Section 2 Applications

**Chapter 3**

*Qiang Liu*

**Abstract**

2–3% more power for a power unit.

house gases from the burning of fossil fuels.

**1. Introduction**

**47**

Waste Heat Recovery from

Organic Rankine Cycles

Fossil-Fired Power Plants by

More than 60% of the world's electricity is still produced from fossil-fired power plants. Recovering heat from flue gas, drained water, and exhaust steam which are discharged in power plants by organic Rankine cycles (ORCs) to generate power is an efficient approach to reduce fossil fuel consumption and greenhouse gas emissions. This chapter proposes conceptual ORC systems for heat recovery of drain from continuous blowdown systems, exhaust flue gas from boilers, and exhaust steam from turbines. The waste heat source temperatures range from 30 to 200°C. Environmentally friendly and nonflammable working fluids including R134a, R1234ze, R236ea, R245fa, R1233zd, and R123 were selected as the working fluids. The parameters of ORC systems were optimized, and the thermodynamic performance was analyzed. The suitable ORC layouts for various kinds of heat sources including drained water, flue gas, and steam were discussed with selecting the proper working fluids. The gross power output of a coal-fired power plant can be increased up to 0.4% by an ORC using the waste heat from the boiler flue gas. The ORCs using turbine exhaust steam with the cooling water as low as 5°C can generate

**Keywords:** coal-fired power plant, waste heat recovery, organic Rankine cycle,

In 2018, the total world electricity generation reached 26,672 TWh [1]. About 64.15% of the total world electricity generation is still from fossil fuels (coal, natural gas, and oil) [1], especially in China, the USA, Japan, Russia, and India. China accounted for 26.2% of the world electricity generation [2]. The installed fossil-fired power capacity is now increased to 1143.67 GW in China which accounts for about 60% of the total installed power capacity; however, 70.4% of electricity was generated from fossil fuels [2]. Now, China is also the world's greatest emitter of green-

The development of large-capacity and high-temperature ultra-supercritical fossil-fired power unit is the trend in the world. However, a large number of waste heat are discharge in a power plant. **Figure 1** shows waste heat sources in a typical coal-fired power unit. Therefore, recovering the waste heat is a main approach to

parametric optimization, working fluid, thermodynamic analysis

#### **Chapter 3**

## Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles

*Qiang Liu*

#### **Abstract**

More than 60% of the world's electricity is still produced from fossil-fired power plants. Recovering heat from flue gas, drained water, and exhaust steam which are discharged in power plants by organic Rankine cycles (ORCs) to generate power is an efficient approach to reduce fossil fuel consumption and greenhouse gas emissions. This chapter proposes conceptual ORC systems for heat recovery of drain from continuous blowdown systems, exhaust flue gas from boilers, and exhaust steam from turbines. The waste heat source temperatures range from 30 to 200°C. Environmentally friendly and nonflammable working fluids including R134a, R1234ze, R236ea, R245fa, R1233zd, and R123 were selected as the working fluids. The parameters of ORC systems were optimized, and the thermodynamic performance was analyzed. The suitable ORC layouts for various kinds of heat sources including drained water, flue gas, and steam were discussed with selecting the proper working fluids. The gross power output of a coal-fired power plant can be increased up to 0.4% by an ORC using the waste heat from the boiler flue gas. The ORCs using turbine exhaust steam with the cooling water as low as 5°C can generate 2–3% more power for a power unit.

**Keywords:** coal-fired power plant, waste heat recovery, organic Rankine cycle, parametric optimization, working fluid, thermodynamic analysis

#### **1. Introduction**

In 2018, the total world electricity generation reached 26,672 TWh [1]. About 64.15% of the total world electricity generation is still from fossil fuels (coal, natural gas, and oil) [1], especially in China, the USA, Japan, Russia, and India. China accounted for 26.2% of the world electricity generation [2]. The installed fossil-fired power capacity is now increased to 1143.67 GW in China which accounts for about 60% of the total installed power capacity; however, 70.4% of electricity was generated from fossil fuels [2]. Now, China is also the world's greatest emitter of greenhouse gases from the burning of fossil fuels.

The development of large-capacity and high-temperature ultra-supercritical fossil-fired power unit is the trend in the world. However, a large number of waste heat are discharge in a power plant. **Figure 1** shows waste heat sources in a typical coal-fired power unit. Therefore, recovering the waste heat is a main approach to

10–30 MW heat can be recovered. The turbine exhaust steam at 30°C contains more than 300 MW heat for a large-capacity power unit. Therefore, this chapter focuses on the thermodynamic performance of ORCs using the waste heat from coal-fired power plants. The suitable cycle layouts are discussed with considering the waste heat source characteristics. The evaporation temperature and pressure of the working fluid will be optimized to maximize the cycle net power output. The proper working fluid is selected according to the thermodynamic performance. The power generation potentials are also evaluated for various kinds of waste heat sources.

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

The ORCs using low-boiling organic working fluids are the unequaled technology for producing electricity from medium-low-temperature heat sources [16]. The ORCs have distinctive features of being simple layouts, being efficient for offdesign conditions, and being reasonable cost-effectiveness. Therefore, ORC systems have been widely used to convert geothermal, solar, biomass, waste heat, and ocean thermal energy to electricity. The theoretical basis for ORCs has been described elsewhere [16–22]. This section briefly describes the cycle layouts and working

The ORCs can be classified into subcritical and supercritical cycles according to the working fluid evaporation pressure. Considering the saturated vapor line shapes of working fluids, the subcritical ORC with saturated vapor at the turbine inlet is called o2 cycle, and that with superheated vapor at the turbine inlet is called o3 cycle which uses a dry working fluid or b3 cycle which uses a wet working fluid, and the supercritical ORC is called s2 cycle [17]. The cycle layouts are selected according to the heat source characteristics and the working fluid critical temperature. **Figure 2(a)** shows a typical ORC system with a wet cooling system with the *T-s* diagrams for o2, o3, b3, and s2 cycles shown in **Figure 3**. The working fluid is heated to a saturated or a

superheated vapor from a subcooled liquid by the heat source fluid in heat

the subscript O represents the organic fluid.

*Schematics of (a) a typical ORC and (b) an ORC with an IHE.*

exchangers. Thus, the energy balance for working fluid evaporation can be uniformly

where *h* is the specific enthalpy, the subscript H represents the heat source, and

*m*\_ <sup>H</sup>ð Þ¼ *h*Hin � *h*Hout *m*\_ <sup>O</sup>ð Þ *h*<sup>0</sup> � *h*<sup>3</sup> (1)

**2. ORC model**

fluids selection.

**2.1 Cycle layouts**

expressed as

**Figure 2.**

**49**

**Figure 1.** *Schematic of the waste heat sources can be recovered by ORCs in a typical coal-fired power unit.*

further improve thermal efficiencies and reduce greenhouse gas emissions for fossil-fired power plants [3].

Continuous blowdown systems are used to purify the steam by removing suspended solids in drum boilers for subcritical power plants as shown in **Figure 1**. A part of the boiler water with temperatures up to 360°C from the drum is discharge to the flash tank. In the flash tank, some of the water flashes to the steam, and then the steam enters into a deaerator or a feedwater heater for feedwater preheating, while the drain is generally discharge to the sewer. The drain temperature is higher than 150°C; however, the heat in the drained water is lost through direct discharge. The waste heat in the drained water can be recovered by an organic Rankine cycle to generate electric power [4].

A modern power station boiler produces large quantities of flue gas. The exhaust flue gas temperatures generally range from 120–140°C for power station boilers. Heat loss due to the exhaust flue gas is the largest heat loss in a power station boiler which significantly affects the boiler thermal efficiency. For large- and mediumcapacity boilers, the exhaust flue gas heat loss accounts for 4–8% of the total boiler heat input from fossil burning [3]. In actual operation, the flue gas temperature may be higher than the design value which results in a lower boiler thermal efficiency and a higher fuel consumption. The heat in the exhaust flue gas can be used to preheat feedwater by a low-pressure economizer [5–7] and pre-dry brown coal [8–11]. The waste heat of exhaust flue gas can also be used to drive an organic Rankine cycle to produce electric power [5, 12–14].

The latent heat of the exhaust steam from a condensing turbine is released to the cooling water at the condenser as shown in **Figure 1**. Even for a modern steam turbine, 50–60% of the heat input is discharge to the heat sink. Seawater is generally used as the cooling water for the offshore power plants. The seawater at highlatitude regions or at deep sea (depth down to 1000 m) has a temperature as low as 5°C. However, the condensing turbine exhaust pressures are generally higher than 3.5kPa with condensation temperatures not less than 26.7°C due to the limits from the wetness loss, last-stage blade length, and droplet erosion. Thus, the temperature difference between the exhaust steam and seawater can be used to drive an ORC to produce electricity. An ORC can also use a part of the extracted steam from lowpressure turbine and boiler flue gas [14] or only the coldest extraction steam [15] to generate additional electricity.

In coal-fired power plants, there are different kinds of heat sources which can be utilized by ORCs. The waste heat sources involve different heat transfer fluids (gas, water, and steam) and large scales of capacity and temperature. The exhaust flue gas with temperatures up to 140°C from a supercritical boiler which contains

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

10–30 MW heat can be recovered. The turbine exhaust steam at 30°C contains more than 300 MW heat for a large-capacity power unit. Therefore, this chapter focuses on the thermodynamic performance of ORCs using the waste heat from coal-fired power plants. The suitable cycle layouts are discussed with considering the waste heat source characteristics. The evaporation temperature and pressure of the working fluid will be optimized to maximize the cycle net power output. The proper working fluid is selected according to the thermodynamic performance. The power generation potentials are also evaluated for various kinds of waste heat sources.

#### **2. ORC model**

further improve thermal efficiencies and reduce greenhouse gas emissions for

*Schematic of the waste heat sources can be recovered by ORCs in a typical coal-fired power unit.*

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

Continuous blowdown systems are used to purify the steam by removing suspended solids in drum boilers for subcritical power plants as shown in **Figure 1**. A part of the boiler water with temperatures up to 360°C from the drum is discharge to the flash tank. In the flash tank, some of the water flashes to the steam, and then the steam enters into a deaerator or a feedwater heater for feedwater preheating, while the drain is generally discharge to the sewer. The drain temperature is higher than 150°C; however, the heat in the drained water is lost through direct discharge. The waste heat in the drained water can be recovered by an

A modern power station boiler produces large quantities of flue gas. The exhaust flue gas temperatures generally range from 120–140°C for power station boilers. Heat loss due to the exhaust flue gas is the largest heat loss in a power station boiler which significantly affects the boiler thermal efficiency. For large- and mediumcapacity boilers, the exhaust flue gas heat loss accounts for 4–8% of the total boiler heat input from fossil burning [3]. In actual operation, the flue gas temperature may be higher than the design value which results in a lower boiler thermal efficiency and a higher fuel consumption. The heat in the exhaust flue gas can be used to preheat feedwater by a low-pressure economizer [5–7] and pre-dry brown coal [8–11]. The waste heat of exhaust flue gas can also be used to drive an organic

The latent heat of the exhaust steam from a condensing turbine is released to the

In coal-fired power plants, there are different kinds of heat sources which can be utilized by ORCs. The waste heat sources involve different heat transfer fluids (gas, water, and steam) and large scales of capacity and temperature. The exhaust flue gas with temperatures up to 140°C from a supercritical boiler which contains

cooling water at the condenser as shown in **Figure 1**. Even for a modern steam turbine, 50–60% of the heat input is discharge to the heat sink. Seawater is generally used as the cooling water for the offshore power plants. The seawater at highlatitude regions or at deep sea (depth down to 1000 m) has a temperature as low as 5°C. However, the condensing turbine exhaust pressures are generally higher than 3.5kPa with condensation temperatures not less than 26.7°C due to the limits from the wetness loss, last-stage blade length, and droplet erosion. Thus, the temperature difference between the exhaust steam and seawater can be used to drive an ORC to produce electricity. An ORC can also use a part of the extracted steam from lowpressure turbine and boiler flue gas [14] or only the coldest extraction steam [15] to

fossil-fired power plants [3].

**Figure 1.**

organic Rankine cycle to generate electric power [4].

Rankine cycle to produce electric power [5, 12–14].

generate additional electricity.

**48**

The ORCs using low-boiling organic working fluids are the unequaled technology for producing electricity from medium-low-temperature heat sources [16]. The ORCs have distinctive features of being simple layouts, being efficient for offdesign conditions, and being reasonable cost-effectiveness. Therefore, ORC systems have been widely used to convert geothermal, solar, biomass, waste heat, and ocean thermal energy to electricity. The theoretical basis for ORCs has been described elsewhere [16–22]. This section briefly describes the cycle layouts and working fluids selection.

#### **2.1 Cycle layouts**

The ORCs can be classified into subcritical and supercritical cycles according to the working fluid evaporation pressure. Considering the saturated vapor line shapes of working fluids, the subcritical ORC with saturated vapor at the turbine inlet is called o2 cycle, and that with superheated vapor at the turbine inlet is called o3 cycle which uses a dry working fluid or b3 cycle which uses a wet working fluid, and the supercritical ORC is called s2 cycle [17]. The cycle layouts are selected according to the heat source characteristics and the working fluid critical temperature. **Figure 2(a)** shows a typical ORC system with a wet cooling system with the *T-s* diagrams for o2, o3, b3, and s2 cycles shown in **Figure 3**. The working fluid is heated to a saturated or a superheated vapor from a subcooled liquid by the heat source fluid in heat exchangers. Thus, the energy balance for working fluid evaporation can be uniformly expressed as

$$
\dot{m}\_{\rm H}(h\_{\rm film} - h\_{\rm Hout}) = \dot{m}\_{\rm O}(h\_0 - h\_3) \tag{1}
$$

where *h* is the specific enthalpy, the subscript H represents the heat source, and the subscript O represents the organic fluid.

**Figure 2.** *Schematics of (a) a typical ORC and (b) an ORC with an IHE.*

**Figure 3.** *Temperature-entropy diagrams for basic ORCs: (a) o2 cycle, (b) o3 cycle, (c) b3 cycle, and (d) s2 cycle.*

The power generated by the turbine was

$$
\dot{W}\_{\rm T} = \dot{m}\_{\rm O} (h\_0 - h\_1) \tag{2}
$$

The use of an IHE can improve the preheater inlet temperature by the turbine exhaust vapor. For the heat source without the limitation of the outlet temperature [18, 23], the heat source fluid outlet temperature increases because the working fluid is preheated as shown in **Figure 4(a)** and the cycle net power output will maintain constant or increase slightly [24]. For the heat source with a limited outlet temperature as shown in **Figure 4(b)**, the increase in the preheater inlet temperature results in an increase in the working fluid mass flow rate or the optimal

The matching characteristics between the heat source fluid and the working fluid affect the ORC net power output. In subsequent sections, the working fluid evaporation temperature or pressure will be optimized to maximize the cycle net power output. The operating parameters and boundary conditions for the ORCs are listed in **Table 1**. The parametric optimization and thermodynamic analyses are based on the parameters listed in **Table 1** except the ORC in Section 5.2. The heat

The pinch point temperature difference is an important parameter for system design and optimization [25, 26]. During the working fluid evaporation, the pinch point may locate at the bubble point, a subcooled liquid state near the critical point and a superheated state [26] which depends on the heat source temperature drop

*Temperature-entropy diagrams of ORCs with an IHE for heat source fluids: (a) without outlet temperature*

**Parameters Values Unit** ORC turbine isentropic efficiency 85 % Working fluid pump efficiency 80 % Generator efficiency 98 % Turbine mechanical efficiency 98 % Evaporator pinch point temperature difference 10 °C IHE pinch point temperature difference 10 °C Condenser pinch point temperature difference 5 °C Condensation temperature 30 °C

evaporation temperature; thus, the net power output increases.

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

losses and pressure drops are neglected except instructions.

**2.2 Systematic parameters**

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

**Figure 4.**

**Table 1.**

**51**

*limitation and (b) with limited outlet temperature.*

*Parameters and boundary conditions for the ORC systems.*

The power consumed by the working fluid feed pump was

$$
\dot{W}\_{\rm FP} = \dot{m}\_{\rm O} (h\_3 - h\_2) \tag{3}
$$

The net power output by the ORC system was defined as the turbine power output which subtracts the parasitic power consumed by the working fluid feed pump:

$$
\dot{W}\_{\text{net}} = \dot{W}\_{\text{T}} \eta\_{\text{m}} \eta\_{\text{g}} - \dot{W}\_{\text{FP}} \tag{4}
$$

The cycle thermal efficiency is defined as

$$\eta\_{\rm th} = \frac{\dot{W}\_{\rm net}}{\dot{Q}\_{\rm in}} \tag{5}$$

The turbine exhaust vapor temperature could still be relative high; the superheated vapor could be worth preheating the working fluid in an internal heat exchanger (IHE), also referred to as a recuperator or regenerator, before the preheater [16–18] as shown in **Figure 2(b)**. Eq. (1) for an ORC with an IHE can be rewritten as

$$
\dot{m}\_{\rm H}(h\_{\rm film} - h\_{\rm Hout}) = \dot{m}\_{\rm O}(h\_{\rm O} - h\_{\rm 3a}) \tag{6}
$$

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

The use of an IHE can improve the preheater inlet temperature by the turbine exhaust vapor. For the heat source without the limitation of the outlet temperature [18, 23], the heat source fluid outlet temperature increases because the working fluid is preheated as shown in **Figure 4(a)** and the cycle net power output will maintain constant or increase slightly [24]. For the heat source with a limited outlet temperature as shown in **Figure 4(b)**, the increase in the preheater inlet temperature results in an increase in the working fluid mass flow rate or the optimal evaporation temperature; thus, the net power output increases.

#### **2.2 Systematic parameters**

The matching characteristics between the heat source fluid and the working fluid affect the ORC net power output. In subsequent sections, the working fluid evaporation temperature or pressure will be optimized to maximize the cycle net power output. The operating parameters and boundary conditions for the ORCs are listed in **Table 1**. The parametric optimization and thermodynamic analyses are based on the parameters listed in **Table 1** except the ORC in Section 5.2. The heat losses and pressure drops are neglected except instructions.

The pinch point temperature difference is an important parameter for system design and optimization [25, 26]. During the working fluid evaporation, the pinch point may locate at the bubble point, a subcooled liquid state near the critical point and a superheated state [26] which depends on the heat source temperature drop

#### **Figure 4.**

The power generated by the turbine was

The cycle thermal efficiency is defined as

rewritten as

**50**

**Figure 3.**

The power consumed by the working fluid feed pump was

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

The net power output by the ORC system was defined as the turbine power output

*<sup>η</sup>*th <sup>¼</sup> *<sup>W</sup>*\_ net *Q*\_ in

The turbine exhaust vapor temperature could still be relative high; the superheated vapor could be worth preheating the working fluid in an internal heat exchanger (IHE), also referred to as a recuperator or regenerator, before the preheater [16–18] as shown in **Figure 2(b)**. Eq. (1) for an ORC with an IHE can be

which subtracts the parasitic power consumed by the working fluid feed pump:

*Temperature-entropy diagrams for basic ORCs: (a) o2 cycle, (b) o3 cycle, (c) b3 cycle, and (d) s2 cycle.*

*<sup>W</sup>*\_ <sup>T</sup> <sup>¼</sup> *<sup>m</sup>*\_ <sup>O</sup>ð Þ *<sup>h</sup>*<sup>0</sup> � *<sup>h</sup>*<sup>1</sup> (2)

*<sup>W</sup>*\_ FP <sup>¼</sup> *<sup>m</sup>*\_ <sup>O</sup>ð Þ *<sup>h</sup>*<sup>3</sup> � *<sup>h</sup>*<sup>2</sup> (3)

*<sup>W</sup>*\_ net <sup>¼</sup> *<sup>W</sup>*\_ <sup>T</sup>*η*m*η*<sup>g</sup> � *<sup>W</sup>*\_ FP (4)

*m*\_ <sup>H</sup>ð Þ¼ *h*Hin � *h*Hout *m*\_ <sup>O</sup>ð Þ *h*<sup>0</sup> � *h*3a (6)

(5)

*Temperature-entropy diagrams of ORCs with an IHE for heat source fluids: (a) without outlet temperature limitation and (b) with limited outlet temperature.*


#### **Table 1.**

*Parameters and boundary conditions for the ORC systems.*


**Table 2.**

*Physical and environmental properties of the six working fluids [29].*

and heat capacity and the working fluid thermophysical properties. Thus, each heat exchanger is divided into 100 sections with equal heat-flow interval [25, 27] for determining the pinch point during the parameter optimization.

The energy balance for the flash tank can be expressed as

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

The mass balance for the flash tank can be expressed as

flow rate, *m*\_ <sup>d</sup> is the drained water flow rate, *h*<sup>00</sup>

collected, and the flow rate may be reached 80 t/h.

ture to the ORC system is set to be 180°C.

**3.2 Thermodynamic performance**

**53**

saturated steam, and *h*<sup>0</sup>

**Figure 5.**

*m*\_ bl*h*bl*η*<sup>f</sup> ¼ *m*\_ <sup>f</sup>*h*<sup>00</sup>

*Schematic of an ORC power generation system using the blowdown waste heat from a subcritical boiler [4].*

where *m*\_ bl is the boiler water flow rate, *h*bl is the specific enthalpy of the boiler water, *η*<sup>f</sup> is the flash tank efficiency which is set to be 0.98, *m*\_ <sup>f</sup> is the saturated steam

The flow rate of the saturated steam flashed from the boiler water increases as the flash tank pressure, *p*f, decreases. The flash tank pressure generally is set to be the deaerator pressure which ranges from 0.8 to 1.1 MPa for subcritical steam turbines. Thus, 40–50% of the boiler water is flashed into saturated steam.

The main-steam flow rates of typical subcritical 600 MW power units range from 1600 to 2000 t/h for the BRL (boiler rated load) condition. About 13–16 t/h boiler water enters the flash tank if the blowdown rate is 0.8% which means that the drained water flow rate can reach 6–10 t/h. A power plant generally has 2–8 power units; thus, the drain water from the continuous blown down systems can be

A typical subcritical 600 MW boiler is analyzed in this section. The boiler parameters for the BRL condition are listed in **Table 3**. The boiler water flow rate in a blowdown system is set to be 17.1 t/h. At the flash tank outlet, the drain flow rate is 9 t/h according to Eqs. (7) and (8). Assume a power plant has same four power units; thus, the total drain flow rate is 10 kg/s (36 t/h). The collected drain is used to drive an ORC as shown in **Figure 5**. Considering the heat loss, the drain tempera-

The turbine inlet temperatures (pressure) were optimized to maximize the cycle net power output for the ORCs with saturated vapor at the turbine inlet (o2 cycle)

<sup>f</sup> is the specific enthalpy of the drained water.

<sup>f</sup> þ *m*\_ <sup>d</sup>*h*<sup>0</sup>

<sup>f</sup> (7)

<sup>f</sup> is the specific enthalpy of the

*m*\_ bl ¼ *m*\_ <sup>f</sup> þ *m*\_ <sup>d</sup> (8)

#### **2.3 Working fluids**

Considering the operating safety, nonflammable working fluids are selected for the ORCs using the waste heat in coal-fired power plants. Moreover, the working fluids should be environmentally friendly with a lower ozone depletion potential (ODP) and global working potential (GWP). The critical temperature has been regarded as a criterion for working fluid selection [17, 28]; thus, the studied fluids should cover a wide range of critical temperatures. The selected working fluids here are R134a (1,1,1,2-tetrafluoroethane), R1234ze (1,3,3,3-tetrafluoropropene), R236ea (1,1,1,2,3,3-hexafluoropropane), R245fa (1,1,1,3,3-pentafluoropropane), R1233zd (trans-1-chloro-3,3,3-trifluoropropene), and R123 (2,2-dichloro-1,1,1 trifluoroethane). The physical and environmental properties of the selected working fluids are listed in **Table 2**. The thermophysical properties for the working fluids were calculated using REFPROP 9.1 [29].

#### **3. Waste heat recovery from a boiler blowdown system**

#### **3.1 System description**

A continuous blowdown system is generally used to purify the steam to maintain acceptable levels of total dissolved solids for a subcritical drum boiler. Part of the saturated boiler water at a higher salt concentration is discharge from the drum to the flash tank [3, 30]. The pressure decreases as the saturated boiler water enters the flash tank. The excess energy in the boiler water is given up as some of the saturated water flashes to the saturated steam. The steam then enters into the deaerator to preheat the boiler feedwater, and the drain is generally discharge to the sewer. The energy in the drained liquid is lost through direct discharge.

The drain temperature is generally higher than 150°C. Here, an organic Rankine cycle (ORC) is designed to recover the waste heat of the drained water from a continuous blowdown system for power generation and then to improve the overall thermal efficiency of the power plant [4]. The organic working fluid is heated by the waste heat in the discharged drain and then generates power by expansion through a turbine as shown in **Figure 5**.

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

**Figure 5.** *Schematic of an ORC power generation system using the blowdown waste heat from a subcritical boiler [4].*

The energy balance for the flash tank can be expressed as

$$
\dot{m}\_{\rm bl} h\_{\rm bl} \eta\_{\rm f} = \dot{m}\_{\rm f} h\_{\rm f}^{\prime \prime} + \dot{m}\_{\rm d} h\_{\rm f}^{\prime} \tag{7}
$$

where *m*\_ bl is the boiler water flow rate, *h*bl is the specific enthalpy of the boiler water, *η*<sup>f</sup> is the flash tank efficiency which is set to be 0.98, *m*\_ <sup>f</sup> is the saturated steam flow rate, *m*\_ <sup>d</sup> is the drained water flow rate, *h*<sup>00</sup> <sup>f</sup> is the specific enthalpy of the saturated steam, and *h*<sup>0</sup> <sup>f</sup> is the specific enthalpy of the drained water.

The mass balance for the flash tank can be expressed as

$$
\dot{m}\_{\rm bl} = \dot{m}\_{\rm f} + \dot{m}\_{\rm d} \tag{8}
$$

The flow rate of the saturated steam flashed from the boiler water increases as the flash tank pressure, *p*f, decreases. The flash tank pressure generally is set to be the deaerator pressure which ranges from 0.8 to 1.1 MPa for subcritical steam turbines. Thus, 40–50% of the boiler water is flashed into saturated steam.

The main-steam flow rates of typical subcritical 600 MW power units range from 1600 to 2000 t/h for the BRL (boiler rated load) condition. About 13–16 t/h boiler water enters the flash tank if the blowdown rate is 0.8% which means that the drained water flow rate can reach 6–10 t/h. A power plant generally has 2–8 power units; thus, the drain water from the continuous blown down systems can be collected, and the flow rate may be reached 80 t/h.

A typical subcritical 600 MW boiler is analyzed in this section. The boiler parameters for the BRL condition are listed in **Table 3**. The boiler water flow rate in a blowdown system is set to be 17.1 t/h. At the flash tank outlet, the drain flow rate is 9 t/h according to Eqs. (7) and (8). Assume a power plant has same four power units; thus, the total drain flow rate is 10 kg/s (36 t/h). The collected drain is used to drive an ORC as shown in **Figure 5**. Considering the heat loss, the drain temperature to the ORC system is set to be 180°C.

#### **3.2 Thermodynamic performance**

The turbine inlet temperatures (pressure) were optimized to maximize the cycle net power output for the ORCs with saturated vapor at the turbine inlet (o2 cycle)

and heat capacity and the working fluid thermophysical properties. Thus, each heat exchanger is divided into 100 sections with equal heat-flow interval [25, 27] for

**Working fluid** *T***cr/°C** *p***cr/MPa Safety class ODP GWP** R134a 101.06 4.059 A1 0 1430 R1234ze(E) 109.36 3.635 A2L 0 6 R236ea 139.29 3.420 n.a 0 1410 R245fa 154.01 3.651 B1 0 1030 R1233zd(E) 165.60 3.572 A1 0.003 1 R123 183.68 3.662 B1 0.01 77

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

Considering the operating safety, nonflammable working fluids are selected for the ORCs using the waste heat in coal-fired power plants. Moreover, the working fluids should be environmentally friendly with a lower ozone depletion potential (ODP) and global working potential (GWP). The critical temperature has been regarded as a criterion for working fluid selection [17, 28]; thus, the studied fluids should cover a wide range of critical temperatures. The selected working fluids here are R134a (1,1,1,2-tetrafluoroethane), R1234ze (1,3,3,3-tetrafluoropropene), R236ea (1,1,1,2,3,3-hexafluoropropane), R245fa (1,1,1,3,3-pentafluoropropane), R1233zd (trans-1-chloro-3,3,3-trifluoropropene), and R123 (2,2-dichloro-1,1,1 trifluoroethane). The physical and environmental properties of the selected working fluids are listed in **Table 2**. The thermophysical properties for the working

A continuous blowdown system is generally used to purify the steam to maintain acceptable levels of total dissolved solids for a subcritical drum boiler. Part of the saturated boiler water at a higher salt concentration is discharge from the drum to the flash tank [3, 30]. The pressure decreases as the saturated boiler water enters the flash tank. The excess energy in the boiler water is given up as some of the saturated water flashes to the saturated steam. The steam then enters into the deaerator to preheat the boiler feedwater, and the drain is generally discharge to the sewer. The

The drain temperature is generally higher than 150°C. Here, an organic Rankine

cycle (ORC) is designed to recover the waste heat of the drained water from a continuous blowdown system for power generation and then to improve the overall thermal efficiency of the power plant [4]. The organic working fluid is heated by the waste heat in the discharged drain and then generates power by expansion

determining the pinch point during the parameter optimization.

*Physical and environmental properties of the six working fluids [29].*

fluids were calculated using REFPROP 9.1 [29].

**3. Waste heat recovery from a boiler blowdown system**

energy in the drained liquid is lost through direct discharge.

through a turbine as shown in **Figure 5**.

