Fundamental experiment of pumpless Rankine-type cycle for low-temperature heat recovery

Energy 36 (2011) 1010e1017

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Fundamental experiment of pumpless Rankine-type cycle for low-temperature heat recovery Noboru Yamada a, *, Takahiro Minami b, Md Nor Anuar Mohamad b, c a

Graduate School of Energy and Environment Science, Nagaoka University of Technology, 1603-1 Kamitomioka, Nagaoka, Niigata 940-2188, Japan Department of Mechanical Engineering, Nagaoka University of Technology, 1603-1 Kamitomioka, Nagaoka, Niigata 940-2188, Japan c Faculty of Mechanical and Manufacturing Engineering, Universiti Tun Hussein Onn Malaysia, 86400 Parit Raja, Batu Pahat, Johor, Malaysia b

a r t i c l e i n f o

a b s t r a c t

Article history: Received 19 May 2010 Received in revised form 29 November 2010 Accepted 7 December 2010 Available online 19 January 2011

This paper proposes a new pumpless Rankine-type cycle for power generation from low-temperature heat sources. The new cycle mainly consists of an expander, two heat exchangers, and switching valves for the expander and heat exchangers. Instead of using a working fluid pump, the switching valves method (SVM) is employed to control the cycle. The SVM makes each heat exchanger switch between functioning as an evaporator and functioning as a condenser. In this arrangement, the working fluid flows back and forth between the two heat exchangers without a working fluid pump. Therefore, the new cycle does not involve problems caused by a pump. In the first basic experiment carried out to clarify the feasibility of the proposed cycle, the function of the expander was emulated by using an expansion nozzle. HFC245fa was used as the working fluid. The experimental results confirm that the proposed cycle works and that it has the potential to produce power. Fundamental time-varying characteristics of the proposed cycle are also shown and discussed. Ó 2010 Elsevier Ltd. All rights reserved.

Keywords: Rankine cycle Organic Rankine cycle Waste heat recovery Working fluid pump Net cycle efficiency Pump efficiency

1. Introduction Recently, various methods to recover low-temperature waste heat have been proposed and developed. One of the feasible and commonly used techniques is the organic Rankine cycle (ORC); in this technique, an organic fluid with a low boiling point is used as the working fluid. Yamamoto et al. [1] described the effect of the thermal properties of an organic working fluid on the expander power output of the system. Numerous manufacturers have introduced commercial ORC systems that convert waste heat into electricity [2]. Yamaguchi et al. [3] reported a unique Rankine cycle system that uses supercritical carbon dioxide (CO2) as the working fluid and elucidated its potential as a solar thermal energy conversion system. In order to generate power from a small temperature difference between hot and cold sources (e.g., 15e25 K in ocean thermal energy conversion), advanced cycles such as the Kalina cycle [4] and the Uehara cycle [5] have been developed. Desai and Bandyopadhyay [6] investigated the ORC by using different types of working fluids and modified the basic ORC by incorporating both regeneration and turbine bleeding in order to improve its thermal efficiency. However, these advanced cycles tend to be large systems with complicated components. Because almost all previous studies and developments have been carried out * Corresponding author. Tel./fax: þ81 258 47 9762. E-mail address: [email protected] (N. Yamada). 0360-5442/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2010.12.007

in order to achieve a power output of more than 10 kW, it would be advantageous if a Rankine cycle system with a power output of 10 kW or less could be achieved for low-temperature heat recovery. Thus far, few studies have focused on the development of a waste heat recovery system (WHRS) for application to automobiles. The Rankine cycle is considered to be a promising technique for a WHRS in automobiles [7e11]. However, few issues must first be overcome in this regard. Unstable heat conditions of the automobile’s waste heat limit the conventional Rankine cycle’s operational range and hence affect the net cycle efficiency. That is, it is difficult to maintain highly efficient expanders, working fluid pumps, and heat exchangers throughout the duration of the operation of an actual automobile engine because the temperature level and the quantity of waste heat from the internal combustion engine (ICE) vary considerably with the engine speed depending on the vehicle’s running condition. For example, Kahraman et al. [12] reported that the temperature level of a four-cylinder gasoline ICE’s (displacement volume: 1197 cc) exhaust gas at the exhaust manifold varies from approximately 653 K to 733 K because the engine speed varies from 2500 rpm to 3500 rpm. In addition, the quantity of waste heat from an ICE (four-cylinder gasoline ICE with a displacement volume of 1796 cc) from the exhaust and the cooling system varies from approximately 44.9 kW to 51.7 kW within the same engine speed range [13]. Under actual running conditions, the engine speed generally varies from the idling speed to more than 4000 rpm. Therefore, the conventional Rankine cycle

