Design and testing of the Organic Rankine Cycle

Design and testing of the Organic Rankine Cycle

Energy 26 (2001) 239–251 www.elsevier.com/locate/energy Design and testing of the Organic Rankine Cycle Takahisa Yamamoto b a,* , Tomohiko Furuhata...

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Energy 26 (2001) 239–251 www.elsevier.com/locate/energy

Design and testing of the Organic Rankine Cycle Takahisa Yamamoto b

a,*

, Tomohiko Furuhata b, Norio Arai b, Koichi Mori

c

a Department of Chemical Engineering, Nagoya University, Chikusa-ku, Nagoya, Japan 464-8603 Research Center for Advanced Energy Conversion, Nagoya University, Chikusa-ku, Nagoya, Japan 464-8603 c Toyo Technica Co. Ltd., Sakae, Naka-ku, Nagoya, Japan 460-0008

Received 16 March 2000

Abstract We propose a new type of environmentally friendly system called the “Organic Rankine Cycle” (ORC) in which low-grade heat sources are utilized. This system combines a circulated thermosyphon with a turbine system. The working fluid used in this study is an organic substance which has a low boiling point and a low latent heat for using low-grade heat sources. A numerical simulation model of the ORC is made in order to estimate its optimum operating conditions. An experimental apparatus is also made in this study. From the numerical simulation, it is suggested that HCFC-123 gives higher turbine power than water which is a conventional working fluid, and operating conditions where saturated vapor at the turbine inlet would give the best performance. From the experimental results, HCFC-123 improves the cycle performance drastically. In addition, the turbine made for trial use in this study gives good performance.  2001 Elsevier Science Ltd. All rights reserved.

1. Background In recent years, accelerated consumption of fossil fuels has caused many serious environmental problems such as global warming, ozone layer destruction and atmospheric pollution. New energy conversion technologies are required to utilize energy resources suitable for power generation without causing environmental pollution. Low-grade heat sources are considered as candidates for the new energy sources. Solar heat, waste heat and geothermal energy are typical examples for low-grade heat sources with their available temperatures ranging between 60 and 200°C. The use of such low-grade heat sources as an alternative energy source generating electricity has long been investigated using power turbine cycles [1–6]. Little attention has been given to these systems because of the low thermal efficiency of the Rankine cycles which are operated at low * Corresponding author. Tel.: +81-52-789-3914; fax: +81-52-789-3910. E-mail address: [email protected] (T. Yamamoto).

0360-5442/01/$ - see front matter  2001 Elsevier Science Ltd. All rights reserved. PII: S 0 3 6 0 - 5 4 4 2 ( 0 0 ) 0 0 0 6 3 - 3

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Nomenclature Cp h m n Pin Pout Qin Qout T Tin Tout T1 T2 Wp Wt Wt(exp) p heff hpump ht g r

Isobaric specific heat capacity at each stage [kJ/kg·K] Specific enthalpy at each stage [kJ/kg] Mass flow rate [kg/s] Rotation speed of turbine shaft [rpm] Turbine inlet pressure [MPa] Turbine outlet pressure [MPa] Evaporator input [kW] Condenser output [kW] Torque [Nm] Turbine inlet temperature [K] Turbine outlet temperature [K] Temperature at stage 1 [K] Temperature at stage 2 [K] Pump work [kW] Turbine output (numerical simulation) [kW] Turbine output (experiment) [W] Pressure ratio, Pin/Pout [—] Effective efficiency of turbine [%] Pump efficiency [%] Turbine efficiency [%] Specific heat ratio of working fluid at each state [—] Density at each stage [kg/m3]

