporator where it expands to give the desired work output. Finally, the working fluid is released into the condenser from the turbine where it is condensed by the cooling water to reach the saturated liquid state. The working fluid is then pumped back to the evaporator to start the new cycle.

The power consumed by working fluid pump can be expressed as:

Wp =

mwf (h 4s−h3) = mwf (h 4−h3) ηp

The amount of heat absorbed by the working fluid from the source in the evaporator is given as:

Q in = mwf (h1−h 4) = m s (h5−h6) = m s c ps (T5−T6) The net power output of the ORC system can be given as:

Wnet = Wt−Wp The thermal efficiency of the cycle is defined as the ratio of net power output to the amount of heat added to the system: Fig. 2. T-s diagram of toluene with simple organic Rankine cycle.

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η=

working fluid, but no fluid can satisfy all the required criterion. It mainly depends on the type of application for which working fluid is required. The current study comes under the category of high temperature ORC, therefore to achieve the best thermodynamic performance, high critical temperature working fluids must be selected. Lai et al. [29] suggested to use linear siloxanes, alkanes and aromatics for a high temperature ORC. For the current study, the condensation temperature of ORC is always set at 100 °C to satisfy the CHP network [30,31]. The corresponding condensation pressure must always be higher than the lowest acceptable pressure for the condenser i.e. 0.05 bar [30,32]. The linear siloxanes need a very high vacuum to condense to a saturated liquid state and the auto-ignition temperatures of alkanes are less than 380 °C [33]. This makes linear siloxanes and alkanes improper for the current application. Therefore, high critical temperature organic fluids from the aromatics family are suitable for the considered ORC systems. Hence, toluene was selected as the working fluid for the proposed system.

Wnet Q in

2.3. Exergy analysis The thermal efficiency of an ORC is an indicator based only on the amount of heat input and the work output, however, it does not reflect the ability of the system to convert energy from heat source to useful work. Therefore, system performance cannot be evaluated based on only energy analysis rather an exergy analysis must be considered which takes into account the measure of irreversibility in the system that is the source of performance loss [25]. Exergy is a measure of maximum theoretical work that can be produced as the system proceeds to a specified final state in equilibrium with its surroundings. The assessment of exergy destruction in various components of the system identifies the magnitude, location and source of thermodynamic inefficiencies in the system [26,27]. The specific exergy at any point can be given as:

2.5. Heat transfer fluid

e= h−h o−To (s−so) This research work aims to investigate the potential of human excreta so that a domestic CHP system could be developed. Human waste when heated in a biogas digester produces biogas. The purified biogas when goes through combustion process results in high temperature flue gases which act as a heat source for an ORC. Depending on the methane content of biogas, the flue gases temperatures ranging from 350 to 450 °C could be achieved. This heat from the flue gases can be recovered in two different ways:

where the subscript ‘o’ refers to the dead state. The exergy destruction in all the components referring to state points shown in T-s diagram (Fig. 2) can be given as:

Evaporator: Idevap = m s (e5−e6) + mwf (e4−e1) Turbine: Idtur = mwf (e1−e2)−Wt Condenser: Idcond = mw (e7−e8) + mwf (e2−e3) Pump: Idpum = Wp−mwf (e4−e3)

(1) direct heat exchange between the flue gases and working fluid (2) an intermediate heat transfer fluid loop which is used to transfer heat from flue gases to the working fluid in the evaporator, usually using a thermal oil

The exergy losses taken off by the source exhaust and cooling water exhaust can be given as:

Source exhaust: Ids = ms (e6) Cooling water exhaust: Idw = mw (e8)

Although the direct evaporation method is more efficient and simpler, it involves a number of issues such as deterioration of the working fluid at high temperature when it approaches its maximum chemical stability temperature. Secondly in case of direct evaporation, it becomes very hard to achieve stability and controllability of the system. Compared to this, the use of a heat transfer fluid loop allows a smoother cycle operation by damping the fast variations of the high temperature heat source. For this reason, most commercial ORC systems make use of an intermediate heat transfer loop. An efficient and cost effective heat transfer fluid selection addresses one of the important design parameters which affects the overall performance of the cycle. In general, a heat transfer fluid must have the following properties: low melting temperature, high temperature stability, high thermal conductivity, high thermal capacity, low coefficient of expansion, low viscosity, limited hazard and environmental issues, low corrosion and low cost [34]. Therminol VP-1 has been selected as the heat transfer fluid for the proposed system. Apart from high thermal conductivity and low viscosity, the auto-ignition temperature of Therminol VP-1 is very high as compared to other thermal oils. The properties of Therminol VP-1 can be found at [35]. This feature makes it most suitable for this particular application as the flame temperatures from biogas with 54.5% methane content may reach up to 460 °C [19].