**52**

**2.3 Working fluids**

**Table 2.**

**3.1 System description**

#### *Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*


supercritical ORCs. The turbine inlet temperature and pressure are optimized simultaneously to maximize the cycle net power output considering the pinch point temperature difference limit. Compared with subcritical ORCs, a supercritical ORC produces 13–37% more net cycle power as shown in **Table 4**. The optimal turbine inlet temperatures for supercritical ORCs with the selected working fluids are very close; however, the working fluid with a lower critical temperature has a higher optimal turbine inlet pressure. The high operating pressure and heat transfer deterioration due to the large specific heat near the critical point must be considered in the system design [4]. The supercritical ORC using R236ea generates the highest net power (764 kW) among the considered ORCs with the selected working fluids.

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

The heat loss in the exhaust flue gas is the largest heat loss in a power station boiler which typically accounts 70–80% of the total boiler heat losses [30]. The temperature of the exhaust flue gas from a power station boiler generally ranges from 120–140°C. Unfortunately, the flue gas temperature may be higher than the design value in operation which results in more heat losses and lower boiler thermal efficiencies. The exhaust flue gas can be cooled by a low-boiling organic working fluid with the fluid then passing through a turbine to generate power. The potential decrease in the exhaust flue gas temperature is 10–30°C. Supercritical 1000 MW boilers for power plants consume about 350 tons of bituminous coal and generate about 3 <sup>10</sup><sup>6</sup> Nm<sup>3</sup> flue gas per hour for the BRL condition. More than 10 MW waste heat can be recovered to drive an ORC when the flue gas temperature is reduced by 10°C. This section discusses the thermodynamic performance of ORCs

More than 90 supercritical 1000 MW coal-fired power units are being operated in China with more 1000 MW power units planned to replace many current smallcapacity and low-efficiency power units [31]. Therefore, a modern supercritical 1000 MW power unit is taken here as a case study. The boiler operating parameters for the BRL condition are listed in **Table 5**. The main-steam temperature is 605°C with a flow rate of 800.5kg/s. The exhaust steam from the high-pressure steam turbine is then reheated to 603°C in the reheater with a flow rate of 649kg/s. The feedwater temperature at the economizer inlet is 299.4°C with no water and steam assumed to be lost in the boiler. The exhaust flue gas temperature is 135°C, and the

The ultimate analysis (UA) of the fuel is useful for the calculation of air and flue gas quantities and other combustion calculations [3, 30]. The coal consists of five elements (carbon, hydrogen, nitrogen, oxygen, and sulfur), moisture, and ash. A typical bituminous coal is used here as the boiler fuel. The elemental weights of the coal are determined on as-received basis: Car = 61.7%, Har = 3.67%, Oar = 8.56%, Nar = 1.12%, Sar = 0.6%, Aar = 8.8%, and Mar = 15.55%. The lower calorific value (LCV) of the bituminous coal is 23.47 MJ/kg. The boiler consumes 353.5 tons of bituminous

Sulfur in the coal on combustion forms SO2 and partly SO3. Sulfuric acid will be formed because of the reaction between SO3 and H2O in the flue gas at lower temperatures, which then condenses on the tube surface below the acid dew point and leads to the corrosion of the tubes [3, 30]. Therefore, the flue gas was assumed to be cooled to 110°C by the organic working fluids to avoid low-temperature corrosion.

coal and generates 3.055 106 Nm3 flue gas per hour for the BRL condition.

**4. Waste heat recovery from boiler exhaust flue gas**

using the flue gas waste heat from a supercritical boiler.

**4.1 Coal-fired boiler**

**55**

boiler thermal efficiency is 92.57%.

#### **Table 3.**

*Blowdown system parameters of a 600 MW boiler for the BRL condition.*


#### **Table 4.**

*Parameters of the ORC for waste heat recovery from a blowdown system.*

using R245fa, R1233zd(E), and R123. The optimized turbine inlet temperatures and the maximum net cycle power outputs are listed in **Table 4**. The results show that the optimal turbine inlet temperature decreases as the critical temperature of the working fluid increases. The net power output as well as the waste heat utilization rate decreases for working fluids with higher critical temperatures. The net cycle power output is 660 kW for the o2 cycle using R245fa, while the net cycle power output is only 599 kW for the o2 cycle with R123.

Here, the superheated ORC uses a working fluid with a lower critical temperature (R134a, R1234ze, and R236ea) for recovering the waste heat from the drained water. The turbine inlet temperature and pressure are optimized simultaneously to obtain the maximum cycle net power output using the generalized reduced gradient (GRG) method which has been successfully used in previous work [4, 27]. Both the optimized turbine inlet temperature and pressure are higher for the working fluid with lower critical temperature as shown in **Table 4**. However, the working fluid with higher critical temperature generates a higher net power due to the higher working fluid flow rate. The net cycle power output with R236ea is higher than that with R134a or R1234ze.

A supercritical ORC provides a better match between the working fluid and the heat source fluid temperature profiles. The working fluids with much lower critical temperatures, such as R134a, R1234ze, and R236ea, are considered here for

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

supercritical ORCs. The turbine inlet temperature and pressure are optimized simultaneously to maximize the cycle net power output considering the pinch point temperature difference limit. Compared with subcritical ORCs, a supercritical ORC produces 13–37% more net cycle power as shown in **Table 4**. The optimal turbine inlet temperatures for supercritical ORCs with the selected working fluids are very close; however, the working fluid with a lower critical temperature has a higher optimal turbine inlet pressure. The high operating pressure and heat transfer deterioration due to the large specific heat near the critical point must be considered in the system design [4]. The supercritical ORC using R236ea generates the highest net power (764 kW) among the considered ORCs with the selected working fluids.

#### **4. Waste heat recovery from boiler exhaust flue gas**

The heat loss in the exhaust flue gas is the largest heat loss in a power station boiler which typically accounts 70–80% of the total boiler heat losses [30]. The temperature of the exhaust flue gas from a power station boiler generally ranges from 120–140°C. Unfortunately, the flue gas temperature may be higher than the design value in operation which results in more heat losses and lower boiler thermal efficiencies. The exhaust flue gas can be cooled by a low-boiling organic working fluid with the fluid then passing through a turbine to generate power. The potential decrease in the exhaust flue gas temperature is 10–30°C. Supercritical 1000 MW boilers for power plants consume about 350 tons of bituminous coal and generate about 3 <sup>10</sup><sup>6</sup> Nm<sup>3</sup> flue gas per hour for the BRL condition. More than 10 MW waste heat can be recovered to drive an ORC when the flue gas temperature is reduced by 10°C. This section discusses the thermodynamic performance of ORCs using the flue gas waste heat from a supercritical boiler.

#### **4.1 Coal-fired boiler**

More than 90 supercritical 1000 MW coal-fired power units are being operated in China with more 1000 MW power units planned to replace many current smallcapacity and low-efficiency power units [31]. Therefore, a modern supercritical 1000 MW power unit is taken here as a case study. The boiler operating parameters for the BRL condition are listed in **Table 5**. The main-steam temperature is 605°C with a flow rate of 800.5kg/s. The exhaust steam from the high-pressure steam turbine is then reheated to 603°C in the reheater with a flow rate of 649kg/s. The feedwater temperature at the economizer inlet is 299.4°C with no water and steam assumed to be lost in the boiler. The exhaust flue gas temperature is 135°C, and the boiler thermal efficiency is 92.57%.

The ultimate analysis (UA) of the fuel is useful for the calculation of air and flue gas quantities and other combustion calculations [3, 30]. The coal consists of five elements (carbon, hydrogen, nitrogen, oxygen, and sulfur), moisture, and ash. A typical bituminous coal is used here as the boiler fuel. The elemental weights of the coal are determined on as-received basis: Car = 61.7%, Har = 3.67%, Oar = 8.56%, Nar = 1.12%, Sar = 0.6%, Aar = 8.8%, and Mar = 15.55%. The lower calorific value (LCV) of the bituminous coal is 23.47 MJ/kg. The boiler consumes 353.5 tons of bituminous coal and generates 3.055 106 Nm3 flue gas per hour for the BRL condition.

Sulfur in the coal on combustion forms SO2 and partly SO3. Sulfuric acid will be formed because of the reaction between SO3 and H2O in the flue gas at lower temperatures, which then condenses on the tube surface below the acid dew point and leads to the corrosion of the tubes [3, 30]. Therefore, the flue gas was assumed to be cooled to 110°C by the organic working fluids to avoid low-temperature corrosion.

using R245fa, R1233zd(E), and R123. The optimized turbine inlet temperatures and the maximum net cycle power outputs are listed in **Table 4**. The results show that the optimal turbine inlet temperature decreases as the critical temperature of the working fluid increases. The net power output as well as the waste heat utilization rate decreases for working fluids with higher critical temperatures. The net cycle power output is 660 kW for the o2 cycle using R245fa, while the net cycle power

**Working fluid Cycle** *T***T,in/°C** *P***T,in/MPa Power output/kW Heat utilization ratio/%**

R245fa o2 116.3 1.79 660 70.99 R1233zd(E) o2 109.8 1.28 622 68.08 R123 o2 106.4 0.90 599 65.88 R134a b1 169.6 3.06 548 75.81 R1234ze o3 153.0 2.63 567 76.47 R236ea o3 121.4 2.06 677 77.81 R134a S2 163.4 8.57 749 77.10 R1234ze S2 163.7 7.48 746 77.10 R236ea S2 160.7 4.95 764 77.10

**Parameters Values Unit** Main-steam flow rate 2011 t/h Steam drum pressure 18.5 MPa Boiler water temperature in drum 359.3 °C Flash tank pressure 1.03 MPa Drain temperature 181.2 °C Flash tank efficiency 98 % Drain flow rate 10 kg/s Drain pressure 1.03 MPa Drain temperature at the evaporator inlet 180 °C

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

Here, the superheated ORC uses a working fluid with a lower critical temperature (R134a, R1234ze, and R236ea) for recovering the waste heat from the drained water. The turbine inlet temperature and pressure are optimized simultaneously to obtain the maximum cycle net power output using the generalized reduced gradient (GRG) method which has been successfully used in previous work [4, 27]. Both the optimized turbine inlet temperature and pressure are higher for the working fluid with lower critical temperature as shown in **Table 4**. However, the working fluid with higher critical temperature generates a higher net power due to the higher working fluid flow rate. The net cycle power output with R236ea is higher than that

A supercritical ORC provides a better match between the working fluid and the heat source fluid temperature profiles. The working fluids with much lower critical

temperatures, such as R134a, R1234ze, and R236ea, are considered here for

output is only 599 kW for the o2 cycle with R123.

*Parameters of the ORC for waste heat recovery from a blowdown system.*

*Blowdown system parameters of a 600 MW boiler for the BRL condition.*

with R134a or R1234ze.

**54**

**Table 3.**

**Table 4.**

#### *Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*


#### **Table 5.**

*Operating parameters of a supercritical 1000 MW boiler for the BRL condition.*

#### **4.2 ORC system driven by waste heat from boiler flue gas**

**Figure 6** shows an ORC system using waste heat of flue gas from a boiler for power generation. A counterflow heat exchanger is used here to recover the waste heat because the flue gas temperature is much lower. The heat exchanger tubes are placed in an in-line arrangement for a lower flue gas pressure drop. The flue gas longitudinal flows outside the tubes. The working fluid is heated to a vapor from a subcooled liquid in the tubes of the heat exchanger.

The heat source temperature is a key parameter which influences the choice of cycle types and working fluids [16–19, 23]. The flue gas temperature should be higher than the acidic dew point to avoid low-temperature external corrosion; thus, the temperature drop of the exhaust flue gas from power station boiler is much lower. Considering the lower temperature drop and the higher flue gas temperature at the preheater outlet, an internal heat exchanger (IHE) is used in the ORC system to improve the working fluid temperature at the preheater inlet by the turbine exhaust vapor and then decrease the temperature difference between the flue gas and the working fluid.

The energy balance in the heat exchanger can be expressed as

$$
\dot{B}V\_{\rm fg} \left( I\_{\rm fg}^{\rm in} - I\_{\rm fg}^{\rm out} \right) \eta\_{\rm HE} = \dot{m}\_{\rm O} (h\_0 - h\_{3\rm a}) \tag{9}
$$

working fluids. The heat losses and pressure drops in the ORC system are neglected

Both the o2 cycle and the o3 cycle are studied with the working fluids of R123, R1233zd, R245fa, and R236ea. The evaporation temperatures (pressures) of the selected working fluids were optimized for the o2 cycle to maximize the net cycle power output. The optimized parameters and system performance are listed in **Table 6**. **Figure 7(a)** shows the T-Q diagrams of the selected working fluids for o2 cycle. The optimal evaporation temperature of the o2 cycle is lower for the working fluid with a higher critical temperature which results in a higher heat load for the isothermal evaporation process and a better match between the working fluid and the flue gas temperatures as shown in **Figure 7(a)**. The working fluid with a higher critical temperature also has a lower evaporation pressure which leads to a lower working fluid pump power consumption. For example, 61.7% of the total heat flow is used for R123 boiling process, while 54.3% of the total heat flow is needed for R236fa preheating which process has a large temperature difference. Therefore, the

to simplify the calculations.

**Figure 6.**

**57**

**4.3 Thermodynamic performance**

*Schematic of an ORC system using waste heat of flue gas from a boiler.*

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

where *B*\_ is the coal consumption rate, *V*fg is the specific volume flue gas based on 1 kg coal (unit flue gas volume), *I* in fg is the flue gas enthalpy at the heat exchanger inlet, and *I* out fg is the flue gas enthalpy at the heat exchanger outlet, *η*HE is the heat exchanger efficiency which considering the heat loss, *m*\_ <sup>o</sup> is the organic working fluid flow rate, *h*<sup>0</sup> is the working fluid specific enthalpy at the heat exchanger outlet, and *h*3a is the specific enthalpy at the IHE outlet.

The unit gas weights (volumes) for the BRL condition were calculated using the ultimate analysis [3, 30]. The excess air coefficient in the flue gas was assumed to be 1.3. The heat-flow rate is 29.26 MW when the flue gas temperature is reduced from 135 to 110°C. The total heat loss is assumed to be 5% during the heat transfer due to the radiation and convection; thus, 27.8 MW heat is transferred to the organic

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

**Figure 6.** *Schematic of an ORC system using waste heat of flue gas from a boiler.*

working fluids. The heat losses and pressure drops in the ORC system are neglected to simplify the calculations.

#### **4.3 Thermodynamic performance**

Both the o2 cycle and the o3 cycle are studied with the working fluids of R123, R1233zd, R245fa, and R236ea. The evaporation temperatures (pressures) of the selected working fluids were optimized for the o2 cycle to maximize the net cycle power output. The optimized parameters and system performance are listed in **Table 6**. **Figure 7(a)** shows the T-Q diagrams of the selected working fluids for o2 cycle. The optimal evaporation temperature of the o2 cycle is lower for the working fluid with a higher critical temperature which results in a higher heat load for the isothermal evaporation process and a better match between the working fluid and the flue gas temperatures as shown in **Figure 7(a)**. The working fluid with a higher critical temperature also has a lower evaporation pressure which leads to a lower working fluid pump power consumption. For example, 61.7% of the total heat flow is used for R123 boiling process, while 54.3% of the total heat flow is needed for R236fa preheating which process has a large temperature difference. Therefore, the

**4.2 ORC system driven by waste heat from boiler flue gas**

*Operating parameters of a supercritical 1000 MW boiler for the BRL condition.*

The energy balance in the heat exchanger can be expressed as

in

*BV*\_ fg *I* in fg � *I* out fg 

outlet, and *h*3a is the specific enthalpy at the IHE outlet.

subcooled liquid in the tubes of the heat exchanger.

and the working fluid.

out

inlet, and *I*

**56**

**Table 5.**

1 kg coal (unit flue gas volume), *I*

**Figure 6** shows an ORC system using waste heat of flue gas from a boiler for power generation. A counterflow heat exchanger is used here to recover the waste heat because the flue gas temperature is much lower. The heat exchanger tubes are placed in an in-line arrangement for a lower flue gas pressure drop. The flue gas longitudinal flows outside the tubes. The working fluid is heated to a vapor from a

**Parameters Values Unit** Main-steam temperature at superheater outlet 605 °C Main-steam pressure at superheater outlet 26.13 MPa Main-steam flow rate 800.49 kg/s Steam temperature at reheater outlet 603 °C Reheated steam pressure at reheater outlet 4.57 MPa Steam temperature at reheater inlet 347.7 °C Reheated steam pressure at reheater inlet 4.80 MPa Reheated steam flow rate 648.96 kg/s Feedwater temperature 299.4 °C Boiler thermal efficiency 92.57 % Exhaust flue gas temperature 135 °C Lower calorific value (LCV) of coal 23469.7 kJ/kg

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

The heat source temperature is a key parameter which influences the choice of cycle types and working fluids [16–19, 23]. The flue gas temperature should be higher than the acidic dew point to avoid low-temperature external corrosion; thus, the temperature drop of the exhaust flue gas from power station boiler is much lower. Considering the lower temperature drop and the higher flue gas temperature at the preheater outlet, an internal heat exchanger (IHE) is used in the ORC system to improve the working fluid temperature at the preheater inlet by the turbine exhaust vapor and then decrease the temperature difference between the flue gas

where *B*\_ is the coal consumption rate, *V*fg is the specific volume flue gas based on

The unit gas weights (volumes) for the BRL condition were calculated using the ultimate analysis [3, 30]. The excess air coefficient in the flue gas was assumed to be 1.3. The heat-flow rate is 29.26 MW when the flue gas temperature is reduced from 135 to 110°C. The total heat loss is assumed to be 5% during the heat transfer due to the radiation and convection; thus, 27.8 MW heat is transferred to the organic

exchanger efficiency which considering the heat loss, *m*\_ <sup>o</sup> is the organic working fluid flow rate, *h*<sup>0</sup> is the working fluid specific enthalpy at the heat exchanger

fg is the flue gas enthalpy at the heat exchanger outlet, *η*HE is the heat

*η*HE ¼ *m*\_ <sup>O</sup>ð Þ *h*<sup>0</sup> � *h*3a (9)

fg is the flue gas enthalpy at the heat exchanger

#### *Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*


and the working fluid decreases, and the o3 cycle produces more power than the o2

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

In this case study, the working fluid with a higher critical temperature needs a lower heat load for superheating and a higher heat load for boiling as shown in **Figure 7(b)**. For instance, the heat load for R123 superheating accounts for 7.4%, and the heat load for R123 boiling accounts for 61.8% of the total heat load; however, the heat load for R134a superheating accounts for 28.8%, and the heat load for R134a boiling only accounts for 22.4% of the total heat load because the boiling temperature is close to its critical temperature. The working fluids with lower critical temperatures have higher optimal evaporation pressures which also results in more power consumed by the working fluid pump. The ORC with superheating using R123 as the working fluid can generate 17.6% more power than that using

The supercritical ORCs using R1234ze and R134a which has a lower critical temperature are analyzed and compared here. The turbine inlet temperature is also set to be 125°C with the turbine inlet pressure of 1.05 *p*cr because the turbine inlet temperature is very close to the critical temperature. The supercritical ORCs with R134a and R1234ze give a better matching of the temperature profiles between the working fluid and the flue gas compared to the subcritical ORCs as shown in **Figure 7(b)** which results in a higher turbine power output; however, the power consumed by the working fluid pump increases with increasing evaporation pressure which offsets the advantages of the supercritical ORCs. For example, the net power output of the supercritical ORC with R134a is even lower than the subcritical ORC for this case study. The subcritical ORC with superheating (o3 cycle) shows a better thermodynamic performance for lower temperature drop heat sources. The maximum power is produced by o3 cycle with R123 for waste heat recovery from the flue gas at a temperature of 135 °C, followed by o3 cycle with R1233zd and

The volume flow rate at the turbine inlet as well as the turbine outlet/inlet volumetric flow ratio is important for the turbine design [17, 26]. **Figure 8(a)** shows the thermal efficiencies as a function of the turbine inlet volume flow rates of the working fluids for the optimal conditions. The mass flow rates of the selected working fluids are very close which range from 119 to 146 kg/s. However, a lower turbine inlet pressure leads to a lower density; thus, the working fluids evaporated at lower pressures have a much higher volume flow rates. The volume flow rate of R123 is the highest for the o2 cycle among the working fluids. Compared to the o2

*Thermal efficiencies versus (a) volume flow rate, V0, at the turbine inlet and (b) turbine outlet/inlet volumetric flow rate ratios for the optimal condition:* ● *saturated vapor,* ○ *superheated vapor, and* ☉ *supercritical fluid.*

cycle as shown in **Table 6**.

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

R245fa in this case study.

R134a.

**Figure 8.**

**59**

**Table 6.**

*Comparison of ORCs using flue gas waste heat from a 1000 MW boiler.*

**Figure 7.**

*T-Q diagrams for (a) the o2 cycles and (b) the cycles with superheated vapor or supercritical fluid at the turbine inlet.*

o2 cycle using a working fluid with a higher critical temperature produces more net power as shown in **Table 6**. The o2 cycle using R123 generates about 4% more net power than that using R236ea.

Increasing the heat load of the isothermal evaporation process can better match the temperature profiles between the flue gas and the working fluid due to the lower temperature drop of the flue gas. The subcritical cycle using superheating process can further improve the temperature matching between the flue gas and the working fluid. The turbine inlet temperatures of the selected working fluids were set here to be 125°C for the subcritical cycles, and then the evaporation pressures were optimized to obtain the maximum net cycle power with the minimum temperature difference between the flue gas and the working fluid not less than 10°C. Compared to the o2 cycle, the use of superheating results in a slight decrease in the evaporation pressure and then reduces the power consumption of the working fluid pump. The turbine exhaust vapor temperature increases about 18°C due to the superheat degree at the turbine inlet; however, the working fluid temperature in the IHE can be increased by 12–13°C by the waste heat from the turbine exhaust vapor. Therefore, the average heat transfer temperature difference between the flue gas

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

and the working fluid decreases, and the o3 cycle produces more power than the o2 cycle as shown in **Table 6**.

In this case study, the working fluid with a higher critical temperature needs a lower heat load for superheating and a higher heat load for boiling as shown in **Figure 7(b)**. For instance, the heat load for R123 superheating accounts for 7.4%, and the heat load for R123 boiling accounts for 61.8% of the total heat load; however, the heat load for R134a superheating accounts for 28.8%, and the heat load for R134a boiling only accounts for 22.4% of the total heat load because the boiling temperature is close to its critical temperature. The working fluids with lower critical temperatures have higher optimal evaporation pressures which also results in more power consumed by the working fluid pump. The ORC with superheating using R123 as the working fluid can generate 17.6% more power than that using R134a.

The supercritical ORCs using R1234ze and R134a which has a lower critical temperature are analyzed and compared here. The turbine inlet temperature is also set to be 125°C with the turbine inlet pressure of 1.05 *p*cr because the turbine inlet temperature is very close to the critical temperature. The supercritical ORCs with R134a and R1234ze give a better matching of the temperature profiles between the working fluid and the flue gas compared to the subcritical ORCs as shown in **Figure 7(b)** which results in a higher turbine power output; however, the power consumed by the working fluid pump increases with increasing evaporation pressure which offsets the advantages of the supercritical ORCs. For example, the net power output of the supercritical ORC with R134a is even lower than the subcritical ORC for this case study. The subcritical ORC with superheating (o3 cycle) shows a better thermodynamic performance for lower temperature drop heat sources. The maximum power is produced by o3 cycle with R123 for waste heat recovery from the flue gas at a temperature of 135 °C, followed by o3 cycle with R1233zd and R245fa in this case study.

The volume flow rate at the turbine inlet as well as the turbine outlet/inlet volumetric flow ratio is important for the turbine design [17, 26]. **Figure 8(a)** shows the thermal efficiencies as a function of the turbine inlet volume flow rates of the working fluids for the optimal conditions. The mass flow rates of the selected working fluids are very close which range from 119 to 146 kg/s. However, a lower turbine inlet pressure leads to a lower density; thus, the working fluids evaporated at lower pressures have a much higher volume flow rates. The volume flow rate of R123 is the highest for the o2 cycle among the working fluids. Compared to the o2

#### **Figure 8.**

*Thermal efficiencies versus (a) volume flow rate, V0, at the turbine inlet and (b) turbine outlet/inlet volumetric flow rate ratios for the optimal condition:* ● *saturated vapor,* ○ *superheated vapor, and* ☉ *supercritical fluid.*

o2 cycle using a working fluid with a higher critical temperature produces more net power as shown in **Table 6**. The o2 cycle using R123 generates about 4% more net

*T-Q diagrams for (a) the o2 cycles and (b) the cycles with superheated vapor or supercritical fluid at the*

**Working fluid Cycle** *T***T,in/°C** *P***T,in/MPa Power out/kW Thermal efficiency/%**

R236ea o2 113.67 2.09 3783.0 13.61 R245fa o2 111.84 1.63 3845.9 13.84 R1233zd(E) o2 110.69 1.31 3876.9 13.95 R123 o2 109.60 0.97 3937.7 14.17 R134a b3 125 3.85 3502.0 12.60 R1234ze o3 125 2.74 3570.6 12.85 R236ea o3 125 1.97 4025.5 14.48 R245fa o3 125 1.55 4070.6 14.65 R1233zd(E) o3 125 1.25 4074.6 14.66 R123 o3 125 0.93 4118.6 14.82 R134a s2 125 4.26 3484.8 12.54 R1234ze s2 125 3.82 3576.5 12.87

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

*Comparison of ORCs using flue gas waste heat from a 1000 MW boiler.*

Increasing the heat load of the isothermal evaporation process can better match

the temperature profiles between the flue gas and the working fluid due to the lower temperature drop of the flue gas. The subcritical cycle using superheating process can further improve the temperature matching between the flue gas and the working fluid. The turbine inlet temperatures of the selected working fluids were set here to be 125°C for the subcritical cycles, and then the evaporation pressures were optimized to obtain the maximum net cycle power with the minimum temperature difference between the flue gas and the working fluid not less than 10°C. Compared to the o2 cycle, the use of superheating results in a slight decrease in the evaporation pressure and then reduces the power consumption of the working fluid pump. The turbine exhaust vapor temperature increases about 18°C due to the superheat degree at the turbine inlet; however, the working fluid temperature in the IHE can be increased by 12–13°C by the waste heat from the turbine exhaust vapor. Therefore, the average heat transfer temperature difference between the flue gas

power than that using R236ea.

**Table 6.**

**Figure 7.**

**58**

*turbine inlet.*

cycle, the volume flow rate is increased by 10–20% for the o3 cycle due to the lower turbine inlet pressure as shown in **Figure 8(a)**, while the thermal efficiency can be increased by 4.6–6.4%. Both the R134a and R1234ze have lower volume flow rates with lower thermal efficiencies. The working fluid has a higher volume flow rate at the turbine inlet, also somewhat has a higher thermal efficiency as shown in **Figure 8(a)**.

mitigated if the back-pressure steam turbine can be operated at the optimum condition (design condition) and the excess steam which beyond the steam/heat demand is used to generate electricity by an ORC which exports electricity directly into the grid [32]. The back-pressure turbine will operate with constant steam flow rate, but the steam flow rate to the ORC system varies with the steam/heat demand. The ORCs have the distinctive features of being simple cycle layouts, being low cost and especially being efficient for partial loads. Therefore, an ORC can generate power using the excess steam with high thermal efficiency and availability for complex operating conditions. In this conceptual system, the back-pressure steam turbine is operating efficiently at the optimal condition, and the extra electricity generated from the ORC system is sold to the grid which improves the economics

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

A typical B25–8.83/1.5 extraction back-pressure turbine is taken as a case study. The parameters for typical conditions of the back-pressure turbine are listed in **Table 7**, and the schematic is shown in **Figure 9**. The superheated steam from the boiler enters the back-pressure steam turbine to generate power. One stage steam is extracted from the steam turbine for the feedwater heater 1 (FWH1) to preheat the feedwater. In the original system, the exhaust steam of the back-pressure turbine is divided into four stages of steam flow: one stage extracted from the exhaust steam for FWH2 to preheat the feedwater, one stage extracted for the deaerator after throttling, one stage extracted for the mixer after throttling to a lower pressure to

**Parameters THA 75%THA 50%THA 30%THA** Main-steam temperature/°C 535 535 535 535 Main-steam pressure/MPa 8.83 8.83 8.83 8.83 Main-steam flow rate/t<sup>h</sup><sup>1</sup> <sup>245</sup> <sup>195</sup> <sup>148</sup> <sup>109</sup> Power output/kW 25,203 18,862 12,637 7545 Exhaust steam temperature/°C 316.2 327 345.3 370.6 Exhaust steam pressure/MPa 1.5 1.5 1.5 1.5 Steam demand/t<sup>h</sup><sup>1</sup> 174.5 140.9 108.6 81.0 Isentropic efficiency/% 79.5 74.77 66.77 55.87

*Schematic of an ORC for recovering the excess exhaust steam heat from a back-pressure steam turbine [34].*

*Operating parameters for a B25–8.83/1.5 back-pressure turbine [34].*

considerably.

**Table 7.**

**Figure 9.**

**61**

**Figure 8(b)** shows the thermal efficiencies as a function of the VR (turbine outlet/inlet volumetric flow rate ratios, V1/V0) at the optimum turbine parameters for the maximum power output. Compared to the o2 cycle, the expansion ratio in the turbine decreases for the o3 cycle with the same condensation pressure which also leads to a decrease in the VR. Both the expansion ratios of R134a and R1234ze are much lower in the subcritical ORC because of the higher condensation pressures which also results in lower VRs. The VRs of R134a and R1234ze are increased in the supercritical ORC. Among the selected working fluids, R123, R1233zd, and R245fa have high thermal efficiencies with high-turbine-inlet volume flow rates and turbine outlet/inlet volumetric flow rate ratios. R1233zd is considered to be the most suitable working fluid for the ORC which recovers waste heat from the boiler flue gas with taking the safety and environment protection into account.