N. Yamada et al. / Energy 36 (2011) 1010e1017

installed for the automobile WHRS has to limit its ICE operational speed to a small range; that is, the conventional Rankine cycle is operated only when the vehicle runs at a stable engine speed. Among the components of the conventional Rankine cycle for the low-temperature WHRS, the working fluid pump tends to cause the most critical decline in the net cycle efficiency. A pump’s efficiency drops sharply under off-design conditions, such as high or low pressures at the inlet/outlet and mass flow rates that are greater or lesser than the designed value. This trend is pronounced in systems whose power output is less than 10 kW and whose hot-source temperature level is less than 200  C (473 K). In the worst case, the net cycle efficiency can be zero or negative depending upon the pumping power consumption [14,15]. Furthermore, the working fluid pump limits the compactness of the system arrangement because the pump must be placed at a level lower than that of the condenser (i.e., a net positive suction head of the pump) in order to maintain sufficient inlet pressure to prevent the occurrence of cavitation, which results in considerable power consumption by the pump. These adverse effects of the working fluid pump appear not only in the automobile WHRS but also in any Rankine cycle system with a small power output and a low-temperature hot source whose temperature level and heat quantity vary during operation. It should be noted that heat obtained from renewable energy sources, such as solar thermal energy, geothermal energy, waste heat from factory processes, biomass thermal energy, and heat obtained from the automobile engine have similar characteristics; that is, a distributed energy results in a small power output and a low energy density results in a low-temperature level. Further, unstable heat is obtained when there is no heat storage system. Motivated by the aforementioned facts, this study has been carried out to examine the possibility of a “pumpless” Rankine-type cycle in order to mitigate the drawbacks of the working fluid pump and to achieve a better net cycle efficiency with a more compact arrangement. The thermosiphon Rankine engine reported by Nugyen et al. [16], and Johnson and Akbarzadeh [17] is a Rankine cycle that operates without a working fluid pump to recover lowtemperature heat from geothermal and solar energy. Although the thermosiphon Rankine engine was reported to be an excellent system for energy conversion, it requires a large system [16] because the thermosiphon effect requires many vertical intervals. The new “pumpless” concept, which is introduced and experimentally examined in this paper, is based on a simple operational principle that is capable of mitigating the drawbacks of the working fluid pump by eliminating the pump from the system design. 2. Pumpless Rankine-type cycle 2.1. Rankine cycle Fig. 1 shows the schematic representation of the conventional Rankine cycle. The Rankine cycle is a fundamental vaporeliquid cycle. It consists of five main components: a working fluid pump, an evaporator, an expander, a condenser, and a working fluid. The cycle begins when the pump pushes the working fluid to the evaporator. In the evaporator, the hot source heats the working fluid up to the saturated or superheated vapor state by isobaric heating. The vapor expands adiabatically and rotates the expander to produce power. After the vapor leaves the expander, the cold source cools and condenses the vapor into a liquid state in the condenser by isobaric cooling before the fluid is re-circulated by the pump. The net cycle efficiency of the Rankine cycle is determined as the net power produced by the cycle divided by the heat gained by the working fluid in the evaporator. A conventional low-temperature Rankine cycle, which is frequently exposed to unstable heat sources, requires a pump that can vary the discharged pressure and mass flow rate

1011

Fig. 1. Schematic representation of conventional Rankine cycle.