temperatures. Such systems, however, have a simple structure at low cost, and existing technologies are applicable. We propose a new type of an environmentally friendly system called an Organic Rankine Cycle (ORC) to utilize low-grade heat sources [7]. This system uses an organic substance as a working fluid in order to utilize low-grade heat sources and consists of an evaporator (heating area), a turbine and a condenser (cooling area). The features of this system are its small size, no emissions of exhaust gases such as CO, CO2, NOX and other atmospheric pollutants. However, the most important feature is utilization of various kinds of heat sources when low-grade heat sources are used for power generation. The working fluid is heated at the evaporator and is converted into a high-pressure vapor. The vapor is changed into a low pressure vapor via the turbine. At the same time, the thermal energy of the high pressure vapor is converted into mechanical energy. Subsequently, the resultant mechanical energy is converted into electricity via the generator which is connected to the turbine shaft. The vapor emerging from the turbine outlet is fed to the condenser and then returns to the liquid phase. Generally, the turbine inlet temperature (TIT), the pressure ratio in the turbine and the mass flow rate in the turbine inlet are among the most important factors in the turbine system. In the conventional turbine systems, TIT is increased to improve the turbine output. However, this is not the case when low-grade

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Table 1 Thermal properties of working fluids Working fluid

Molecular weight

Boiling point [K]

Liquid density [kg/m3]

Latent heat [kJ/kg]

Water (H2O) HCFC-123 (CHCl2– CF3)

18.0 152.9

373.15 300.85

997.0 1462.2

2257.00 168.41

Specific heat ratio 1.33 1.11

heat sources are used. In addition, the thermal efficiency is low for Rankine cycles operating at low temperatures. These are serious and difficult problems to overcome. Therefore, the organic substance selected for the working fluid must have low latent heat and high density. Such properties are preferable to increase the turbine inlet mass flow rate. Heretofore, we have selected the organic substance which is suitable for this system. Consequently, it has been determined that HCFC-123 gives the best characteristics over other candidates such as water and methanol [7–9]. In this study, an attempt to find the optimum conditions needed for the working fluid to yield higher turbine output of ORC is discussed from a theoretical stand point. An experimental apparatus has also been made and tested. HCFC-123 and water are chosen as working fluids. Table 1 lists the thermal properties of these working fluids. 2. Numerical simulations 2.1. Numerical simulation model In order to determine the optimum operating conditions, we have carried out a thermodynamic analysis of the ORC using a process simulator HYSYS (Hyprotech Co., Canada). This simulator is useful for thermodynamic analysis, especially steady state condition. The simulation flow diagram is shown in Fig. 1. This simulation requires such conditions as the evaporator input, Qin,

Fig. 1. Simulation flow diagram.

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the turbine inlet pressure Pin (=P2) and the turbine outlet pressure Pout. The turbine inlet vapor is set to be a superheated or a saturated phase. The working fluid passing through the condenser is assumed to be a saturated liquid (T1, P1). The circulating mass flow rate (m) was determined by the above conditions. The cycle efficiency is kept constant under the condition of each pressure ratio (p=Pin/Pout). In this simulation, each phase of the working fluid is expressed by the equations given below [10]. 2.1.1. Pump phase (stages 1–2) The circulation pump is the driving mechanism of the proposed system. The working fluid (saturated liquid) leaving the condenser at low pressure P1 regains high pressure here to P2. The working fluid is pumped back into the evaporator. The circulation pump Wp is calculated by the following equation: (P2−P1)·m ˙ W p⫽ r·hpump

(1)

where r and hpump denote the density of working fluid (saturated condition) and the adiabatic efficiency of the circulation pump, respectively. The specific enthalpy of the working fluid at the circulation pump outlet, h2, is h2⫽h1⫹Wp/m ˙

(2)

where, h1 is the specific enthalpy of the working fluid at the circulation pump inlet. 2.1.2. Evaporator phase (stages 2–3) The evaporator heats the working fluid at the pump outlet to the turbine inlet condition. Then the working fluid is at superheated or saturated vapor state. The evaporator outlet condition of the working fluid is given by the following equation: hin⫽h2⫹(Qin/m ˙)

(3)

where Qin, hin and h2 are the evaporator input, the specific enthalpy at the evaporator inlet and at the outlet, respectively.

Fig. 2. Effects of the evaporator input on turbine output for (a) HCFC-123 and (b) water.

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Fig. 3.

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Effects of TIT on the mass flow rate at the turbine inlet and output for (a) HCFC-123 and (b) water.