#### **5. Heat recovery from turbine exhaust steam**

#### **5.1 ORCs driven by exhaust steam from a back-pressure turbine**

Back-pressure steam turbines supply not only electricity but also the steam and heat for processes. The exhaust steam of the back-pressure steam turbine is directly used to supply heat or steam to the facilities without condensation. Ideally, the exhaust steam and the electricity from the back-pressure steam turbine are supplied to the same users [32]. The back-pressure steam turbine does not have any cold source loss (heat loss released directly to the environment). Therefore, the backpressure steam turbines are efficient and have been widely used in industrial applications such as oil refineries, petrochemical plants, and cogeneration [33].

The exhaust steam pressure generally is set to be the demand pressure from the facility or outside needs [33]; thus, a lower expansion ratio results in a lower enthalpy drop and small power output in a back-pressure steam turbine. Only fewer turbine stages are used due to the lower enthalpy drop which leads to a simple structure and a lower cost for a back-pressure steam turbine.

Process steam/heat demand and electricity demand change independently according to the season or facility production demand [33]. An imbalance between process steam/heat and electricity demands is one of the most common problems in actual operation. The steam/heat demand is the primary requirement with electricity demand a secondary consideration to solve the imbalance demands because the steam and heat cannot economically and conveniently be transported over a long distance. Adjusting the main-steam flow rate of a back-pressure turbine is the major solution to meet the steam/heat demand. Thus, the back-pressure turbine power (electricity) output varies with the steam/heat demand. However, the thermal efficiency of the back-pressure steam turbine is significantly decreased with decreasing electricity output at partial loads due to the lower enthalpy drop. For example, the turbine isentropic efficiency (relative internal efficiency) can be decreased from 85 to 60% for lower steam flow rates which leads to an increase in the heat consumption of power generation [34].

The imbalance between steam/heat and electricity demands reduces the economic performance of back-pressure steam turbines. Part of this risk will be

#### *Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

mitigated if the back-pressure steam turbine can be operated at the optimum condition (design condition) and the excess steam which beyond the steam/heat demand is used to generate electricity by an ORC which exports electricity directly into the grid [32]. The back-pressure turbine will operate with constant steam flow rate, but the steam flow rate to the ORC system varies with the steam/heat demand. The ORCs have the distinctive features of being simple cycle layouts, being low cost and especially being efficient for partial loads. Therefore, an ORC can generate power using the excess steam with high thermal efficiency and availability for complex operating conditions. In this conceptual system, the back-pressure steam turbine is operating efficiently at the optimal condition, and the extra electricity generated from the ORC system is sold to the grid which improves the economics considerably.

A typical B25–8.83/1.5 extraction back-pressure turbine is taken as a case study. The parameters for typical conditions of the back-pressure turbine are listed in **Table 7**, and the schematic is shown in **Figure 9**. The superheated steam from the boiler enters the back-pressure steam turbine to generate power. One stage steam is extracted from the steam turbine for the feedwater heater 1 (FWH1) to preheat the feedwater. In the original system, the exhaust steam of the back-pressure turbine is divided into four stages of steam flow: one stage extracted from the exhaust steam for FWH2 to preheat the feedwater, one stage extracted for the deaerator after throttling, one stage extracted for the mixer after throttling to a lower pressure to


#### **Table 7.**

cycle, the volume flow rate is increased by 10–20% for the o3 cycle due to the lower turbine inlet pressure as shown in **Figure 8(a)**, while the thermal efficiency can be increased by 4.6–6.4%. Both the R134a and R1234ze have lower volume flow rates with lower thermal efficiencies. The working fluid has a higher volume flow rate at the turbine inlet, also somewhat has a higher thermal efficiency as shown in

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

**Figure 8(b)** shows the thermal efficiencies as a function of the VR (turbine outlet/inlet volumetric flow rate ratios, V1/V0) at the optimum turbine parameters for the maximum power output. Compared to the o2 cycle, the expansion ratio in the turbine decreases for the o3 cycle with the same condensation pressure which also leads to a decrease in the VR. Both the expansion ratios of R134a and R1234ze are much lower in the subcritical ORC because of the higher condensation pressures which also results in lower VRs. The VRs of R134a and R1234ze are increased in the supercritical ORC. Among the selected working fluids, R123, R1233zd, and R245fa have high thermal efficiencies with high-turbine-inlet volume flow rates and turbine outlet/inlet volumetric flow rate ratios. R1233zd is considered to be the most suitable working fluid for the ORC which recovers waste heat from the boiler flue

gas with taking the safety and environment protection into account.

**5.1 ORCs driven by exhaust steam from a back-pressure turbine**

cations such as oil refineries, petrochemical plants, and cogeneration [33].

structure and a lower cost for a back-pressure steam turbine.

the heat consumption of power generation [34].

**60**

facility or outside needs [33]; thus, a lower expansion ratio results in a lower enthalpy drop and small power output in a back-pressure steam turbine. Only fewer turbine stages are used due to the lower enthalpy drop which leads to a simple

Process steam/heat demand and electricity demand change independently according to the season or facility production demand [33]. An imbalance between process steam/heat and electricity demands is one of the most common problems in actual operation. The steam/heat demand is the primary requirement with electricity demand a secondary consideration to solve the imbalance demands because the steam and heat cannot economically and conveniently be transported over a long distance. Adjusting the main-steam flow rate of a back-pressure turbine is the major solution to meet the steam/heat demand. Thus, the back-pressure turbine power (electricity) output varies with the steam/heat demand. However, the thermal efficiency of the back-pressure steam turbine is significantly decreased with decreasing electricity output at partial loads due to the lower enthalpy drop. For example, the turbine isentropic efficiency (relative internal efficiency) can be decreased from 85 to 60% for lower steam flow rates which leads to an increase in

The imbalance between steam/heat and electricity demands reduces the economic performance of back-pressure steam turbines. Part of this risk will be

Back-pressure steam turbines supply not only electricity but also the steam and heat for processes. The exhaust steam of the back-pressure steam turbine is directly used to supply heat or steam to the facilities without condensation. Ideally, the exhaust steam and the electricity from the back-pressure steam turbine are supplied to the same users [32]. The back-pressure steam turbine does not have any cold source loss (heat loss released directly to the environment). Therefore, the backpressure steam turbines are efficient and have been widely used in industrial appli-

The exhaust steam pressure generally is set to be the demand pressure from the

**5. Heat recovery from turbine exhaust steam**

**Figure 8(a)**.

*Operating parameters for a B25–8.83/1.5 back-pressure turbine [34].*

#### **Figure 9.**

*Schematic of an ORC for recovering the excess exhaust steam heat from a back-pressure steam turbine [34].*

preheat the makeup water, and the rest most of exhaust steam transported to the heat consumer. The heat consumer steam demand accounts for 70% of the turbine main-steam flow. The main-steam flow rate of the back-pressure turbine decreases with decreasing steam demand, which results in a serious decrease in the isentropic efficiency (relative internal efficiency) as shown in **Table 7**.

A conceptual system had been proposed to improve the thermodynamic performance of the back-pressure turbine as shown in **Figure 9** [34]. The back-pressure turbine is always operating at the turbine heat acceptance (THA) condition. When the heat consumer steam demand decreases, the excess exhaust steam is provided to an ORC for organic working fluid heating. The steam is condensed to the saturated water and then returns back to the mixer after throttling to preheat makeup water. Compared to the conditions with lower main steam flow rates, the back-pressure turbine generates additional electricity, *ΔW*\_ T, due to the high isentropic efficiency and a constant steam flow rate. The amount of electricity produced by the ORC system using the excess exhaust steam is defined as *W*\_ O. Thus, the total additional electricity (increase in electricity) from the system can be expressed as

$$
\Delta\dot{W}\_{\text{Sys}} = \Delta\dot{W}\_{\text{T}} + \dot{W}\_{\text{O}} \tag{10}
$$

with R123 and R1233zd are still used here. Subcritical ORC using R123 with the lowest evaporation pressure shows a better matching of the steam and the working fluid temperature profiles as shown in **Figure 10(a)** and the highest thermal efficiency among the working fluids. The working fluid with a lower critical temperature has a higher evaporation pressure but shows a lower thermal efficiency. The

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

*T-Q diagrams for working fluid evaporating at (a) the optimal pressure and (b) both the optimized pressure*

In this conceptual system, the back-pressure turbine outputs 6.3, 12.6, and 17.7 MWh additional electricity for the steam demands of 140.9, 108.6, and 81 t/h, respectively. The additional electricity generation for ORCs using the excess exhaust steam is listed in **Table 9**. As the steam demand decreases to 140.9 t/h, the ORCs generate about 6 MWh electricity with the excess steam flow rate of 44.8 t/h. The ORCs can generate 16.7 MWh electricity when the steam demand is decreased to 81 t/h. Thus, the total additional electricity output of the system, *ΔW*\_ Sys, can

At last, this section gives a further theoretical discussion on the cycle and working fluid choices for the heat source type with isothermal condensation. Both the temperature and the pressure at the ORC turbine inlet were optimized simultaneously to obtain the maximum power output. **Figure 10(b)** shows the T-Q diagrams for the working fluids evaporating at the optimal condition. The exhaust steam from the back-pressure turbine has a much large superheat degree, but the sensible heat is still much lower than the latent heat. The pinch point occurs at the dew point (saturated vapor state) of the steam as shown in **Figure 10(b)**. The working fluid may be heated to a temperature higher than the steam condensation temperature by the sensible heat. In this case study, the optimal temperatures of the working fluids at the evaporator outlet range from 205–213°C. The working fluid with a higher critical temperature has a higher optimal evaporator outlet temperature and then results in a higher thermal efficiency. The subcritical ORC using the

**R134a R1234ze R236ea R245fa R1233zd R123**

140.9 5292 5409 5741 5835 5835 5986 108.6 10,536 10,769 11,429 11,617 11,617 11,918 81.0 14,744 15,069 15,992 16,256 16,256 16,677

thermal efficiencies for R1233zd and R245fa are very close.

**Steam demand/ t<sup>h</sup><sup>1</sup>** *<sup>W</sup>*\_ **O/kW<sup>h</sup>**

*Electricity generated by ORCs using excess exhaust steam at various steam demands.*

reach 12–34 MWh for partial steam demands.

**Figure 10.**

**Table 9.**

**63**

*and temperature.*

The additional electricity can be provided to the facilities or the grid. The steam/ heat and electricity demands can then both be regulated by adjusting the mainsteam flow rate. The additional electricity from the back-pressure turbine, *ΔW*\_ T, mainly depends on the turbine operating parameters and the steam demand. The electricity produced from the ORC system, *W*\_ O, depends on the heat input from the excess exhaust steam and the ORC thermal efficiency [34].

The flow rates of the excess exhaust steam for various steam demands are determined according to the mass and the energy balances in the back-pressure turbine system. The steam in the ORC evaporator is condensed to the saturated liquid. The pressure drop between the back-pressure turbine outlet and the ORC evaporator outlet is set to be 5%. The superheat degree of the exhaust steam reaches 120°C, but the sensible heat is much lower than the latent heat. Thus, the matching characteristics between the organic working fluid evaporation and the steam condensation temperatures affect the ORC thermal efficiency.

The ORC turbine inlet temperature was assumed to be 195°C. The ORC turbine inlet pressures of the selected working fluids were optimized to maximize the net cycle power output with the boundary conditions listed in **Table 1**. The parameters and thermal efficiencies of ORCs are listed in **Table 8**. **Figure 10(a)** shows the *T-Q* diagrams of the evaporation process for the selected working fluids. Supercritical cycles are adopted for R134a, R1234ze, R236ea, and R245fa due to their lower critical temperatures and higher heat source temperature, while subcritical ORCs


#### **Table 8.**

*Comparison of ORCs using exhaust steam from a back-pressure turbine.*

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

**Figure 10.**

preheat the makeup water, and the rest most of exhaust steam transported to the heat consumer. The heat consumer steam demand accounts for 70% of the turbine main-steam flow. The main-steam flow rate of the back-pressure turbine decreases with decreasing steam demand, which results in a serious decrease in the isentropic

A conceptual system had been proposed to improve the thermodynamic performance of the back-pressure turbine as shown in **Figure 9** [34]. The back-pressure turbine is always operating at the turbine heat acceptance (THA) condition. When the heat consumer steam demand decreases, the excess exhaust steam is provided to an ORC for organic working fluid heating. The steam is condensed to the saturated water and then returns back to the mixer after throttling to preheat makeup water. Compared to the conditions with lower main steam flow rates, the back-pressure turbine generates additional electricity, *ΔW*\_ T, due to the high isentropic efficiency and a constant steam flow rate. The amount of electricity produced by the ORC system using the excess exhaust steam is defined as *W*\_ O. Thus, the total additional

The additional electricity can be provided to the facilities or the grid. The steam/ heat and electricity demands can then both be regulated by adjusting the mainsteam flow rate. The additional electricity from the back-pressure turbine, *ΔW*\_ T, mainly depends on the turbine operating parameters and the steam demand. The electricity produced from the ORC system, *W*\_ O, depends on the heat input from the

The flow rates of the excess exhaust steam for various steam demands are determined according to the mass and the energy balances in the back-pressure turbine system. The steam in the ORC evaporator is condensed to the saturated liquid. The pressure drop between the back-pressure turbine outlet and the ORC evaporator outlet is set to be 5%. The superheat degree of the exhaust steam reaches 120°C, but the sensible heat is much lower than the latent heat. Thus, the matching characteristics between the organic working fluid evaporation and the steam con-

The ORC turbine inlet temperature was assumed to be 195°C. The ORC turbine inlet pressures of the selected working fluids were optimized to maximize the net cycle power output with the boundary conditions listed in **Table 1**. The parameters and thermal efficiencies of ORCs are listed in **Table 8**. **Figure 10(a)** shows the *T-Q* diagrams of the evaporation process for the selected working fluids. Supercritical cycles are adopted for R134a, R1234ze, R236ea, and R245fa due to their lower critical temperatures and higher heat source temperature, while subcritical ORCs

**Working fluid Cycle** *T***T,in/°C** *p***T,in/MPa** *p***T,out/MPa Thermal efficiency/%**

R134a s2 195 7.15 0.77 18.99 R1234ze s2 195 6.09 0.58 19.40 R236ea s2 195 4.37 0.24 20.59 R245fa s2 195 3.87 0.18 20.93 R1233zd o3 195 3.23 0.15 20.93 R123 o3 195 2.80 0.11 21.48

<sup>Δ</sup>*W*\_ Sys <sup>¼</sup> <sup>Δ</sup>*W*\_ <sup>T</sup> <sup>þ</sup> *<sup>W</sup>*\_ <sup>O</sup> (10)

efficiency (relative internal efficiency) as shown in **Table 7**.

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

electricity (increase in electricity) from the system can be expressed as

excess exhaust steam and the ORC thermal efficiency [34].

densation temperatures affect the ORC thermal efficiency.

*Comparison of ORCs using exhaust steam from a back-pressure turbine.*

**Table 8.**

**62**

*T-Q diagrams for working fluid evaporating at (a) the optimal pressure and (b) both the optimized pressure and temperature.*

with R123 and R1233zd are still used here. Subcritical ORC using R123 with the lowest evaporation pressure shows a better matching of the steam and the working fluid temperature profiles as shown in **Figure 10(a)** and the highest thermal efficiency among the working fluids. The working fluid with a lower critical temperature has a higher evaporation pressure but shows a lower thermal efficiency. The thermal efficiencies for R1233zd and R245fa are very close.

In this conceptual system, the back-pressure turbine outputs 6.3, 12.6, and 17.7 MWh additional electricity for the steam demands of 140.9, 108.6, and 81 t/h, respectively. The additional electricity generation for ORCs using the excess exhaust steam is listed in **Table 9**. As the steam demand decreases to 140.9 t/h, the ORCs generate about 6 MWh electricity with the excess steam flow rate of 44.8 t/h. The ORCs can generate 16.7 MWh electricity when the steam demand is decreased to 81 t/h. Thus, the total additional electricity output of the system, *ΔW*\_ Sys, can reach 12–34 MWh for partial steam demands.

At last, this section gives a further theoretical discussion on the cycle and working fluid choices for the heat source type with isothermal condensation. Both the temperature and the pressure at the ORC turbine inlet were optimized simultaneously to obtain the maximum power output. **Figure 10(b)** shows the T-Q diagrams for the working fluids evaporating at the optimal condition. The exhaust steam from the back-pressure turbine has a much large superheat degree, but the sensible heat is still much lower than the latent heat. The pinch point occurs at the dew point (saturated vapor state) of the steam as shown in **Figure 10(b)**. The working fluid may be heated to a temperature higher than the steam condensation temperature by the sensible heat. In this case study, the optimal temperatures of the working fluids at the evaporator outlet range from 205–213°C. The working fluid with a higher critical temperature has a higher optimal evaporator outlet temperature and then results in a higher thermal efficiency. The subcritical ORC using the


**Table 9.**

*Electricity generated by ORCs using excess exhaust steam at various steam demands.*

working fluid with a higher critical temperature is more suitable for the kind of isothermal heat source due to the better matches of the temperature profiles of the working fluid and the heat source fluid. The temperature at the evaporator outlet (turbine inlet) strongly affects the cycle thermal efficiency. The cycle efficiency can be relatively increased by about 4% for a 10°C increase in the turbine inlet temperature. Compared to the ORCs with the fixed turbine inlet temperature of 195°C, the cycle efficiencies are relatively improved by 3.7–5.7% for the optimal turbine inlet temperatures and pressures as shown in **Figure 11**. Among the six working fluids, the cycle efficiency with R123 is the highest with the highest turbine inlet temperature, followed by R1233zd. The ORC thermal efficiency can exceed 20% for a steam heat source at a temperature of 200°C. However, the thermal stabilities of the working fluids should be considered primarily [16, 18, 35].

circulating seawater flow rate to the exhaust steam flow rate can be reduced to 30–

Organic Rankine cycles have been studied to utilize the ocean thermal energy with the temperature difference between warm seawater and cold seawater as low as 20°C [36–39]. Therefore, the temperature difference between the turbine exhaust steam and the cooling water can be used to drive an ORC for power generation as shown in **Figure 12**. The exhaust steam from the low-pressure turbine (LPT) is condensed by the organic working fluid in the condenser. The working fluid is heated to a saturated vapor and then enters into the ORC turbine to expand. After expansion, the organic working fluid is condensed by the cold seawater in the ORC condenser. The cycle thermal efficiency of the ORC is lower due to the lower temperature difference between the heat source and sink. However, the ORC can also provide substantial power because the discharged heat from the exhaust steam

The parameters and boundary conditions for the ORC using exhaust steam as the heat source and cold seawater as the heat sink are listed in **Table 10**. A subcritical 600 MW condensing steam turbine is taken as a case study. The exhaust pressure of the steam turbine is 3.5 kPa with a flow rate of 300 kg/s. The subcritical ORC with saturated vapor at the turbine inlet is only studied here due to the exhaust steam isothermal condensation. The organic working fluid evaporation temperature is only related to the steam condensation temperature with the pinch point temperature difference, *ΔT*P; thus, the organic working fluid condensation temperature is the key parameter which affects the thermodynamic performance as shown in **Figure 13**. The ORC thermal efficiency increases as the condensation temperature decreases. A lower condensation temperature needs a lower cooling water (seawater) temperature rise with a higher flow rate; however, the cooling water pumps consume a significant fraction of the ORC output power [27, 36–39]. The cooling water (seawater) flow rate was assumed here to be 45,000 kg/s with a temperature

**Figure 14** shows the turbine power generation and the net power output of the ORC system using the exhaust steam as heat source and the cold seawater as heat sink. The ORC turbine output can exceed 23.5 MW in this case study. The turbine

*Schematic of an ORC using the exhaust steam from a condensing turbine and the cold seawater.*

50 because of the higher cooling water temperature rise.

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

in a power plant is much huge.

rise about 3.58°C.

**Figure 12.**

**65**

#### **5.2 ORCs driven by exhaust steam form a condensing steam turbine**

The majority of heat loss in a steam Rankine cycle is the exhaust steam latent heat released to the coolant at the condenser. The cold source loss generally accounts for 50–60% of the total heat input for an extraction condensing turbine. The steam condensation temperature should be close to the coolant (environment) temperature to reduce the turbine exhaust pressure and then increase the thermal efficiency (absolute internal efficiency). The exhaust pressures generally range 4.9– 11.8 kPa for condensing turbines using closed cooling systems with circulating water as the coolant; thus, the exhaust steam is condensed at temperatures of 32.5– 49°C with the circulating water flow rate up to 50–100 times the exhaust steam flow rate. The condensation temperature (turbine exhaust pressure) is mainly affected by the environment temperature and the operating conditions. The seawater is generally used as the coolant for offshore power plants. The seawater temperature can reach as low as 5 °C in the cold season or the seawater from a depth up to 1000 m; thus, the turbine exhaust pressures range 3.5–5 kPa with the condensation temperatures between 26.7 and 32.9°C. Considering the issues including wetness fraction, volume flow rate, and blade length of the last stage of the low-pressure turbine (LPT), the exhaust pressure generally does not need to be less than 3.5 kPa even through the cooling water temperature is much lower. The ratio of the

#### *Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

circulating seawater flow rate to the exhaust steam flow rate can be reduced to 30– 50 because of the higher cooling water temperature rise.

Organic Rankine cycles have been studied to utilize the ocean thermal energy with the temperature difference between warm seawater and cold seawater as low as 20°C [36–39]. Therefore, the temperature difference between the turbine exhaust steam and the cooling water can be used to drive an ORC for power generation as shown in **Figure 12**. The exhaust steam from the low-pressure turbine (LPT) is condensed by the organic working fluid in the condenser. The working fluid is heated to a saturated vapor and then enters into the ORC turbine to expand. After expansion, the organic working fluid is condensed by the cold seawater in the ORC condenser. The cycle thermal efficiency of the ORC is lower due to the lower temperature difference between the heat source and sink. However, the ORC can also provide substantial power because the discharged heat from the exhaust steam in a power plant is much huge.

The parameters and boundary conditions for the ORC using exhaust steam as the heat source and cold seawater as the heat sink are listed in **Table 10**. A subcritical 600 MW condensing steam turbine is taken as a case study. The exhaust pressure of the steam turbine is 3.5 kPa with a flow rate of 300 kg/s. The subcritical ORC with saturated vapor at the turbine inlet is only studied here due to the exhaust steam isothermal condensation. The organic working fluid evaporation temperature is only related to the steam condensation temperature with the pinch point temperature difference, *ΔT*P; thus, the organic working fluid condensation temperature is the key parameter which affects the thermodynamic performance as shown in **Figure 13**. The ORC thermal efficiency increases as the condensation temperature decreases. A lower condensation temperature needs a lower cooling water (seawater) temperature rise with a higher flow rate; however, the cooling water pumps consume a significant fraction of the ORC output power [27, 36–39]. The cooling water (seawater) flow rate was assumed here to be 45,000 kg/s with a temperature rise about 3.58°C.

**Figure 14** shows the turbine power generation and the net power output of the ORC system using the exhaust steam as heat source and the cold seawater as heat sink. The ORC turbine output can exceed 23.5 MW in this case study. The turbine

**Figure 12.** *Schematic of an ORC using the exhaust steam from a condensing turbine and the cold seawater.*

working fluid with a higher critical temperature is more suitable for the kind of isothermal heat source due to the better matches of the temperature profiles of the working fluid and the heat source fluid. The temperature at the evaporator outlet (turbine inlet) strongly affects the cycle thermal efficiency. The cycle efficiency can be relatively increased by about 4% for a 10°C increase in the turbine inlet temperature. Compared to the ORCs with the fixed turbine inlet temperature of 195°C, the cycle efficiencies are relatively improved by 3.7–5.7% for the optimal turbine inlet temperatures and pressures as shown in **Figure 11**. Among the six working fluids, the cycle efficiency with R123 is the highest with the highest turbine inlet temperature, followed by R1233zd. The ORC thermal efficiency can exceed 20% for a steam heat source at a temperature of 200°C. However, the thermal stabilities of the

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

working fluids should be considered primarily [16, 18, 35].

**Figure 11.**

**64**

**5.2 ORCs driven by exhaust steam form a condensing steam turbine**

*Cycle thermal efficiencies for the ORCs using exhaust steam from back-pressure turbine.*

heat released to the coolant at the condenser. The cold source loss generally accounts for 50–60% of the total heat input for an extraction condensing turbine. The steam condensation temperature should be close to the coolant (environment) temperature to reduce the turbine exhaust pressure and then increase the thermal efficiency (absolute internal efficiency). The exhaust pressures generally range 4.9– 11.8 kPa for condensing turbines using closed cooling systems with circulating water as the coolant; thus, the exhaust steam is condensed at temperatures of 32.5– 49°C with the circulating water flow rate up to 50–100 times the exhaust steam flow rate. The condensation temperature (turbine exhaust pressure) is mainly affected by the environment temperature and the operating conditions. The seawater is generally used as the coolant for offshore power plants. The seawater temperature can reach as low as 5 °C in the cold season or the seawater from a depth up to 1000 m; thus, the turbine exhaust pressures range 3.5–5 kPa with the condensation temperatures between 26.7 and 32.9°C. Considering the issues including wetness fraction, volume flow rate, and blade length of the last stage of the low-pressure turbine (LPT), the exhaust pressure generally does not need to be less than 3.5 kPa even through the cooling water temperature is much lower. The ratio of the

The majority of heat loss in a steam Rankine cycle is the exhaust steam latent

#### *Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*


#### **Table 10.**

*Operating parameters and boundary conditions for the ORC using exhaust steam as heat source and seawater as sink.*

> power output. Thus, the system net power output (subtracting the power consumed by the working fluid pump and the cooling water pumps from the turbine power output) with R134a is the lowest due to the largest parasitic power consumption, while the ORC with R123 outputs the maximum net power among the six working fluids. When the cold seawater is directly used to condense the exhaust steam from LPTs, the cooling water flow rate for the steam turbine is about 8,500 kg/s, but the power consumed by the circulating water pumps is still more than 1 MW. When the exhaust steam is used as an ORC heat source with the cold seawater as the heat sink, the system can provide net power output up to 18 MW equal to 3% increase in the

*Turbine power output and net cycle power output for ORCs with exhaust steam as heat source and cold*

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

In coal-fired power plants, there are a large number of medium-low-grade waste

heat which is originally discharge to the environment. Conceptual systems and thermodynamic performance for the ORCs recovering waste heat from coal-fired power plants are studied in this chapter. The optimal cycle layouts and proper working fluids for the ORCs are discussed. The conclusions are summarized as

1.The amount of the drained water that directly discharges has been toward fewer and smaller in modern power plants. The original discharge or leaked drain can be collected to drive a small-capacity ORC. A supercritical ORC matches well with the kind of heat source with a large temperature drop.

2.The flue gas flow rate is much high for a large-capacity boiler. The amount of waste heat in exhaust flue gas can be reached 10 MW even through the flue gas is only cooled by 10°C. The subcritical ORC with an IHE (a recuperator or regenerator) using a working fluid with a higher critical temperature is more

gross power of the coal-fired power unit.

**6. Conclusions**

**Figure 14.**

*seawater as heat sink.*

follow.

**67**

#### **Figure 13.**

*Temperature-entropy diagram of an ORC using R1234ze as the working fluid driven by exhaust steam from a condensing turbine with cold seawater as heat sink.*

using R134a produces the maximum power among the selected six working fluids. The cooling water pumps consume a significant fraction of the turbine output power. The pump power consumption accounts for about 23.3% of the turbine power generation when the cooling water pump head is 10 m. The pressure increase in the working fluid pump is much higher for the working fluid with a lower critical temperature and leads to a higher power consumption. The parasitic power consumed by the working fluid feed pump with R134a accounts for 3.6% of the turbine *Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

#### **Figure 14.**

*Turbine power output and net cycle power output for ORCs with exhaust steam as heat source and cold seawater as heat sink.*

power output. Thus, the system net power output (subtracting the power consumed by the working fluid pump and the cooling water pumps from the turbine power output) with R134a is the lowest due to the largest parasitic power consumption, while the ORC with R123 outputs the maximum net power among the six working fluids. When the cold seawater is directly used to condense the exhaust steam from LPTs, the cooling water flow rate for the steam turbine is about 8,500 kg/s, but the power consumed by the circulating water pumps is still more than 1 MW. When the exhaust steam is used as an ORC heat source with the cold seawater as the heat sink, the system can provide net power output up to 18 MW equal to 3% increase in the gross power of the coal-fired power unit.

#### **6. Conclusions**

In coal-fired power plants, there are a large number of medium-low-grade waste heat which is originally discharge to the environment. Conceptual systems and thermodynamic performance for the ORCs recovering waste heat from coal-fired power plants are studied in this chapter. The optimal cycle layouts and proper working fluids for the ORCs are discussed. The conclusions are summarized as follow.


using R134a produces the maximum power among the selected six working fluids. The cooling water pumps consume a significant fraction of the turbine output power. The pump power consumption accounts for about 23.3% of the turbine power generation when the cooling water pump head is 10 m. The pressure increase in the working fluid pump is much higher for the working fluid with a lower critical temperature and leads to a higher power consumption. The parasitic power consumed by the working fluid feed pump with R134a accounts for 3.6% of the turbine

*Temperature-entropy diagram of an ORC using R1234ze as the working fluid driven by exhaust steam from a*

**Parameters Values Unit** Turbine isentropic efficiency [36] 80 % Working fluid pump efficiency [36] 75 % Cooling water pump efficiency 80 % Generator efficiency 98 % Turbine mechanical efficiency 98 % Evaporator pinch point temperature difference [37] 2 °C Condenser pinch point temperature difference 2 °C Exhaust steam temperature 26.67 °C Exhaust steam dryness fraction 92 % Exhaust steam flow rate 300 kg/s Cold seawater temperature 5 °C Cooling water (seawater) flow rate 45,000 kg/s Cooling water pump head 10 m Seawater specific heat capacity [37] 4.025 kJ/(kg °C)

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

*Operating parameters and boundary conditions for the ORC using exhaust steam as heat source and seawater*

**Table 10.**

**Figure 13.**

**66**

*condensing turbine with cold seawater as heat sink.*

*as sink.*

suitable for the kind of heat source with a higher limited outlet temperature. The gross power output of the coal-fired power unit can be relatively increased by 0.4% by the ORC using the waste heat from boiler exhaust flue gas.