according to the heat condition while maintaining high efficiency. However, no such pump has been developed thus far. 2.2. Pumpless Rankine-type cycle Fig. 2 shows the operation principles of the proposed pumpless Rankine-type cycle (PRC). The PRC consists of an expander; two heat exchangers, heat exchanger 1 (HX1) and heat exchanger 2 (HX2); valves at the expander inlet and/or outlet (one of the valves may not be needed); switching valves for hot and cold sources; and a working fluid. Fig. 3 shows the basic pressureeenthalpy (peh) diagram of the PRC. This cycle is called a “Rankine-type” cycle because the thermodynamic process of the PRC is not exactly the same as that of the conventional Rankine cycle. The basic operational steps of the PRC shown in Fig. 2(a)e(d) can be described as follows: (a) All valves at the expander inlet/outlet are closed, which creates enclosed conditions for both HX1 and HX2. The hot fluid from the hot source enters HX1 and performs an evaporation process; that is, it evaporates the working fluid inside HX1. When the heating process is taking place in HX1, cold fluid from the cold source enters HX2 and performs a condensation process; that is, it condenses the working fluid inside HX2. As a result, the isochoric heating process in HX1 increases the working fluid’s temperature and pressure, whereas the isochoric cooling process in HX2 reduces the fluid’s temperature and pressure. (b) The valves open when the pressure difference reaches a preset value (given in Table 1). Isobaric heating and flashing cause the working fluid in HX1 to vaporize. The working fluid then flows into the expander and expands adiabatically. It should be noted that the adiabatic expansion is an ideal process without any losses, that is, it is an isentropic process; however, in an actual operation, the expansion will be non-isentropic. This expansion process in the expander produces a mechanical power output WE that is expressed as follows:

_ 1  h2 j WE ¼ mjh

(1)

_ is the mass flow rate of the working fluid, here, h is the enthalpy, m and the subscripts 1 and 2 represent the heat exchanger corresponding to the state numbers shown in Fig. 1. The expander power output changes in a time-dependent manner until no pressure difference exists between the expander inlet and the expander outlet, or until the pressure difference becomes smaller than a preset value. The vapor that passes through the expander is

1012

N. Yamada et al. / Energy 36 (2011) 1010e1017 Table 1 Principle specifications of experimental apparatus. Component

Specification

Expander

Emulated by expansion nozzle Nozzle discharge area: 1.8 mm2 Brazed-plate heat exchanger Volume: 0.18 L Heat conduction area: 0.252 m2 Bi-directional solenoid valve Power consumption: 6 W Preset pressure for on/off valve opening: Evaporator side (pH): 0.6e0.9 MPa Condenser side (pC): 0.15 MPa HFC245fa (CF3CH2CHF2) Molecular weight: 134.05 Boiling temperature: 14.9  C Water Temperature range: 76e96  C Pressure: 0.5 MPa Flow rate: 2.4 L/min Electrical heater (Adjustable by variable resistor) Output capability: 10 kW Water Temperature: 26  C Pressure: 0.5 MPa Flow rate: 8 L/min 3-wayetype solenoid valve Number of valves: 4 Power consumption: 12 W/piece

Heat exchangers (Evaporator/Condenser) On/Off valve

Working fluid

Hot source

Heater

Cold source

Switching valve

Fig. 2. Operation principle of proposed pumpless Rankine-type cycle (PRC). (a) Isochoric process. (b) Flashing and adiabatic expansion. (c) Switching heat source and isochoric process. (d) Flashing and adiabatic expansion.

condensed, and the phase is returned to the saturated liquid state by isobaric cooling. (c) After the working fluid has flowed from HX1 to HX2, the circulation flows from the heat and cold sources are switched

between HX1 and HX2 by using the switching valves. The cold fluid from the cold source enters HX1, whereas the hot fluid from the hot source enters HX2. HX1 and HX2 then operate as a condenser and an evaporator, respectively. Instead of using the pump, the SVM alternately switches the functions of HX1 and HX2 from a condenser to an evaporator and vice versa. The heating and cooling processes here are the same as those described in (a), but they are carried out in opposite components. (d) The valves at the expander’s inlet and outlet open, and the process is the same as the one described in (b); however, it is carried out in a reverse direction. All the processes are then repeated; through the SVM, the working fluid flows back and forth between the two heat exchangers without the need of a pump. The proposed cycle operates in the following order: isochoric heating, isobaric heating, flashing, adiabatic expansion, isobaric cooling, and isochoric cooling. The isochoric cooling process tends to occur for a very short time between the closing of the valves at the expander inlet/outlet and switching of the hot and cold fluid flows; this is not shown in Fig. 3. The average cycle power WAve in processes (a)e(d) can be expressed as follows:

Zt WAve ¼

Pressure

Isobaric heating Pumping process of conventional Rankine cycle

Isochoric heating

WE dt=t

(2)

0

Adiabatic expansion

Isobaric cooling

Enthalpy Fig. 3. Basic pressureeenthalpy (peh) diagram of proposed PRC.

here, t is the operation time of the cycle. The “internal” cycle efficiency hin, which only accounts for the heat successfully transferred to the working fluid Qin, is given by

hin ¼

WAve Qin

(3)

The heat exchanger efficiency hHX between the heat source and the working fluid is given by

hHX ¼

Qin Qex

(4)

N. Yamada et al. / Energy 36 (2011) 1010e1017

1013

here, Qex is the heat from the heat source. The “external” cycle efficiency hex, which includes the effect of the heat exchanger efficiency, is given by

hex ¼

WAve W ¼ hHX Ave Qex Qin

(5)

In this cycle, the working fluid pump’s power consumption is replaced by the power/electricity consumed by switching the valves. Therefore, the net cycle power Wnet for the entire process is given by

Zt Wnet ¼

ðWE  WV Þdt=t

(6)

0

Here, WV is the power consumption of the valves. The net cycle efficiency hnet is given by

hnet ¼

Wnet Qex

(7)

If the power/electricity consumed by the switching valves is kept lower than the power consumption of the working fluid pump in the conventional Rankine cycle, the PRC can achieve a higher net cycle efficiency than the conventional Rankine cycle. Unlike a steady conventional Rankine cycle, the PRC is a nonsteady cycle where the flow condition of the working fluid occurs intermittently. Because the working fluid that flows in the PRC is intermittent in a reversible flow direction, a suitable expander is required. Displacement-type expanders that include a screw, a scroll, a rotary vane, an axial piston, and gear expanders would be suitable candidates because these expanders tend to maintain a high expander efficiency over a wide range of pressure differences and mass flow rates [15,18,19]. The expander can be installed even in a reversible flow direction with small modifications if the switching valves are installed in the piping among the expander and heat exchangers. Bi-rotational velocity-type expanders such as the impulse turbine [20] and Wells turbine [21] that have been developed for wave-energy conversion systems can also be used in the PRC. 3. Experimental setup and methods The objectives of the present experiment are to clarify that the PRC works as explained above (shown in Fig. 2) and to elucidate the fundamental time-varying characteristics of the PRC. The experimental apparatus was built to have as small a size as possible for laboratory demonstration. The experimental apparatus is shown in Fig. 4. The measured temperature T and pressure p points of the working fluid are also shown in Fig. 4. Table 1 lists the specifications of the experimental apparatus. HFC245fa was chosen as the working fluid. HFC245fa has a wide range of applications: it is used as a foaming agent, refrigerant, filling for thermosiphons, and working fluid for an ORC for low-temperature heat recovery [22]. The critical temperature of HFC245fa is 427.16 K (approximately 154  C), which is considerably higher than the highest temperature expected in the present system. The hot source was water with Twh ¼ 76e96  C that is heated by an electric heater and circulated throughout the hot-side heat exchanger (i.e., evaporator). Similarly, the cold source was water with Twc ¼ 26  C that is cooled by a chiller and circulated throughout the cold-side heat exchanger (i.e., condenser). Therefore, the range of the temperature difference between the hot source and the cold source (DT ¼ Twh  Twc) was DT ¼ 50, 60, and 70  C. It should be noted that the actual average temperature of both the sources fluctuated

Fig. 4. Schematic representation of experimental apparatus for PRC (a) HX1 ¼ Evaporator and HX2 ¼ Condenser. (b) HX1 ¼ Condenser and HX2 ¼ Evaporator.