Fig. 4. Effects of the pressure ratio on cycle efficiency.

2.1.3. Turbine phase (stages 3–4) The superheated or saturated vapor of the working fluid passes through the turbine to generate mechanical power. After the vapor expands, it is depressurized by the turbine blades. The vapor comes out of the turbine at lower pressure Pout and at low temperature Tout. Then, the turbine output is given by: 1−g

Wt⫽m ˙ ·Cp·ht·Tin[1⫺p g ]⫽m ˙ ·ht·(hin⫺hout)

(4)

where Cp and Tin denote the isobaric specific heat capacity and the turbine inlet temperature, respectively.

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2.1.4. Condenser phase (stages 4–1) The vapor of the working fluid goes through a constant pressure phase change process in the condenser into a state of saturated liquid, rejecting the latent heat into the environment or the condenser coolant. The pressure of the working fluid within the condenser is equal to the Rankine cycle lower pressure, P1, and the temperature is equal to the saturation temperature of the pressure, P1. The condenser load, Qout, which is the rate of latent heat rejection from condensing working fluid, can be calculated from the following equation; ˙ ·(hout⫺h1) Qout⫽m

(5)

2.1.5. Cycle efficiency Generally, various parameters are used to evaluate system performance, such as thermal efficiency and the coefficient of performance (COP). In this study the thermal efficiency was adopted for system evaluation. The cycle efficiency is calculated using the following equation. Wt−Wp ⫻100 h⫽ Qin

(6)

In fact, the pressure drops occur in the turbine and condenser processes, and the pressure rises occur in the evaporator process. However, as explained above, the pressure rise occurs in the circulation pump only, and the pressure drop occurs in the turbine only in order to be clear of the thermodynamic analysis and the characteristics of the working fluids. Therefore this analysis assumed steady state, no heat loss and pressure drop in the entire system. This means that the temperature at the evaporator outlet is equal to that at the turbine inlet. The adiabatic efficiencies of the turbine and the pump are both assumed to be 85%. 2.2. Numerical simulation results Fig. 2 shows the dependence of the turbine output on the evaporator input under the saturated vapor conditions at the turbine inlet. In the case of each working fluid, the turbine output increases proportionally with the increase in the evaporator input. These results can be understood as follows. The turbine output is calculated by Eq. (4). As for this equation, turbine inlet temperature Tin, pressure ratio p, specific heat ratio g and isobaric specific heat capacity Cp are constant, and only the mass flow rate increases proportionally as the evaporator input increases. Accordingly the relationship between the turbine output and the evaporator input is expressed as a straight line. These simulation results tell us that HCFC-123 gives the best performance, even though the ORC uses low-grade heat sources. Fig. 3 shows the effects of TIT on the turbine inlet mass flow rate and the turbine output, while the evaporator input is fixed at 12 kW. As for the case of water, the turbine output increases as the mass flow rate decreases. Conversely, it only slightly decreases in the case of HCFC-123. The reason for the decreased turbine output in the case of HCFC-123 is as follows. The latent heat of HCFC-123 is about one-tenth of that of water. Thus, TIT and the turbine inlet mass flow

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Fig. 5. Schematic drawing of the experimental apparatus.

rate should be increased to raise the turbine output. In this simulation, however, TIT increases as the mass flow rate decreases, or the mass flow rate increases as TIT decreases. It is therefore understood that the turbine output decreases when the mass flow rate in the turbine inlet decreases rapidly. If a working fluid with a low latent heat is used, the saturated vapor at the turbine inlet would give the best operating conditions. Fig. 4 presents the dependency of the pressure ratio on the cycle efficiency for both water and HCFC-123. From the point of view of the actual system, these values of the cycle efficiencies are not strict, because this analysis is only an ideal calculation. However, these results tell us the characteristics of each working fluid and the effect of the pressure ratio. The cycle efficiency of each working fluid increases as pressure ratio rises, and HCFC-123 has a higher efficiency compared with water. Therefore raising the pressure ratio and HCFC-123 give us much higher system performance. 3. Design and test of the experimental apparatus 3.1. Experimental apparatus and experimental conditions We have made an experimental apparatus and tested it. Fig. 5 shows the schematic diagram of the apparatus which consists of a tank, a circulation pump, an evaporator, a turbine and a condenser. Electric heaters (10 kW×2) are installed instead of the low-grade heat sources in the

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Fig. 6.

