Energetic and exergetic investigation of an organic Rankine cycle at different heat source temperatures

Energetic and exergetic investigation of an organic Rankine cycle at different heat source temperatures

Energy 38 (2012) 85e95 Contents lists available at SciVerse ScienceDirect Energy journal homepage: www.elsevier.com/locate/energy Energetic and exe...

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Energy 38 (2012) 85e95

Contents lists available at SciVerse ScienceDirect

Energy journal homepage: www.elsevier.com/locate/energy

Energetic and exergetic investigation of an organic Rankine cycle at different heat source temperatures Jing Li, Gang Pei*, Yunzhu Li, Dongyue Wang, Jie Ji* Department of Thermal Science and Energy Engineering, University of Science and Technology of China, Jinzhai Road 96, Hefei City, Anhui Province 230026, People’s Republic of China

a r t i c l e i n f o

a b s t r a c t

Article history: Received 24 August 2011 Received in revised form 18 December 2011 Accepted 22 December 2011 Available online 16 January 2012

The energetic and exergetic performance of an updated ORC (organic Rankine cycle) is investigated. The thermal efficiencies of the ORC at different heat source temperatures of about 100, 90, 80, and 70  C are explored. The thermodynamic irreversibility that takes place in the evaporator, condenser, turbine, pump, and separator is revealed. The ORC feasibility for low-temperature applications is demonstrated. With a hot side temperature of around 80  C, a thermal efficiency of 7.4% and a turbine isentropic efficiency of 0.68 can be achieved. The present research further indicates that exergy destruction caused by heat transfer through an appreciable temperature difference in the evaporator is the largest in the energy conversion process, followed by that in the condenser. The exergy destroyed in the heat exchangers amounts to 74% of the overall exergy loss. The total system exergy efficiency is approximately 40%; thus, ways to improve exergy efficiency are required. HCFC-123, a dry fluid, is experimentally confirmed to be highly superheated after expansion in this study. A regenerator should be used to preheat HCFC-123 prior to entering the evaporator. Meanwhile the heat-transfer configuration with two oil cycles can be a good solution to overcome the thermodynamic disadvantage of a one-stage evaporator. Ó 2011 Elsevier Ltd. All rights reserved.

Keywords: Organic Rankine cycle Irreversibility Energy conversion Exergy loss

1. Introduction About 80% of the electric power used worldwide is generated by the steam Rankine cycle [1]. Little work is required to drive the pump during the compression stage in this cycle. However, common heat sources for the steam Rankine cycle are coal, natural gas, oil combustion, and nuclear fission, which provide high-grade heat sources at about 565  C. Given the high-energy density of these heat sources, the steam Rankine cycle favorably operates in the power range above 100 MW. Fossil fuels on Earth are finite, and their use has led to air, water, and land pollution. Conventional nuclear power poses unanswered questions, such as potential hazards of accidents and explosions, and it should not be considered a reliable, long-term solution either [2]. To lessen the fossil fuel and nuclear power dependence for energy needs, many technologies, such as waste heat recovery, biomass, solar power, and geothermal power, are being developed to generate enough electricity for global use and to help reach a sustainable energy system. In these technology fields, the heat sources are generally available at low temperatures, and the power * Corresponding authors. Tel./fax: þ86 551 3607367. E-mail addresses: [email protected] (G. Pei), [email protected] (J. Ji). 0360-5442/$ e see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2011.12.032

plants have limited sizes. The steam Rankine cycle seems no longer applicable based on the following. (1) There is great difficulty in developing efficient steam turbines at low power. A water molecule has polar covalent bonds between oxygen and hydrogen atoms. The heat capacity and the latent heat of vaporization of water are very large. For example, the latent heat at 100  C is 2257 kJ/kg, which is 16 times larger than that of HCFC-123 at the same temperature. Compared with most organic fluids, water offers a much smaller mass flow rate through a low-power turbine. A small mass flow rate leads to great technological difficulties in designing and manufacturing turbine injectors, blades, etc. (2) The expansion ratio of water at the turbine outlet and inlet is also large in lowtemperature applications. For example, the ratio of watersaturated pressure at 100  C to that at 20  C is about 43.35, whereas it is only 10.39 for HCFC-123. A large expansion ratio requires multistage expansion and a complicated turbine. (3) Due to the need for superheat, the efficiency of steam Rankine cycle is inadequate in low-temperature applications. Water is a typically wet fluid with a negative slope on its saturation vapor temperatureeentropy (Tes) curve. To prevent water droplets from hitting the turbine blades at a high speed, superheat is required. Table 1 presents the required degree of superheat in the lowtemperature steam Rankine cycle. The required degree of

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J. Li et al. / Energy 38 (2012) 85e95

Table 1 The degree of superheat variation with turbine inlet pressure in steam Rankine cycle. Condensation temperature/ C Turbine efficiency Turbine inlet pressure/bar Saturated temperature/ C Required turbine inlet temperature/ C Degree of superheat Dt/ C