within 1  C. Brazed-plate heat exchangers with capacities of 0.18 L and a height of 55 mm were used as both the evaporator and the condenser. Currently, there is no suitable expander that can function efficiently under the conditions of the present system. Therefore, an expansion nozzle (discharge area: 1.8 mm2) is installed in order to emulate the expansion condition of the system. A similar method has been reported in the literature [14,23,24]. The simulated expander power WE, average cycle power WAve, net cycle power Wnet, internal efficiency hin, external efficiency hex, and net cycle efficiency hnet were calculated by using Eqs. (1)e(7). Fig. 5 shows the details of the pressure and temperature measurement points on the HX1 side. The piping of the heat-source fluids and the inside structures of the plate heat exchanger are not shown in this figure. The designs of HX1 and HX2 are similar, and the arrangements of their measurement points are symmetric to each other. The solenoid valve (diameter: 7.8 mm) was located next to the nozzle as an on/off (open/close) valve, and pressure-tight rubber tubes (length: 150 mm; diameter: 6 mm) and copper tubes (length: 255 mm; nominal diameter: 14 mm) were used to connect the ports of the heat exchangers (inlet and outlet) to the expansion nozzle and the on/off valve. HX1 was filled with the working fluid after closing the on/off valve and vacuuming the heat exchangers. As mentioned above, the cycle operation was regulated by the preset pressure difference between the evaporator and the expander. The preset pressure ratio g is defined by

g¼

pH pC

(8)

N. Yamada et al. / Energy 36 (2011) 1010e1017

25

• •T=50 ••C = 50ºC Δ

••C = 60ºC Δ• T• =60

• •T=70 ••C = 70ºC Δ

• •T=50 ••C = 50ºC Δ

••C = 60ºC Δ• T• =60

• •T=70 ••C = 70ºC Δ

• •T=50 ••C = 50ºC Δ

••C = 60ºC Δ• T• =60

• •T=70 ••C = 70ºC Δ

20 h2| [kJ/kg]

where pH and pC are the preset pressures of the evaporator and the condenser, respectively. When the hot-side pressure (p1 or p2) and the cold-side pressure (p1 or p2) reached the preset pressures of the evaporator pH (hot-side) and the condenser pC (cold-side), the on/ off valve was opened; further, the working fluid expanded through the expansion nozzle. Then, the on/off valve was closed when the pressure difference between the hot and cold sides dropped to 0.01 MPa. The circulations of hot and cold water were switched between HX1 and HX2 through the SVM by two pieces of 3-way solenoid valves. Before the measurements were taken, the expansion process was repeated several times to preheat the nozzle, valves, and tubes in order to prevent the working fluid from condensing inside these components. Temperature measurements were carried out using type-T thermocouples having an uncertainty of 1.0  C and response speed of approximately 0.36 s for the range from 20  C to 100  C. Pressure measurements were carried out using pressure transducers having an uncertainty of 0.5% of the full-scale load and a response speed of approximately 5 ms. All measured data were recorded and monitored using a data-acquisition PC with a speed of one datum per second. Enthalpies and internal energies were calculated from the measured pressure and temperature data by using REFPROP ver. 8 developed by the NIST [25]. The mass flow rate of the working fluid was estimated from the changes in the weights of the heat exchangers during the experiment. Two digital weighing scales were placed beneath each heat exchanger HX1 and HX2 to detect the time-dependent weight changes and were connected to the data-acquisition PC. The values measured by the weighing scales were calibrated by the reference weight before the experiment. The pressure-tight flexible rubber tubes that connected the heat exchangers to the expander contributed to an accurate weight measurement.

15 10

|h1

Fig. 5. Details of measurement points for HX1. (HX2 has a symmetric configuration).

estimated expander power WE was also intermittently produced. Hence, it was confirmed that the function of the heat exchanger and the flow direction of the working fluid were successfully reversed by the SVM and that the PRC basically worked according to the conceptual operation principles. Here, it should be noted that the expander efficiency was assumed to be 100% in Eq. (1). The required time from the on/off valve closing to reach the preset pressure (time required for processes (a) or (c) in Fig. 2) decreased with an increase in DT, which resulted in an expansion frequency of 7.5 expansion processes every 10 min for DT ¼ 50  C, 10 for DT ¼ 60  C, and 11 for DT ¼ 70  C. On the other hand, the cycle constantly required approximately 40 s from the on/off valve opening to reach the last pressure difference (time required for process (b) or (d) in Fig. 2) for all values of DT. The enthalpy difference reached a peak 20 s after the valve opened and then decreased. The maximum enthalpy difference increased with an increase in DT. The mass flow rate reached its peak 1 s after the valve opened and then dropped. Approximately 150 g of the working fluid was moved from the evaporator to the condenser by a single expansion process in the condition. The expansion process from HX1 to HX2 and the expansion process from HX2 to HX1 were not exactly the same because the heat-source supply pipings between the two heat exchangers were not perfectly symmetric, which resulted in dissimilar fluid flows, different fluid frictions, heat losses,