**References**

June 2019]

[1] Available from: https://www.iea.org/ geco/data/ [Accessed: 08 June 2019]

*DOI: http://dx.doi.org/10.5772/intechopen.89354*

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

[9] Liu M, Zhang XW, Han XQ, Li G, Yan JJ. Using pre-drying technology to improve the exergetic efficiency of bioenergy utilization process with combustion: A case study of a power plant. Applied Thermal Engineering. 2017;**127**:1416-1426. DOI: 10.1016/j.

[10] Han XQ, Liu M, Wu KL, Chen WX, Xiao F, Yan JJ. Exergy analysis of the flue gas pre-dried lignite-fired power system based on the boiler with open pulverizing system. Energy. 2016;**106**:285-300. DOI:

applthermaleng.2017.08.156

10.1016/j.energy.2016.03.047

egypro.2017.03.518

e20020089

07.496

[11] Han XQ, Karellas S, Liu M, Braimakis K, Chen WX, Yan JJ, et al. Integration of organic Rankine cycle with lignite flue gas pre-drying for waste heat and water recovery from dryer exhaust gas: Thermodynamic and economic analysis. Energy Procedia. 2017;**105**:1614-1621. DOI: 10.1016/j.

[12] Huang SW, Li CZ, Tan TY, Fu P, Wang LG, Yang YP. Comparative evaluation of integrated waste heat utilization systems for coal-fired power plants based on in-depth boiler-turbine integration and organic Rankine cycle. Entropy. 2018;**20**:89. DOI: 10.3390/

[13] Xi XM, Zhou YY, Guo C, Yang LJ, Du XZ. Characteristics of organic Rankine cycles with zeotropic mixture for heat recovery of exhaust gas of boiler. Energy Procedia. 2015;**75**: 1093-1101. DOI: 10.1016/j.egypro.2015.

Ziółkowski P, Mikielewicz J. Utilisation of waste heat from the power plant by use of the ORC aided with bleed steam and extra source of heat. Energy. 2016;

[14] Mikielewicz D, Wajs J,

**97**:11-19. DOI: 10.1016/j. energy.2015.12.106

[2] Available from: https://github.com/c itation-style-language/styles/blob/maste r/vancouver-brackets.csl [Accessed: 08

[3] Che DF. Boilers - Theory, Design and

[4] Liu Q, Duan YY, Wan XC. Power generation system using continuous blowdown waste heat from drum boiler driving an organic Rankine cycle. Proceedings of the CSEE. 2013;**33**(35): 1-7. DOI: 10.13334/j.0258-8013.

[5] Yang YP, Xu C, Xu G, Han Y, Fang YX, Zhang DK. A new conceptual cold-end design of boilers for coal-fired power plants with waste heat recovery. Energy Conversion and Management. 2015;**89**:137-146. DOI: 10.1016/j.

[6] Liu M, Zhang XW, Ma YG, Yan JJ. Thermo-economic analyses on a new conceptual system of waste heat

recovery integrated with an S-CO2 cycle for coal-fired power plants. Energy Conversion and Management. 2018;**161**: 243-253. DOI: 10.1016/j.enconman.2018.

[7] Wang CJ, He BS, Yan LB, Pei XH, Chen SN. Thermodynamic analysis of a low-pressure economizer based waste heat recovery system for a coal-fired power plant. Energy. 2014;**65**:80-90. DOI: 10.1016/j.energy.2013.11.084

[8] Łukowicz H, Kochaniewicz A. Analysis of the use of waste heat obtained from coal-fired units in organic Rankine cycles and for brown coal drying. Energy. 2012;**45**:203-212. DOI: 10.1016/j.energy.2012.03.035

Operation. Xi'an: Xi'an Jiaotong

University Press; 2008

pcsee.2013.35.006

enconman.2014.09.065

01.049

**69**

3.More than 50% of the heat input to the turbine is released in the condenser. When the cooling water temperature is much lower, the use of ORCs to recover the waste heat from the turbine exhaust steam becomes attractive. Although the cooling water pumps consume a significant fraction of the ORC output power, the net efficiencies of ORCs still reach 2.6%. The power output increase potential for the coal-fired power unit can exceed 2–3% when all of the turbine exhaust steam is used to drive ORCs with lower temperature seawater as the heat sink.

### **Acknowledgements**

This work was supported by the National Natural Science Foundation of China (Grant Nos. 51506223 and 51736005) and the National Basic Research Program of China (973 program) (Grant No. 2015CB251502).

### **Conflict of interest**

The authors declare no competing financial interest.

### **Author details**

Qiang Liu College of Mechanical and Transportation Engineering, China University of Petroleum, Beijing, P.R. China

\*Address all correspondence to: qliu@cup.edu.cn

© 2019 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

#### **References**

suitable for the kind of heat source with a higher limited outlet temperature. The gross power output of the coal-fired power unit can be relatively increased

by 0.4% by the ORC using the waste heat from boiler exhaust flue gas.

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

seawater as the heat sink.

China (973 program) (Grant No. 2015CB251502).

The authors declare no competing financial interest.

College of Mechanical and Transportation Engineering, China University of

© 2019 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

**Acknowledgements**

**Conflict of interest**

**Author details**

Petroleum, Beijing, P.R. China

\*Address all correspondence to: qliu@cup.edu.cn

provided the original work is properly cited.

Qiang Liu

**68**

3.More than 50% of the heat input to the turbine is released in the condenser. When the cooling water temperature is much lower, the use of ORCs to recover the waste heat from the turbine exhaust steam becomes attractive. Although the cooling water pumps consume a significant fraction of the ORC output power, the net efficiencies of ORCs still reach 2.6%. The power output increase potential for the coal-fired power unit can exceed 2–3% when all of the turbine exhaust steam is used to drive ORCs with lower temperature

This work was supported by the National Natural Science Foundation of China (Grant Nos. 51506223 and 51736005) and the National Basic Research Program of

[1] Available from: https://www.iea.org/ geco/data/ [Accessed: 08 June 2019]

[2] Available from: https://github.com/c itation-style-language/styles/blob/maste r/vancouver-brackets.csl [Accessed: 08 June 2019]

[3] Che DF. Boilers - Theory, Design and Operation. Xi'an: Xi'an Jiaotong University Press; 2008

[4] Liu Q, Duan YY, Wan XC. Power generation system using continuous blowdown waste heat from drum boiler driving an organic Rankine cycle. Proceedings of the CSEE. 2013;**33**(35): 1-7. DOI: 10.13334/j.0258-8013. pcsee.2013.35.006

[5] Yang YP, Xu C, Xu G, Han Y, Fang YX, Zhang DK. A new conceptual cold-end design of boilers for coal-fired power plants with waste heat recovery. Energy Conversion and Management. 2015;**89**:137-146. DOI: 10.1016/j. enconman.2014.09.065

[6] Liu M, Zhang XW, Ma YG, Yan JJ. Thermo-economic analyses on a new conceptual system of waste heat recovery integrated with an S-CO2 cycle for coal-fired power plants. Energy Conversion and Management. 2018;**161**: 243-253. DOI: 10.1016/j.enconman.2018. 01.049

[7] Wang CJ, He BS, Yan LB, Pei XH, Chen SN. Thermodynamic analysis of a low-pressure economizer based waste heat recovery system for a coal-fired power plant. Energy. 2014;**65**:80-90. DOI: 10.1016/j.energy.2013.11.084

[8] Łukowicz H, Kochaniewicz A. Analysis of the use of waste heat obtained from coal-fired units in organic Rankine cycles and for brown coal drying. Energy. 2012;**45**:203-212. DOI: 10.1016/j.energy.2012.03.035

[9] Liu M, Zhang XW, Han XQ, Li G, Yan JJ. Using pre-drying technology to improve the exergetic efficiency of bioenergy utilization process with combustion: A case study of a power plant. Applied Thermal Engineering. 2017;**127**:1416-1426. DOI: 10.1016/j. applthermaleng.2017.08.156

[10] Han XQ, Liu M, Wu KL, Chen WX, Xiao F, Yan JJ. Exergy analysis of the flue gas pre-dried lignite-fired power system based on the boiler with open pulverizing system. Energy. 2016;**106**:285-300. DOI: 10.1016/j.energy.2016.03.047

[11] Han XQ, Karellas S, Liu M, Braimakis K, Chen WX, Yan JJ, et al. Integration of organic Rankine cycle with lignite flue gas pre-drying for waste heat and water recovery from dryer exhaust gas: Thermodynamic and economic analysis. Energy Procedia. 2017;**105**:1614-1621. DOI: 10.1016/j. egypro.2017.03.518

[12] Huang SW, Li CZ, Tan TY, Fu P, Wang LG, Yang YP. Comparative evaluation of integrated waste heat utilization systems for coal-fired power plants based on in-depth boiler-turbine integration and organic Rankine cycle. Entropy. 2018;**20**:89. DOI: 10.3390/ e20020089

[13] Xi XM, Zhou YY, Guo C, Yang LJ, Du XZ. Characteristics of organic Rankine cycles with zeotropic mixture for heat recovery of exhaust gas of boiler. Energy Procedia. 2015;**75**: 1093-1101. DOI: 10.1016/j.egypro.2015. 07.496

[14] Mikielewicz D, Wajs J, Ziółkowski P, Mikielewicz J. Utilisation of waste heat from the power plant by use of the ORC aided with bleed steam and extra source of heat. Energy. 2016; **97**:11-19. DOI: 10.1016/j. energy.2015.12.106

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[16] Macchi E. Theoretical basis of the organic Rankine cycle. In: Macchi E, Astolfi M, editors. Organic Rankine Cycle (ORC) Power Systems. London: Woodhead Publishing; 2017. pp. 3-24. DOI: 10.1016/B978-0-08-100510- 1.00001-6

[17] Saleh B, Koglbauer G, Wendland M, Fischer J. Working fluids for lowtemperature organic Rankine cycles. Energy. 2007;**32**:1210-1221. DOI: 10.1016/j.energy.2006.07.001

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[19] Astolfi M. Technical options for organic Rankine cycle systems. In: Macchi E, Astolfi M, editors. Organic Rankine Cycle (ORC) Power Systems. London: Woodhead Publishing; 2017. pp. 67-89. DOI: 10.1016/B978-0- 08-100510-1.00003-X

[20] Madhawa Hettiarachchi HD, Golubovic M, Worek WM, Ikegami Y. Optimum design criteria for an organic Rankine cycle using low-temperature geothermal heat sources. Energy. 2007; **32**:1698-1706. DOI: 10.1016/j. energy.2007.01.005

[21] Maraver D, Royo J, Lemort V, Quoilin S. Systematic optimization of subcritical and transcritical organic Rankine cycles (ORCs) constrained by technical parameters in multiple applications. Applied Energy. 2014;**117**: 11-29. DOI: 10.1016/j.apenergy.2013. 11.076

[22] Quoilin S, Broek MVD, Declaye S, Dewallef P, Lemort V. Technoeconomic survey of organic Rankine cycle (ORC) systems. Renewable and Sustainable Energy Review. 2013;**22**: 168-186. DOI: 10.1016/j.rser.2013. 01.028

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08-100314-5.00002-6

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[35] Dai XY, Shi L, Qian WZ. Review of the working fluid thermal stability for organic Rankine cycles. Journal of Thermal Science. 2019;**28**:597-607. DOI:

[36] Yamada N, Hoshi A, Ikegami Y. Performance simulation of solarboosted ocean thermal energy conversion plant. Renewable Energy. 2009;**34**:1752-1758. DOI: 10.1016/j.

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Group; 2009

10.078

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*DOI: http://dx.doi.org/10.5772/intechopen.89354*

*Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles*

of a closed-cycle ocean thermal energy conversion system with solar thermal

[38] Khan N, Kalair A, Abas N, Haider A. Review of ocean tidal, wave and thermal energy technologies. Renewable and Sustainable Energy Review. 2017;**72**: 590-604. DOI: 10.1016/j.rser.2017.

[39] Bernardoni C, Binotti M, Giostri A. Techno-economic analysis of closed OTEC cycles for power generation. Renewable Energy. 2019;**132**:1018-1033. DOI: 10.1016/j.renene.2018.08.007

preheating and superheating. Renewable Energy. 2014;**72**:154-163. DOI: 10.1016/j.renene.2014.07.001

01.079

[23] Zhai HX, An QS, Shi L, Lemort V, Quoilin S. Categorization and analysis of heat sources for organic Rankine cycle systems. Renewable and Sustainable Energy Review. 2016;**64**:790-805. DOI: 10.1016/j.rser.2016.06.076

[24] Sun J, Liu Q, Duan YY. Effects of evaporator pinch point temperature difference on thermo-economic performance of geothermal organic Rankine cycle systems. Geothermics. 2018;**75**:249-258. DOI: 10.1016/j. geothermics.2018.06.001

[25] Pan LS, Shi WX. Investigation on the pinch point position in heat exchangers. Journal of Thermal Science. 2016;**25**:258-265. DOI: 10.1007/ s11630-016-0858-7

[26] Liu Q, Duan YY, Yang Z. Performance analyses of geothermal organic Rankine cycles with selected hydrocarbon working fluids. Energy. 2013;**63**:123-132. DOI: 10.1016/j. energy.2013.10.035

[27] Liu Q, Shen AJ, Duan YY. Parametric optimization and performance analyses of geothermal organic Rankine cycles using R600a/ R601a mixtures as working fluids. Applied Energy. 2015;**148**:410-420. DOI: 10.1016/j.apenergy.2015.03.093

[28] Xu JL, Yu C. Critical temperature criterion for selection of working fluids for subcritical pressure organic Rankine cycles. Energy. 2014;**74**:719-733. DOI: 10.1016/j.energy.2014.07.038

[29] Lemmon EW, Huber ML, McLinden MO. NIST Standard Reference Database 23: Reference Fluid *Waste Heat Recovery from Fossil-Fired Power Plants by Organic Rankine Cycles DOI: http://dx.doi.org/10.5772/intechopen.89354*

Thermodynamic and Transport Properties-REFPROP, Version 9.1. Gaithersburg: National Institute of Standards and Technology, Standard Reference Data Program; 2013

[15] Angelino G, Invernizzi C,

0957650991537446

1.00001-6

Molteni G. The potential role of organic bottoming Rankine cycles in steam power stations. Proceedings of the Institution of Mechanical Engineers, Part A: Journal of Power and Energy. 1999;**213**:75-81. DOI: 10.1243/

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

[22] Quoilin S, Broek MVD, Declaye S, Dewallef P, Lemort V. Technoeconomic survey of organic Rankine cycle (ORC) systems. Renewable and Sustainable Energy Review. 2013;**22**: 168-186. DOI: 10.1016/j.rser.2013.

[23] Zhai HX, An QS, Shi L, Lemort V, Quoilin S. Categorization and analysis of heat sources for organic Rankine cycle systems. Renewable and Sustainable Energy Review. 2016;**64**:790-805. DOI:

[24] Sun J, Liu Q, Duan YY. Effects of evaporator pinch point temperature difference on thermo-economic performance of geothermal organic Rankine cycle systems. Geothermics. 2018;**75**:249-258. DOI: 10.1016/j. geothermics.2018.06.001

[25] Pan LS, Shi WX. Investigation on the pinch point position in heat

2016;**25**:258-265. DOI: 10.1007/

[26] Liu Q, Duan YY, Yang Z. Performance analyses of geothermal organic Rankine cycles with selected hydrocarbon working fluids. Energy. 2013;**63**:123-132. DOI: 10.1016/j.

[27] Liu Q, Shen AJ, Duan YY. Parametric optimization and

performance analyses of geothermal organic Rankine cycles using R600a/ R601a mixtures as working fluids. Applied Energy. 2015;**148**:410-420. DOI: 10.1016/j.apenergy.2015.03.093

[28] Xu JL, Yu C. Critical temperature criterion for selection of working fluids for subcritical pressure organic Rankine cycles. Energy. 2014;**74**:719-733. DOI:

Reference Database 23: Reference Fluid

10.1016/j.energy.2014.07.038

[29] Lemmon EW, Huber ML, McLinden MO. NIST Standard

s11630-016-0858-7

energy.2013.10.035

exchangers. Journal of Thermal Science.

10.1016/j.rser.2016.06.076

01.028

[16] Macchi E. Theoretical basis of the organic Rankine cycle. In: Macchi E, Astolfi M, editors. Organic Rankine Cycle (ORC) Power Systems. London: Woodhead Publishing; 2017. pp. 3-24. DOI: 10.1016/B978-0-08-100510-

[17] Saleh B, Koglbauer G, Wendland M, Fischer J. Working fluids for lowtemperature organic Rankine cycles. Energy. 2007;**32**:1210-1221. DOI: 10.1016/j.energy.2006.07.001

[18] Invernizzi CM. Closed Power Cycles - Thermodynamic Fundamentals and Applications. London: Springer-Verlag; 2013. DOI: 10.1007/978-1-4471-5140-1

[19] Astolfi M. Technical options for organic Rankine cycle systems. In: Macchi E, Astolfi M, editors. Organic Rankine Cycle (ORC) Power Systems. London: Woodhead Publishing; 2017. pp. 67-89. DOI: 10.1016/B978-0-

[20] Madhawa Hettiarachchi HD, Golubovic M, Worek WM, Ikegami Y. Optimum design criteria for an organic Rankine cycle using low-temperature geothermal heat sources. Energy. 2007;

**32**:1698-1706. DOI: 10.1016/j.

[21] Maraver D, Royo J, Lemort V, Quoilin S. Systematic optimization of subcritical and transcritical organic Rankine cycles (ORCs) constrained by technical parameters in multiple applications. Applied Energy. 2014;**117**: 11-29. DOI: 10.1016/j.apenergy.2013.

energy.2007.01.005

11.076

**70**

08-100510-1.00003-X

[30] Rayaprolu K. Boilers for Power and Process. Boca Raton: Taylor & Francis Group; 2009

[31] Liu Q, Shang LL, Duan YY. Performance analyses of a hybrid geothermal–fossil power generation system using low-enthalpy geothermal resources. Applied Energy. 2016;**162**: 149-162. DOI: 10.1016/j.apenergy.2015. 10.078

[32] Breeze P. Combined Heat and Power. London: Elsevier; 2018. DOI: B978-0-12-812908-1.00003-1

[33] Ohji A, Haraguchi M. Steam turbine cycles and cycle design optimization: The Rankine cycle, thermal power cycles, and IGCC power plants. In: Tanuma T, editor. Advances in Steam Turbines for Modern Power Plants. London: Woodhead Publishing; 2017. pp. 11-40. DOI: 10.1016/B978-0- 08-100314-5.00002-6

[34] Liu Q, Duan YY. Cogeneration system coupled with a back pressure steam turbine unit and an organic Rankine cycle (ORC). Proceedings of the CSEE. 2013;**33**(23):29-35. DOI: 10.13334/j.0258-8013.pcsee.2013.23.014

[35] Dai XY, Shi L, Qian WZ. Review of the working fluid thermal stability for organic Rankine cycles. Journal of Thermal Science. 2019;**28**:597-607. DOI: 10.1007/s11630-019-1119-3

[36] Yamada N, Hoshi A, Ikegami Y. Performance simulation of solarboosted ocean thermal energy conversion plant. Renewable Energy. 2009;**34**:1752-1758. DOI: 10.1016/j. renene.2008.12.028

[37] Aydin H, Lee HS, Kim HJ, Shin SK, Park K. Off-design performance analysis of a closed-cycle ocean thermal energy conversion system with solar thermal preheating and superheating. Renewable Energy. 2014;**72**:154-163. DOI: 10.1016/j.renene.2014.07.001

[38] Khan N, Kalair A, Abas N, Haider A. Review of ocean tidal, wave and thermal energy technologies. Renewable and Sustainable Energy Review. 2017;**72**: 590-604. DOI: 10.1016/j.rser.2017. 01.079

[39] Bernardoni C, Binotti M, Giostri A. Techno-economic analysis of closed OTEC cycles for power generation. Renewable Energy. 2019;**132**:1018-1033. DOI: 10.1016/j.renene.2018.08.007

**73**

**Chapter 4**

**Abstract**

**1. Introduction**

The Development and Application

Power conversion systems based on organic Rankine cycles have been identified as a potential technology especially in converting low-grade waste heat into electricity as well as in small-scale biomass, solar, or geothermal power plants. The theoretical analysis can guide the ORC design, but cannot predict accurately the system performance. Actually, the operation characteristics of every component have a vital effect on the system performance. This chapter presents the detailed operation characteristic of a small-scale ORC. The effects of the operation parameters, the mixture working fluid and the operation strategy on system overall performance are addressed. It can be concluded that improving the system overall performance should give priority to increase the pressure drop. Whether the mixtures exhibit better thermodynamic performance than the pure working fluids depend on the operation parameters and mass fraction of mixtures. The mixture working fluids obtain a higher expander shaft power but a relatively higher BWR. The expander rotating speed for standalone operation strategy keeps rising from 2320 to 2983 rpm,

whereas that of grid connect operation strategy keeps constant of 3600 rpm.

Energy is an indispensable resource for human progress and social development, improving energy efficiency has become a global research hot spot. Meanwhile, waste heat resource utilization problem has received widespread attention. If those waste heats can be effectively utilized, it will not only provide important technical support for energy conservation, emission reduction and environmental protection, but also generate certain economic benefits. Organic Rankine cycle (ORC) was adapted as a new technology to utilize waste heat [1–3]. The principle of ORC is similar to that of Rankine cycle. The main difference is that the low-boiling organic fluid can be used to replace the water of the Rankine cycle, which can significantly reduce the final discharge temperature, thus achieving the purpose of waste heat recovery [4–8]. Recently, many scholars have conducted in-depth research on ORC system concerning on the working fluids selection, ORC performance optimization and component development. However, the theoretical analysis can guide the ORC design, but cannot predict accurately the system performance. Therefore, some

**Keywords:** organic Rankine cycle (ORC), operation characteristic,

researchers devoted main effort on ORC experimental studies.

mixture working fluids, system generating efficiency

of a Small-Scale Organic Rankine

Cycle for Waste Heat Recovery

*Tzu-Chen Hung and Yong-qiang Feng*

#### **Chapter 4**

## The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat Recovery

*Tzu-Chen Hung and Yong-qiang Feng*

#### **Abstract**

Power conversion systems based on organic Rankine cycles have been identified as a potential technology especially in converting low-grade waste heat into electricity as well as in small-scale biomass, solar, or geothermal power plants. The theoretical analysis can guide the ORC design, but cannot predict accurately the system performance. Actually, the operation characteristics of every component have a vital effect on the system performance. This chapter presents the detailed operation characteristic of a small-scale ORC. The effects of the operation parameters, the mixture working fluid and the operation strategy on system overall performance are addressed. It can be concluded that improving the system overall performance should give priority to increase the pressure drop. Whether the mixtures exhibit better thermodynamic performance than the pure working fluids depend on the operation parameters and mass fraction of mixtures. The mixture working fluids obtain a higher expander shaft power but a relatively higher BWR. The expander rotating speed for standalone operation strategy keeps rising from 2320 to 2983 rpm, whereas that of grid connect operation strategy keeps constant of 3600 rpm.

**Keywords:** organic Rankine cycle (ORC), operation characteristic, mixture working fluids, system generating efficiency

#### **1. Introduction**

Energy is an indispensable resource for human progress and social development, improving energy efficiency has become a global research hot spot. Meanwhile, waste heat resource utilization problem has received widespread attention. If those waste heats can be effectively utilized, it will not only provide important technical support for energy conservation, emission reduction and environmental protection, but also generate certain economic benefits. Organic Rankine cycle (ORC) was adapted as a new technology to utilize waste heat [1–3]. The principle of ORC is similar to that of Rankine cycle. The main difference is that the low-boiling organic fluid can be used to replace the water of the Rankine cycle, which can significantly reduce the final discharge temperature, thus achieving the purpose of waste heat recovery [4–8].

Recently, many scholars have conducted in-depth research on ORC system concerning on the working fluids selection, ORC performance optimization and component development. However, the theoretical analysis can guide the ORC design, but cannot predict accurately the system performance. Therefore, some researchers devoted main effort on ORC experimental studies.

Mathias et al. [9] compared the operation characteristics using piston pump and gear pump, expressing that the piston pump outperformed than the gear pump. Carraro et al. [10] integrated a multi-diaphragm positive displacement pump into a 4 kW ORC experimental prototype, and found that the pump global efficiency was about 45–48%. Xu et al. [11] used R123 to study the matching degree of the piston pump and the expander. They stated that the low pump frequency was applicable to all expander torques, while the high pump frequency was only applicable to low torques. Zhao et al. [12] studied the diaphragm pump performance under various conditions using four different working fluids. The experimental results showed that a higher volume flow rate and pressure difference led to a higher pump isentropic efficiency. Lei et al. [13] tested a roto-jet pump at different rotating speeds using R123. They illustrated that the pump efficiency was in range of 11–23%, and an increase in pump rotating speed or a decrease in mass flow rate can cause the decrease in pump efficiency. Zeleny et al. [14] proposed a modified gear pump used in micro ORC and discussed the contribution to pump different losses. They stated that as the pressure increases, the effect of mechanical losses decreased, while the volume loss was reversed. Bianchi et al. [15] applied a sliding vane pump on the ORC using R236fa. When the mass flow rate increased from 52 to 119 g/s, the pump's mechanical power increased by 289 W. Wu et al. [16] used multistage gas-liquid booster pump to improve the performance of ORC system, revealing that the maximum conversion efficiency from high-pressure air to water was 0.72. Yang et al. [17] compared experimental characteristics of three pumps, demonstrating that the maximum actual pump efficiency of multistage centrifugal pump was 58.76%. Landelle et al. [18] discussed the operation characteristic of the reciprocating pump, reporting that the reciprocating pump had eminent volumetric efficiency and the ORC efficiency decreased as cavitation margin increased. Meng et al. [19] analyzed the multistage centrifugal pump on the ORC, expressing that the maximum overall pump efficiency was 65.7%. Bianchi et al. [20] changed the heat source temperature and feed pump speed to explore the performance of micro ORC. They found that the achieved pump efficiency was ranging from 10 to 20% and the net efficiency of the system was 2.2%. Yang et al. [21] studied the key parameters of the hydraulic diaphragm metering pump. The tested efficiency and BWR were 88.7% and 0.93, respectively. Bianchi et al. [22] used the CFD model to design and analyze a sliding vane pump. The pump performance was found to be optimal at 1250 rpm and 9.7 bar. Sun et al. [23] compared three different pumps from the viewpoint of the actual cycle and the ideal cycle. The experimental results showed that the combination of a centrifugal pump and a scroll expander can maximize the isentropic efficiency of the expander. Xi et al. [24] analyzed the influence of the plunger stroke of the working fluid pump on the system performance by orthogonal analysis. The results showed that the plunger stroke had a great influence on pump power consumption and working fluid flow. Aleksandra Borsukiewicz-Gozdur [25] found that a higher pumping power required a higher cycle pressure.

Compared with the pure working fluids, the main advantage of mixtures as ORC working fluids stems from their non-isothermal phase transitions during vaporization and condensation, and hence effectively match the temperature change of heat source and cooling water. Therefore, great attention has been drawn to the mixture working fluids. Dong et al. [26] compared the thermal efficiency of hightemperature ORC system between mixture and pure working fluids, and found that the range of options for working fluids was widened by the mixture working fluids. Garg et al. [27] investigated the thermodynamic analysis using isopentane, R245fa and their mixtures, reporting that 0.7isopentane/0.3R245fa was the preferred candidate working fluid. Lecompte et al. [28] discussed the exergy efficiency of ORC

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*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

[34] compared the system behaviors using R245fa and R245fa/R601a.

In this chapter, the effect of mass flow rate, pressure drop, degree of superheating and condenser temperature on thermal efficiency and system generating efficiency are examined [35–36]. Several experimental investigations using pure working fluids (R123 and R245fa) of small-scale ORC test rig have been performed. However, few of them tried to fulfill the experimental comparison between pure and mixture working fluids. And therefore, two pure working fluids (R245fa, R123) and two mixtures working fluids (0.67R245fa/0.33R123 and 0.33R245fa/0.67R123) are tested and compared [37]. The system behaviors at two different operation

A 3 kW ORC experimental prototype is adopted, as shown in **Figure 1**. It includes three loops: heating loop, ORC loop and cooling loop. The photo of experi-

An electric heater using conduction oil is used as the simulated heat source, within four electrical heating rods having the capacity of 80 kW. An axial pump adjusted the mass flow rates of conductive oil, ensuring the heat source temperatures ranging from 110 to 140°C. Meanwhile, the evaporator heat transfer rate can be changed by adjusting the input power of electric heater, which is controlled by

The ORC loop is made up of four major components: a plunger pump, an evaporator, a scroll-type expander and a condenser. It should be noted that R123 is used in this study because of its better thermal efficiency and environmental performance. The pump supplies the working fluid to the evaporator where the working fluid is heated and vaporized by the conductive oil. The high pressure vapor flows into the expander and its enthalpy is converted into work. The low pressure vapor exits the expander and is led to the condenser where it is liquefied by water. The liquid is available at the condenser outlet, and then it is pumped back to the evaporator and a

mental layout and the main facilities of ORC system are plotted in **Figure 2**.

system using mixture working fluids, stating that the mixture presented 7.1–14.2% higher second law efficiency than the pure working fluids. Zhao et al. [29] analyzed the effect of composition shift of mixture working fluids on system performance, demonstrating that the composition shift significantly influenced the performance of ORC system. Liu et al. [30] investigated the effect of condensation temperature glide using mixture on ORC performance, reporting that when the increase of cooling water temperature is greater than the condensation temperature glide, an optimal working fluid mole fraction can be obtained to maximum thermodynamic performance. Furthermore, great attention has been drawn to the experimental comparison between mixture and pure working fluids. Pu et al. [31] compared the system performance using R245fa and HFE7100, showing that R245fa obtained the maximum net power output of 1.98 kW, which is 0.95 kW higher than that of HEF7100. Molés et al. [32] proposed using HCFO-1233zd-E to replace HFC-245fa, and found that the net electrical efficiency was in rage of 5–9.7%. Jung et al. [33] studied the dynamic behavior of a kW ORC test rig using R245fa/365mfc. Li et al.