30 0.80 0.47 80 200

30 0.80 1.01 100 265

30 0.80 4.76 150 410

30 0.80 15.54 200 530

30 0.80 39.73 250 635

120

165

260

330

385

superheat to avoid the binary phase in the steam expansion process is high. Therefore, the average temperature for the heating process is reduced, and the steam Rankine cycle efficiency is low. For example, the steam Rankine cycle efficiency at the required expander inlet temperature of 200  C is only 11.7%, which is less than the cycle efficiency at the hot side temperature of around 120  C upon using HCFC-123 as the working fluid [3]. In low-temperature fields such as waste heat recovery, solar thermal electric generation, and biomass and geothermal plants, the ORC (organic Rankine cycle) is preferably adopted. The working principle of the ORC is the same as that of the steam Rankine cycle: vapor generated in heat exchangers expands through an expander, exporting power due to the enthalpy drop. The ORC can use fluids that boil at a temperature lower than that of water. It can overcome the limited efficiency of the steam Rankine cycle at low temperature. In the case of a dry fluid, the ORC does not require superheat. This non-requirement of superheat raises the average temperature for the heating process, and the thermal efficiency is significantly improved, especially when the temperature difference between the hot and cold sides is small. Moreover, the development of an efficient ORC turbomachinery at low power is easier. ORC research is growing fast. However, most of the recent studies deal with theoretical aspects. Information on the practical usage of the ORC is limited. James et al. conducted an experimental study on gerotor and scroll expanders, which produced 2.07 and 2.96 kW, respectively. Both expanders had significant potential to produce power from low-grade energy [4]. Lemort et al. introduced the prototype of an open-drive, oil-free scroll expander integrated into an ORC working with refrigerant HCFC-123. The experimental results show that a delivered shaft power of 1.82 kW and an overall isentropic effectiveness of 68% could be achieved [5,6]. Riffat et al. studied a 1.34 kW ORC-based CHP system assisted by fuel gas. The electrical efficiency was 16%, and the overall efficiency was about 59% [7]. Peterson et al. presented a study on the performance of a small-scale regenerative Rankine power cycle employing a scroll expander. System efficiency was about 7.2% [8]. Manolakos et al. presented detailed laboratory experimental results of a lowtemperature ORC engine coupled with a reverse osmosis desalination unit. The results indicate that the efficiency of the Rankine cycle fluctuated from 3.5% to 5.0% [9]. Liu et al. developed and evaluated a biomass-fired microscale CHP system. An electric efficiency of 1.34% and an overall CHP efficiency of 88% were achieved [10]. Wang et al. tested a prototype low-temperature solar ORC system. With a 1.73 kW rolling-piston expander, the overall power generation efficiency was estimated at 4.2% and 3.2% for evacuated and flat plate collectors, respectively [11]. Yamada et al. proposed a new pumpless micro-ORC for power generation from low-temperature heat sources. The experimental results confirmed that this cycle worked and that it had the potential to produce power [12]. The construction of a 3.5 kW ORC system using a carefully designed and manufactured turbine was introduced by the authors [13]. The design chart and some problem-solving techniques, such as the avoidance of cavitation in the pump, were presented. A preliminary test of the constructed ORC was also conducted. Considering the poor inner pressure distribution in the self-

designed separator, a thorough and comprehensive investigation of the system has not been done yet. An updated ORC system is presented in the current paper. The limitation of the system working pressure is overcome, and a more sufficient expansion of HCFC-123 through the turbine is made possible. ORC experiments at different heat source temperatures of about 100, 90, 80, and 70  C are conducted. The exergy loss in the evaporator, turbine, separator, condenser, and pump is investigated. The ORC feasibility in low-temperature applications is demonstrated. HCFC-123, a typical dry fluid, is used in this ORC system. HCFC-123 should qualify for a superior environmental benefit exemption because it has a small ODP (0.020), small GWP (76), and short half-life (1.3 years) compared with CFC-12, HCFC-22, and HFC-134a [14]. Currently, HFC-245fa seems a more promising ORC fluid. However, HFC-245fa was not widely used in 2007 when the design of this ORC began. Its thermodynamic properties may not match the turbine, pump, heat exchanger, etc., which were designed based on HCFC-123. 2. Thermodynamic analysis Fig. 1 is the design chart of the updated ORC system. In this system, an electric heater is set around the tank to maintain a pressure near 1 atm in cold winter. Therefore, high-purity HCFC-123 in the closed system is more easily guaranteed when the ORC is off work. More by-pass conduits are also added to enable greater flexibility and reliability of the ORC. These conduits are represented by dashed lines. The system consists of three units: HCFC-123 circuit, oil circuit, and water circuit. The oil temperature can be set by a temperature controller, and heat is transferred to HCFC-123 in the evaporator. Water is cooled down in a cooling tower and helps condense HCFC-123 in the condenser. HCFC-123 is vaporized under high pressure in the evaporator. The vapor flows into the turbine and expands. The outlet vapor is then condensed to a liquid state. The liquid is pressurized by the pump and is sent back to the evaporator. The low-grade heat to power conversion is conducted in such a way. The line segments, along with T or P in Fig. 1, are the measurement positions for temperature or pressure. The key thermodynamic state points are marked by circles outside the numbers. 2.1. Energy analysis Heat transferred from the oil to HCFC-123 in the evaporator is calculated by the increment in HCFC-123 enthalpy:

_ f ðh4  h3 Þ Q_ in ¼ m

(1)

The power generated by the turbine [Eq. (2)] and that consumed by the pump [Eq. (3)] are approximately calculated by