5 0 30 Mass flow rate m [g/s]

1014

·

25 20 15 10 5 0 150

The main parameters of the present experiment were the filling mass of the working fluid M, preset pressure, and temperature difference DT. HX1 was filled with the working fluid of M ¼ 100, 200, and 300 g. The preset pressures of the evaporator pH and the condenser pC were pH ¼ 0.6, 0.75, and 0.9 MPa, respectively, and pC ¼ 0.15 MPa (i.e., pressure ratio g ¼ 4, 5, and 6, respectively). The temperature differences were DT ¼ 50, 60, and 70  C, respectively. Fig. 6 shows the time-varying characteristics of the enthalpy difference between the expander inlet and outlet jh1  h2j, mass flow rate, and expander power WE calculated by using Eq. (1) for DT ¼ 50, 60, and 70  C; g ¼ 4; and M ¼ 200 g. The measurements were repeated for processes (a)e(d). Obviously, both the enthalpy difference and the mass flow rate were intermittently generated; therefore, the

Expander power WE [W]

4. Experimental results with PRC 125 100 75 50 25 0 0

0.5

1

1.5

2 2.5 3 Time [min]

3.5

4

4.5

Fig. 6. Time variations of enthalpy, mass flow rate, and emulated expander power. (DT ¼ 50e70  C, g ¼ 4, M ¼ 200 g).

N. Yamada et al. / Energy 36 (2011) 1010e1017

Inlet pressure pin [MPa]

etc. The result shows that intermittently produced power is a particular characteristic of the PRC. Coupling with a generator and power electric devices and/or employing a multiple-cycle system may be able to smoothen the intermittently produced power of the PRC if one needs stable output power. Fig. 7(a)e(c) show more detailed time-varying characteristics of the emulated expander inlet pressure pin and power WE during a single expansion process; the time scale in this figure starts when the on/off valve is opened and ends at the last of the pressure difference. The effect of the heat-source temperature difference DT is shown in Fig. 7(a), which was obtained under the same condition as Fig. 6. The inlet pressure increased with DT. For DT ¼ 60  C, the inlet pressure was constant at 0.02 MPa from pH after the valve opened until half the period of the expansion process, following which it decreased. For DT ¼ 70  C, the inlet pressure reached 0.7 MPa after the valve opened. These differences were mainly caused by the differences in the balance between the amount of vapor flowing out of the evaporator and the amount of vapor emitted from the liquid surface in the evaporator. For DT ¼ 70  C, the amount of vapor emitted from the liquid surface in the evaporator was larger than that flowing out. For DT ¼ 50  C, the maximum expander power WEmax was generated 2 s after the valve opened, but for DT ¼ 70  C, WEmax was generated 20 s after jh1  h2j reached its peak. The effect of the preset pressure ratio g is shown in Fig. 7(b). As g increased, the time-varying characteristics of the pressure drop changed and the duration of expansion decreased. The maximum expander output WEmax increased with g. For g ¼ 6, WEmax was obtained immediately after the valve opened, and the increased amount was twice that for g ¼ 4. The effect of the filling mass of the working fluid is shown in Fig. 7(c). The expansion process time increased with M. The inlet pressure decrement of M ¼ 300 g after the valve opened was smaller than that in the other cases because the condenser pressure at the expander outlet (shown as pout(300 g) in Fig. 7(c)) increased during expansion, whereas the pout values of the other cases (not shown in Fig. 7(c)) were almost constant. This was caused by insufficient condensation, i.e., the amount of working fluid that flowed into the condenser exceeded the condensation ability and the volume of the heat exchanger. Consequently, the expander power of M ¼ 300 g decreased with the pressure difference. These results indicate that the dynamic behavior of the PRC is considerably influenced by the transient phase change and the mass transfer.