The total exergy input to the system is the sum of exergy input by heat source and by the cooling water:

Ein = m s (e5) + mw (e7) The exergy efficiency of the system is defined as the ratio of net power output to the total exergy input to the system:

ηex =

Wnet Ein

2.4. Choice of working fluid The selection of an efficient and safe working fluid is an important factor to achieve an energy efficient ORC system. Compared to a conventional steam Rankine cycle, the working pressures and temperatures in an ORC are low ensuring better safety of the system. The working fluids are categorized into three types depending on the slope of saturated vapour curve i.e. dT/ds [28]. The wet fluids having a negative slope requires the working fluid to be superheated prior to entering the turbine. If the fluid enters the turbine as a saturated vapour, it falls in the two phase region after expansion which can damage turbine blades due to corrosion. However, in case of isentropic and dry fluids having a non-negative slope, superheating is not required as the fluid is in the superheated vapour state at the turbine outlet. Therefore, dry and isentropic fluids are recommended for an ORC so that superheating equipment need not be installed and also the possibility of turbine blade damage due to corrosion may be avoided. These factors can significantly reduce the overall cost of the system by reducing the evaporator size and turbine maintenance issues. Many studies have been carried out by researchers for an optimal

2.6. Design and simulation The modelling of an ORC system requires that the thermodynamic properties of the working fluid must be known at each state point of the cycle. This is accomplished by use of software packages which are developed specifically to calculate different thermodynamic properties of various organic fluids. The refrigerant properties have been taken from NIST REFPROP 9.1 database [36], which is used in MATLAB R2017a 630

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transfer between different fluids in these heat exchangers. In the second part, all above mentioned performance parameters are examined in the estimated pressure range to find the optimum working conditions for the cycle. Pinch point constraints connected with fluids, transferring heat at evaporator and condenser, restricts the temperature profile of the fluids as shown in Figs. 3–5 which governs the cycle possibility. In Fig. 3, temperature design constraints are illustrated to understand the cycle possibility against evaporation pressure keeping pinch point and source temperature constant at 5 °C and 335 °C, respectively. If the evaporator pinch point temperature is set at 5 °C, then thermal oil exit temperature cannot exceed 175 °C to satisfy the pinch point constraint. This is because the flue gases outlet temperature is already set at 180 °C to avoid the acidic dew point temperature. While considering the heat transfer between thermal oil and working fluid in evaporator, the thermal oil exit temperature cannot go below pump exit temperature. Moreover, to ensure perfect heat transfer, lower limit for thermal oil exit temperature has to be 110 °C as the condenser temperature is fixed at 100 °C to compensate for the CHP requirement. A small temperature rise in the pumping process is neglected here while setting the lower limit for thermal oil exit temperature. This entails that the condenser temperature and pump exit temperature is considered same for the investigation purposes. Similarly, the cooling water exit temperature in the condenser must be between 95 and 100 °C to meet the 5 °C pinch point condition. The pressures at which the exit temperatures do not satisfy the pinch point temperature constraints are discarded from the analysis. Looking at Fig. 3 it can be observed that, at 5 bar, the thermal oil exit temperature is below the prescribed lower limit although the cooling water exit temperature is within the desired range. Therefore, cycle is not possible to operate at 5 bar. Specifically, from 6 to 11 bar, cycle is satisfying both evaporator and condenser pinch point conditions. Although, thermal oil exit temperature fulfils the temperature constraints for evaporator relative to flue gases and pump until 25 bar, however, cycle is still not possible to function due to the cooling water exit temperature exceeding the upper limit. In general, the plot highlights a considerable rise in thermal oil exit temperature and a minute progressive rise in cooling water exit temperature. Furthermore, Fig. 4 presents the noticeable influence of pinch point conditions when set to ΔTpp = 10/10. The positive influence is evident in terms of the availability of wide applicable pressure range. While considering thermal oil exit temperature, cycle is applicable from 5 to 20 bar, beyond which it rises above the exhaust temperature of flue gases i.e. 180 °C. A negligible hike is observed in the cooling water exit temperature throughout the pressure range ensuring cycle functionality relative to condenser pinch point condition. Hence, it is concluded that setting pinch point to 10 °C justifies the heat transfer at condenser but still limits the evaporator to 20 bar. It is nearly impossible to maintain low pinch point during the industrial processes, therefore, considerable temperature difference should be taken for thermodynamic estimations. It is noteworthy from Fig. 5, that relatively high evaporation