_ t ¼ m _ f ðh5  h6 Þ  Q_ loss W

(2)

_ p ¼ m _ f ðh3  h2 Þ W

(3)

where Q_ loss is the external heat loss from the turbine. In the ORC system, the evaporator, separator, and turbine work at higher temperatures than the other components. The evaporator and separator are wrapped by a mattress and the external heat losses can be negligible. Thermal insulation for the turbine is highly complicated because of its irregular shape and the connections between it and the pipes. Repair and maintenance are frequently required, such as adding oil to the turbine. Outer cladding is wasteful, or it can make the process inconvenient. Mattress insulation for a compact turbine can be eliminated when the operation temperature is around 100  C [15]. In this case, Q_ loss is small but _ t . Q_ should be considered for a precise calculation of W loss for the noninsulated turbine can be calculated by

J. Li et al. / Energy 38 (2012) 85e95

87

Fig. 1. Design chart of the updated ORC system.

Q_ loss ¼ Q_ cond þ Q_ conv þ Q_ rad

(4)

The external heat loss of the current turbine in the temperature range of 60  Ce100  C has been experimentally evaluated. Conductive heat loss is almost directly proportional to the difference between the average turbine working temperature and the environment temperature, Q_ cond fDT. Radiative heat loss is quadratically proportional to DT, Q_ rad fðDT þ 3DT 2 =2Ta Þ. Convective heat loss is approximately a power function of DT, Q_ conv fDT 1:25 [15]. Turbine isentropic efficiency is defined by 3t

¼

h5  h6 h5  h6s

(5)

The ORC output based on the shaft power of the turbine is calculated by

_ t W _ p _ out ¼ W W

(6)

The net electric power output is calculated by

_ out;n ¼ W _ t $3 $3 g  W _ p W b

(7)

where 3 b and 3 g are the gearbox and generator efficiency, respectively. The overall ORC efficiency is defined by the ratio of the ORC output [Eq. (8)] or the net electric power output [Eq. (9)] to the total heat supplied:

_ W

hORC ¼ _ out Q in hORC;n ¼

_ out;n W Q_

(8)

(9)

in

2.2. Exergy analysis The exergy loss associated with element i is the difference between the exergy Eiin flowing into i and the exergy Eiout leaving it [16]:

xi ¼ Eiin  Eiout ¼ Eia  Eiu

(10)

Eia is the available exergy for i, and Eiu is the used exergy. The exergy efficiency is determined as

hiexergy ¼

Eiu Eia

(11)

The ORC is mainly divided into five regions: Region One from position 2 to 3, Region Two from 3 to 4, Region Three from 4 to 5, Region Four from 5 to 6, and Region Five from 6 to 1. HCFC-123 in the tank is in a liquid state and flows orderly. The exergy

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J. Li et al. / Energy 38 (2012) 85e95

destruction in the region from position 1 to 2 is much lower than that in other regions. Therefore, it is negligible. The main elements of the five regions are associated with the evaporator, separator, turbine, condenser, and pump, respectively. Eia and Eiu for each region in the ORC are expressed as [17]

_ t Etu ¼ W

(17)

Eca ¼ E6  E1

(18)

Eea ¼ E9  E10

(12)

_ p Epa ¼ W

Eeu ¼ E4  E3

(13)

Esa ¼ E4  E5

(14)

The total exergy loss of the system is the sum of the exergy losses in the five regions:

Esu ¼ 0

(15)

xORC ¼ xe þ xs þ xt þ xc þ xs

Eta ¼ E5  E6

(16)

The system exergy efficiency is the ratio of the total used exergy to the total available exergy of the system:

Ecu ¼ E7  E8

Epu ¼ E3  E2

(19) (20) (21)

(22)

Fig. 2. (a) Layout of the updated ORC system; (b) insulated ORC: (1) turbine; (2) gearbox; (3) generator; (4) condenser; (5) tank; (6) pump; (7) flow meter; (8) evaporator ; (9) separator.

J. Li et al. / Energy 38 (2012) 85e95

hexergy;ORC ¼

u Etotal Eu z a t a a Etotal Ee þ Ep

(23)

The reference temperature and pressure are 0  C and 101 kPa, respectively, in the thermodynamic analysis. 3. System update and experimental setup Fig. 2 shows the layout of the updated ORC system. The updated system offers several benefits in comparison with the previously constructed one. First, a redesigned separator is adopted. The scheme of the redesigned separator is shown in Fig. 3. Compared with the previous one, this separator has two independent channels (a) for the indicator. The vapor outlet (b) is no longer shared or connected with the indicator (top a). Static pressure at the top of the indicator is influenced more slightly now by the vapor flow through the separator and is approximately equal to that at the bottom. The false liquid level associated with the liquid indicator is avoided. The diameters of the inlet and outlet vapor pipes are also increased. The flow irreversibility through the separator is efficiently reduced, and the turbine inlet pressure is no longer restricted to a low range. Therefore, the ORC can operate in a much wider range of pressures, and more comprehensive information about its performance is available. Second, a larger pressure increment by gravity is guaranteed for HCFC-123 prior to entering the pump. The height difference between the tank and the pump is increased by about 0.8 m. A super-cooled liquid state is more easily