1.0

ΔT:

0.8

Fig. 8(a)e(c) show the average cycle power WAve, the internal cycle efficiency hin, and the external cycle efficiency hex corresponding to the experimental results shown in Fig. 7(a)e(c). From Fig. 8(a) and (b), it was confirmed that the performance of the proposed cycle improved by increasing the heat-source temperature difference DT and pressure ratio g, as in the case of conventional thermodynamic cycles. However, the external cycle efficiency hex defined by Eq. (5) was around 1% because of the significantly low heat exchanger efficiency (i.e., hHX ¼ approximately 20% in most cases in the present experiments). In the conventional ORC, an average hHX value of around 70% has been reported [14]. Thus, in the present PRC, half of the heat source energy was not transferred to the working fluid. The cycle switched the functions of the heat exchanger from a condenser to an evaporator and vice versa. Therefore, the heat source had to heat and cool not only the working fluid but also the material of the heat exchanger itself. Consequently, heat loss due to the heat capacity of the material of heat exchanger occurred. Moreover, in the present PRC, the heat transfer coefficient could be lower than that of the conventional ORC because the working fluid does not flow during the isochoric heating process as shown in Fig. 2(a) and (c). It is expected that the installation of a heat exchanger with a minimum heat capacity and maximum heat transfer area would contribute to an improvement in the heat exchanger efficiency and external cycle efficiency. The filling mass of the working fluid, M, also affected system performance. When M was varied (Fig. 8 (c)), the cycle average power declined when the filling mass was greater than 200 g. A possible reason for this is the limited space available for the working fluid to expand inside the heat exchanger (condenser). Although the pressure at the evaporator inlet Pin is higher compared to that at M ¼ 100 g or 200 g, the expansion process could not take place in a manner similar to that at M ¼ 100 g and 200 g owing to the fact that the volume of the fluid was greater than the space available for expansion. Actually, for M ¼ 300 g, the differential between Pin and Pout shown in the upper figure of Fig. 7(c) became zero in as short a time as that for M ¼ 200 g. In other words, the expansion process for M ¼ 300 g ended sooner than expected. The resultant internal cycle efficiency and external cycle efficiency are approximately 6% and less than 1%, respectively. These results indicate that there is an optimum working fluid mass for the PRC system and the use of excess filling mass of the working fluid must be avoided in order to maintain the cycle efficiency.

c

b

a 50ºC

60ºC

1015

γ:

70ºC

4

5

6

M:

100g

200g

300g

200g

300g

0.6 0.4 0.2

pout (300g)

Expander power WE [W]

0.0

ΔT:

200

50ºC

60ºC

γ:

70ºC

4

5

M:

6

100g

150 100 50 0 0

5

10 15 20 25 30 35 40

0

5

10 15 20 25 30 35 40

Time [s]

Time [s]

ΔT = 50~70 ºC, γ = 4, M =200g

ΔT = 70 ºC, γ = 4~6, M =200g

0

5

10 15 20 25 30 35 40 Time [s]

ΔT = 70ºC, γ = 6, M =100~300g

Fig. 7. Details of time variations of pin and WE with expanded time scale. (The valve is open at 0 s.) (a) DT ¼ 50e70  C, g ¼ 4, M ¼ 200 g. (b) DT ¼ 70  C, g ¼ 4e6, M ¼ 200 g. (c) DT ¼ 70  C, g ¼ 6, M ¼ 100e300 g.

1016

N. Yamada et al. / Energy 36 (2011) 1010e1017

a

b

c

Fig. 8. Average cycle power WAve and efficiencies hin and hex at a given condition. (a) DT ¼ 50e70  C, g ¼ 4, M ¼ 200 g. (b) DT ¼ 70  C, g ¼ 4e6, M ¼ 200 g. (c) DT ¼ 70  C, g ¼ 6, M ¼ 100e300 g.