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

strategies are addressed [38].

the four electrical heating rods.

**2.1 Heating loop**

**2.2 ORC loop**

new cycle begins.

**2. Experimental setup of ORC system**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

system using mixture working fluids, stating that the mixture presented 7.1–14.2% higher second law efficiency than the pure working fluids. Zhao et al. [29] analyzed the effect of composition shift of mixture working fluids on system performance, demonstrating that the composition shift significantly influenced the performance of ORC system. Liu et al. [30] investigated the effect of condensation temperature glide using mixture on ORC performance, reporting that when the increase of cooling water temperature is greater than the condensation temperature glide, an optimal working fluid mole fraction can be obtained to maximum thermodynamic performance. Furthermore, great attention has been drawn to the experimental comparison between mixture and pure working fluids. Pu et al. [31] compared the system performance using R245fa and HFE7100, showing that R245fa obtained the maximum net power output of 1.98 kW, which is 0.95 kW higher than that of HEF7100. Molés et al. [32] proposed using HCFO-1233zd-E to replace HFC-245fa, and found that the net electrical efficiency was in rage of 5–9.7%. Jung et al. [33] studied the dynamic behavior of a kW ORC test rig using R245fa/365mfc. Li et al. [34] compared the system behaviors using R245fa and R245fa/R601a.

In this chapter, the effect of mass flow rate, pressure drop, degree of superheating and condenser temperature on thermal efficiency and system generating efficiency are examined [35–36]. Several experimental investigations using pure working fluids (R123 and R245fa) of small-scale ORC test rig have been performed. However, few of them tried to fulfill the experimental comparison between pure and mixture working fluids. And therefore, two pure working fluids (R245fa, R123) and two mixtures working fluids (0.67R245fa/0.33R123 and 0.33R245fa/0.67R123) are tested and compared [37]. The system behaviors at two different operation strategies are addressed [38].

#### **2. Experimental setup of ORC system**

A 3 kW ORC experimental prototype is adopted, as shown in **Figure 1**. It includes three loops: heating loop, ORC loop and cooling loop. The photo of experimental layout and the main facilities of ORC system are plotted in **Figure 2**.

#### **2.1 Heating loop**

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

Mathias et al. [9] compared the operation characteristics using piston pump and gear pump, expressing that the piston pump outperformed than the gear pump. Carraro et al. [10] integrated a multi-diaphragm positive displacement pump into a 4 kW ORC experimental prototype, and found that the pump global efficiency was about 45–48%. Xu et al. [11] used R123 to study the matching degree of the piston pump and the expander. They stated that the low pump frequency was applicable to all expander torques, while the high pump frequency was only applicable to low torques. Zhao et al. [12] studied the diaphragm pump performance under various conditions using four different working fluids. The experimental results showed that a higher volume flow rate and pressure difference led to a higher pump isentropic efficiency. Lei et al. [13] tested a roto-jet pump at different rotating speeds using R123. They illustrated that the pump efficiency was in range of 11–23%, and an increase in pump rotating speed or a decrease in mass flow rate can cause the decrease in pump efficiency. Zeleny et al. [14] proposed a modified gear pump used in micro ORC and discussed the contribution to pump different losses. They stated that as the pressure increases, the effect of mechanical losses decreased, while the volume loss was reversed. Bianchi et al. [15] applied a sliding vane pump on the ORC using R236fa. When the mass flow rate increased from 52 to 119 g/s, the pump's mechanical power increased by 289 W. Wu et al. [16] used multistage gas-liquid booster pump to improve the performance of ORC system, revealing that the maximum conversion efficiency from high-pressure air to water was 0.72. Yang et al. [17] compared experimental characteristics of three pumps, demonstrating that the maximum actual pump efficiency of multistage centrifugal pump was 58.76%. Landelle et al. [18] discussed the operation characteristic of the reciprocating pump, reporting that the reciprocating pump had eminent volumetric efficiency and the ORC efficiency decreased as cavitation margin increased. Meng et al. [19] analyzed the multistage centrifugal pump on the ORC, expressing that the maximum overall pump efficiency was 65.7%. Bianchi et al. [20] changed the heat source temperature and feed pump speed to explore the performance of micro ORC. They found that the achieved pump efficiency was ranging from 10 to 20% and the net efficiency of the system was 2.2%. Yang et al. [21] studied the key parameters of the hydraulic diaphragm metering pump. The tested efficiency and BWR were 88.7% and 0.93, respectively. Bianchi et al. [22] used the CFD model to design and analyze a sliding vane pump. The pump performance was found to be optimal at 1250 rpm and 9.7 bar. Sun et al. [23] compared three different pumps from the viewpoint of the actual cycle and the ideal cycle. The experimental results showed that the combination of a centrifugal pump and a scroll expander can maximize the isentropic efficiency of the expander. Xi et al. [24] analyzed the influence of the plunger stroke of the working fluid pump on the system performance by orthogonal analysis. The results showed that the plunger stroke had a great influence on pump power consumption and working fluid flow. Aleksandra Borsukiewicz-Gozdur [25] found that a higher pumping power required a higher

Compared with the pure working fluids, the main advantage of mixtures as ORC working fluids stems from their non-isothermal phase transitions during vaporization and condensation, and hence effectively match the temperature change of heat source and cooling water. Therefore, great attention has been drawn to the mixture working fluids. Dong et al. [26] compared the thermal efficiency of hightemperature ORC system between mixture and pure working fluids, and found that the range of options for working fluids was widened by the mixture working fluids. Garg et al. [27] investigated the thermodynamic analysis using isopentane, R245fa and their mixtures, reporting that 0.7isopentane/0.3R245fa was the preferred candidate working fluid. Lecompte et al. [28] discussed the exergy efficiency of ORC

**74**

cycle pressure.

An electric heater using conduction oil is used as the simulated heat source, within four electrical heating rods having the capacity of 80 kW. An axial pump adjusted the mass flow rates of conductive oil, ensuring the heat source temperatures ranging from 110 to 140°C. Meanwhile, the evaporator heat transfer rate can be changed by adjusting the input power of electric heater, which is controlled by the four electrical heating rods.

#### **2.2 ORC loop**

The ORC loop is made up of four major components: a plunger pump, an evaporator, a scroll-type expander and a condenser. It should be noted that R123 is used in this study because of its better thermal efficiency and environmental performance. The pump supplies the working fluid to the evaporator where the working fluid is heated and vaporized by the conductive oil. The high pressure vapor flows into the expander and its enthalpy is converted into work. The low pressure vapor exits the expander and is led to the condenser where it is liquefied by water. The liquid is available at the condenser outlet, and then it is pumped back to the evaporator and a new cycle begins.

#### *Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

**Figure 1.** *Schematic diagram of an ORC system.*

**Figure 2.** *The photos of (a) experimental setup, (b) the expander generator set, and (c) the electrical load.*

**77**

**Figure 3.**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

The plunger pump is controlled by a frequency converter, with the achievable maximum delivered pressure head and flow rate are 50 bars and 18 L/min, respectively. The pump power consumption can be obtained by measuring the voltage and current, while the pump shaft power can be calculated by the thermodynamic

The evaporator and condenser are plate heat exchangers with the same heat

A scroll-type expander, which was modified from commercial oil-free scroll

expander shaft power is transferred to the three-phase permanent-magnet generator by pulley and belt. The alternating current produced by permanent-magnet generator is converted to direct current using three bridge rectifier. The frequency of direct current is adjusted by the electrical resistance and capacitance to meet the

The cooling tower is installed at roof and extras the heat from condenser to air environment. Therefore, the cooling inlet water is fluctuated by the environmental temperature. The needle value is used to adjust the mass flow rate of cooling water.

It should be noted that the thermodynamic properties of pure and mixture working fluids are obtained based on NIST Refprop [39]. Based on the measured temperatures and pressures, the corresponding enthalpies and entropies for every state can be obtained. **Figure 3** shows *T-s* plot of the thermodynamic cycle. For the expander, the vapor working fluid (state 1) enters the expander to generate power output, and then exits (state 2). The evaporator heat transfer rate (*Q*eva), which is

type air compressor with a built-in volume ratio of three, is employed. The

. Asbestos board and insulating foam are equipped around

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

parameters at the inlet and outlet of pump.

the evaporator and condenser to avoid the heat loss.

transfer area of 4.157 m2

grid frequency requirements.

**3. Thermodynamic analysis method**

*T-s plot of the ORC cycle, as well as the evaporator and condenser model.*

heated by the conductive oil, is

**2.3 Cooling loop**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

The plunger pump is controlled by a frequency converter, with the achievable maximum delivered pressure head and flow rate are 50 bars and 18 L/min, respectively. The pump power consumption can be obtained by measuring the voltage and current, while the pump shaft power can be calculated by the thermodynamic parameters at the inlet and outlet of pump.

The evaporator and condenser are plate heat exchangers with the same heat transfer area of 4.157 m2 . Asbestos board and insulating foam are equipped around the evaporator and condenser to avoid the heat loss.

A scroll-type expander, which was modified from commercial oil-free scroll type air compressor with a built-in volume ratio of three, is employed. The expander shaft power is transferred to the three-phase permanent-magnet generator by pulley and belt. The alternating current produced by permanent-magnet generator is converted to direct current using three bridge rectifier. The frequency of direct current is adjusted by the electrical resistance and capacitance to meet the grid frequency requirements.

#### **2.3 Cooling loop**

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

**76**

**Figure 2.**

**Figure 1.**

*Schematic diagram of an ORC system.*

*The photos of (a) experimental setup, (b) the expander generator set, and (c) the electrical load.*

The cooling tower is installed at roof and extras the heat from condenser to air environment. Therefore, the cooling inlet water is fluctuated by the environmental temperature. The needle value is used to adjust the mass flow rate of cooling water.

#### **3. Thermodynamic analysis method**

It should be noted that the thermodynamic properties of pure and mixture working fluids are obtained based on NIST Refprop [39]. Based on the measured temperatures and pressures, the corresponding enthalpies and entropies for every state can be obtained. **Figure 3** shows *T-s* plot of the thermodynamic cycle. For the expander, the vapor working fluid (state 1) enters the expander to generate power output, and then exits (state 2). The evaporator heat transfer rate (*Q*eva), which is heated by the conductive oil, is

**Figure 3.** *T-s plot of the ORC cycle, as well as the evaporator and condenser model.*

$$Q\_{\text{eva}} = m \left( h\_1 - h\_6 \right) \tag{1}$$

The expander power output is calculated according to the thermodynamic state at expander inlet and outlet, while the expander shaft power is measured by torque meter, which can be obtained by torque and rotating speed. Meanwhile, the ideal isentropic expansion process in expander is from 1 to 2 s, while the real one is from 1 to 2. The expander power output (*W*p,exp) and shaft power (*W*sh,exp) can be expressed as follows:

$$\mathcal{W}\_{\text{p,exp}} = m \left( h\_1 - h\_2 \right) \tag{2}$$

$$
\hbar \mathcal{W}\_{\text{sh,exp}} = \frac{2\pi}{60} \mathcal{M}\_{\text{exp}} n\_{\text{exp}} \tag{3}
$$

where *h*1 and *h*2 are the inlet and outlet enthalpy of expander, which is determined by the measured pressure and temperature; *M*exp is the torque, and *n*exp is the rotating speed, which is measured by tachometer.

The electrical power produced by generator can be calculated by measuring the current and voltage. The pump shaft power can be calculated by the measured pressure and temperature, while the power consumption is measured by pump frequency converter. To better understand the portion of the electricity output for driving the pump, back work ratio (BWR) is proposed to represent the ratio between pump consumption and electricity output, which can be expressed as: BWR = \_

$$\text{BWR} = \frac{\mathcal{W}\_{\text{ele,pump}}}{\mathcal{W}\_{\text{ele,exp}}} \tag{4}$$

The thermal efficiency and system generating efficiency can be expressed as:

$$
\eta\_{\text{th,cal}} = \frac{\text{Mass generating efficiency can be expressed as:}}{h\_1 - h\_6} \tag{5}
$$

$$
\eta\_{\text{th,est}} = \frac{(h\_1 - h\_2) - (h\_6 - h\_5)}{h\_1 - h\_6} \tag{6}
$$

$$
\eta\_{\text{th,est}} = \frac{\mathcal{W}\_{\text{sh}} - \mathcal{W}\_{\text{p}}}{Q\_{\text{gva}}} \tag{6}
$$

$$
\eta\_{\rm th,test} = \frac{\mathcal{W}\_{\rm sh} - \mathcal{W}\_{\rm p}}{Q\_{\rm sva}} \tag{6}
$$

$$
\ln \eta\_{\rm ele} = \frac{\mathcal{W}\_{\rm ele,exp} - \mathcal{W}\_{\rm ele,p}}{Q\_{\rm sys}} \tag{7}
$$

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*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

of 0.124–0.222 kg/s and heat source temperature is in range of 110–140°C, the variation of heat source temperatures and mass flow rates with time are plotted in **Figure 4**. Obviously, the mass flow rate and heat source temperature have a rela-

Taking the mass flow rate of 0.124 kg/s and heat source temperature of 110°C as an example, the detailed steady-state operation characteristic for 20 min steady operation is shown in **Figure 5**. It can be seen that the pump outlet pressure (**Figure 5b**), expander inlet pressure (**Figure 5c**), and mass flow rate of working fluids (**Figure 5f**) have a relativelṢy strong variation, indicating that a control

The working fluid mass flow rate is adjusted by the pump frequencies. **Figure6** plots the variation of thermal efficiency with mass flow rate at different heat source temperatures. It can be seen that the thermal efficiencies for different heat source temperatures have a similar behavior of a decreasing trend with mass flow rate, which may be attributed to the increasing heat input. Meanwhile, the thermal efficiency keeps rising with the heat source temperature, which may be attributed to the increasing power output. It also can be found that when the mass flow rate is 0.124 kg/s and the heat source tem-

**Figure 7** shows the variation of system generating efficiency with mass flow rate at different heat source temperatures. As the mass flow rate increases, the system generating efficiency keeps decreasing for a heat source temperature smaller than 120°C, whereas represents a parabolic trend for a heat source temperature higher than 130°C. Meanwhile, the system generating efficiency keeps rising with the heat source temperature. A highest system generating efficiency of 3.25% is appeared for a mass flow rate of 0.198 kg/s and a heat source temperature of 140°C. The system

The pressure drop denotes the expander inlet pressure minus the pump inlet pressure. The variation of electrical power, thermal efficiency and system generating

perature is 140°C, the thermal efficiency owns a maximum value of 5.14%.

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

tively stable variation.

**Figure 4.**

**4.2 Effect of mass flow rate**

strategy for the pump and expander is necessary.

*Variation of heat source temperatures and mass flow rates with time.*

generating efficiency is in range of 0.94–3.25%.

**4.3 Effect of pressure drop**

#### **4. Effects of different operation parameters**

The operation parameters have a significant influence on system performance. Based on the 3 kW ORC experimental prototype, the steady-state operation test is discussed at first, and then the four operation parameters on system behavior are examined.

#### **4.1 Steady-state operation test**

The heat input is used to control the heat source temperature, while the working fluid pump frequency is adopted to adjust the working fluid mass flow rate. The experimental data is collected at every 5 s. The test is recorded when the heat source temperatures are within a small fluctuation below ±0.5°C. One point is the average value of a 20 min steady operation. When the mass flow rate is in range

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

#### **Figure 4.**

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

The expander power output is calculated according to the thermodynamic state at expander inlet and outlet, while the expander shaft power is measured by torque meter, which can be obtained by torque and rotating speed. Meanwhile, the ideal isentropic expansion process in expander is from 1 to 2 s, while the real one is from 1 to 2. The expander power output (*W*p,exp) and shaft power (*W*sh,exp) can be expressed

*<sup>W</sup>*sh,exp = \_2*<sup>π</sup>*

rotating speed, which is measured by tachometer.

**4. Effects of different operation parameters**

where *h*1 and *h*2 are the inlet and outlet enthalpy of expander, which is determined by the measured pressure and temperature; *M*exp is the torque, and *n*exp is the

The electrical power produced by generator can be calculated by measuring the current and voltage. The pump shaft power can be calculated by the measured pressure and temperature, while the power consumption is measured by pump frequency converter. To better understand the portion of the electricity output for driving the pump, back work ratio (BWR) is proposed to represent the ratio between pump consumption and electricity output, which can be

BWR = \_

The thermal efficiency and system generating efficiency can be expressed as: <sup>η</sup>th,cal = (*h*<sup>1</sup> <sup>−</sup>*h*2) − (*h*<sup>6</sup> <sup>−</sup>*h*5) \_\_\_\_\_\_\_\_\_\_\_\_\_\_\_ *<sup>h</sup>*<sup>1</sup> <sup>−</sup>*h*<sup>6</sup>

ηth,test = \_

<sup>η</sup>ele = *W*ele,exp <sup>−</sup>*W*ele,p \_\_\_\_\_\_\_\_\_\_\_\_ *<sup>Q</sup>*eva

The operation parameters have a significant influence on system performance. Based on the 3 kW ORC experimental prototype, the steady-state operation test is discussed at first, and then the four operation parameters on system behavior are

The heat input is used to control the heat source temperature, while the working fluid pump frequency is adopted to adjust the working fluid mass flow rate. The experimental data is collected at every 5 s. The test is recorded when the heat source temperatures are within a small fluctuation below ±0.5°C. One point is the average value of a 20 min steady operation. When the mass flow rate is in range

*W*ele,pump *W*ele,exp

*W*sh − *W*<sup>p</sup> *Q*eva

*Q*eva = *m*(*h*<sup>1</sup> − *h*6) (1)

*W*p,exp = *m*(*h*<sup>1</sup> − *h*2) (2)

60 *M*exp *n*exp (3)

(4)

(5)

(6)

(7)

**78**

examined.

**4.1 Steady-state operation test**

as follows:

expressed as:

*Variation of heat source temperatures and mass flow rates with time.*

of 0.124–0.222 kg/s and heat source temperature is in range of 110–140°C, the variation of heat source temperatures and mass flow rates with time are plotted in **Figure 4**. Obviously, the mass flow rate and heat source temperature have a relatively stable variation.

Taking the mass flow rate of 0.124 kg/s and heat source temperature of 110°C as an example, the detailed steady-state operation characteristic for 20 min steady operation is shown in **Figure 5**. It can be seen that the pump outlet pressure (**Figure 5b**), expander inlet pressure (**Figure 5c**), and mass flow rate of working fluids (**Figure 5f**) have a relativelṢy strong variation, indicating that a control strategy for the pump and expander is necessary.

#### **4.2 Effect of mass flow rate**

The working fluid mass flow rate is adjusted by the pump frequencies. **Figure6** plots the variation of thermal efficiency with mass flow rate at different heat source temperatures. It can be seen that the thermal efficiencies for different heat source temperatures have a similar behavior of a decreasing trend with mass flow rate, which may be attributed to the increasing heat input. Meanwhile, the thermal efficiency keeps rising with the heat source temperature, which may be attributed to the increasing power output. It also can be found that when the mass flow rate is 0.124 kg/s and the heat source temperature is 140°C, the thermal efficiency owns a maximum value of 5.14%.

**Figure 7** shows the variation of system generating efficiency with mass flow rate at different heat source temperatures. As the mass flow rate increases, the system generating efficiency keeps decreasing for a heat source temperature smaller than 120°C, whereas represents a parabolic trend for a heat source temperature higher than 130°C. Meanwhile, the system generating efficiency keeps rising with the heat source temperature. A highest system generating efficiency of 3.25% is appeared for a mass flow rate of 0.198 kg/s and a heat source temperature of 140°C. The system generating efficiency is in range of 0.94–3.25%.

#### **4.3 Effect of pressure drop**

The pressure drop denotes the expander inlet pressure minus the pump inlet pressure. The variation of electrical power, thermal efficiency and system generating

#### **Figure 5.**

*Variation of system parameters with time. (a) pump inlet pressure and temperature, (b) pump outlet pressure and temperature, (c) expander inlet pressure and temperature, (d) expander outlet pressure and temperature, (e) cooling water inlet and outlet temperature and heat source inlet and outlet temperature, (f) mass flow rates of working fluid and cooling water.*

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**Figure 8.**

**Figure 7.**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

*Variation of system generating efficiency with mass flow rate at different heat source temperature.*

efficiency are demonstrated in **Figure 8**. As the pressure drop increases, the system generating efficiency and thermal efficiency own a similar behavior of an increase first and then a slightly decrease, whereas the electrical power goes up. The increasing heat input causes the parabolic trend for thermal efficiency and system generating efficiency. A higher pressure drop denotes a higher investment cost of heat exchanger for pressure-bearing requirement. And therefore, the optimum pressure drop is 8.16 bar, with the corresponding thermal efficiency of 5.89% and system generating efficiency of 3.86%. It also can be found that the maximum thermal efficiency of 5.92% and maximum system generating efficiency of 3.93% are obtained

*Variation of electrical power, thermal efficiency and system generating efficiency with pressure drop.*

with the corresponding pressure drop of 9.33 bar.

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

**Figure 6.** *Variation of thermal efficiency with mass flow rate at different heat source temperatures.*

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

**Figure 7.** *Variation of system generating efficiency with mass flow rate at different heat source temperature.*

**Figure 8.** *Variation of electrical power, thermal efficiency and system generating efficiency with pressure drop.*

efficiency are demonstrated in **Figure 8**. As the pressure drop increases, the system generating efficiency and thermal efficiency own a similar behavior of an increase first and then a slightly decrease, whereas the electrical power goes up. The increasing heat input causes the parabolic trend for thermal efficiency and system generating efficiency. A higher pressure drop denotes a higher investment cost of heat exchanger for pressure-bearing requirement. And therefore, the optimum pressure drop is 8.16 bar, with the corresponding thermal efficiency of 5.89% and system generating efficiency of 3.86%. It also can be found that the maximum thermal efficiency of 5.92% and maximum system generating efficiency of 3.93% are obtained with the corresponding pressure drop of 9.33 bar.

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

**80**

**Figure 6.**

**Figure 5.**

*of working fluid and cooling water.*

*Variation of thermal efficiency with mass flow rate at different heat source temperatures.*

*Variation of system parameters with time. (a) pump inlet pressure and temperature, (b) pump outlet pressure and temperature, (c) expander inlet pressure and temperature, (d) expander outlet pressure and temperature, (e) cooling water inlet and outlet temperature and heat source inlet and outlet temperature, (f) mass flow rates* 

#### **4.4 Effect of degree of superheating**

The variation of electrical power, thermal efficiency and system generating efficiency with degree of superheating are displayed in **Figure 9**. It can be seen that the electrical power, thermal efficiency and system generating efficiency yield a small variation with the degree of superheating, which is in line with the theoretical study. The main reason is that the degree of superheating has a negligible effect on the power output. The electrical power of 1.35 kW, thermal efficiency of 6.48% and system generating efficiency of 3.91% can be owned.

#### **4.5 Effect of condenser temperature**

The needle value is used to adjust the cooling water mass flow rate, while the working fluid mass flow rate is set to be 0.10 kg/s. The variation of electrical power, thermal efficiency and system generating efficiency with condenser temperature are

**Figure 9.** *Variation of electrical power, thermal efficiency and system generating efficiency with degree of superheating.*

**Figure 10.**

*Variation of electrical power, thermal efficiency and system generating efficiency with condenser temperature.*

**83**

**Figure 11.**

*Variation of BWR with mass flow rate for R245fa, R123 and their mixtures.*

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

demonstrated in **Figure 10**. As the condenser temperature increases, the electrical power, thermal efficiency and system generating efficiency own a similar behavior of a decrease trend. The main reason is that the rising condenser temperature decreases the expander enthalpy difference, resulting in the decrease in the power output. When the condenser temperature increases from 22 to 43°C, the electrical power decreases from 0.83 to 0.50 kW, while the thermal efficiency decreases from 3.02 to 1.34%.

Based on the parametric analysis, the pressure drop has a relatively effect on the system performance, indicating that improving the system overall performance should give priority to increase the pressure drop. Meanwhile, the optimum electrical power and thermal efficiency are 1.89 kW and 5.92%, respectively. The maximum thermal efficiency does not represent the highest electrical power, which is in

**5. Operation characteristics comparison between mixture and pure** 

To better compare the operation characteristics between the pure working fluids and mixture working fluids, two pure working fluids (R245fa, R123) and two mixtures working fluids (0.67R245fa/0.33R123 and 0.33R245fa/0.67R123) are tested. BWR denotes the ratio between pump consumption and electricity output. For theoretical study, the pump consumption is always ignore, whereas it accounts for a large proportion for ORC experimental prototype. The BWR for R245fa, R123 and their mixtures are displayed in **Figure 11**. As the mass flow rate increases, the BWR for different working fluids own a parabolic trend. There is an optimum mass flow rate to the lowest BWR. Meanwhile, the mixture working fluids yield a relatively higher BWR than the pure working fluids. 0.67R245fa/0.33R123 owns the highest BWR, while R123 yields the lowest BWR. BWR is in range of 11.86–23.22%, indicating that improving the pump operation characteristics is one way to enhance the

**Figure 12** shows the expander shaft power with mass flow rate for R245fa, R123 and their mixtures. It can be seen that as the mass flow rate increases, the expander shaft power for R245fa, 0.67R245fa/0.33R123 and 0.33R245fa/0.67R123 present a

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

line with the theoretical study.

**working fluids**

ORC performance.

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

demonstrated in **Figure 10**. As the condenser temperature increases, the electrical power, thermal efficiency and system generating efficiency own a similar behavior of a decrease trend. The main reason is that the rising condenser temperature decreases the expander enthalpy difference, resulting in the decrease in the power output. When the condenser temperature increases from 22 to 43°C, the electrical power decreases from 0.83 to 0.50 kW, while the thermal efficiency decreases from 3.02 to 1.34%.

Based on the parametric analysis, the pressure drop has a relatively effect on the system performance, indicating that improving the system overall performance should give priority to increase the pressure drop. Meanwhile, the optimum electrical power and thermal efficiency are 1.89 kW and 5.92%, respectively. The maximum thermal efficiency does not represent the highest electrical power, which is in line with the theoretical study.

#### **5. Operation characteristics comparison between mixture and pure working fluids**

To better compare the operation characteristics between the pure working fluids and mixture working fluids, two pure working fluids (R245fa, R123) and two mixtures working fluids (0.67R245fa/0.33R123 and 0.33R245fa/0.67R123) are tested.

BWR denotes the ratio between pump consumption and electricity output. For theoretical study, the pump consumption is always ignore, whereas it accounts for a large proportion for ORC experimental prototype. The BWR for R245fa, R123 and their mixtures are displayed in **Figure 11**. As the mass flow rate increases, the BWR for different working fluids own a parabolic trend. There is an optimum mass flow rate to the lowest BWR. Meanwhile, the mixture working fluids yield a relatively higher BWR than the pure working fluids. 0.67R245fa/0.33R123 owns the highest BWR, while R123 yields the lowest BWR. BWR is in range of 11.86–23.22%, indicating that improving the pump operation characteristics is one way to enhance the ORC performance.

**Figure 12** shows the expander shaft power with mass flow rate for R245fa, R123 and their mixtures. It can be seen that as the mass flow rate increases, the expander shaft power for R245fa, 0.67R245fa/0.33R123 and 0.33R245fa/0.67R123 present a

**Figure 11.** *Variation of BWR with mass flow rate for R245fa, R123 and their mixtures.*

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

system generating efficiency of 3.91% can be owned.

The variation of electrical power, thermal efficiency and system generating efficiency with degree of superheating are displayed in **Figure 9**. It can be seen that the electrical power, thermal efficiency and system generating efficiency yield a small variation with the degree of superheating, which is in line with the theoretical study. The main reason is that the degree of superheating has a negligible effect on the power output. The electrical power of 1.35 kW, thermal efficiency of 6.48% and

The needle value is used to adjust the cooling water mass flow rate, while the working fluid mass flow rate is set to be 0.10 kg/s. The variation of electrical power, thermal efficiency and system generating efficiency with condenser temperature are

**4.4 Effect of degree of superheating**

**4.5 Effect of condenser temperature**

**82**

**Figure 10.**

**Figure 9.**

*Variation of electrical power, thermal efficiency and system generating efficiency with condenser temperature.*

*Variation of electrical power, thermal efficiency and system generating efficiency with degree of superheating.*

#### **Figure 12.**

*Variation of expander shaft power with mass flow rate for R245fa, R123 and their mixtures.*

#### **Figure 13.**

*Variation of thermal efficiency with mass flow rate for R245fa, R123 and their mixtures.*

similar trend of increase first and then decrease, while that of R123 almost has no change. The highest expander shaft power is obtained by 0.67R245fa/0.33R123, while the lowest one is got by R123. One optimum mass flow rate is existed to ensure the highest expander shaft power. The maximum expander shaft power for R245fa of 2.76 kW, 0.67R245fa/0.33R123 of 2.85 kW, 0.33R245fa/0.67R12 of 2.46 kW and R123 of 1.82 kW are obtained.