89

maintained during the pumping process. The previous problem, i.e., that cavitation is easily facilitated because of the irreversible pumping process and the low height of the tank, has been addressed. Third, the turbine, gearbox, and generator are set at a place 1.3 m higher. In the operating period, the turbine works at a relatively high temperature. When the operation is completed, the turbine and pipes are cooled down, and HCFC-123 is condensed. Liquid is likely to accumulate inside the turbine and to cause inadequate lubrication damage. A higher position for the turbine is an efficient solution to this problem. Fourth, more bypass conduits are added to enhance the flexibility and reliability of the ORC. These conduits have been proven useful during the equipment commissioning process, and they can help determine the exact cause once the ORC does not work well. A gearbox with reduction ratio of 20:1 is used to ensure a lower rotation speed of the generator. The generator is 8SC3110 (Prestolite), which requires excitation with an input voltage of 24 V. Compared with the permanent magnetic generator, this one costs less. The generator should not be excited when the rotation speed is in low range or when it is off according to product usage directions. Otherwise, it will get hot and eventually damage the rotor. Load to the generator can be changed by the number of working bulbs. Notably, this generator is suitable for an air-conditioned bus, but it was not designed for application in the ORC. It can be replaced by a high-speed qualified generator to provide superior efficiency for converting shaft power into electricity. An oil temperature controller is used to heat HCFC-123. The oil temperature is set to around 100, 90, 80, and 70  C in this experiment. The pump used is CR1-30 (GRUNDFOS). Its input power is

Fig. 3. The redesigned separator: a-indicator interface; b-vapor outlet; c-vapor inlet; d-liquid outlet.

J. Li et al. / Energy 38 (2012) 85e95

The experimental details are summarized in this section. Both energetic and exergetic analyses are carried out. The energy conversion and the irreversible thermodynamics in the evaporator, separator, turbine, condenser, and pump are examined. 4.1. Experimental results The tests of the ORC were performed respectively on February 14, 15, 16, and 17, 2011. The experimental conditions for each of the tests are listed in Table 2. In the following discussion, the timevariations in turbine boundary pressure and temperature are first shown. The corresponding variations in turbine rotation speed, HCFC-123 specific enthalpy drop, mass flow rate, and load to generator are subsequently presented. Fig. 4 shows the pressure and temperature at the turbine inlet varying with time under a heat source temperature of about 100  C. At the beginning, there was a flow rate through the by-pass conduit parallel to the turbine. The by-pass conduit was used to warm up the ORC properly without damaging the turbine. The inlet pressure dropped as the opening degree of the turbine inlet valve increased. At 12:59, the by-pass conduit was quickly shut off, and there was a sudden increment in the inlet pressure. The turbine inlet opening degree was further adjusted until 13:21. The inlet pressure and temperature fluctuated slightly after the inlet valve was fully opened. Fig. 5 shows the variations in the turbine outlet pressure and temperature with time. The outlet pressure varied slightly because of a steady condensing condition. The outlet temperature was

Operational variable

Value in each of the tests

Oil temperature at evaporator inlet ( C) Average flow rate of oil (kg/s) Water temperature at condenser inlet ( C) Average flow rate of water (kg/s) Converter frequency for pumping (Hz) Average flow rate of HCFC-123 with full-open valve (kg/s)

#1 (02/14)

#2 (02/15)

#3 (02/15)

#4 (02/16)

106.0

96.0

85.0

74.0

1.80

1.81

1.81

1.83

4.1

4.0

2.8

2.8

2.12

2.12

2.13

2.11

30

27

24

20

0.11

0.096

0.079

0.057

correlated with the turbine rotation speed. The observed fluctuation was about 14  C. Fig. 6 shows the corresponding turbine rotation speed and HCFC-123 specific enthalpy drop variations with time. The rotation speed and specific enthalpy drop were determined by the opening degree of the turbine inlet valve and load to the generator. The valve opening degree and load were adjusted by hand and by bulbs, respectively. In the startup stage, the bulbs were turned on one by one to ensure a gentle change in turbine rotation speed. At 13:09, the number of bulbs at work was seven. During the period from 13:09 to 13:26, the number of working bulbs remained. The valve opening degree was adjusted gradually, and its maximum was reached at 13:21. Then, ORC ran on a relatively steady condition until 13:26. More bulbs were turned on after 13:26. The rotation speed increased with the increment in the valve opening degree and the decrement in the generator load. Several peak points formed on the turbine rotation speed curve in the startup stage because of the alternating adjustments in the valve opening degree and the generator load. A sudden increment and a sudden decrement occurred in the rotation speed at 13:21 and 13:26, respectively. The turbine rotation speed varied sensitively with the HCFC-123 mass flow rate and the generator load when it was about 25,000 rpm. The rotation speed was relatively high from 13:21 to 13:26, and the average value was about 42,000 rpm. The specific enthalpy drop of HCFC-123 in the expansion process varied similarly with the rotation speed. A higher turbine rotation speed

104 6 100

5

o

4. Results and discussion

Table 2 The ORC test conditions.