Schematic drawing of (a) the micro-turbine and (b) the nozzle.

evaporator which heats the working fluid to saturated conditions. A shell and tube type heat exchanger is used as a condenser. Water level sensors are installed in both the evaporator and the tank. Thermocouples and pressure sensors are set up at the inlet and outlet of each unit. A computer sums up data from these sensors, and controls the electric heaters and the circulation pump as the water level of the evaporator is always constant from its base. The turbine has not been developed at the low temperature mentioned above, using a large density working fluid, until now. A micro-turbine and a nozzle are made in order to discuss the optimum design of the turbine blade shape. Fig. 6 shows the schematic diagrams of the microturbines and the nozzle. The turbine made from aluminum is radial type with 18 blades, 30 mm in diameter and 4.5 mm thick. The nozzle is 46 mm in external diameter and 31 mm in inside diameter. Fig. 7 shows the efficiency of this turbine which is a typical and useful parameter to evaluate turbine performance in the design stage [11]. This parameter is calculated by means of the turbine blade. The peak point of the diagram efficiency shifts to a higher rotation speed as

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Fig. 7.

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Diagram efficiency of the micro-turbine.

the pressure ratio rises. A turbine shaft is directly connected to both a tachometer and a torquemeter. Then the turbine output is calculated as follows; Wt(exp)⫽

2pn ·T 60

(7)

where n denotes the rotation speed of the turbine shaft which is measured by the tachometer, and T denotes the turbine torque which is measured by the torque-meter. This study carried out experiments with four conditions of the evaporator input (16.5, 17.3, 18.2, 19.9 kW) for water, and three conditions of the evaporator input (9.7, 11.4, 13.0 kW) for HCFC-123. When the rotation speed of the turbine shaft is stable, then measurements of temperature, pressure, torque, rotation speed and volume flow rate at each point are started. We have evaluated the performance of this apparatus using experimental data such as measured turbine output, pressure, volume flow rate and temperature at the turbine inlet and outlet with each evaporator input condition. 3.2. Experimental results We have carried out the experiments using two kinds of working fluids, namely, HCFC-123 and water. This experimental apparatus has been tested for more than 6 h a day, for about 1 week, without any mechanical problems. Then maximum cycle efficiency was 1.25% in the case of using HCFC-123. Fig. 8 shows HCFC-123 and water temperatures, pressures and volume flow rates at the turbine inlet and outlet, respectively. For both HCFC-123 and water, the temperature, pressure and volume flow rate at the turbine inlet and outlet gradually increase. However, only turbine inlet pressure, in the case of HCFC-123, increases rapidly because the latent heat of HCFC-123 (see Table 1) is one-tenth that of water. HCFC-123 is able to run this system under low evaporator input conditions, and it also shows a high performance compared with water. The turbine outlet volume flow rate is larger than the turbine inlet for both working fluids. The main reason is an expansion

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Fig. 8. (a) Temperature, (b) volume flow rate, and (c) pressure of the experimental data at the turbine inlet and outlet.

of the working fluid, when it passes through the turbine. In spite of the volume flow rate of HCFC-123 being smaller than water, the maximum mass flow rates of HCFC-123 and water are estimated to be 100 kg/h and 18 kg/h, respectively. Figs. 9 and 10 show the characteristics of turbine output when the working fluids are water and HCFC-123, respectively. For both cases, there are optimum operating conditions in the relationship between the turbine output and the rotation speed. These optimum operating conditions shift to a higher rotation speed as the evaporator input rises. In the case of water with an evaporator input of 19.9 kW, the maximum rotation speed and the turbine output are 45,000 rpm and 150 W, respectively. As for HCFC-123, when the evaporator input is 13.0 kW, the maximum rotation speed and the turbine output are 35,000 rpm and 150 W, respectively. The turbine performance is discussed by means of the experimental results and process simulator. In order to evaluate the turbine, this study focuses effective efficiency heff which is calculated as the following equation: heff⫽Wt(exp)/Wt