The thermal efficiencies for R245fa, R123 and their mixtures are plotted in **Figure 13**. As observed, as the mass flow rate increases, the thermal efficiencies for 0.33R245fa/0.67R123, 0.67R245fa/0.33R123 and R123 keep decreasing, whereas that of R245 presents a parabolic trend. The thermal efficiency is determined by the net power output and heat input. The main reason is the comprehensive effect of the increasing net power output and heat input. It also can be found that the mixture working fluids have a relatively higher thermal efficiency than the

**85**

**Figure 14.**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

pure working fluids. The maximum thermal efficiencies for R245fa of 6.70%, 0.67R245fa/0.33R123 of 7.33%, 0.33R245fa/0.67R123 of 6.99% and R123 of 5.06% are obtained. 0.67R245fa/0.33R123 owns the highest maximum thermal efficiency of 7.33, 9.4% higher than that of R245fa and 44.86% higher than that of R123. **Figure 14** shows the system generating efficiencies with mass flow rate for R245fa, R123 and their mixtures. The system generating efficiencies for different working fluids have a same trend with the thermal efficiency (as shown in **Figure 13**). The comprehensive influence of heat input and net electricity output enables the parabolic trend in system generating efficiency. The maximum system generating efficiencies for R245fa of 4.03%, 0.67R245fa/0.33R123 of 4.53%, 0.33R245fa/0.67R123 of 4.18% and R123 of 3.10% are yielded. Meanwhile, the corresponding mass flow rate for maximum system generating efficiencies of R245fa, 0.67R245fa/0.33R123, 0.33R245fa/0.67R123 and R123 are 0.130, 0.149, 0.125 and 0.124 kg/s, respectively. 0.67R245fa/0.33R123 owns the highest system generating

*Variation of system generating efficiency with mass flow rate for R245fa, R123 and their mixtures.*

efficiency of 4.53%, which is 12.41% higher than that of R245fa.

**6. System behaviors at different operation strategies**

pump is necessary for ORC application.

Through the experimental comparison between the pure and mixture working fluids, it indicates that whether the mixtures exhibit better thermodynamic performance than the pure working fluids depend on the operation parameters and mass fraction of mixtures. Meanwhile, the mixture working fluids obtain a higher expander shaft power but a relatively higher BWR, indicating enhancement in

To ensure that ORC can operate at different zone, two operation strategies are proposed: stand alone and gird connect operation strategies. The mass flow rates and expander rotating speed with heat input for standalone and gird connect operation strategies are displayed in **Figure 15**. As the heat input increases, the mass flow rate for standalone operation strategy is in range of 0.153–0.359 kg/s, while that of grid connect operation strategy increases from 0.230 to 0.420 kg/s. Meanwhile, the expander rotating speed for standalone operation strategy keeps rising from 2320 to 2983 rpm,

whereas that of grid connect operation strategy keeps constant of 3600 rpm.

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

#### **Figure 14.**

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

similar trend of increase first and then decrease, while that of R123 almost has no change. The highest expander shaft power is obtained by 0.67R245fa/0.33R123, while the lowest one is got by R123. One optimum mass flow rate is existed to ensure the highest expander shaft power. The maximum expander shaft power for R245fa of 2.76 kW, 0.67R245fa/0.33R123 of 2.85 kW, 0.33R245fa/0.67R12 of 2.46 kW and

*Variation of thermal efficiency with mass flow rate for R245fa, R123 and their mixtures.*

*Variation of expander shaft power with mass flow rate for R245fa, R123 and their mixtures.*

The thermal efficiencies for R245fa, R123 and their mixtures are plotted in **Figure 13**. As observed, as the mass flow rate increases, the thermal efficiencies for 0.33R245fa/0.67R123, 0.67R245fa/0.33R123 and R123 keep decreasing, whereas that of R245 presents a parabolic trend. The thermal efficiency is determined by the net power output and heat input. The main reason is the comprehensive effect of the increasing net power output and heat input. It also can be found that the mixture working fluids have a relatively higher thermal efficiency than the

**84**

**Figure 13.**

**Figure 12.**

R123 of 1.82 kW are obtained.

*Variation of system generating efficiency with mass flow rate for R245fa, R123 and their mixtures.*

pure working fluids. The maximum thermal efficiencies for R245fa of 6.70%, 0.67R245fa/0.33R123 of 7.33%, 0.33R245fa/0.67R123 of 6.99% and R123 of 5.06% are obtained. 0.67R245fa/0.33R123 owns the highest maximum thermal efficiency of 7.33, 9.4% higher than that of R245fa and 44.86% higher than that of R123.

**Figure 14** shows the system generating efficiencies with mass flow rate for R245fa, R123 and their mixtures. The system generating efficiencies for different working fluids have a same trend with the thermal efficiency (as shown in **Figure 13**). The comprehensive influence of heat input and net electricity output enables the parabolic trend in system generating efficiency. The maximum system generating efficiencies for R245fa of 4.03%, 0.67R245fa/0.33R123 of 4.53%, 0.33R245fa/0.67R123 of 4.18% and R123 of 3.10% are yielded. Meanwhile, the corresponding mass flow rate for maximum system generating efficiencies of R245fa, 0.67R245fa/0.33R123, 0.33R245fa/0.67R123 and R123 are 0.130, 0.149, 0.125 and 0.124 kg/s, respectively. 0.67R245fa/0.33R123 owns the highest system generating efficiency of 4.53%, which is 12.41% higher than that of R245fa.

Through the experimental comparison between the pure and mixture working fluids, it indicates that whether the mixtures exhibit better thermodynamic performance than the pure working fluids depend on the operation parameters and mass fraction of mixtures. Meanwhile, the mixture working fluids obtain a higher expander shaft power but a relatively higher BWR, indicating enhancement in pump is necessary for ORC application.

#### **6. System behaviors at different operation strategies**

To ensure that ORC can operate at different zone, two operation strategies are proposed: stand alone and gird connect operation strategies. The mass flow rates and expander rotating speed with heat input for standalone and gird connect operation strategies are displayed in **Figure 15**. As the heat input increases, the mass flow rate for standalone operation strategy is in range of 0.153–0.359 kg/s, while that of grid connect operation strategy increases from 0.230 to 0.420 kg/s. Meanwhile, the expander rotating speed for standalone operation strategy keeps rising from 2320 to 2983 rpm, whereas that of grid connect operation strategy keeps constant of 3600 rpm.

#### **Figure 15.**

*Variation of working fluid mass flow rates and expander rotating speed with heat input for standalone and gird connect operation strategies.*

#### **Figure 16.**

*Variation of thermal efficiencies with heat input for standalone and gird connect operation strategies.*

**Figure 16** shows the thermal efficiencies with heat input for standalone and gird connect operation strategies. When the heat input increases from 34.18 to 87.40 kW, the thermal efficiency for standalone operation strategy decreases 4.04–3.64%, which may be attributed to the comprehensive effect of increasing heat input. However, for the gird connect operation strategy, the thermal efficiency increases from 0.96–4.89% with heat input. Meanwhile, when the heat input is lower than 74.20 kW, the stand alone operation strategy owns a relatively higher thermal efficiency than the gird connect operation strategy, but an opposite trend for the heat input higher than 74.20 kW. When the heat input is 102.62 kW, the gird connect operation strategy yields the highest thermal efficiency of 4.89%.The system generating efficiencies with heat input for standalone and gird connect operation strategies are demonstrated in **Figure 17**. The system generating efficiencies for standalone and gird connect operation strategies present a similar trend with the thermal efficiencies (as shown in **Figure 16**). The maximum system generating

**87**

**Acknowledgements**

**7. Conclusion**

**Figure 17.**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

efficiency for gird connect operation strategy is 4.59%, which is 35% higher than

*Variation of system generating efficiencies with heat input for standalone and gird connect operation strategies.*

Power conversion systems based on organic Rankine cycles have been identified as a potential technology especially in converting low-grade waste heat into electricity. This chapter presents the detailed operation characteristic of a small-scale ORC. The effect of mass flow rate, pressure drop, degree of superheating and condenser temperature on thermal efficiency and system generating efficiency are examined. Two pure working fluids (R245fa, R123) and two mixtures working fluids (0.67R245fa/0.33R123 and 0.33R245fa/0.67R123) are tested and compared. The system behaviors at two different operation strategies are addressed. The pressure drop has a relatively effect on the system performance, indicating that improving the system overall performance should give priority to increase the pressure drop. The maximum thermal efficiency does not represent the highest electrical power, which is in line with the theoretical study. The degree of superheating exhibits insensitive on the electrical power. There is an optimum mass flow rate to ensure the minimum BWR. Whether the mixtures exhibit better thermodynamic performance than the pure working fluids depend on the operation parameters and mass fraction of mixtures. Meanwhile, the mixture working fluids obtain a higher expander shaft power but a relatively higher BWR. The expander rotating speed for standalone operation strategy keeps rising from 2320 to 2983 rpm, whereas that of grid connect operation strategy keeps constant of 3600 rpm. When the heat input is lower than 74.20 kW, the stand alone operation strategy owns a relatively higher thermal efficiency than the gird connect operation

that of standalone operation strategy approaching of 3.40%.

strategy, but an opposite trend for the heat input higher than 74.20 kW.

This research work has been supported by the Ministry of Science and Technology, Taiwan under the grants of Contract No. MOST 107-2221-E-027-091, and by "Research Center of Energy Conservation for New Generation of Residential,

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

#### **Figure 17.**

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

*Variation of thermal efficiencies with heat input for standalone and gird connect operation strategies.*

*Variation of working fluid mass flow rates and expander rotating speed with heat input for standalone and* 

**Figure 16** shows the thermal efficiencies with heat input for standalone and gird connect operation strategies. When the heat input increases from 34.18 to 87.40 kW, the thermal efficiency for standalone operation strategy decreases 4.04–3.64%, which may be attributed to the comprehensive effect of increasing heat input. However, for the gird connect operation strategy, the thermal efficiency increases from 0.96–4.89% with heat input. Meanwhile, when the heat input is lower than 74.20 kW, the stand alone operation strategy owns a relatively higher thermal efficiency than the gird connect operation strategy, but an opposite trend for the heat input higher than 74.20 kW. When the heat input is 102.62 kW, the gird connect operation strategy yields the highest thermal efficiency of 4.89%.The system generating efficiencies with heat input for standalone and gird connect operation strategies are demonstrated in **Figure 17**. The system generating efficiencies for standalone and gird connect operation strategies present a similar trend with the thermal efficiencies (as shown in **Figure 16**). The maximum system generating

**86**

**Figure 16.**

**Figure 15.**

*gird connect operation strategies.*

*Variation of system generating efficiencies with heat input for standalone and gird connect operation strategies.*

efficiency for gird connect operation strategy is 4.59%, which is 35% higher than that of standalone operation strategy approaching of 3.40%.

#### **7. Conclusion**

Power conversion systems based on organic Rankine cycles have been identified as a potential technology especially in converting low-grade waste heat into electricity. This chapter presents the detailed operation characteristic of a small-scale ORC. The effect of mass flow rate, pressure drop, degree of superheating and condenser temperature on thermal efficiency and system generating efficiency are examined. Two pure working fluids (R245fa, R123) and two mixtures working fluids (0.67R245fa/0.33R123 and 0.33R245fa/0.67R123) are tested and compared. The system behaviors at two different operation strategies are addressed. The pressure drop has a relatively effect on the system performance, indicating that improving the system overall performance should give priority to increase the pressure drop. The maximum thermal efficiency does not represent the highest electrical power, which is in line with the theoretical study. The degree of superheating exhibits insensitive on the electrical power. There is an optimum mass flow rate to ensure the minimum BWR. Whether the mixtures exhibit better thermodynamic performance than the pure working fluids depend on the operation parameters and mass fraction of mixtures. Meanwhile, the mixture working fluids obtain a higher expander shaft power but a relatively higher BWR. The expander rotating speed for standalone operation strategy keeps rising from 2320 to 2983 rpm, whereas that of grid connect operation strategy keeps constant of 3600 rpm. When the heat input is lower than 74.20 kW, the stand alone operation strategy owns a relatively higher thermal efficiency than the gird connect operation strategy, but an opposite trend for the heat input higher than 74.20 kW.

#### **Acknowledgements**

This research work has been supported by the Ministry of Science and Technology, Taiwan under the grants of Contract No. MOST 107-2221-E-027-091, and by "Research Center of Energy Conservation for New Generation of Residential, Commercial, and Industrial Sector" from The Featured Areas Research Center Program within the framework of the Higher Education Sprout Project by the Ministry of Education (MOE) in Taiwan. The authors are grateful for National Natural Science Foundation of China (51806081), the Natural Science Foundation of Jiangsu Province (BK20180882), the China Postdoctoral Science Foundation (2018 M632241) and the Open Foundation Program of Key Laboratory of Efficient Utilization of Low and Medium Grade Energy (Tianjin University), the Ministry of Education of China (201806-402).

### **Conflict of interest**

The author declared that there is no conflict of interest.

### **Author details**

Tzu-Chen Hung1 \* and Yong-qiang Feng2

1 Department of Mechanical Engineering, National Taipei University of Technology, Taipei, Taiwan

2 School of Energy and Power Engineering, Jiangsu University, Zhenjiang, China

\*Address all correspondence to: tchung@ntut.edu.tw

© 2019 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

**89**

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[10] Carraro G, Pallis P, Leontaritis AD, Karellas S, Vourliotis P, Rech S, et al. Experimental performance evaluation of a multi-diaphragm pump of a micro-ORC system. Energy Procedia. 2017;**129**:1018-1025. DOI: 10.1016/j.

[11] Yang XF, Xu JL, Miao Z, Zou JH, Yu C. Operation of an organic Rankine cycle dependent on pumping flow rates and expander torques. Energy.

[12] Xu W, Zhang JY, Zhao L, Deng S, Zhang Y. Novel experimental research

organic Rankine cycle (ORC). Energy

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*DOI: http://dx.doi.org/10.5772/intechopen.88208*

[2] Feng YQ, Zhang YN, Li BX, Yang JF, Shi Y. Comparison between regenerative organic Rankine cycle (RORC) and basic organic Rankine cycle (BORC) based on thermoeconomic multiobjective optimization considering exergy efficiency and levelized energy cost (LEC). Energy Conversion and Management. 2015;**96**:58-71. DOI: 10.1016/j.enconman.2015.02.045

[3] Feng YQ, Hung TC, Greg K, Zhang YN, Li BX, Yang JF.

[4] Tian H, Shu GQ, Wei HQ, Liang XY, Liu LN. Fluids and

[5] Tchanche B, Papadakis G,

[6] Dong BS, Xu GQ, Luo X,

Lambrinos G, Frangoudakis A. Fluid selection for a low-temperature solar organic Rankine cycle. Applied Thermal Engineering. 2009;**29**:2468-2476. DOI: 10.1016/j.applthermaleng.2008.12.025

Zhuang LH, Quan YK. Potential of low temperature organic Rankine cycle with zeotropic mixtures as working fluid. Energy Procedia. 2017;**105**:1489-1494. DOI: 10.1016/j.egypro.2017.03.444

Thermoeconomic comparison between pure and mixture working fluids for low-grade organic Rankine cycles (ORCs). Energy Conversion and Management. 2015;**106**:859-872. DOI: 10.1016/j.enconman.2015.09.042

parameters optimization for the organic Rankine cycles (ORCs) used in exhaust heat recovery of internal combustion engine (ICE). Energy. 2012;**47**:125-136. DOI: 10.1016/j.energy.2012.09.021

[1] Feng YQ, Hung TC, Zhang Y, Li BX, Yang JF, Shi Y. Performance comparison of low-grade organic Rankine cycles (ORCs) using R245fa, pentane and their mixtures based on the thermoeconomic multi-objective optimization and decision makings. Energy. 2015;**99**:2018-2029. DOI: 10.1016/j.energy.2015.10.065

**References**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

#### **References**

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

The author declared that there is no conflict of interest.

\* and Yong-qiang Feng2

\*Address all correspondence to: tchung@ntut.edu.tw

provided the original work is properly cited.

1 Department of Mechanical Engineering, National Taipei University of

2 School of Energy and Power Engineering, Jiangsu University, Zhenjiang, China

© 2019 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

Education of China (201806-402).

**Conflict of interest**

**Author details**

Tzu-Chen Hung1

Technology, Taipei, Taiwan

Commercial, and Industrial Sector" from The Featured Areas Research Center Program within the framework of the Higher Education Sprout Project by the Ministry of Education (MOE) in Taiwan. The authors are grateful for National Natural Science Foundation of China (51806081), the Natural Science Foundation of Jiangsu Province (BK20180882), the China Postdoctoral Science Foundation (2018 M632241) and the Open Foundation Program of Key Laboratory of Efficient Utilization of Low and Medium Grade Energy (Tianjin University), the Ministry of

**88**

[1] Feng YQ, Hung TC, Zhang Y, Li BX, Yang JF, Shi Y. Performance comparison of low-grade organic Rankine cycles (ORCs) using R245fa, pentane and their mixtures based on the thermoeconomic multi-objective optimization and decision makings. Energy. 2015;**99**:2018-2029. DOI: 10.1016/j.energy.2015.10.065

[2] Feng YQ, Zhang YN, Li BX, Yang JF, Shi Y. Comparison between regenerative organic Rankine cycle (RORC) and basic organic Rankine cycle (BORC) based on thermoeconomic multiobjective optimization considering exergy efficiency and levelized energy cost (LEC). Energy Conversion and Management. 2015;**96**:58-71. DOI: 10.1016/j.enconman.2015.02.045

[3] Feng YQ, Hung TC, Greg K, Zhang YN, Li BX, Yang JF. Thermoeconomic comparison between pure and mixture working fluids for low-grade organic Rankine cycles (ORCs). Energy Conversion and Management. 2015;**106**:859-872. DOI: 10.1016/j.enconman.2015.09.042

[4] Tian H, Shu GQ, Wei HQ, Liang XY, Liu LN. Fluids and parameters optimization for the organic Rankine cycles (ORCs) used in exhaust heat recovery of internal combustion engine (ICE). Energy. 2012;**47**:125-136. DOI: 10.1016/j.energy.2012.09.021

[5] Tchanche B, Papadakis G, Lambrinos G, Frangoudakis A. Fluid selection for a low-temperature solar organic Rankine cycle. Applied Thermal Engineering. 2009;**29**:2468-2476. DOI: 10.1016/j.applthermaleng.2008.12.025

[6] Dong BS, Xu GQ, Luo X, Zhuang LH, Quan YK. Potential of low temperature organic Rankine cycle with zeotropic mixtures as working fluid. Energy Procedia. 2017;**105**:1489-1494. DOI: 10.1016/j.egypro.2017.03.444

[7] Chang JC, Hung TC, He YL, Zhang WP. Experimental study on low-temperature organic Rankine cycle utilizing scroll type expander. Applied Energy. 2015;**155**:150-159. DOI: 10.1016/j.apenergy.2015.05.118

[8] Muhammad U, Imran M, Lee DH, Park BS. Design and experimental investigation of a 1 kW organic Rankine cycle system using R245fa as working fluid for low-grade waste heat recovery from steam. Energy Conversion and Management. 2015;**103**:1089-1100. DOI: 10.1016/j.enconman.2015.07.045

[9] Mathias JA, Johnston JR, Cao JM, Priedeman DK, Christensen RN. Experimental testing of gerotor and scroll expanders used in, and energetic and exergetic modeling of, an organic Rankine cycle. Journal of Energy Resources Technology Transactions ASME. 2009;**131**:21-24. DOI: 10.1115/1.3066345

[10] Carraro G, Pallis P, Leontaritis AD, Karellas S, Vourliotis P, Rech S, et al. Experimental performance evaluation of a multi-diaphragm pump of a micro-ORC system. Energy Procedia. 2017;**129**:1018-1025. DOI: 10.1016/j. egypro.2017.09.232

[11] Yang XF, Xu JL, Miao Z, Zou JH, Yu C. Operation of an organic Rankine cycle dependent on pumping flow rates and expander torques. Energy. 2015;**90**:1-15. DOI: 10.1016/j. energy.2015.07.121

[12] Xu W, Zhang JY, Zhao L, Deng S, Zhang Y. Novel experimental research on the compression process in organic Rankine cycle (ORC). Energy Conversion and Management. 2017;**137**:1-11. DOI: 10.1016/j. enconman.2017.01.025

[13] Lei B, Wang JF, Wu YT, Ma CF, Wang W, Zhang L, et al. Experimental study and theoretical analysis of a Roto-Jet pump in small scale organic Rankine cycles. Energy Conversion and Management. 2016;**111**:198-204. DOI: 10.1016/j.enconman.2015.12.062

[14] Zeleny Z, Vodicka V, Novotny V, Mascuch J. Gear pump for low power output ORC—An efficiency analysis. Energy Procedia. 2017;**129**:1002-1009. DOI: 10.1016/j.egypro.2017.09.227

[15] Bianchi G, Fatigati F, Murgia S, Cipollone R, Contaldi G. Modeling and experimental activities on a small-scale sliding vane pump for ORC-based waste heat recovery applications. Energy Procedia. 2016;**101**:1240-1247. DOI: 10.1016/j.egypro.2016.11.139

[16] Wu TM, Liu JH, Zhang L, Xu XJ. Experimental study on multistage gas-liquid booster pump for working fluid pressurization. Applied Thermal Engineering. 2017;**126**:9-16. DOI: 10.1016/j. applthermaleng.2017.07.159

[17] Yang YX, Zhang HG, Xu YH, Yang FB, Wu YT, Lei B. Matching and operating characteristics of working fluid pumps with organic Rankine cycle system. Applied Thermal Engineering. 2018;**142**:622-631. DOI: 10.1016/j. applthermaleng.2018.07.039

[18] Landelle A, Tauveron N, Revellin R, Haberschill P, Colasson S, Roussel V. Performance investigation of reciprocating pump running with organic fluid for organic Rankine cycle. Applied Thermal Engineering. 2017;**113**:962-969. DOI: 10.1016/j. applthermaleng.2016.11.096

[19] Meng FX, Zhang HG, Yang FB, Hou XC, Lei B, Zhang L. Study of efficiency of a multistage centrifugal pump used in engine waste heat recovery application. Applied Thermal Engineering. 2017;**110**:779-786. DOI: 10.1016/j.applthermaleng.2016.08.226

[20] Bianchi M, Branchini L, Casari N, De PA, Melino F, Ottaviano S, et al. Experimental analysis of a micro-ORC driven by piston expander for lowgrade heat recovery. Applied Thermal Engineering. 2019;**148**:1278-1291. DOI: 10.1016/j.applthermaleng.2018.12.019

[21] Yang YX, Zhang HG, Xu YH, Zhao R, Hou XC, Liu Y. Experimental study and performance analysis of a hydraulic diaphragm metering pump used in organic Rankine cycle system. Applied Thermal Engineering. 2018;**132**:605-612. DOI: 10.1016/j. applthermaleng.2018.01.001

[22] Bianchi G, Fatigati F, Murgia S, Cipollone R. Design and analysis of a sliding vane pump for waste heat to power conversion systems using organic fluids. Applied Thermal Engineering. 2017;**124**:1038-1048. DOI: 10.1016/j. applthermaleng.2017.06.083

[23] Sun HC, Qin J, Hung TC, Hua HY, Ya PG, Lin CH. Effect of flow losses in heat exchangers on the performance of organic Rankine cycle. Energy. 2019;**172**:391-400. DOI: 10.1016/j. energy.2019.01.131

[24] Xi H, Zhang HH, He YL, Huang ZH. Sensitivity analysis of operation parameters on the system performance of organic Rankine cycle system using orthogonal experiment. Energy. 2019;**172**:435-442. DOI: 10.1016/j.energy.2019.01.072

[25] Borsukiewicz-Gozdur A. Pumping work in the organic Rankine cycle. Applied Thermal Engineering. 2013;**51**:781-786. DOI: 10.1016/j. applthermaleng.2012.10.033

[26] Dong BS, Xu GQ, Cai Y, Li HW. Analysis of zeotropic mixtures used in high-temperature organic Rankine cycle. Energy Conversion and Management. 2014;**84**:253-260. DOI: 10.1016/j.enconman.2014.04.026

**91**

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat…*

2015;**81**:601-614. DOI: 10.1016/j.

Experimental comparison of R245fa and R245fa/R601a for organic Rankine cycle using scroll expander. International

[35] Feng YQ, Hung TC, Wu SL, Lin CH, Li BX, Huang KC, et al. Operation characteristic of a R123-based organic Rankine cycle depending on working fluid mass flow rates and heat source temperature. Energy Conversion and Management. 2017;**131**:55-68. DOI: 10.1016/j.enconman.2016.11.004

[36] Yang SC, Hung TC, Feng YQ, Wu CJ, Wong KW, Huang KC.

Experimental investigation on a 3 kW organic Rankine cycle for low grade waste heat under different operation parameters. Applied Thermal Engineering. 2017;**113**:756-764. DOI: 10.1016/j.applthermaleng.2016.11.032

[37] Feng YQ, Hung TC, He YL, Wang Q, Wang S, Li BX, et al. Operation characteristic and performance comparison of organic Rankine cycle (ORC) for low-grade waste heat using R245fa, R123 and their mixtures. Energy Conversion and Management. 2017;**144**:153-163. DOI: 10.1016/j.

[38] Feng YQ, Hung TC, Su TY, Wang S, Wang Q, Yang SC, et al. Experimental investigation of a R245fa-based organic Rankine cycle adapting two operation strategies: Stand alone and grid connect. Energy. 2017;**141**:1239-1253. DOI: 10.1016/j.energy.2017.09.119

[39] NIST Standard Reference Database

transport properties of refrigerants and refrigerant mixtures REFROP, Version

23. NIST thermodynamic and

9.0. 2010

enconman.2017.04.048

[34] Li T, Zhu J, Fu W, Hu K.

Journal of Energy Research. 2015;**39**(2):202-214. DOI: 10.1002/

energy.2015.01.003

er.3228

*DOI: http://dx.doi.org/10.5772/intechopen.88208*

[27] Garg P, Kumar P, Srinivasan K, Dutta P. Evaluation of isopentane, R-245fa and their mixtures as working fluids for organic Rankine cycles. Applied Thermal Engineering. 2013;**51**(1):292-300. DOI: 10.1016/j. applthermaleng.2012.08.056

[28] Lecompte S, Ameel B, Ziviani D, van den Broek M, De Paepe M. Exergy analysis of zeotropic mixtures as working fluids in organic Rankine cycles. Energy Conversion and Management. 2014;**85**:727-739. DOI: 10.1016/j.enconman.2014.02.028

[29] Zhao L, Bao JJ. The influence of composition shift on organic Rankine cycle (ORC) with zeotropic mixtures. Energy Conversion and Management. 2014;**83**:203-211. DOI: 10.1016/j.

[30] Liu Q, Duan Y, Yang Z. Effect of condensation temperature glide on the performance of organic Rankine cycles with zeotropic mixture working fluids. Applied Energy. 2014;**115**:394-404. DOI:

10.1016/j.apenergy.2013.11.036

[31] Pu W, Yue C, Han D, He W, Liu X, Zhang Q, et al. Experimental study on organic Rankine cycle for low grade thermal energy recovery. Applied Thermal Engineering. 2016;**94**:221-227. DOI: 10.1016/j. applthermaleng.2015.09.120

[32] Molés F, Navarro-Esbrí J,

evaluation of HCFO-1233zd-E as HFC-245fa replacement in an organic Rankine cycle system for low temperature heat sources. Applied Thermal Engineering. 2016;**98**:954-961. DOI: 10.1016/j. applthermaleng.2016.01.011

Peris B, Mota-Babiloni A. Experimental

[33] Jung HC, Taylor L, Krumdieck S. An experimental and modelling study of a 1 kW organic Rankine cycle unit with mixture working fluid. Energy.

enconman.2014.03.072

*The Development and Application of a Small-Scale Organic Rankine Cycle for Waste Heat… DOI: http://dx.doi.org/10.5772/intechopen.88208*

[27] Garg P, Kumar P, Srinivasan K, Dutta P. Evaluation of isopentane, R-245fa and their mixtures as working fluids for organic Rankine cycles. Applied Thermal Engineering. 2013;**51**(1):292-300. DOI: 10.1016/j. applthermaleng.2012.08.056

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

[20] Bianchi M, Branchini L, Casari N, De PA, Melino F, Ottaviano S, et al. Experimental analysis of a micro-ORC driven by piston expander for lowgrade

heat recovery. Applied Thermal Engineering. 2019;**148**:1278-1291. DOI: 10.1016/j.applthermaleng.2018.12.019

[21] Yang YX, Zhang HG, Xu YH, Zhao R, Hou XC, Liu Y. Experimental study and performance analysis of a hydraulic diaphragm metering pump used in organic Rankine cycle system. Applied Thermal Engineering. 2018;**132**:605-612. DOI: 10.1016/j. applthermaleng.2018.01.001

[22] Bianchi G, Fatigati F, Murgia S, Cipollone R. Design and analysis of a sliding vane pump for waste heat to power conversion systems using organic fluids. Applied Thermal Engineering. 2017;**124**:1038-1048. DOI: 10.1016/j.

[23] Sun HC, Qin J, Hung TC, Hua HY, Ya PG, Lin CH. Effect of flow losses in heat exchangers on the performance of organic Rankine cycle. Energy. 2019;**172**:391-400. DOI: 10.1016/j.