inlet temperature C

related to the frequency of the converter. The flow meter for HCFC123 is MFM2081K-60P/DN25 (KROHNE). The zero point stability is 0.012 kg/min, and the measurement accuracy is 0.15% MV þ Cz. Copper-constantan thermocouples are used to measure the temperature in the ORC, with an accuracy of .0.5  C. Two kinds of ceramic pressure transmitters manufactured by Huba Control Co., which range from 1 to 9 bar and from 0 to 25 bar, respectively, with an accuracy of 1.0%, are employed in the experiment to measure the pressure. The measurement errors of the temperature, pressure, and mass flow rate are dependent on the thermocouple, pressure transmitter, and flow meter, respectively. The absolute error between measured temperature and its true value is 0.5  C. The first type of pressure transmitter ranging from 1 to 9 bar is used for the measurement at thermodynamic state points 1, 2, and 6. The other type is used for the measurement at thermodynamic state points 3, 4, and 5. The absolute errors for the two types of pressure transmitters are 0.1 and 0.25 bar, respectively. The absolute error for mass flow rate is 0.00225 kg/s. All the measurement data are recorded and stored on a disk through the computer data-acquisition system Agilent Bench Link Date Logger. The time interval for data acquisition is 5 s. Temperature and pressure informations are transmitted by voltage signal. Mass flow rate and rotation speed informations are transmitted by the current and frequency signals, respectively. Current measurement error is caused by the series burden voltage. Input frequency signal conditioning is performed by Agilent’s ac voltage measurement section. The Agilent features 0.004% basic DC voltage accuracy [18]. The specific enthalpy of HCFC-123 at the liquid or vapor state is mainly determined by temperature. Heat capacities for the liquid and vapor HCFC-123 from 0 to 100  C are about 1.0 kJ/kg/ C and 0.80 kJ/kg/ C, respectively. With a measured temperature error of 0.5  C, the errors for the calculated liquid and vapor specific enthalpies are about 0.5 kJ/kg and 0.4 kJ/kg, respectively.

inlet pressure bar

90

4

96

3 92 2

pressure temperature

1

0

88

84 13:02:00

13:14:30

13:27:00

13:39:30

13:52:00

time Fig. 4. Turbine inlet pressure and temperature varying with time under a heat source temperature of 100  C.

J. Li et al. / Energy 38 (2012) 85e95 90

pressure temperature

3.6

91

flow rate electricity

0.12

900

75 0.10

0.0 13:14:30

13:27:00

13:39:30

V=27.3 N=7 V=27.1 N=8 V=22.4

0.06

N=9

450

300

V=15.7 N=16 V=12.8 N=19 V=10.2 N=23

150

0

0.02

0 13:02:00

600 0.08

0.04

15

0.6

750

electric power output W

30

1.2

mass flow rate kg/s

45 1.8

o

60 2.4

outlet temperature C

outlet pressure bar

3.0

13:02:00

13:52:00

13:14:30

13:27:00

13:39:30

13:52:00

time

time

Fig. 5. Turbine outlet pressure and temperature varying with time under a heat source temperature of 100  C.

Fig. 7. Mass flow rate and electric power output varying with time under a heat source temperature of 100  C.

resulted in a more sufficient expansion and a larger enthalpy drop. The average specific enthalpy drop from 13:21 to 13:26 was about 21.8 kJ/kg. Fig. 7 shows the variations in the HCFC-123 mass flow rate and generator load with time. V and N represent the generator output voltage and the number of bulbs at work, respectively. The bulb model is 24 V, 90/100 W. The electric power output increased as the output voltage and the number of bulbs increased. The output voltage increased as the rotation speed increased. For instance, the output voltage was about 10.2, 15.4, 18.9, 22.0, 25.0, 27.3 and 27.4 V for 15,000, 20,000, 22,000, 24,000, 25,500, 28,100, 44,000 rpm, respectively, when the number of working bulbs was six. The generator output voltage was limited to 28.0 V (the generator rated voltage). With the relationship between the output voltage and the rotation speed, the sudden increment and decrement in Fig. 6 can be more easily understood. The generator output voltage was approximately 27 V when the turbine rotation speed was higher than 25,500 rpm. When the inlet valve opening degree was adjusted to its maximum at 13:21, the mass flow rate increased and the driving force for turbine became larger. As the generator output voltage and the working bulbs remained, load to the generator varied slightly. Therefore, the turbine accelerated greatly until

a new mechanical balance was established, which was attributed to a larger energy transmission loss. Due to more friction and more strenuous motions the energy transmission loss became larger when the rotation speed increased. After the inlet valve was fully opened, the electric output varied step by step as the working bulbs changed. The maximum electric output was about 760 W with 8 bulbs. Notably, the time for the maximum electric output was not the same as that for the highest rotation speed and specific enthalpy drop. Thus, a large proportion of exergy was destructed in the power transmission process when the turbine operated at high rotation speed (around 42,000 rpm) in this ORC. The mass flow rate increased as the turbine inlet valve opening degree increased. The flow rate was measured by the flow meter, and it was equal to that through the pump. The measured flow rate was larger than that through the turbine when the by-pass conduit was open. Based on the variations in HCFC-123 specific enthalpy drop and mass flow rate, a relationship between the turbine shaft power output and the rotation speed can be established. The shaft power output is proportional to the product of the HCFC-123 specific enthalpy drop and mass flow rate through the turbine. In the test, the HCFC-123 mass flow rate was mainly determined by the inlet valve opening degree and was only slightly influenced by

22

12

24000

TD 8

16000

ratation speed rpm

16

32000

enthalpy drop kJ/kg

20

40000

ratation speed rpm

48000

rotation speed enthalpy drop

rotation speed enthalpy drop

40000

18

32000

14

24000

10

TD

16000 6

8000

4

8000

0

0

0

enthalpy drop kJ/kg

24 48000

2 13:02:00

13:14:30

13:27:00

13:39:30

13:52:00

time Fig. 6. HCFC-123 specific enthalpy drop during expansion and turbine rotation speed varying with time under a heat source temperature of 100  C.