(9.2) simulation code [37] to solve the thermodynamic processes of the ORC system. MATLAB provides many built-in mathematical and thermophysical property functions which can efficiently evaluate any thermodynamic property of a fluid from a built-in function call from REFPROP database in terms of any two other state properties. The design conditions for the current work are summarised in Table 1. Specifically, the thermal energy input from the biogas combustion through thermal oil is 1000 kW. The turbine and pump isentropic efficiencies of 85% and 65% respectively have been selected according to literature [38]. The efficiency of the heating process (from biogas to working fluid through thermal oil) is considered as 85% [32]. Finally, the thermal efficiencies of evaporator and condenser are assumed as 96% and 98%, respectively [39]. Mago et al. [28] and Hung et al. [40] have suggested that superheating is not required when using dry fluids in an organic Rankine cycle, as it increases the cycle irreversibility and hence decreases the exergy efficiency. However, Usman et al. [41] emphasized that in real practise super heating value of exactly zero cannot be ensured for stable cycle operation. Therefore, super heating value of 2 °C is set for the current analysis. The flame temperature of biogas has been experimentally investigated by Abdel Hadi [19]. He suggested that the flame temperature is a strong function of methane content of biogas. As the methane content of biogas rises, so does the flame temperature. Flame temperature of 460 °C is attained when the methane concentration in biogas is 54.5%. The flame temperature rises up to 631 °C for 68.1% methane concentration in biogas. In the proposed analysis, as the heat from biogas combustion is delivered to the organic fluid through the thermal oil, therefore thermal oil inlet temperature is considerably affected by the methane concentration of biogas. In order to further investigate this effect, five different thermal oil inlet temperatures are selected for the cycle analysis i.e. 325 °C, 335 °C, 345 °C, 355 °C, and 365 °C. The exit temperature of flue gases is a significant parameter to be considered. In general, the flue gases are required to be cooled down to lowest possible temperature to extract the maximum thermal power. However, as the flue gases contain water and sulphur oxides, cooling it below the acid dew point leads to corrosion damage in heat exchanger due to condensation of sulphuric acid vapour. So a nominal value of 180 °C has been taken to avoid acid dew point of the flue gases [42]. The condenser cooling water inlet temperature is considered as 75 °C. Three different pinch point conditions for the evaporator and condenser have been chosen for the analysis; 5 °C pinch point temperature difference in both evaporator and condenser ΔTpp = 5/5, 10 °C for both evaporator and condenser ΔTpp = 10/10, and 5 °C for evaporator and 10 °C for condenser ΔTpp = 5/10. 3. Results and discussions The current work examines the energetic and exergetic performance of a slightly superheated subcritical organic Rankine cycle with Toluene as a working fluid. The heat source temperature has a significant impact on different parameters which contribute towards overall performance of the cycle. As discussed earlier, the flame temperature of flue gases from biogas combustion mainly depends on methane concentration of biogas, thereby affecting the heat source inlet temperature. Hence, the cycle performance parameters are analysed at different heat source temperatures i.e. 325 °C, 335 °C, 345 °C, 355 °C, and 365 °C. Various performance parameters namely electrical efficiency, net power output, component exergy destruction, and cycle exergetic efficiency are analysed. The results and discussion analysis mainly consists of two parts. In the first part, for a particular heat source temperature, an applicable range of evaporating pressures is estimated at different pinch point temperature conditions. The applicable evaporation pressure range is defined on the basis of exit temperatures of thermal oil and condenser cooling water. The pinch point temperature condition has to be maintained throughout the evaporator and condenser to ensure proper heat

Table 1 Major design considerations for biogas based ORC.

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Parameter

Units

Values

Critical temperature of toluene Critical pressure of toluene Biogas thermal power Flue gases inlet temperature Flue gases exit temperature Biogas combustion to thermal oil efficiency Evaporator efficiency Turbine isentropic efficiency Degree of superheating Condenser efficiency Cooling water inlet temperature Pump isentropic efficiency

°C bar kW °C °C % % % °C % °C %

318.6 41.26 1000 450 180 85 96 85 2 98 75 65

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Fig. 3. ORC design constraints; pinch point at evaporator and condenser is 5 °C while source temperature is 335 °C (ΔTpp = 5/10).

Fig. 4. ORC design constraints; pinch point at evaporator and condenser is 10 °C while source temperature is 335 °C (ΔTpp = 10/10).

point of 5 °C at condenser doesn’t benefit the cycle, as this is difficult to maintain in real processes. Moreover, it provides limited applicable pressure range. This implies that for effective cogeneration with ORC, condenser pinch point condition should be at least 10 °C as it satisfies the complete pressure range as shown in Figs. 4 and 5. Contrary to this, pinch point of 5 °C at evaporator gives larger applicable pressure range compared to 10 °C. In addition, it provides the higher values of applicable evaporation pressure. Table 2 describes the applicable evaporation pressure ranges for individual source temperatures having three different pinch point

pressure range is attained when evaporator is dedicated with 5 °C pinch point and condenser with 10 °C. To satisfy the pinch point at evaporator, thermal oil exit temperature must be between 105 °C and 175 °C whereas the water exit temperature at the condenser should be between 90 °C and 100 °C. These both desired temperature ranges are achieved for a wide range of evaporation pressure i.e. 6–25 bar. Although, cooling water exit temperature is fulfilling the thermal limits beyond 25 bar, but cycle remains non-functional due to the thermal oil exit temperature crossing over the temperature bound of 180 °C. By comparing Figs. 3–5, it can be rationally concluded that pinch 632