(8)

where Wt denotes theoretical turbine output which is calculated using both experimental data and Eq. (4). Fig. 11 shows the characteristics of effective efficiency. In the case of using water, the effective efficiency gradually decreases as evaporator input rises. However, it increases in the

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Fig. 9. Characteristics of the micro-turbine in the case of water [evaporator input: (a) 16.5 kW, (b) 17.3 kW, (c) 18.2 kW, (d) 19.9 kW].

case of HCFC-123. This is the reason why the volume flow rate of water is three times as large as that of HCFC-123. Thus, in the case of using water, friction loss may be dominant when the working fluid goes through the turbine. From the above results, HCFC-123 improves the cycle performance drastically. The turbine made for trial use has good performance under the conditions using HCFC-123.

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Fig. 10. Characteristics of the micro-turbine in the case of HCFC-123 [evaporator input: (a) 9.7 kW, (b) 11.4 kW, (c) 13.0 kW].

4. Conclusions The performance and characteristics of the closed type Organic Rankine Cycle (ORC) using working fluids such as HCFC-123 and water have been investigated theoretically and experimentally in this study. From the numerical simulations, it is apparent that for optimum operating conditions of water, increasing the turbine inlet temperature gives high turbine power. Conversely, the best operating conditions for HCFC-123 exist when the turbine inlet temperature is as low as possible above the boiling point of the working fluid. If a working fluid with a low latent heat is used, the saturated vapor at the turbine inlet would give the best operating condition. From the

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Fig. 11.

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Effective efficiency of the micro-turbine [(a) HCFC-123, (b) water].

experiments, there are optimum operating conditions between the rotation speed and the turbine output. In addition, it is demonstrated that HCFC-123 improves the cycle performance drastically. The turbine made for trial use in this study gives good performance under the conditions for HCFC-123. It may be concluded from the above, that the ORC can be applied to low-grade heat sources and HCFC-123 is able to improve the ORC performance significantly. References [1] Giampaolo M, Shukuru M. Energy control for a flat plate collector/Rankine cycle solar power system. J Solar Energy Engng 1991;113(2):89–97. [2] Nugyen T, Johnson P, Mochizuki M. Design, manufacture and testing of a closed cycle thermosyphon Rankine engine. Heat Recovery Syst CHP 1995;5:333–8. [3] Johnson P, Akbarzadeh A. Thermosyphon Rankine engine for energy and waste heat applications. Heat Pipe Technol 1993;2:254–8. [4] Wolpert JL, Riffat SB. Solar powered Rankine engine for domestic application. Appl Thermal Engng 1996;16:281–9. [5] Selahattin G. Design parameters of a solar-driven heat engine. Energy Sources 1996;18(1):37–42. [6] Best FG, Riffat SB. Miniature combined heat and power system. Renewable Energy 1995;6:49–51. [7] Yamamoto T, Furuhata T, Arai N, Mori K. Characteristics of Organic Rankine Cycle for solar and waste heat. In: Arai N, editor. RAN’98 International Symposium Nagoya. Nagoya: Seikodo, 1998:326–7. [8] Yamamoto T, Furuhata T, Arai N, Mori K. Simulation of Organic Rankine Cycle. In: Ishida N, editor. ECOS’99 International Symposium Tokyo. Tokyo: Sanyodo, 1999:254–7. [9] Kuo C-H, Yang W-J, Arai N, Mori K. Solar-powered Organic Rankine System for domestic electric-power generation. In: Dincer I, editor. TIEES’99 International Symposium Turkey. New York: Begell House, 1999:67–74. [10] Adrian B. Advanced engineering thermodynamics. New York: Wiley, 1997. [11] Hyoudou T, Ishida Y. Design of turbine. Tokyo: Power Co, 1969.