[25] Borsukiewicz-Gozdur A. Pumping work in the organic Rankine cycle. Applied Thermal Engineering. 2013;**51**:781-786. DOI: 10.1016/j. applthermaleng.2012.10.033

[26] Dong BS, Xu GQ, Cai Y, Li HW. Analysis of zeotropic mixtures used in high-temperature organic Rankine

cycle. Energy Conversion and Management. 2014;**84**:253-260. DOI: 10.1016/j.enconman.2014.04.026

applthermaleng.2017.06.083

[24] Xi H, Zhang HH, He YL, Huang ZH. Sensitivity analysis of operation parameters on the system performance of organic Rankine cycle system using orthogonal experiment. Energy. 2019;**172**:435-442. DOI: 10.1016/j.energy.2019.01.072

energy.2019.01.131

study and theoretical analysis of a Roto-Jet pump in small scale organic Rankine cycles. Energy Conversion and Management. 2016;**111**:198-204. DOI: 10.1016/j.enconman.2015.12.062

[14] Zeleny Z, Vodicka V, Novotny V, Mascuch J. Gear pump for low power output ORC—An efficiency analysis. Energy Procedia. 2017;**129**:1002-1009. DOI: 10.1016/j.egypro.2017.09.227

[15] Bianchi G, Fatigati F, Murgia S, Cipollone R, Contaldi G. Modeling and experimental activities on a small-scale sliding vane pump for ORC-based waste heat recovery applications. Energy Procedia. 2016;**101**:1240-1247. DOI:

10.1016/j.egypro.2016.11.139

[16] Wu TM, Liu JH, Zhang L, Xu XJ. Experimental study on multistage gas-liquid booster pump for working fluid pressurization. Applied Thermal Engineering. 2017;**126**:9-16. DOI: 10.1016/j. applthermaleng.2017.07.159

[17] Yang YX, Zhang HG, Xu YH, Yang FB, Wu YT, Lei B. Matching and operating characteristics of working fluid pumps with organic Rankine cycle system. Applied Thermal Engineering. 2018;**142**:622-631. DOI: 10.1016/j. applthermaleng.2018.07.039

[18] Landelle A, Tauveron N,

Revellin R, Haberschill P, Colasson S, Roussel V. Performance investigation of reciprocating pump running with organic fluid for organic Rankine cycle. Applied Thermal Engineering. 2017;**113**:962-969. DOI: 10.1016/j. applthermaleng.2016.11.096

[19] Meng FX, Zhang HG, Yang FB, Hou XC, Lei B, Zhang L. Study of efficiency of a multistage centrifugal pump used in engine waste heat recovery application. Applied Thermal Engineering. 2017;**110**:779-786. DOI: 10.1016/j.applthermaleng.2016.08.226

**90**

[28] Lecompte S, Ameel B, Ziviani D, van den Broek M, De Paepe M. Exergy analysis of zeotropic mixtures as working fluids in organic Rankine cycles. Energy Conversion and Management. 2014;**85**:727-739. DOI: 10.1016/j.enconman.2014.02.028

[29] Zhao L, Bao JJ. The influence of composition shift on organic Rankine cycle (ORC) with zeotropic mixtures. Energy Conversion and Management. 2014;**83**:203-211. DOI: 10.1016/j. enconman.2014.03.072

[30] Liu Q, Duan Y, Yang Z. Effect of condensation temperature glide on the performance of organic Rankine cycles with zeotropic mixture working fluids. Applied Energy. 2014;**115**:394-404. DOI: 10.1016/j.apenergy.2013.11.036

[31] Pu W, Yue C, Han D, He W, Liu X, Zhang Q, et al. Experimental study on organic Rankine cycle for low grade thermal energy recovery. Applied Thermal Engineering. 2016;**94**:221-227. DOI: 10.1016/j. applthermaleng.2015.09.120

[32] Molés F, Navarro-Esbrí J, Peris B, Mota-Babiloni A. Experimental evaluation of HCFO-1233zd-E as HFC-245fa replacement in an organic Rankine cycle system for low temperature heat sources. Applied Thermal Engineering. 2016;**98**:954-961. DOI: 10.1016/j. applthermaleng.2016.01.011

[33] Jung HC, Taylor L, Krumdieck S. An experimental and modelling study of a 1 kW organic Rankine cycle unit with mixture working fluid. Energy.

2015;**81**:601-614. DOI: 10.1016/j. energy.2015.01.003

[34] Li T, Zhu J, Fu W, Hu K. Experimental comparison of R245fa and R245fa/R601a for organic Rankine cycle using scroll expander. International Journal of Energy Research. 2015;**39**(2):202-214. DOI: 10.1002/ er.3228

[35] Feng YQ, Hung TC, Wu SL, Lin CH, Li BX, Huang KC, et al. Operation characteristic of a R123-based organic Rankine cycle depending on working fluid mass flow rates and heat source temperature. Energy Conversion and Management. 2017;**131**:55-68. DOI: 10.1016/j.enconman.2016.11.004

[36] Yang SC, Hung TC, Feng YQ, Wu CJ, Wong KW, Huang KC. Experimental investigation on a 3 kW organic Rankine cycle for low grade waste heat under different operation parameters. Applied Thermal Engineering. 2017;**113**:756-764. DOI: 10.1016/j.applthermaleng.2016.11.032

[37] Feng YQ, Hung TC, He YL, Wang Q, Wang S, Li BX, et al. Operation characteristic and performance comparison of organic Rankine cycle (ORC) for low-grade waste heat using R245fa, R123 and their mixtures. Energy Conversion and Management. 2017;**144**:153-163. DOI: 10.1016/j. enconman.2017.04.048

[38] Feng YQ, Hung TC, Su TY, Wang S, Wang Q, Yang SC, et al. Experimental investigation of a R245fa-based organic Rankine cycle adapting two operation strategies: Stand alone and grid connect. Energy. 2017;**141**:1239-1253. DOI: 10.1016/j.energy.2017.09.119

[39] NIST Standard Reference Database 23. NIST thermodynamic and transport properties of refrigerants and refrigerant mixtures REFROP, Version 9.0. 2010

**93**

**Chapter 5**

**Abstract**

waste heat recovery

**1. Introduction**

A Recent Review in Performance

of Organic Rankine Cycle (ORC)

Increasing emissions of carbon dioxide and fuel prices lead to extra efforts in finding solution to reduce the environment waste heat. One of the solutions emerging is the organic Rankine cycle (ORC) system. It is one of the promising exhaust heat recovery technologies which is widely been used to recover low to mediumgrade heat rather than conventional steam Rankine cycle system. This chapter highlights on the different conditions and configurations of ORCs that are usually been applied. These different configurations have different constraints and usually

**Keywords:** organic Rankine cycle, preheater, supercritical, superheating,

As higher efficiency of industrial technology is in demand, more and latest technologies are needed to produce energy. Increasing of population growth [1] and escalating process of electricity are mostly due to emission gases from the industry, vehicles, deforestation and others. In the aerospace industry, engineers continuously search for new methods to upgrade the efficiency of the engines. Recovery waste heat could increase the engine efficiency [2]. Although society undergoes global issues, social problems or economy crisis, this does not stop aerospace industry from expanding which leads to increase in demand on aircraft. This results in increasing of fuel price since more conventional fuel is needed and causes pollution into the environment [3]. Still, the price in this development is used to optimize the engine. In each flight, greater revenue could be achieved when the number of passengers is greater. And to have more passengers, lighter aircraft is needed. It is important to note that weight is very crucial in changing aircraft engines as it connects linearly with the amount of fuel used in every unit of force powered by engine (specific fuel consumption). Waste heat recovery (WHR) is one of the most important solutions found to lower the emission and fuel consumption [4]. Waste heat or low- grade waste heat is heat energy produced in the atmosphere through internal combustion. However, these low-grade waste heat are not sufficient enough to generate power due to insufficient low temperature. Thus, to recover these waste heat, organic Rankine cycle (ORC) system is one of the beneficial exhaust heat recovery technologies which is widely utilized in the applications of low grade heat recovery rather than conventional Rankine cycle [5]. By combining an ORC with energy system, for instance in power plants, organic fluid of low boiling point is

*Syamimi Saadon and Salmah Md Saiful Islam*

will be considered based on the applications.

#### **Chapter 5**

## A Recent Review in Performance of Organic Rankine Cycle (ORC)

*Syamimi Saadon and Salmah Md Saiful Islam*

#### **Abstract**

Increasing emissions of carbon dioxide and fuel prices lead to extra efforts in finding solution to reduce the environment waste heat. One of the solutions emerging is the organic Rankine cycle (ORC) system. It is one of the promising exhaust heat recovery technologies which is widely been used to recover low to mediumgrade heat rather than conventional steam Rankine cycle system. This chapter highlights on the different conditions and configurations of ORCs that are usually been applied. These different configurations have different constraints and usually will be considered based on the applications.

**Keywords:** organic Rankine cycle, preheater, supercritical, superheating, waste heat recovery

#### **1. Introduction**

As higher efficiency of industrial technology is in demand, more and latest technologies are needed to produce energy. Increasing of population growth [1] and escalating process of electricity are mostly due to emission gases from the industry, vehicles, deforestation and others. In the aerospace industry, engineers continuously search for new methods to upgrade the efficiency of the engines. Recovery waste heat could increase the engine efficiency [2]. Although society undergoes global issues, social problems or economy crisis, this does not stop aerospace industry from expanding which leads to increase in demand on aircraft. This results in increasing of fuel price since more conventional fuel is needed and causes pollution into the environment [3]. Still, the price in this development is used to optimize the engine. In each flight, greater revenue could be achieved when the number of passengers is greater. And to have more passengers, lighter aircraft is needed. It is important to note that weight is very crucial in changing aircraft engines as it connects linearly with the amount of fuel used in every unit of force powered by engine (specific fuel consumption). Waste heat recovery (WHR) is one of the most important solutions found to lower the emission and fuel consumption [4]. Waste heat or low- grade waste heat is heat energy produced in the atmosphere through internal combustion. However, these low-grade waste heat are not sufficient enough to generate power due to insufficient low temperature. Thus, to recover these waste heat, organic Rankine cycle (ORC) system is one of the beneficial exhaust heat recovery technologies which is widely utilized in the applications of low grade heat recovery rather than conventional Rankine cycle [5]. By combining an ORC with energy system, for instance in power plants, organic fluid of low boiling point is

utilized to change heat into electricity. The organic fluids or refrigerants used in air conditioning systems accumulates (collect) heat from a volume of air and release it to different type of heat exchanger which increases the expansion of high vapor pressure in expander. The heat accumulated is transformed into mechanical power or electricity and therefore will help to increase the thermal efficiency and the overall performance of the engine. Thus, higher thrust could be obtained as less electrical power is needed from aircraft engine resulting lower engine bleed air [2, 5, 6]. Because of its thermodynamic properties, organic fluid is the best selection for low quality heat sources with temperatures below than 100°C [2]. By selecting proper working fluid for low waste heat recovery system and modeling an optimum design of heat exchanger configurations, the waste energy recovered through this ORC system could be maximized. Thus, designing a fuel-efficient and cheaper heat exchanger, ORC power plant can effectually utilize the economic and environmental issues especially in aerospace industry.

#### **2. Aspects of recovery waste heat systems**

Based on second law of thermodynamics, the efficiency of a process would not be 100% as there is no process that can entirely transform all amount of heat into work. The energy that is not used to produce work is being dissipated as heat at different temperatures, levels streams. On aircraft, half of the fuel energy lost through this way. Nevertheless, these sources of waste heat are everywhere; from this lost energy, only a part of it can be used to produce mechanical work or other purposes, where around 30% of the total waste heat could be changed to useful work. As the demand of aircraft is increasing vastly, the aviation industry has been the world center of attraction as new technologies are needed and maximum exploitation of fuel is a must. The conversion of heat energy to mechanical or electrical power depends on the characteristics of the source. Let say in an air conditioning system, an external hose is two or three degrees above the ambient temperature, it is a waste to recover that little amount of energy, however, this power leak will be an irreversibility process together with other similar leaks will decrease the thermal efficiency. This is called as waste heat and is an unused heat energy produced as a by-product of process of energy transformation, as a natural consequence on any non-adiabatic process from the thermodynamics law. Most of the available waste heat is low waste heat that can be used by an ORC which utilizes low boiling point organic fluid as working fluid, for example, toluene, hexane or pentane. Presently, there has not been any waste heat recovery (WHR) system added to an aircraft. Nevertheless, researchers suggest on adding WHR system to future engines and propose to make changes in current engines. However, it is a hassle to change the actual design of the engine as more expenses will be used in research, tests and certifications and a lot of heat source needs to be taken into account. Pasini et al. [7] analyzed the possibilities of heat recovery results in overall efficiency of an aircraft engine. A waste heat recovery system is modeled in a jet engine and a turbo propeller engine. Their project takes into account the nozzle works in off design state. The heat emitted influences greatly in the system performance. They also developed a numerical thermodynamic code to evaluate the positive impacts of waste heat recovery in a turboprop, a turbofan and a turbojet. The turbofan engine is of great interest due to large fraction of thrust is provided by cold flow, whilst gas generator supplies needed power. The authors then concluded that the enthalpy level ahead of exhaust nozzle of gas generator could be decreased without losing a lot of thrust. From the results of the calculations, it was found an increase of thermal efficiency about

**95**

and simpler usage of components [11].

*A Recent Review in Performance of Organic Rankine Cycle (ORC)*

4% when heat recovery was done (efficiency of regeneration was 0.5). At the same time, if the efficiency was 0.7, an increase of 10% was achieved. From the numerical simulations, the best place for heat recovery is from hot gas before entering nozzle. Another research done by Li et al. [8] was to study the small-scale ORC system performance with low grade heat sources to provide electricity in various working state. The experiment setup includes normal ORC system components, for instance, turboexpander with high speed generator, finned tube condenser, ORC pump and plate evaporator. Results found that the turbine power and condenser heat output, ORC pump power and evaporator heat input, turbine isentropic, overall efficiencies and system thermal efficiencies rises when heat source temperature rises too. The fluid of ORC during superheat and pressure at turbine inlet were two crucial variables that kept constant with temperature heat source and pump speed of ORC.

ORC utilizes organic compound instead of water as a working fluid, generally, a refrigerant, a hydrocarbon such as pentane, butane, perfluorocarbon or silicon oil. The organic fluid's boiling point is much lesser compared to water and enable heat recovering at lesser temperatures instead of the steam Rankine cycle [9]. ORC's first commercial applications with medium-scale power plants for geothermal and solar applications were developed in the 9 late 1970s and 1980s. These days, over 200 ORC power plants are recognized with more than 1800 MWe installed and the technologies keep on increasing day by day [10]. Mostly, the plants were installed in biomass CHP application, geothermal plants and plants of WHR followed. The layout of ORC is much simpler compared to the steam cycle as there is no water vapor attached to the boiler, and a single heat exchanger could be utilized for the three processes of evaporation including preheating, vaporizing and superheating. ORC is able to use low grade heat sources than steam Rankine cycle. Since it could be utilized in lower temperature at the turbine inlet and reduce thermal stresses in the boiler. In regular steam plant systems, the performance cycle is at risk damage due to gaseous infiltrations that occur in sub atmospheric condensing pressure. In steam-based cycle, the usage of a single tube for evaporation is abstained due to large density difference that exists in between liquid and vapor phases. However, some organic fluids have condensation pressure higher than the atmospheric pressures and this avoids the infiltration of non-condensable gases in the condenser. The small differences in density organic fluid phase of liquid and vapor also enables the use of once-through boilers. This led to avoidance of using stream drums and simplified the operation of the whole plant. A simple plant system can be developed and less cost is needed when uses organic fluid compared to steam based cycle. In ORC, usage of deaerator is unnecessary but that is not the case for steam base cycle. Due to presence of oxygen, water deaerator or water treatment must be added to avert erosion. Because of low fluid density in the cycle low-pressure part, steam cycle also needs large turbines, heat exchangers and hydraulic diameter for pipes. Meanwhile, since organic fluid has higher density fluid, usage of compact appliances is allowed, especially in marine application, the available space for recovery plant of waste heat is restricted. Other than that, the enthalpy drop in ORC is much lower compared to steam cycle. The process in ORC can be done in a single stage with much simpler turbine compared to steam cycle which requires turbine with some expansion stages. ORC normally operated at much lower pressure levels and rarely exceeding 30 bars. Thus, ORC is beneficial in low to medium power range due to its cycle simplicity, less cost and stress level needed at boiler, easier to control

*DOI: http://dx.doi.org/10.5772/intechopen.89763*

**2.1 Organic Rankine cycle**

#### *A Recent Review in Performance of Organic Rankine Cycle (ORC) DOI: http://dx.doi.org/10.5772/intechopen.89763*

4% when heat recovery was done (efficiency of regeneration was 0.5). At the same time, if the efficiency was 0.7, an increase of 10% was achieved. From the numerical simulations, the best place for heat recovery is from hot gas before entering nozzle. Another research done by Li et al. [8] was to study the small-scale ORC system performance with low grade heat sources to provide electricity in various working state. The experiment setup includes normal ORC system components, for instance, turboexpander with high speed generator, finned tube condenser, ORC pump and plate evaporator. Results found that the turbine power and condenser heat output, ORC pump power and evaporator heat input, turbine isentropic, overall efficiencies and system thermal efficiencies rises when heat source temperature rises too. The fluid of ORC during superheat and pressure at turbine inlet were two crucial variables that kept constant with temperature heat source and pump speed of ORC.

#### **2.1 Organic Rankine cycle**

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

tal issues especially in aerospace industry.

**2. Aspects of recovery waste heat systems**

utilized to change heat into electricity. The organic fluids or refrigerants used in air conditioning systems accumulates (collect) heat from a volume of air and release it to different type of heat exchanger which increases the expansion of high vapor pressure in expander. The heat accumulated is transformed into mechanical power or electricity and therefore will help to increase the thermal efficiency and the overall performance of the engine. Thus, higher thrust could be obtained as less electrical power is needed from aircraft engine resulting lower engine bleed air [2, 5, 6]. Because of its thermodynamic properties, organic fluid is the best selection for low quality heat sources with temperatures below than 100°C [2]. By selecting proper working fluid for low waste heat recovery system and modeling an optimum design of heat exchanger configurations, the waste energy recovered through this ORC system could be maximized. Thus, designing a fuel-efficient and cheaper heat exchanger, ORC power plant can effectually utilize the economic and environmen-

Based on second law of thermodynamics, the efficiency of a process would not be 100% as there is no process that can entirely transform all amount of heat into work. The energy that is not used to produce work is being dissipated as heat at different temperatures, levels streams. On aircraft, half of the fuel energy lost through this way. Nevertheless, these sources of waste heat are everywhere; from this lost energy, only a part of it can be used to produce mechanical work or other purposes, where around 30% of the total waste heat could be changed to useful work. As the demand of aircraft is increasing vastly, the aviation industry has been the world center of attraction as new technologies are needed and maximum exploitation of fuel is a must. The conversion of heat energy to mechanical or electrical power depends on the characteristics of the source. Let say in an air conditioning system, an external hose is two or three degrees above the ambient temperature, it is a waste to recover that little amount of energy, however, this power leak will be an irreversibility process together with other similar leaks will decrease the thermal efficiency. This is called as waste heat and is an unused heat energy produced as a by-product of process of energy transformation, as a natural consequence on any non-adiabatic process from the thermodynamics law. Most of the available waste heat is low waste heat that can be used by an ORC which utilizes low boiling point organic fluid as working fluid, for example, toluene, hexane or pentane. Presently, there has not been any waste heat recovery (WHR) system added to an aircraft. Nevertheless, researchers suggest on adding WHR system to future engines and propose to make changes in current engines. However, it is a hassle to change the actual design of the engine as more expenses will be used in research, tests and certifications and a lot of heat source needs to be taken into account. Pasini et al. [7] analyzed the possibilities of heat recovery results in overall efficiency of an aircraft engine. A waste heat recovery system is modeled in a jet engine and a turbo propeller engine. Their project takes into account the nozzle works in off design state. The heat emitted influences greatly in the system performance. They also developed a numerical thermodynamic code to evaluate the positive impacts of waste heat recovery in a turboprop, a turbofan and a turbojet. The turbofan engine is of great interest due to large fraction of thrust is provided by cold flow, whilst gas generator supplies needed power. The authors then concluded that the enthalpy level ahead of exhaust nozzle of gas generator could be decreased without losing a lot of thrust. From the results of the calculations, it was found an increase of thermal efficiency about

**94**

ORC utilizes organic compound instead of water as a working fluid, generally, a refrigerant, a hydrocarbon such as pentane, butane, perfluorocarbon or silicon oil. The organic fluid's boiling point is much lesser compared to water and enable heat recovering at lesser temperatures instead of the steam Rankine cycle [9]. ORC's first commercial applications with medium-scale power plants for geothermal and solar applications were developed in the 9 late 1970s and 1980s. These days, over 200 ORC power plants are recognized with more than 1800 MWe installed and the technologies keep on increasing day by day [10]. Mostly, the plants were installed in biomass CHP application, geothermal plants and plants of WHR followed. The layout of ORC is much simpler compared to the steam cycle as there is no water vapor attached to the boiler, and a single heat exchanger could be utilized for the three processes of evaporation including preheating, vaporizing and superheating. ORC is able to use low grade heat sources than steam Rankine cycle. Since it could be utilized in lower temperature at the turbine inlet and reduce thermal stresses in the boiler. In regular steam plant systems, the performance cycle is at risk damage due to gaseous infiltrations that occur in sub atmospheric condensing pressure. In steam-based cycle, the usage of a single tube for evaporation is abstained due to large density difference that exists in between liquid and vapor phases. However, some organic fluids have condensation pressure higher than the atmospheric pressures and this avoids the infiltration of non-condensable gases in the condenser. The small differences in density organic fluid phase of liquid and vapor also enables the use of once-through boilers. This led to avoidance of using stream drums and simplified the operation of the whole plant. A simple plant system can be developed and less cost is needed when uses organic fluid compared to steam based cycle. In ORC, usage of deaerator is unnecessary but that is not the case for steam base cycle. Due to presence of oxygen, water deaerator or water treatment must be added to avert erosion. Because of low fluid density in the cycle low-pressure part, steam cycle also needs large turbines, heat exchangers and hydraulic diameter for pipes. Meanwhile, since organic fluid has higher density fluid, usage of compact appliances is allowed, especially in marine application, the available space for recovery plant of waste heat is restricted. Other than that, the enthalpy drop in ORC is much lower compared to steam cycle. The process in ORC can be done in a single stage with much simpler turbine compared to steam cycle which requires turbine with some expansion stages. ORC normally operated at much lower pressure levels and rarely exceeding 30 bars. Thus, ORC is beneficial in low to medium power range due to its cycle simplicity, less cost and stress level needed at boiler, easier to control and simpler usage of components [11].

#### *2.1.1 Difference between steam Rankine cycle and organic Rankine cycle*

Traditional steam Rankine Cycle utilizes water and higher pressure vapor as the main flow fluid in the cycle and mainly used in high temperature more than 500°C. However, ORC uses organic pure less boiling point or working fluid mixture mostly used in lower temperature process less than 500°C. Thus, ORC is more advantageous in recovering less temperature heat energy for various temperature ranges. Although steam Rankine cycle is the most common technology used in recovering heat process that converts heat into electricity, it is unfit for conditions of low temperature and pressure. This results from the need for high temperature and pressure in operation. If exhaust temperature and pressure are similar, the SRC exhaust steam enthalpy is greater and the heat from the cold sources increases [12].

Based on the **Figure 1** which is the T-s diagram above, two important differences could be found. Firstly, the curve of the organic fluid is abundantly vertical, meanwhile, water has negative curve of saturated vapor slope. Hence, when process of expansion finish, the limitation of vapor quality invisibles in ORC cycle and superheating vapor is unnecessary before the turbine inlet. Second, the gap in entropy between saturated liquid and saturated vapor for organic fluids is lesser. Thus, the enthalpy vaporization is lesser. The organic working fluid mass flow rate should therefore be greater than water to absorb equal thermal power in the evaporator, resulting in more pump consumptions [13].

#### *2.1.2 Supercritical organic Rankine cycle*

ORC can be operated in a subcritical or supercritical cycle. In supercritical cycle, the working fluid evaporation ends in supercritical area and the heat rejection in condenser occurs in the subcritical area. Many studies have been performed on the supercritical ORC. **Figure 2** provides the temperature and entropy changes of supercritical ORC.

Yagli et al. [14] modeled subcritical and supercritical ORC to recuperate exhaust gas waste heat of biogas fuelled CHP engine. Comparing with subcritical condition,

#### **Figure 1.**

*T-s diagram of the water saturation curves of water and a few typical ORC organic fluids applications [13].*

**97**

**Figure 2.**

is to modify the mass flow rate.

tions from industrial waste heat.

*2.1.3 Application of organic Rankine cycle*

*2.1.3.1 The combined heat and power (CHP) of biomass*

*A Recent Review in Performance of Organic Rankine Cycle (ORC)*

supercritical ORC shows greater performance. At constant pressure, supercritical ORC performance rise as turbine inlet temperature rises. The most excellent performed cycle net power, thermal efficiency and exergy efficiency are evaluated as 79.23 kW, 15.51 and 27.20% for subcritical and 81.52 kW, 15.93 and 27.76% for supercritical ORC, respectively. Guo et al. [15] studied the subcritical and transcritical ORC performance in regards to the evaporator pinch point locations. Found that transcritical ORCs gives higher performance as the heat source outlet temperatures lessen. Ran et al. [16] studied the impact of big transformation in the thermophysical properties of pseudocritical region. Utilizing network output, thermal efficiency and total vapor area an optimization method was found. The results showed that in transcritical ORC's, the thermophysical properties of the working fluid work at supercritical coefficient and logarithmic mean temperature difference (LMTD). Moloney et al. [17] analyzed the pressure effect to optimize the first law efficiency, second law's efficiency and net power of a supercritical ORC with 170–240°C turbine inlet temperature suitable for geothermal reservoirs of medium temperature. Found that supercritical cycle is much more efficient than subcritical cycle to optimize the plant efficiency. Chowdury et al. [18] presented an ORC simulation with different source of heat from the actual vehicle exhaust in supercritical state. The simulation shows that the key in transforming the operating temperature at the evaporator outlet

The ORC technology has been utilized broadly and applied in various industrial activities especially in biomass and geothermal application. Nevertheless, ORC technology has been increasing in solar thermal system and heat recovery applica-

There is widespread use of agricultural or industrial processes such as lumber or agricultural waste in biomass due to low energy density than the fossil fuels

*DOI: http://dx.doi.org/10.5772/intechopen.89763*

*Supercritical ORC temperature and entropy change [14].*

*A Recent Review in Performance of Organic Rankine Cycle (ORC) DOI: http://dx.doi.org/10.5772/intechopen.89763*

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

sources increases [12].

supercritical ORC.

rator, resulting in more pump consumptions [13].

*2.1.2 Supercritical organic Rankine cycle*

*2.1.1 Difference between steam Rankine cycle and organic Rankine cycle*

Traditional steam Rankine Cycle utilizes water and higher pressure vapor as the main flow fluid in the cycle and mainly used in high temperature more than 500°C. However, ORC uses organic pure less boiling point or working fluid mixture mostly used in lower temperature process less than 500°C. Thus, ORC is more advantageous in recovering less temperature heat energy for various temperature ranges. Although steam Rankine cycle is the most common technology used in recovering heat process that converts heat into electricity, it is unfit for conditions of low temperature and pressure. This results from the need for high temperature and pressure in operation. If exhaust temperature and pressure are similar, the SRC exhaust steam enthalpy is greater and the heat from the cold

Based on the **Figure 1** which is the T-s diagram above, two important differences could be found. Firstly, the curve of the organic fluid is abundantly vertical, meanwhile, water has negative curve of saturated vapor slope. Hence, when process of expansion finish, the limitation of vapor quality invisibles in ORC cycle and superheating vapor is unnecessary before the turbine inlet. Second, the gap in entropy between saturated liquid and saturated vapor for organic fluids is lesser. Thus, the enthalpy vaporization is lesser. The organic working fluid mass flow rate should therefore be greater than water to absorb equal thermal power in the evapo-

ORC can be operated in a subcritical or supercritical cycle. In supercritical cycle, the working fluid evaporation ends in supercritical area and the heat rejection in condenser occurs in the subcritical area. Many studies have been performed on the supercritical ORC. **Figure 2** provides the temperature and entropy changes of

Yagli et al. [14] modeled subcritical and supercritical ORC to recuperate exhaust gas waste heat of biogas fuelled CHP engine. Comparing with subcritical condition,

*T-s diagram of the water saturation curves of water and a few typical ORC organic fluids applications [13].*

**96**

**Figure 1.**

**Figure 2.** *Supercritical ORC temperature and entropy change [14].*

supercritical ORC shows greater performance. At constant pressure, supercritical ORC performance rise as turbine inlet temperature rises. The most excellent performed cycle net power, thermal efficiency and exergy efficiency are evaluated as 79.23 kW, 15.51 and 27.20% for subcritical and 81.52 kW, 15.93 and 27.76% for supercritical ORC, respectively. Guo et al. [15] studied the subcritical and transcritical ORC performance in regards to the evaporator pinch point locations. Found that transcritical ORCs gives higher performance as the heat source outlet temperatures lessen. Ran et al. [16] studied the impact of big transformation in the thermophysical properties of pseudocritical region. Utilizing network output, thermal efficiency and total vapor area an optimization method was found. The results showed that in transcritical ORC's, the thermophysical properties of the working fluid work at supercritical coefficient and logarithmic mean temperature difference (LMTD). Moloney et al. [17] analyzed the pressure effect to optimize the first law efficiency, second law's efficiency and net power of a supercritical ORC with 170–240°C turbine inlet temperature suitable for geothermal reservoirs of medium temperature. Found that supercritical cycle is much more efficient than subcritical cycle to optimize the plant efficiency. Chowdury et al. [18] presented an ORC simulation with different source of heat from the actual vehicle exhaust in supercritical state. The simulation shows that the key in transforming the operating temperature at the evaporator outlet is to modify the mass flow rate.

#### *2.1.3 Application of organic Rankine cycle*

The ORC technology has been utilized broadly and applied in various industrial activities especially in biomass and geothermal application. Nevertheless, ORC technology has been increasing in solar thermal system and heat recovery applications from industrial waste heat.

#### *2.1.3.1 The combined heat and power (CHP) of biomass*

There is widespread use of agricultural or industrial processes such as lumber or agricultural waste in biomass due to low energy density than the fossil fuels

and availability of heat and electricity, where biomass is suitable on off-grid case or unreliable grid connection. Local generation results in smaller power plants that exclude traditional steam cycles that in this power range are not profit-making.