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Fig. 5. ORC design constraints; pinch point at evaporator is 5 °C and at condenser is 10 °C while source temperature is 335 °C (ΔTpp = 5/10).

temperature, a narrow range of applicable pressures appeared i.e. 23–28 bar. A thermodynamic cycle analysis is carried out involving multiple simulations focusing on getting maximum work output. The net power is noticed to vary with the ORC evaporation pressure. It may reach a maximum value at some particular evaporation pressure or increase linearly. When highest power is achieved, the corresponding pressure is regarded as the optimum pressure. This also implies that if pressure changes monotonously, then a specific pressure will be selected as optimum pressure considering other parameters. The pinch point condition and source temperature corresponding to the optimum pressure are regarded as optimum working conditions. Table 3 explains the energy analysis of simple ORC with toluene as working fluid. For each source temperature, three different pinch point conditions are chosen to study the variations on the work output and cycle efficiency. It is observed that every pinch point condition with specific source temperature possesses an optimum pressure, which is the pressure within the applicable range given in Table 2 for which the work output is maximum. The optimum pressure value is mentioned next to the maximum work output and corresponding cycle efficiency. A general trend from the results furnish that work output and cycle efficiency increases with increase in evaporation pressure, except when temperature difference between heat source and Tcritical of working fluid is much higher as in case of Ts = 355 °C and 365 °C. It can be seen from Table 3, for ΔTpp = 5/5, Ts = 325 °C, 335 °C, and 345 °C showed same maximum work output and maximum efficiency at 11 bar, i.e. 127.7 kW and 15.6%, respectively. When source temperature is increased further to Ts = 355 °C and 365 °C, the cycle does not satisfy the pinch point condition ΔTpp = 5/5, hence they are discarded. For ΔTpp = 10/10, highest cycle efficiency of 19.1% is obtained at 34 bar

Table 2 Applicable ranges of evaporation pressure corresponding to source temperature and pinch point temperature difference in evaporator and condenser. Source temperature [°C]

Pinch point condition ΔTpp = 5/5

325 335 345 355 365

ΔTpp = 10/10

ΔTpp = 5/10

Applicable range of evaporation pressure [bar] 5–11 5–17 5–21 6–11 5–20 6–25 9–11 7–26 9–36 – 11–34 14–32 – 17–30 23–28

conditions. By comparison, it is worthy to notice that the ORC applicable evaporation pressures did not appear with quite a wide range when pinch point is ΔTpp = 5/5 and source temperature is increasing. Specifically, none of evaporation pressures appeared to be satisfactory to the temperature constraints for source temperatures 355 °C and 365 °C. In addition, comparatively wide applicable pressure ranges are observed when the allocated pinch point is 10 °C with maximum 34 bar for source temperature 355 °C. This is imperative to the fact that the heat source temperature should not be much higher than the Tcritical of the working fluid i.e. 318.6 °C in case of toluene. A twofold behaviour is found while checking the results from individual source temperatures keeping pinch point at ΔTpp = 5/5 and ΔTpp = 10/10, respectively. As the source temperature increases from Ts = 325 °C, 335 °C, 345 °C, and 355 °C particularly, the applicable evaporation pressure also increases giving values in the range; 5–11 bar (low), 17–25 bar (medium), and 30–36 bar (high). Especially, source temperature 345 °C showed highest evaporation pressure of 36 bar. Conversely, in case of 365 °C source

Table 3 Energetic analysis of ORC while changing pinch point conditions corresponding to individual source temperature. Source temperature (°C)

325 335 345 355 365

Pinch point (ΔTpp = 5/5)

Pinch point (ΔTpp = 10/10)

Pinch point (ΔTpp = 5/10)

Maximum work output, Wmax [kW]

Maximum ORC efficiency, ηmax [%]

Maximum work output, Wmax [kW]

Maximum ORC efficiency, ηmax [%]

Maximum work output, Wmax [kW]

Maximum ORC efficiency, ηmax [%]