10:22:52

10:35:22

10:47:52

11:00:22

11:12:52

11:25:22

time Fig. 8. HCFC-123 specific enthalpy drop during expansion and turbine rotation speed varying with time under a heat source temperature of 90  C.

J. Li et al. / Energy 38 (2012) 85e95

84

pressure temperature

5

5

74

62

3

50

TD 38

2

26

1

inlet pressure bar

4

o

C

72

outlet temperature

inlet pressure bar

6 86

4 60 3

48

2

36

TD 24

1

14

12

0

0 10:22:52

10:35:22

10:47:52

11:00:22

11:12:52

C

pressure temperature

o

6

temperature

92

16:40:30

11:25:22

16:48:50

16:57:10

17:05:30

17:13:50

17:22:10

time

time Fig. 9. Turbine inlet pressure and outlet temperature varying with time under a heat source temperature of 90  C.

Fig. 11. Turbine inlet pressure and outlet temperature varying with time under a heat source temperature of 80  C.

the rotation speed after the by-pass conduit was closed. Therefore, the turbine shaft power output increased as the rotation speed increased. Figs. 4e7 show the major variations in ORC parameters under a heat source temperature of about 100  C. Once the turbine inlet valve was fully open, the inlet temperature, outlet pressure, and HCFC-123 mass flow rate fluctuated slightly. The variation in generator output was appreciable. However, this generator, together with the gearbox, was not qualified for the ORC power transmission. For example, with a specific enthalpy drop of 21 kJ/kg and a flow rate of about 0.11 kg/s, the generated electricity was only 760 W. The ratio of generated electricity to the shaft power delivered from the turbine was only 33%. The gearbox or generator, or both, needed improvement. To present a quick understanding of the energetic and exergetic performance of the thermodynamic cycle, the current paper focuses on the ORC power conversion rather than on the power transmission. After excluding the inlet temperature, outlet pressure, mass flow rate, and load to the generator, only the HCFC-123 specific enthalpy drop, turbine rotation speed, inlet pressure, and outlet temperature are shown for the following tests. And the ORC efficiency is based on the shaft power output. Fig. 8 shows the turbine rotation speed and the HCFC-123 specific enthalpy drop varying with time under a heat source

temperature of about 90  C. From 10:18 to 10:43, the number of bulbs at work was kept constant. The turbine rotation speed increased because of the increment in the valve opening degree. At 10:43, the rotation speed reached about 41,000 rpm, and a sufficient expansion was available. Then, the ORC ran steadily from 10:43 to 10:48. The average specific enthalpy drop was about 19.6 kJ/kg. At 10:48, more bulbs were turned on, which resulted in a decrement in the rotation speed. The valve opening degree was further adjusted, and its maximum was reached at 10:57. More bulbs were turned on continuously. Fig. 9 shows the corresponding turbine inlet pressure and outlet temperature varying with time. The inlet pressure fluctuated around 4.7 bar after the inlet valve was fully opened. The outlet temperature decreased as the rotation speed increased. During the period from 10:43 to 11:00, the outlet temperature fluctuated slightly at around 51.5  C. Fig. 10 shows the turbine rotation speed and the HCFC-123 specific enthalpy drop varying with time under a heat source temperature of about 80  C. From 16:53 to 17:01, the number of bulbs at work remained constant. The turbine rotation speed was about 40,000 rpm at 17:01. The ORC ran steadily from 17:01 to 17:06. The average specific enthalpy drop was about 18.7 kJ/kg. Then more bulbs were turned on continuously. Fig. 11 shows the variations in turbine inlet pressure and outlet temperature. The

20

48000

enthalpy drop

40000

32000 12 24000

TD

8

16000 4 8000

0 16:48:50

16:57:10

17:05:30

17:13:50

0 17:22:10

time Fig. 10. HCFC-123 specific enthalpy drop during expansion and turbine rotation speed varying with time under a heat source temperature of 80  C.

ratation speed rpm

16

enthalpy drop kJ/kg

ratation speed rpm

40000

16:40:30

15

rotation speed enthalpy drop

rotation speed

12

32000 9 24000 6 16000

TD

enthalpy drop kJ/kg

48000

3

8000

0

0 11:15:08

11:19:18

11:23:28

11:27:38

11:31:48

time Fig. 12. HCFC-123 specific enthalpy drop during expansion and turbine rotation speed varying with time under a heat source temperature of 70  C.