The net cycle power Wnet and net cycle efficiency hnet are important indexes for clarifying the advantages of the proposed PRC. However, the resultant values calculated by using Eqs. (6) and (7) were low: Wnet ¼ 10.0 W to 7.1 W and hnet ¼ 0.62% to 0.24%. Although one may think that these values are so low that the feasibility of the proposed PRC cannot be emphasized, the conventional Rankine cycle with an ordinary working fluid pump under the same heat-source condition and the similar output power level tends to exhibit rather negative values depending on the pump efficiency. Moreover, it should be noted that this small or negative net cycle efficiency can be overcome if the proposed cycle is applied to a larger output power system than the present experimental system because the electricity consumption by the solenoid valves can be small as compared to the expander power in the case of a large output power system. The output power level of the present experimental system was considerably small (around 30 W) as compared to the capacity of the solenoid valves. Furthermore, it is expected that the proposed cycle has an advantage when the pressure and mass flow rate fluctuate widely because the electricity consumption by the valves is basically proportional to the on/off (open/close) frequency; in contrast, the electricity consumption by the working fluid pump tends to increase sharply under the pump’s off-design condition. An experimental clarification of this advantage is one of our future research objectives. As mentioned above, a single PRC system produces intermittent power because the state of the working fluid supplied to the expander inlet changes transiently. Therefore, it is important to understand the time-dependent behavior of the expansion process.

1.0

Saturated line

Expander inlet Expander outlet

Pressure [MPa]

0.8 0.6

2s

13 s

0.4

22 s 60°C

32 s

0.2 0.0 250

300

350

400

450

500

Enthalpy [kJ/kg] Fig. 9. Time-dependent peh diagram of the expansion process for DT ¼ 50  C, g ¼ 4, M ¼ 200 g. (Time 2 s, 13 s, 22 s, and 32 s corresponding to the horizontal axis of Fig. 7(a)).

Fig. 9 shows the actual time-dependent peh diagram of the expansion process in the emulated expander for DT ¼ 50  C; the basic peh diagram of the PRC has already been shown in Fig. 3. Fig. 9 illustrates the expander inlet and the outlet state at 2 s, 13 s, 22 s, and 32 s after the on/off valve opened, corresponding to Fig. 7 (a). It is obvious that the change in the state of the working fluid throughout the PRC expander varies significantly with time. Until 22 s after the opening of the valve, the expansion occurred in a twophase region. After 22 s, the expansion occurred in the vapor region. The expander inlet pressure dropped slightly until the working fluid vaporized. After the working fluid vaporized at the expander inlet, the inlet temperature became constant at 60  C and produced a larger enthalpy drop than the two-phase state. This result indicates that the proposed PRC needs an expander that allows a two-phase expansion. Installing a real expander and investigating the dynamic behaviors of the proposed cycle under a fluctuating heat-source condition will be carried out together with a theoretical approach in a future work.

5. Conclusion In this paper, we proposed a new Rankine-type cycle without the working fluid pump for a low-temperature waste heat recovery system. The proposed cycle uses several switching valves to alternate the hot and cold source fluids between the two heat exchangers. This operation makes it possible to transfer a mass of the working fluid between the evaporator and the condenser via the expander without a pump. The first fundamental experimental results based on the emulated expander and the organic working fluid HFC245fa revealed that the proposed cycle could produce an enthalpy difference and a mass flow rate; therefore, it has the potential to produce power although the net cycle efficiency estimated from the present experimental system does not lie in the range of practical values because the system is very small and has unmatched solenoid valves. The expected advantage of the proposed cycle is that it can eliminate the drawbacks caused by the working fluid pump in the conventional Rankine cycle, especially when the cycle is applied to small and low-temperature systems with unstable heat-source conditions. On the other hand, the disadvantages of the proposed cycle are as follows: (1) The heat exchanger efficiency tends to be low because of the heat capacity of the heat exchanger, which results in a low external cycle efficiency. (2) The expander needs to allow a two-phase condition of the working fluid during expansion. (3) The power produced by a single cycle system is intermittent. Mitigating these disadvantages by optimizing the system design will make the proposed cycle an economically efficient thermodynamic cycle. Although the present

N. Yamada et al. / Energy 36 (2011) 1010e1017

study is still in the early stages, we are reporting the fundamental results on this new concept because research and development on technologies to generate power from low-temperature sources are urgently required in order to address energy-related and environmental issues.

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