127.7 @11 bar 127.7 @11 bar 127.7 @11 bar – –

15.6 @11 bar 15.6 @11 bar 15.6 @11 bar – –

140.6 144.8 150.7 155.6 153.5

17.2 17.7 18.5 19.1 18.8

146.0 150.0 156.4 154.5 152.2

17.8 18.3 19.2 18.9 18.6

@17 bar @20 bar @26 bar @34 bar @30 bar

633

@17 bar @20 bar @26 bar @34 bar @30 bar

@21 bar @25 bar @36 bar @32 bar @28 bar

@21 bar @25 bar @36 bar @32 bar @28 bar

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(state 1) and the pump exit (state 4) and the source input power according to the heat and mass transfer correlation Qs = mwf (h4-h1). While enthalpy difference, in this case depends upon the turbine inlet conditions, increases as the temperature rises and consequently decreases the mass flowrate. Pump exit temperature hike is considered as negligible. For Ts = 355 °C, the mass flow rate at ΔTpp = 10/10 is 1.302 kg/sec for optimum pressure of 34 bar, whereas, it is 1.307 kg/ sec at ΔTpp = 5/10 for pressure of 32 bar. However, comparing the source temperatures of 355 °C and 365 °C, the mass flowrate is increasing when considering individual pinch point condition. This upsurge in mass flowrate entails that much higher source temperature is not suitable for ORC operation and result in diverted trends. While dealing with ORC, mass flow rate has significant role in designing the components reflecting the size and safety constraints. Organic fluids have a disadvantage over steam because of the small enthalpy drop i.e. latent heat within the dome, implying to the fact that comparatively higher amount of mass flow rate is required in ORC than steam for the same power output. However, steam cycle usually functions at very high pressures and temperature, requiring large amount of mass flow rate which is still higher than mass flow rate of organic working fluids. As stated earlier, the approach used in this research work aims to find the most optimum working conditions for ORC with toluene as a working fluid. Hence the energy analysis leads to some conclusions by looking at each pinch point condition under study individually. By comparing Figs. 6 and 7, the pinch point condition of ΔTpp = 5/5 gives the lowest work outputs and the highest mass flowrates for source temperatures 325 °C, 335 °C, and 345 °C. Moreover, this pinch point condition operates at considerable low applicable pressures with 11 bar being the maximum turbine inlet pressure. However, it does not perform for the Ts = 355 °C and 365 °C. Taking ΔTpp = 10/10 into account, maximum work output resulted for Ts = 355 °C while corresponding mass flowrate is the lowest of all. It gave diverted trends for Ts = 365 °C. In case of ΔTpp = 5/10, peak value of work output and smallest mass flow rate is noticed for the Ts = 345 °C compared to other source temperatures. For source temperatures of 355 °C and 365 °C, work output starts to decrease and mass flowrate starts to increase. Here it can deduced that parameters; ΔTpp = 5/5 showed reasonable results till Ts = 345 °C but low performances, ΔTpp = 10/10 remained valid till Ts = 355 °C, and ΔTpp = 5/10 demonstrated well till Ts = 345 °C. Beyond this, the pinch point conditions resulted in diverted trends. The energy analysis is based solely on the heat gain and work output

with Ts = 355 °C. It showed a declining trend when source temperature is increased to 365 °C. The same efficiency decreasing pattern is recorded with ΔTpp = 5/10, when source temperature is increased to 355 °C and 365 °C. This slumping effect is a consequence of higher temperature difference between critical temperature of working fluid and the source temperature. Fig. 6 depicts the relation of maximum work output with different source temperatures at optimal evaporation pressures by varying pinch point conditions. The results furnished reveal that work output is lowest for pinch point ΔTpp = 5/5, substantially high for pinch point ΔTpp = 10/10 and highest for pinch point ΔTpp = 5/10. This is valid for Ts = 325 °C, 335 °C, and 345 °C solely. The typical reason behind the phenomena is increasing high pressure which delivers high enthalpy difference ensuing better work performance of the system. Moreover, power appears to give largest value 156.4 kW at Ts = 345 °C when pinch point is set at ΔTpp = 5/10. However, the work output drops for the subsequent source temperatures 355 °C and 365 °C, giving highest value for pinch point ΔTpp = 10/10 and lesser value for pinch point ΔTpp = 5/10. This analysis provides a better pragmatic approach to study the systems while designing the low temperature power generation units compared to steam and gas power cycles. Furthermore, the approach also highlights the optimal pressures which should be considered keeping the safety measures into account. It explains a wide range of pressure ranges, low to medium to high, on which ORC can be operated and guides to further set the design specifications of other components. The characteristics of ORC cycle are mainly determined by the turbine work output, which in turn depends upon the mass flow rate flowing in the turbine. Fig. 7 describes the mass flow rate behaviour inside the cycle. As a general rule of Rankine cycle, turbine inlet temperature increases with evaporation pressure, consequently, mass flow rate also decreases due to high pressure. As observed from the previous analysis, higher source temperature caters larger optimal evaporation pressure. Overall it can be seen, mass flow rate of toluene is showing declining trend with the rise in source temperature and evaporation pressure. Highest mass flow of 1.5 kg/sec is recorded for Ts = 325 °C, 335 °C, and 345 °C among all three pinch point temperature conditions at optimum pressure of 11 bar. As source temperature is augmented from Ts = 325 °C to 345 °C, mass flowrate is decreased, with comparatively lesser value for ΔTpp = 10/10 than ΔTpp = 5/5, and least for ΔTpp = 5/10 compared to other two pinch point conditions. This is mainly due to the rise in enthalpy difference between the turbine inlet

Fig. 6. Maximum work output with respect to source temperature and pinch point conditions at optimum pressures.