J. Li et al. / Energy 38 (2012) 85e95 Table 3 Distribution of thermodynamic parameters on relative steady conditions.

80

5

pressure temperature

68

o

Case Ⅰ

3

56

2

44

TD 1

32

0

outlet temperature

inlet pressure bar

C

4

Case Ⅱ

20 11:15:08

11:19:18

11:23:28

11:27:38

93

11:31:48

time Fig. 13. Turbine inlet pressure and outlet temperature varying with time under a heat source temperature of 70  C.

inlet pressure and outlet temperature fluctuated around 3.8 bar and 44  C, respectively, from 17:01 to 17:09. Fig. 12 shows the turbine rotation speed and the HCFC-123 specific enthalpy drop varying with time under a heat source temperature of about 70  C. Only one bulb was turned on because of the inefficient generator and the irreversible power transmission between the turbine and the gearbox. The turbine rotation speed increased as the valve opening degree increased, but it was about 21,000 rpm even when the valve was fully opened at 11:22. The corresponding specific enthalpy drop was about 12.3 kJ/kg. The valve opening degree was decreased at 11:31. The ORC operated under quite steady inlet pressure and outlet temperature from 11:22 to 11:31 as shown in Fig. 13.

Case Ⅲ

Case Ⅳ

4.2. Energetic and exergetic analyses Table 3 shows the thermodynamic parameter distribution in the ORC. Considering the errors in measuring instantaneous values, the presented temperature and pressure are the mean recorded values. Case I is related to the test carried out under a heat source temperature of about 100  C. With a time interval of 5 s for data acquisition, there are about 12 groups of data averaged from 13:24:00 to 13:24:59. The ORC ran on a relatively steady condition from 13:21 to 13:26 and the values for Case I are typical in that period. The presented enthalpy, entropy, and exergy are calculated by pressure and temperature using the Computer-Aided Thermodynamic Table 2 Version 1.0 software [19]. Cases II, III, and IV are related to the tests carried out under heat source temperatures of about 90, 80 and 70  C, respectively. The presented values are the average recorded values around 10:45, 17:04, and 11:23, respectively. The time intervals for averaging data points are marked with TD in the corresponding figures. The thermodynamic parameters during TD are accompanied by high turbine rotation speeds for each test. At least 5 min are guaranteed for the ORC to operate at high rotation speeds without any adjustment in the inlet valve or generator load. According to the corresponding figures, variations in the thermodynamic parameters during TD are slight. From the point of view of quasi-static process, performing steady state calculations during TD using the equations presented in Section 2 is reasonable, although the entire testing process is dynamic. The entropy at state point 4 is higher than that at state point 5 because of heat loss from the separator and pipes. In Cases II and III, the exergy at state point 2 is slightly higher than that at state point

State point

Fluid

1 2 3 4 5 6 7 8 9 10 1 2 3 4 5 6 7 8 9 10 1 2 3 4 5 6 7 8 9 10 1 2 3 4 5 6 7 8 9 10

R123 R123 R123 R123 R123 R123 water water Oil Oil R123 R123 R123 R123 R123 R123 Water water Oil Oil R123 R123 R123 R123 R123 R123 Water water Oil Oil R123 R123 R123 R123 R123 R123 Water water Oil Oil

p kPa

t

57.8 69.0 610.2 590.3 567.0 90.2 / / / / 55.9 66.6 497.8 478.6 459.7 83.0 / / / / 56.5 63.4 414.8 396.9 381.5 73.1 / / / / 51.4 60.8 301.2 291.5 280.8 68.4 / / / /

6.2 6.3 8.4 99.5 97.6 58.1 6.7 4.1 106.9 100.8 5.8 5.9 8.1 89.7 87.6 51.5 6.2 4.0 95.9 90.8 4.0 4.2 6.3 79.9 78.0 43.9 4.5 2.8 85.2 81.0 3.8 3.9 6.1 70.1 68.5 45.6 4.1 2.8 74.0 71.1



C

h kJ/kg

s kJ/kg/ C

e kJ/kg

44.9 45.0 47.3 281.8 280.6 258.8 28.2 17.3 233.5 218.8 44.5 44.6 46.98 275.7 274.4 254.2 26.2 16.9 207.2 195.1 42.7 42.93 45.15 269.4 268.1 249.0 19.1 11.8 182.1 172.3 42.5 42.6 44.9 263.8 262.8 250.3 17.4 11.8 156.2 149.5

0.1755 0.1759 0.1827 0.8614 0.8602 0.8914 0.1017 0.0625 0.7185 0.6796 0.1741 0.1744 0.1818 0.8550 0.8533 0.8817 0.0942 0.0610 0.6483 0.6154 0.1676 0.1683 0.1755 0.8464 0.8443 0.8722 0.0690 0.0427 0.5792 0.5517 0.1669 0.1673 0.1749 0.8456 0.8445 0.8799 0.0625 0.0427 0.5058 0.4865

0.009 0.001 0.442 49.555 48.683 18.360 0.2983 0.086 37.242 33.167 0.009 0.009 0.368 45.203 44.368 16.410 0.2497 0.080 30.117 27.003 0.0334 0.005 0.259 41.252 40.526 13.805 0.112 0.018 23.891 21.603 0.042 0.051 0.172 35.871 35.171 13.002 0.0861 0.018 18.040 16.612