**Figures 3** and **4** define the working principle of such cogeneration system, at a temperature from 150 to 320°C, heats from combustion is transmitted from the flue gases to the heat transfer fluid in two heat exchangers. When temperature lowers a little bit below 300°C, heat transfer fluid (thermal oil) is sent to the ORC loop to evaporate the working fluid. Then, the evaporated fluid expands, to preheat the liquid using recuperator and when temperature reached 90°C, the fluid condensed to produce hot water.

ORC efficiency is lesser compared to traditional steam cycles and gradually reduces for small scale units. To raise the overall energy conversion efficiency of plant, heat demand is needed and could be met through space heating or industrial processes (wood drying). Load of plant could be managed through on- site heat request or maximize power generation which includes additional wasting heats but increases, the full load operating hours per year.

From **Figure 3**, even though the CHP system's electrical efficiency is somewhat less (18%), the overall system efficiency is 88% greater than centralized power plants where most residual heat is lost. These gases need to be cooled to the least possible value, so that acid dew point could not be achieved and to lower the losses in flue gases. Two heat transfer loops are utilized to achieve this point (high and low temperature). The lower temperature loops are installed after the high flue temperature to lower the outlet temperature. Competitive technology in generating electric out of solid biofuels is biomass gasification where biomass changes into an organic gas mainly consisting H2, CO, CO2 and CH4. In order to remove solid particles, this synthetic gas is treated and filtered and finally burned in an ICE or in a gas turbine. Contrasting Biomass CHP's technology and costs with an ORC with gasification, gasification yields higher investment costs (75%), higher maintenance costs (200%) and more power-to-thermal ratio, where utilization is increase profit-making. ORC is an established technology meanwhile gasification plants are normally used as prototype in operation.

**99**

**Table 1.**

**Figure 4.**

*Biomass CHP ORC system working principle [13].*

*A Recent Review in Performance of Organic Rankine Cycle (ORC)*

Geothermal heat sources ranges from 10 to 300°C. The actual lower technological limit to generate electricity is about 80°C, and became less efficient with temperature less than 80°C and causes uneconomical geothermal plants. The potential of geothermal energy in Europe is shown in **Table 1** and indicates that low tempera-

For better production and injection, boreholes need to be drilled in the ground (**Figure 5**) to recover heat at an acceptable temperature. Then, the hot brine is pumped out of the first one and injected at a lower temperature in the second. Boreholes might be few thousand meters deep which results in working continuously for few months depend on the configuration of the geology and causes increasing share of drilling for geothermal plant cost investment (up to 70%) [19]. High auxiliary consumption is also characterized by low geothermal ORC: the pumps ingest 30–50% of the gross power output [20]. The brine pump together with a significant flow rate has to circulate the brine over large stances is the primary consumer. Working fluid of pump consumption is greater than higher temperature cycles, as the ratio of pump consumption to turbine output power

**Temperature MWth MWe** 65–90°C 147,736 10,462 90–120°C 75,421 7503 120–150°C 22,819 1268 150–225°C 42,703 4745 225–350°C 66,897 11,150

*Geothermal energy potential in Europe for different temperature ranges of heat sources [19].*

*DOI: http://dx.doi.org/10.5772/intechopen.89763*

ture sources have higher potential.

*2.1.3.2 Geothermal energy*

**Figure 3.** *Energy flows in a CHP system of biomass [13].*

#### *2.1.3.2 Geothermal energy*

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

profit-making.

to produce hot water.

increases, the full load operating hours per year.

normally used as prototype in operation.

and availability of heat and electricity, where biomass is suitable on off-grid case or unreliable grid connection. Local generation results in smaller power plants that exclude traditional steam cycles that in this power range are not

**Figures 3** and **4** define the working principle of such cogeneration system, at a temperature from 150 to 320°C, heats from combustion is transmitted from the flue gases to the heat transfer fluid in two heat exchangers. When temperature lowers a little bit below 300°C, heat transfer fluid (thermal oil) is sent to the ORC loop to evaporate the working fluid. Then, the evaporated fluid expands, to preheat the liquid using recuperator and when temperature reached 90°C, the fluid condensed

ORC efficiency is lesser compared to traditional steam cycles and gradually reduces for small scale units. To raise the overall energy conversion efficiency of plant, heat demand is needed and could be met through space heating or industrial processes (wood drying). Load of plant could be managed through on- site heat request or maximize power generation which includes additional wasting heats but

From **Figure 3**, even though the CHP system's electrical efficiency is somewhat less (18%), the overall system efficiency is 88% greater than centralized power plants where most residual heat is lost. These gases need to be cooled to the least possible value, so that acid dew point could not be achieved and to lower the losses in flue gases. Two heat transfer loops are utilized to achieve this point (high and low temperature). The lower temperature loops are installed after the high flue temperature to lower the outlet temperature. Competitive technology in generating electric out of solid biofuels is biomass gasification where biomass changes into an organic gas mainly consisting H2, CO, CO2 and CH4. In order to remove solid particles, this synthetic gas is treated and filtered and finally burned in an ICE or in a gas turbine. Contrasting Biomass CHP's technology and costs with an ORC with gasification, gasification yields higher investment costs (75%), higher maintenance costs (200%) and more power-to-thermal ratio, where utilization is increase profit-making. ORC is an established technology meanwhile gasification plants are

**98**

**Figure 3.**

*Energy flows in a CHP system of biomass [13].*

Geothermal heat sources ranges from 10 to 300°C. The actual lower technological limit to generate electricity is about 80°C, and became less efficient with temperature less than 80°C and causes uneconomical geothermal plants. The potential of geothermal energy in Europe is shown in **Table 1** and indicates that low temperature sources have higher potential.

For better production and injection, boreholes need to be drilled in the ground (**Figure 5**) to recover heat at an acceptable temperature. Then, the hot brine is pumped out of the first one and injected at a lower temperature in the second. Boreholes might be few thousand meters deep which results in working continuously for few months depend on the configuration of the geology and causes increasing share of drilling for geothermal plant cost investment (up to 70%) [19]. High auxiliary consumption is also characterized by low geothermal ORC: the pumps ingest 30–50% of the gross power output [20]. The brine pump together with a significant flow rate has to circulate the brine over large stances is the primary consumer. Working fluid of pump consumption is greater than higher temperature cycles, as the ratio of pump consumption to turbine output power

#### **Figure 4.**

*Biomass CHP ORC system working principle [13].*


#### **Table 1.**

*Geothermal energy potential in Europe for different temperature ranges of heat sources [19].*

**Figure 5.** *Geothermal ORC system working principle [13].*

('back work ratio') rose as evaporation temperature lowered. Geothermal heat sources temperature (>150°C) allow for CHP, where the condensing temperature is restricted to a higher temperature such as 60°C, enabling district heating uses cooled water. Thus, the overall efficiency of energy recovery rises with lower electrical efficiency expenses.

#### *2.1.3.3 Solar power plants*

Solar power concentration is the best technology on a linear or punctual collector that tracks and reflects the sun, transferring heat to high temperature fluid. Electricity is generated as heat is transmitted to a power cycle; electricity is generated. The three primary technologies of concentration are the parabolic platform, solar tower and the parabolic trough. Punctual concentrating technologies consist of parabolic dishes and solar towers, results in more concentration factor and greater temperatures. For solar towers, the Stirling engine (small-scale plants), the steam cycle or even the combined cycle is the best suited power cycles. Parabolic troughs operate at lesser temperature (300–400°C). Till today, they were combined to traditional steam Rankine cycles to generate electricity [21]. Geothermal or biomass power plants for example, steam cycles need higher temperatures, pressure and installed power to be more cost- effective. Organic Rankine cycle is a favorable technology that could lower the small scale of investment costs by working at lesser temperatures and reduce total installed power to kW scale. The working principle of the system is shown in **Figure 6**. As Fresnel linear technology need lower investment costs [22], they are suitable for ORCs but operate at lesser temperature.

Till recently, only a few of CSP plants with ORC are accessible on the market:

• In 2006, at Arizona, a 1MWe solar concentration of ORC power plant was accomplished. The ORC module utilizes n-pentane as the working fluid with 20% efficiency. On design point, the overall solar energy efficiency is 12.1% [23].

**101**

*A Recent Review in Performance of Organic Rankine Cycle (ORC)*

electrification in Lesotho by "STG International". To produce and integrate small size solar thermal technology with medium temperature collectors and an ORC to acquire economics equivalent to big installation of solar thermal is the objective of this project. This design intended to change or adding diesel generators in developing countries at off-grid areas through generating clean

At low temperature, most of the application in manufacturing industry reject load. Normally, the heat is enormous in large-scale plants, and could not be used again for on-site district heating. The heat then discharged into the atmosphere and

• Health/Environmental issues results from pollutants (CO2, NOX, SOX, HC) of

• Unbalance of aquatic equilibrium and negative effector biodiversity due to

These two types of pollution could be diminished by waste heat recovering. Moreover, it could provide on-site electricity to be consumed or sent it back to the grid. Normally, waste heat is recuperated through an intermediate heat transfer loop in such a system and used to evaporate the cycle's working fluid. In USA, power generation from industrial waste heat sources is approximately about 750 MWe [25]. Some industries have greater potential in recovery of waste heat. One of it, the cement industrial loses 40% of flue gas heat. These flue gases are placed at a temperature of 215–315°C after the preheater of limestone or in the clinker cooler [26]. CO2 released from the cement industry is 5% of the world's total CO2 emissions, half of the results from fossil fuels combustion in kilns [24]. Further possible industries include iron and steel industries (for example, 10% of CO2 emissions in China),

*DOI: http://dx.doi.org/10.5772/intechopen.89763*

power at lesser costs.

*Solar ORC system working principle [13].*

**Figure 6.**

*2.1.3.4 Mechanical and industrial heat recovery*

results in two types of pollution [24]:

flue gases.

rejection of heat.

• Few small-scale for the applications of remote-off grid were studied. The only proof of concept obtained is that 1KWe system installed for rural

*A Recent Review in Performance of Organic Rankine Cycle (ORC) DOI: http://dx.doi.org/10.5772/intechopen.89763*

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

('back work ratio') rose as evaporation temperature lowered. Geothermal heat sources temperature (>150°C) allow for CHP, where the condensing temperature is restricted to a higher temperature such as 60°C, enabling district heating uses cooled water. Thus, the overall efficiency of energy recovery rises with lower

Solar power concentration is the best technology on a linear or punctual collector that tracks and reflects the sun, transferring heat to high temperature fluid. Electricity is generated as heat is transmitted to a power cycle; electricity is generated. The three primary technologies of concentration are the parabolic platform, solar tower and the parabolic trough. Punctual concentrating technologies consist of parabolic dishes and solar towers, results in more concentration factor and greater temperatures. For solar towers, the Stirling engine (small-scale plants), the steam cycle or even the combined cycle is the best suited power cycles. Parabolic troughs operate at lesser temperature (300–400°C). Till today, they were combined to traditional steam Rankine cycles to generate electricity [21]. Geothermal or biomass power plants for example, steam cycles need higher temperatures, pressure and installed power to be more cost- effective. Organic Rankine cycle is a favorable technology that could lower the small scale of investment costs by working at lesser temperatures and reduce total installed power to kW scale. The working principle of the system is shown in **Figure 6**. As Fresnel linear technology need lower investment costs [22], they are suitable for ORCs but operate at lesser temperature. Till recently, only a few of CSP plants with ORC are accessible on the market:

• In 2006, at Arizona, a 1MWe solar concentration of ORC power plant was accomplished. The ORC module utilizes n-pentane as the working fluid with 20% efficiency. On design point, the overall solar energy efficiency is 12.1% [23].

• Few small-scale for the applications of remote-off grid were studied. The only proof of concept obtained is that 1KWe system installed for rural

electrical efficiency expenses.

*Geothermal ORC system working principle [13].*

*2.1.3.3 Solar power plants*

**Figure 5.**

**100**

electrification in Lesotho by "STG International". To produce and integrate small size solar thermal technology with medium temperature collectors and an ORC to acquire economics equivalent to big installation of solar thermal is the objective of this project. This design intended to change or adding diesel generators in developing countries at off-grid areas through generating clean power at lesser costs.

#### *2.1.3.4 Mechanical and industrial heat recovery*

At low temperature, most of the application in manufacturing industry reject load. Normally, the heat is enormous in large-scale plants, and could not be used again for on-site district heating. The heat then discharged into the atmosphere and results in two types of pollution [24]:


These two types of pollution could be diminished by waste heat recovering. Moreover, it could provide on-site electricity to be consumed or sent it back to the grid. Normally, waste heat is recuperated through an intermediate heat transfer loop in such a system and used to evaporate the cycle's working fluid. In USA, power generation from industrial waste heat sources is approximately about 750 MWe [25]. Some industries have greater potential in recovery of waste heat. One of it, the cement industrial loses 40% of flue gas heat. These flue gases are placed at a temperature of 215–315°C after the preheater of limestone or in the clinker cooler [26]. CO2 released from the cement industry is 5% of the world's total CO2 emissions, half of the results from fossil fuels combustion in kilns [24]. Further possible industries include iron and steel industries (for example, 10% of CO2 emissions in China),

refineries or chemical industries. Although their potential is higher and cost-effective (1000–2000 €/kWe), ORC recovery waste heat cycles have only 9–10% of the world's installed ORC plants compared to biomass CHP and geothermal units [10].

#### *2.1.3.5 Aircraft engine*

Perullo et al. [27] integrated an ORC to an engine for power generation. They mentioned the problem, as bypass ratio keep on growing and the engine cores becomes effective, the diameter of engine fan increases and the core size decreases which causes pneumatic offset needing greater percentage of the core flow and results in higher performance penalties. They tried to solve the problem by changing the pneumatic off-take to an electrical and used power generated to drive external air to the environmental control system (ECS). With the idea of no-bleed aircraft, performance penalties for shrinking cores and increased fan diameter are supposed to be eliminated and they had demonstrated that a rise in efficiency from 0.9 to 2.5% is possible. Boeing has also applied the no-bleed system; but using generator as the source of energy, not ORC where the generator works with energy taken from the APU and engines. As this application save fuel by 3%, this explain why they put this idea together in ORC rather than extracting energy from fuel, the waste heat could supply the energy needed. ORC is used due to the low quality of the range temperature. The WHR system is placed in the core jet exhaust of a turbofan engine. Conversely to land ORC systems, used in steam power plants for instance, an on-board ORC would suppose operating conditions that may vary continuously in the course of every few hours in external pressure and temperatures. The amount of heat extracted from the engine should be considered to avoid reduction of thrust. The system is distributed in the nozzle, the nose cowl and the Pylon. It uses R245fa as the working fluid having demonstrating highest thermal efficiency in a wide range of operating pressure. The MathCAD 2001 software was used to model the design to govern whether energy is enough to extract our f the exhaust gases to a power of 270 hp. motor. **Figure 7** below describes the ORC schematics.

The model was integrated on a CFM56-7B configuration and cruise conditions were used in it. Some assumptions were done; not analyzing the system with takeoff conditions as working fluid dissociates at high temperatures, heat is taken out of core exhaust flow before expanding in the nozzle, assuming a weight of 430 kg and was used to calculate the fuel burn reduction (0.9%) a TSFC reduced its value in 2% compared to the engine alone. It was also assumed that the ORC could produce

**103**

**Figure 8.**

essential.

*A Recent Review in Performance of Organic Rankine Cycle (ORC)*

greater power which is needed to drive the ECS air compressor and resulted in reduction of TSFC for 22%. Perullo et al. [27] concluded that an ORC WHR system could produce more power on the existing engine and can be utilized to supply sufficient power to a compressor driving air to the ECS. They suggested that the design system should be reconfigured to obtain the best results of fuel burn and take into account the need of an electric starting mechanism if the bleed system was removed in future research. The option of using the engine cowl or the anti-icing system in

Regenerative ORCs are designed where ORCs and turbine bleeding are integrated to a heat exchanger. The cycle heats up the working fluid upon infiltrating the evaporator which is almost similar to the ORC with recuperator. **Figure 8(a)** and **(b)** provides the schematic cycle and T-s diagram of regenerative cycle.

At 139°C of turbine inlet temperature, Le et al. [28] utilized a genetic method to optimize the first law and effectiveness of the system for diverse fluids. When examining, CO2 results the worst while recuperative cycle was discovered for greater efficiencies compared to simple cycle. Moloney et al. [17] studied the environmental fluids with critical temperature below 200°C in regenerative supercritical ORCs to upgrade the geothermal energy efficiency and noted that CO2 operates the lowest. The same purpose was enforced by Muhammad et al. [29] to the basic ORC; single and double stage regenerative ORC for applications of recovering waste heat. Studies showed that the single and double stage regenerative ORC has greater thermal efficiency with lower economic performance rather than the basic ORC.

Found that superheating of dry fluid negatively affects the ORC's efficiency while wet fluid positively affects the ORC's efficiency and isentropic fluid did not really affect ORC. Nevertheless, an experimental observation by [15] indicated that ORC with wet fluid superheat utilizing R245fa at 1.8°C and if the superheat rises to 8.7°C, the system is stable. Thus, even for dry working fluid, superheating is

Li et al. [30] conducted an experimental study to inquire the performance of a small-scale ORC system with low grade heat source to produce electricity at various

the wing s as the condenser of the ORC system was suggested as well.

*2.1.4 Organic Rankine cycle (ORC) with regenerator (RORC)*

*DOI: http://dx.doi.org/10.5772/intechopen.89763*

*2.1.5 Organic Rankine cycle with superheating*

*(a) ORC with regenerator and its (b) T-s diagram [28].*

**Figure 7.** *Aircraft engine ORC system working principle [27].*

*A Recent Review in Performance of Organic Rankine Cycle (ORC) DOI: http://dx.doi.org/10.5772/intechopen.89763*

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

*2.1.3.5 Aircraft engine*

refineries or chemical industries. Although their potential is higher and cost-effective (1000–2000 €/kWe), ORC recovery waste heat cycles have only 9–10% of the world's installed ORC plants compared to biomass CHP and geothermal units [10].

Perullo et al. [27] integrated an ORC to an engine for power generation. They mentioned the problem, as bypass ratio keep on growing and the engine cores becomes effective, the diameter of engine fan increases and the core size decreases which causes pneumatic offset needing greater percentage of the core flow and results in higher performance penalties. They tried to solve the problem by changing the pneumatic off-take to an electrical and used power generated to drive external air to the environmental control system (ECS). With the idea of no-bleed aircraft, performance penalties for shrinking cores and increased fan diameter are supposed to be eliminated and they had demonstrated that a rise in efficiency from 0.9 to 2.5% is possible. Boeing has also applied the no-bleed system; but using generator as the source of energy, not ORC where the generator works with energy taken from the APU and engines. As this application save fuel by 3%, this explain why they put this idea together in ORC rather than extracting energy from fuel, the waste heat could supply the energy needed. ORC is used due to the low quality of the range temperature. The WHR system is placed in the core jet exhaust of a turbofan engine. Conversely to land ORC systems, used in steam power plants for instance, an on-board ORC would suppose operating conditions that may vary continuously in the course of every few hours in external pressure and temperatures. The amount of heat extracted from the engine should be considered to avoid reduction of thrust. The system is distributed in the nozzle, the nose cowl and the Pylon. It uses R245fa as the working fluid having demonstrating highest thermal efficiency in a wide range of operating pressure. The MathCAD 2001 software was used to model the design to govern whether energy is enough to extract our f the exhaust gases to a power of 270 hp. motor. **Figure 7** below describes the ORC schematics. The model was integrated on a CFM56-7B configuration and cruise conditions were used in it. Some assumptions were done; not analyzing the system with takeoff conditions as working fluid dissociates at high temperatures, heat is taken out of core exhaust flow before expanding in the nozzle, assuming a weight of 430 kg and was used to calculate the fuel burn reduction (0.9%) a TSFC reduced its value in 2% compared to the engine alone. It was also assumed that the ORC could produce

**102**

**Figure 7.**

*Aircraft engine ORC system working principle [27].*

greater power which is needed to drive the ECS air compressor and resulted in reduction of TSFC for 22%. Perullo et al. [27] concluded that an ORC WHR system could produce more power on the existing engine and can be utilized to supply sufficient power to a compressor driving air to the ECS. They suggested that the design system should be reconfigured to obtain the best results of fuel burn and take into account the need of an electric starting mechanism if the bleed system was removed in future research. The option of using the engine cowl or the anti-icing system in the wing s as the condenser of the ORC system was suggested as well.

#### *2.1.4 Organic Rankine cycle (ORC) with regenerator (RORC)*

Regenerative ORCs are designed where ORCs and turbine bleeding are integrated to a heat exchanger. The cycle heats up the working fluid upon infiltrating the evaporator which is almost similar to the ORC with recuperator. **Figure 8(a)** and **(b)** provides the schematic cycle and T-s diagram of regenerative cycle.

At 139°C of turbine inlet temperature, Le et al. [28] utilized a genetic method to optimize the first law and effectiveness of the system for diverse fluids. When examining, CO2 results the worst while recuperative cycle was discovered for greater efficiencies compared to simple cycle. Moloney et al. [17] studied the environmental fluids with critical temperature below 200°C in regenerative supercritical ORCs to upgrade the geothermal energy efficiency and noted that CO2 operates the lowest. The same purpose was enforced by Muhammad et al. [29] to the basic ORC; single and double stage regenerative ORC for applications of recovering waste heat. Studies showed that the single and double stage regenerative ORC has greater thermal efficiency with lower economic performance rather than the basic ORC.

#### *2.1.5 Organic Rankine cycle with superheating*

Found that superheating of dry fluid negatively affects the ORC's efficiency while wet fluid positively affects the ORC's efficiency and isentropic fluid did not really affect ORC. Nevertheless, an experimental observation by [15] indicated that ORC with wet fluid superheat utilizing R245fa at 1.8°C and if the superheat rises to 8.7°C, the system is stable. Thus, even for dry working fluid, superheating is essential.

Li et al. [30] conducted an experimental study to inquire the performance of a small-scale ORC system with low grade heat source to produce electricity at various

**Figure 8.** *(a) ORC with regenerator and its (b) T-s diagram [28].*

working state. It was found that the fluid of ORC during superheat and pressure at the turbine inlet were two main variables able to be managed with temperature of heat source and speed of the ORC pump. It was also found that superheat and internal heat exchanger are crucial for ORC from both perspectives of thermodynamic and techno-economic. Roy et al. studied the consequences of superheat and recovering on ORC system at certain degree of superheat [31]. Note that Guo et al. [15] argued if the superheat coupled with an internal heat exchanger, greater development could be done. Zhang et al. inquires the consequences of superheat and internal heat exchanger on three ORC designs' thermos-economic performance from fluid properties and heat sources. It has been discovered that the thermoeconomic performance of internal heat exchanger ORC with dry surpasses the wet fluid as temperature of heat source load increases [32]. Brizard et al. [33] suggested preventing condensation drops during operation of superheating; the inlet of expander must exceed 20°C. Radulovic et al. [34] mentioned that superheat is important in cycle especially in wet fluids. As the temperature of superheater rises, the cycle efficiency also rises and the chance of the working fluid condenses during pressure drop inside turbine, resulting in corrosion and efficiency drop is lesser. To get a higher efficiencies and net power output, superheating is important to prevent wet expansion. Feng et al. [35] found that rises the superheat degree assure the decrease in mean heat transfer temperature difference in superheating area of evaporator causes decreasing of overall heat transfer area, and decrease in the investment cost of the system. It was also found that outlet temperature of evaporator and superheat degree gives good feedback on the efficiency of exergy. Li et al. [30] construct an investigation on the experimental of a small-scale ORC system under designated working state for the recovery application of low-grade thermal energy. The reaction between condenser cooling water temperatures and superheat.

R245fa at turbine inlet were measured and analyzed on the performance of the system. The outcomes show that when evaporating pressure is constant, superheat at the inlet of expander gives negative feedback on the turboexpander and performance of the at some temperatures of cold water. In conclusion, superheat is crucial in assuring an efficient and safe system operation. Bianchi et al. [36] presented an experimental micro-ORC setup for low-temperature application by implementing a test bench to acquire data for the energy system characterization. From the results, it was found that for the tested working points, efficiency is from 2.9 to 4.4% and increases as degree of superheating decreases. Ismail et al. [37] concluded that utilizing superheated vapor in the system with internal heat exchanger results in increasing of thermal efficiency ORC. The mass flow rate required for the system together with superheated vapor is lower than the saturated vapor system. Thus, superheated is essential to lower the mass flow rate, and enhanced the performance of the system with presence of internal heat exchanger.

#### **3. Conclusion**

This chapter presents a comprehensive review on the developments of organic Rankine cycle (ORC) systems that have been used for power generation by using a waste heat source. This review also highlights more on the different configurations of ORCs used, depending on their applications. From here, we could conclude that superheating and the condition of the organic working fluid are crucial in ORC system from the thermodynamic point of views. Therefore, this study plans to investigate the design of an ORC model with better output power by modifying the configurations and adding a superheating device and also to study the effect of using organic dry working fluid (R245fa) at supercritical condition.

**105**

**Author details**

Malaysia, Serdang, Malaysia

Syamimi Saadon1,2\* and Salmah Md Saiful Islam1

University Putra Malaysia, Serdang, Selangor

provided the original work is properly cited.

\*Address all correspondence to: mimisaadon@upm.edu.my

1 Department of Aerospace Engineering, Faculty of Engineering, University Putra

2 Faculty of Engineering, Aerospace Manufacturing Research Centre (AMRC),

© 2019 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

*A Recent Review in Performance of Organic Rankine Cycle (ORC)*

*DOI: http://dx.doi.org/10.5772/intechopen.89763*

*A Recent Review in Performance of Organic Rankine Cycle (ORC) DOI: http://dx.doi.org/10.5772/intechopen.89763*

*Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications*

of the system with presence of internal heat exchanger.

This chapter presents a comprehensive review on the developments of organic Rankine cycle (ORC) systems that have been used for power generation by using a waste heat source. This review also highlights more on the different configurations of ORCs used, depending on their applications. From here, we could conclude that superheating and the condition of the organic working fluid are crucial in ORC system from the thermodynamic point of views. Therefore, this study plans to investigate the design of an ORC model with better output power by modifying the configurations and adding a superheating device and also to study the effect of

using organic dry working fluid (R245fa) at supercritical condition.

working state. It was found that the fluid of ORC during superheat and pressure at the turbine inlet were two main variables able to be managed with temperature of heat source and speed of the ORC pump. It was also found that superheat and internal heat exchanger are crucial for ORC from both perspectives of thermodynamic and techno-economic. Roy et al. studied the consequences of superheat and recovering on ORC system at certain degree of superheat [31]. Note that Guo et al. [15] argued if the superheat coupled with an internal heat exchanger, greater development could be done. Zhang et al. inquires the consequences of superheat and internal heat exchanger on three ORC designs' thermos-economic performance from fluid properties and heat sources. It has been discovered that the thermoeconomic performance of internal heat exchanger ORC with dry surpasses the wet fluid as temperature of heat source load increases [32]. Brizard et al. [33] suggested preventing condensation drops during operation of superheating; the inlet of expander must exceed 20°C. Radulovic et al. [34] mentioned that superheat is important in cycle especially in wet fluids. As the temperature of superheater rises, the cycle efficiency also rises and the chance of the working fluid condenses during pressure drop inside turbine, resulting in corrosion and efficiency drop is lesser. To get a higher efficiencies and net power output, superheating is important to prevent wet expansion. Feng et al. [35] found that rises the superheat degree assure the decrease in mean heat transfer temperature difference in superheating area of evaporator causes decreasing of overall heat transfer area, and decrease in the investment cost of the system. It was also found that outlet temperature of evaporator and superheat degree gives good feedback on the efficiency of exergy. Li et al. [30] construct an investigation on the experimental of a small-scale ORC system under designated working state for the recovery application of low-grade thermal energy. The reaction between condenser cooling water temperatures and superheat. R245fa at turbine inlet were measured and analyzed on the performance of the system. The outcomes show that when evaporating pressure is constant, superheat at the inlet of expander gives negative feedback on the turboexpander and performance of the at some temperatures of cold water. In conclusion, superheat is crucial in assuring an efficient and safe system operation. Bianchi et al. [36] presented an experimental micro-ORC setup for low-temperature application by implementing a test bench to acquire data for the energy system characterization. From the results, it was found that for the tested working points, efficiency is from 2.9 to 4.4% and increases as degree of superheating decreases. Ismail et al. [37] concluded that utilizing superheated vapor in the system with internal heat exchanger results in increasing of thermal efficiency ORC. The mass flow rate required for the system together with superheated vapor is lower than the saturated vapor system. Thus, superheated is essential to lower the mass flow rate, and enhanced the performance

**104**

**3. Conclusion**

### **Author details**

Syamimi Saadon1,2\* and Salmah Md Saiful Islam1

1 Department of Aerospace Engineering, Faculty of Engineering, University Putra Malaysia, Serdang, Malaysia

2 Faculty of Engineering, Aerospace Manufacturing Research Centre (AMRC), University Putra Malaysia, Serdang, Selangor

\*Address all correspondence to: mimisaadon@upm.edu.my

© 2019 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

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**108**

## *Edited by Silvia Lasala*

This book comprises five chapters on developed research activities on organic Rankine cycles. The first section aims to provide researchers with proper modelling (Chapter 1) and experimental (Chapter 2) tools to calculate and empirically validate thermophysical properties of ORC working fluids. The second section introduces some theoretical and experimental studies of organic Rankine cycles for waste heat recovery applications: a review of different supercritical ORC (Chapter 3), ORC for waste heat recovery from fossil-fired power plants (Chapter 4), the experimental detailed characterization of a small-scale ORC of 3 kW operating with either pure fluids or mixtures (Chapter 5).

Published in London, UK © 2020 IntechOpen © penkanya / iStock

Organic Rankine Cycles for Waste Heat Recovery - Analysis and Applications

Organic Rankine Cycles for

Waste Heat Recovery

Analysis and Applications

*Edited by Silvia Lasala*