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Fig. 7. Mass flow rate of toluene relationship at corresponding to source temperature and pinch point bounds at optimal pressures.

taken away by source exhaust from the ORC system is dependent upon the mass flowrate and the temperature of thermal oil flowing out of the evaporator. From Fig. 9, it can be understood regarding the exergy destruction in evaporator. It gave the same value of 29.4 kW for all the source temperatures with ΔTpp = 5/5, which is regarded as the minimum of all. Furthermore, a peak value of evaporator exergy destruction is reached in case of Ts = 355 °C with ΔTpp = 10/10. By comparing Figs. 8 and 9, it can be concluded that exergy destruction rates of source exhaust and evaporator are taking place in single component i.e. evaporator. Hence, combined value of exergy destructions of exhaust and evaporator should be kept into account to assess the system efficiency. The pinch point ΔTpp = 5/5 are attributing the least values but as their work output values are not much higher so we can neglect. Whereas, ΔTpp = 10/10 showed relatively less source exergy destruction in comparison to ΔTpp = 5/10. This is not enough regarding exergy analysis, Fig. 10 examines the exergy destruction in condenser with all working conditions. As the source temperature increases, the exergy destruction in condenser shows a decreasing trend. Among the pinch point conditions,

of the system and does not account for the irreversibilities associated with the components. These irreversibilities are the major source of performance loss of the overall system. Thus, an exergy analysis which identifies the magnitude and location of these irreversibilities must be carried out. As pinch point investigation necessitates the exergy study of heat exchangers, Figs. 8–10 demonstrates the exergy destruction rates of heat source exhaust, evaporator and condenser respectively. The heat source exergy destruction rate in Fig. 8 tends to be decreasing monotonically with increasing heat source temperature. Minimum values are presented by ΔTpp = 5/5, intermediate by ΔTpp = 10/10, and maximum by ΔTpp = 5/10 from Ts = 325–345 °C. When taking Ts = 355 °C and 365 °C into consideration, the heat source exergy destruction is greater with ΔTpp = 10/10 compared to ΔTpp = 5/10. It indicates that higher level of thermodynamic suitability is achieved for Ts = 365 °C with both ΔTpp = 5/10 and ΔTpp = 10/10, and for Ts = 355 °C with ΔTpp = 5/10, but as their work outputs are deviated from normal trend so they cannot be considered. Although, ΔTpp = 5/5 also appears to have lesser values, however, their low work outputs do not qualify them as potential working conditions. In actual, the exergy

Fig. 8. Exergy destruction in heat source exhaust, comparing different working conditions.

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Fig. 9. Exergy destruction in evaporator, comparing different working conditions.

Fig. 10. Exergy destruction in condenser, comparing different working conditions.

ΔTpp = 10/10 gives the most exergy destruction compared to other conditions ΔTpp = 5/5 and ΔTpp = 5/10. The mass flow rate and temperature of water flowing out are the dependent parameters of exergy destruction in condenser. As the temperature of condenser is fixed at 100 °C and the water exit temperature rise is also very small, so exergy destruction entirely depends upon the mass flow rate of water. To understand the complete cycle exergy phenomenon, total exergy inflow and destructing inside the cycle are the essential parameters of evaluation. Exergy is defined as the maximum theoretical work obtainable in disequilibrium with respect to the dead state. The more exergy is flowing into the system, the highest work output would be possible. Fig. 11 explains the total exergy inflow into the cycle. Among all conditions, pinch point ΔTpp = 5/10 showed the best results with Ts = 325 °C, 335 °C, and 345 °C with exergy input values of 647 kW, 631 kW, and 627 kW, respectively. The same pinch point presented low performance with Ts = 355 °C and 365 °C with large percentage decrease. Total exergy input to the cycle have the lowest values for ΔTpp = 5/5. For a specified input condition of the source, exergy input rate varies due to the mass flow rate of source fluid i.e. thermal oil which is varying in every case.

Fig. 12 describes the total exergy destruction of the system. It accounts for the overall destruction of the cycle comprising of the destruction in each component. By comparing both Figs. 11 and 12, it can be seen that the usual trend is same for all the working conditions. In the cases where the exergy input is recorded highest, total exergy destruction happens to be large too. This is logical to understand that when exergy inflow is high, components will be more vulnerable to destroy the exergy. If temperatures are high in each of the component, temperature drop will also be high, hence resulting in high exergy destruction. On the other hand, high exergy also means that system can produce more work output. Least exergy is destructed in case of ΔTpp = 5/10 with Ts = 365 °C and maximum is destructed in case of ΔTpp = 5/10 with Ts = 325 °C. Fig. 13 explains the exergetic efficiency of ORC cycle with different working conditions. In general, the exergetic efficiency monotonically increase with the rise in source temperature. This is due to the more utilization of the high quality heat into useful work. Exergetic efficiency is directly proportional to the work output attained plus the heat taken out from the condenser for CHP and inversely proportional to the heat available with respect to dead state (standard ambient temperature, 636

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Fig. 11. Total exergy input in the cycle, comparing different working conditions.

Fig. 12. Total exergy destruction of the cycle, comparing different working conditions.

trends of work outputs, high and diverted trends of mass flow rates, high exergy destruction, and low exergetic efficiency. This qualification technique of shortlisting best working conditions for the operation of ORC accounts to best fit with the thermodynamic suitability. With the help of such activity, the working fluid can be qualified under different conditions and applicability can be evaluated. For the combined heat and power application (CHP) application, thermal efficiency accounts for the heat available for the cogeneration purposes. Thermal efficiency is described as the thermal power used for cogeneration from the condensation process divided by the thermal input by the evaporator. In the current work, cogeneration is studied for the best seven working conditions to investigate how much thermal power can be extracted out of this to be used for district heating as a CHP application. Fig. 14 well illustrates the thermal efficiency for seven shortlisted set of working conditions. Water has been used for as heat carrier from the condenser, assumed to be entering the condenser at 75 °C. This research highlighted the exit temperature of water as the design constraint of the ORC cycle for cogeneration study, specifically, which has been discussed in Figs. 3–5. Temperature of condenser is set

To = 25 °C and standard atmospheric pressure, Po = 1 bar). Therefore, when considering Ts = 325–345 °C, lowest exergetic efficiency is produced by the ΔTpp = 10/10, subsequently second lowest by ΔTpp = 5/ 10 and the highest exergetic efficiency is attributed by ΔTpp = 5/5. But as ΔTpp = 5/5 operates at low pressure producing less work output so it can be neglected. Furthermore, it can be concluded that Ts = 355 °C and 365 °C with ΔTpp = 5/10 and Ts = 365 °C with ΔTpp = 10/10, although, showed high exergetic performance, but they will be discarded too due to low work output comparatively to other source temperatures. This ambiguous behaviour is opposing the usual trend as described in Table 3 and Fig. 6. Based on the energetic and exergetic analysis to find out best working conditions for toluene, keeping in view the source temperature, evaporation pressure, and pinch point conditions both in evaporator and condenser, Table 4 describes the working conditions and the reason for their qualification and disqualification. Seven different conditions are qualified to be operated to get maximum work output and resulting in less exergy destruction. Other conditions which did not qualify are discarded due to impossibility of cycle, low and diverted

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Fig. 13. Exergetic efficiency of the cycle, comparing different working conditions.

as 100 °C and the ΔTpp = 10 °C is maintained in all the seven sets of working conditions. As evident from Fig. 14, the highest thermal efficiency is available for the set 1 of working conditions i.e. ΔTpp = 10/ 10, Ts = 325 °C, and evaporation pressure as 17 bar. The minimum thermal efficiency is recorded for the set 7 of working conditions i.e. ΔTpp = 5/10, Ts = 345 °C, and evaporation pressure as 36 bar. It is useful to notice while looking at the trend of set of conditions, when source temperature and optimal evaporation pressure is increasing, the thermal efficiency is decreasing. This is due to the reason that work output increases with evaporation pressure and source temperature on one hand. However, on the other hand, difference between evaporator and condenser thermal powers also increases, which consequently

decreases the thermal efficiency of the cycle. Furthermore it can be understood from the thermal efficiency values that the variation is not much among all the seven set of working conditions. Each set of working conditions is suitable to be used for CHP application, because each set of condition is capable of giving thermal efficiency of at least 80%. This indicates that the combined system can be helpful to extract 80 percent of the thermal power to be utilised for power generation and thermal energy available for CHP application. 4. Environmental impact and benefits ORC systems along with CHP can be installed in the rural or urban

Table 4 Qualification and disqualification chart for the working conditions for toluene. Pinch point conditions

Source temperature (°C)

Optimal evaporation pressure (bar)

Not qualified/reason

Qualified/reason

Evaporator and condenser ΔTpp = 5/5

325 335 345 355 365

11 11 11 Cycle not possible Cycle not possible

Low Wnet, High mwf, Low Ein Low Wnet, High mwf, Low Ein Low Wnet, High mwf, Low Ein – –

– – – – –

Evaporator and condenser ΔTpp = 10/10

325

17

–

335

20

–

345

26

–

355

34

–

365

30

Comparatively low Wnet – diverted trend due to high Ts

Comparable Wnet, Low mwf Comparable ηex, High Ein Comparable Wnet, Low mwf Comparable ηex, High Ein Comparable Wnet, Low mwf Comparable ηex, High Ein Highest Wnet, Lowest mwf Highest ηex, Lowest Ein –

325

21

—

335

25

—

345

36

—

355

32

365

28

Comparatively low Wnet – diverted trend due to high Ts Comparatively low Wnet – diverted trend due to high Ts

Evaporator and condenser ΔTpp = 5/10

638

Comparable Wnet, Comparable mwf Comparable ηex, Highest Ein Comparable Wnet, Comparable mwf Comparable ηex, Comparable Ein Highest Wnet, Lowest mwf Highest ηex, Comparable Ein – –

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Fig. 14. Thermal efficiency comparison of seven best working conditions.

83.1

Tpp= 10/10, Ts= 325C, Pevap= 17bar

Thermal eĸĐŝeŶĐLJ [%]

82.8

Tpp= 10/10. Ts= 335C, Pevap= 20bar

82.5

Tpp= 10/10, Ts= 345C, Pevap= 26bar 82.2

Tpp= 10/10, Ts= 355C, Pevap= 34bar

81.9

Tpp= 5/10, Ts= 325C, Pevap= 21bar

81.6

Tpp= 5/10, Ts= 335C, Pevap= 25bar Tpp= 5/10, Ts= 345C, Pevap= 36bar

81.3 81 80.7

0

1

2

3

4

5

6

7

8

^eǀeŶ sets of ǁorŬŝŶŐ ĐoŶĚŝƟoŶs

biogas technology in combination with ORC. Following conclusions are drawn from the study.

sewage plants where sewage and cow dungs are abundantly available. The combined system can become useful to generate power coupled with district heating as an additive advantage to welfare remote areas. If such systems are designed for commercial purposes, cheap power units can be made available on one hand. On the other hand, environmental impacts can also be mitigated through the utilization of biogas from the sewage and cow dung. Otherwise, it is detrimental for the environment in terms of global warming by releasing free methane and carbon dioxide. The usage of biomass, from animals and sewage, is gaining a considerable attention among the energy market worldwide. These raw products which are freely available in environment have huge importance due to its renewable nature. Some reflections regarding the impacts and benefits of biomass and biogas technology are highlighted in this research. Biomass formed by the waste of animals and human is a pragmatic source of energy created by the dead microorganisms, whose potential is lost in surroundings if not taken care of, generating greenhouse gases. Explicit burning of biomass generates hefty clouds of smoke comprising of carbon dioxide and other toxic gases. Generation of biogas from biomass reduces the threat of free methane released to environment from the waste spread in surroundings. Moreover, purification of biogas can remove the carbon dioxide from biogas substantially. The positive influence of biogas technology refers to the essentiality of captivating the uncontrolled methane emanating from the landfills or slum sanitary conditions. Carbon dioxide adopts the higher concentration in the atmosphere and scientific research is prominently working on its control, however, rare attention is being focused towards methane proportion. Methane attributes 20% to the global warming in comparison to the carbon dioxide 62%, but it possess 25 times higher global warming potential than carbon dioxide in a time horizon of 100 years [43]. The reduction of 1 kg methane is equivalent to the reduction of 25 kg carbon dioxide. If this high global warming potential gas is taken heedless, it can pose huge threat to our environment. ORC based on biogas through utilization of biomass can provide an effective solution to the power generation along with environmental benefits.

• Exit temperature of source in evaporator and exit temperature of • • • • • • • •

cooling fluid in condenser has a significant role in devising the organic Rankine cycle. Source temperature should be close to the critical temperature of working fluid. Temperature difference of above 37 °C is not recommended, subsequently to which cycle performance starts declining and gives diverted trends. Pinch point temperature difference played a vital role in improving the performance of the cycle. ΔTpp = 5/5 gave the lowest work outputs and highest mass flow rate with all the source temperatures i.e. 325 °C, 335 °C, 345 °C, 355 °C, and 365 °C, although, their exergetic efficiency was comparable to others set of conditions. Maximum work output of 155.6 kW was achieved for Ts = 355 °C among other heat source temperatures with ΔTpp = 10/10. Its exergetic efficiency was identified as maximum but exergy input was lowest of all. Ts = 345 °C with ΔTpp = 5/10 resulted with highest work output and highest exergetic efficiency. Lowest mass flow rate and comparable value of exergy input was recorded among all other source temperatures with given pinch point condition. Thermal efficiency of the cycle for the cogeneration purposes showed high values for all the seven set of best working conditions. For CHP application, pinch point temperature difference should not be less than 10 °C in condenser. ORC system based on the combustion of biogas coupled with a combined heat and power application can be a recommended remedy to the rural and urban sewage plants for power generation and district heating.

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5. Conclusions The work focused on the biogas based Organic Rankine Cycle (ORC) involving a new approach to study the design constraints and working conditions of the cycle. Apart from energetic analysis, it also involves exergetic analysis based on toluene as working fluid. Furthermore, step by step analysis leads to a shortlisting of seven set of best working conditions out of 15 possible cycle considerations. Specifically, the selected conditions are applied via thermal efficiency analysis for a Combined Heat and Power (CHP) application purposes to integrate the system with sewage plant. In addition, environmental aspects of the biogas technology is also discussed to highlight the importance of 639

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