1 because of the gravity acting on HCFC-123. The pressure loss in the separator has been greatly reduced in the updated ORC. Table 4 shows the energy load for the evaporator, condenser, turbine, and pump based on the thermodynamic parameter distribution in Table 3. The practical ORC efficiencies for Cases I, II, III, and IV are 8.2%, 7.7%, 7.4%, and 4.6%, respectively. The degrees of superheat at the turbine inlet are 11.6, 10.1, 7.6, and 9.2  C, respectively. The ORC efficiency for Case I is highest because of the highest heat source temperature. Although the oil temperature for Case III at the evaporator inlet is about 9.4  C lower than that for Case II, the ORC efficiency is only 0.3% lower. This result is attributed to the smaller degree of superheat for Case III. The ORC efficiency for Case IV is lowest because of the lowest heat source temperature and the insufficient expansion of HCFC-123. A higher rotation speed results in a more sufficient expansion of HCFC-123. The isentropic turbine efficiency for Cases I, II, and III is 0.68, which is about 0.14 higher than that for Case IV. Table 4 Energy analysis of ORC components unit: kW.

Case Case Case Case

Ⅰ Ⅱ Ⅲ Ⅳ

Evaporator mf ðh4  h3 Þ

Condenser mf ðh6  h1 Þ

Turbine mf ðh5  h6 Þ

Pump mf ðh3  h2 Þ

Thermal efficiency

25.79 21.72 17.49 12.25

23.52 19.92 15.92 11.63

2.40 1.92 1.49 0.70

0.25 0.22 0.17 0.12

8.2% 7.7% 7.4% 4.6%

94

J. Li et al. / Energy 38 (2012) 85e95

Table 5 Examination of exergy loss in the low-grade heat to power conversion process unit: kW.

Case Case Case Case

Ⅰ Ⅱ Ⅲ Ⅳ

Evaporator

Separator

Turbine

Condenser

Pump

1.932 1.376 0.944 0.614

0.096 0.079 0.057 0.039

0.937 0.737 0.573 0.541

1.569 1.20 0.847 0.586

0.204 0.192 0.153 0.116

The updated ORC demonstrates adequate expander and cycle efficiencies compared with similar systems tested in previous research [4,8e11]. However, to improve the ORC further, a close examination of the irreversible thermodynamics taking place in the components is necessary. Table 5 shows the exergy loss in the evaporator, separator, turbine, condenser, and pump. The total exergy loss for Cases I, II, III, and IV is 4.7, 3.5, 2.5, and 1.8 kW, respectively. In all cases, the evaporator has the largest exergy loss, followed successively by the condenser, turbine, pump, and separator. The percentage of exergy loss in each component, with respect to the total system exergy loss for Case I, is shown in Fig. 14. About 74% of the total system exergy loss is contributed by the heat exchangers. The heat-transfer irreversibility in the evaporator is large because of the great difference between the oil outlet temperature and the HCFC-123 inlet temperature. To address this problem, two-stage heat exchangers were previously suggested, in which the mass flow rate of the oil in the first-stage heat exchanger was smaller than that in the second-stage heat exchanger. Given a matched heat transfer capacity between the oil and the organic fluid in the first stage heat exchanger, the thermodynamic irreversibility was efficiently reduced [20,21]. The necessity of this configuration is experimentally confirmed in the current paper. The heat-transfer irreversibility in the condenser is also large because of the high degree of superheat for HCFC-123 at the turbine outlet. For example, the degree of superheat for HCFC-123 after expansion is about 35  C, and the temperature difference between the superheated vapor and the condensed liquid is more than 50  C. The heat-transfer irreversibility is appreciable because the temperature difference between the hot and cold sides is only 100  C. To recover the heat from the turbine outlet vapor, an IHE (internal heat exchanger) was proposed [22]. The IHE is useful especially when the vapor leaving the turbine is superheated. Therefore, a regenerator is added when using valves 3, 4, 13, and 14

Table 6 Exergy efficiency of ORC components.

Case Case Case Case

Ⅰ Ⅱ Ⅲ Ⅳ

Evaporator

Turbine

Condenser

Pump

0.737 0.756 0.772 0.765

0.719 0.723 0.723 0.564

0.223 0.231 0.192 0.197

0.192 0.151 0.114 0.097

to improve this ORC further, as shown in Fig. 1. The vapor from the turbine is cooled prior to entering the condenser. The exergy efficiencies of the vaporator, condenser, turbine, pump, and separator are shown in Table 6. The separator, whose exergy efficiency is zero, is not used for power conversion or heattransfer purpose. The evaporator has the highest exergy efficiency, followed successively by the turbine, condenser, and pump. This finding indicates that the largest exergy loss is not necessary because of the lowest exergy efficiency. The available exergy for the evaporator is much larger than that for the condenser. The exergy for the ORC is mainly available from the oil; it is 7.3, 5.6, 4.1, and 2.5 kW for Cases I, II, III, and IV, respectively. The system exergy efficiencies are 0.37, 0.38, 0.40, and 0.31, respectively. 5. Conclusion The updated ORC system has many benefits compared with the previously constructed one. The limitation of the system’s working pressure in the separator is overcome. The cavitation problem caused by the irreversible pumping process and the low height of the tank has been addressed. A close examination of the energy conversion and the irreversible thermodynamics is conducted for the updated ORC at different heat source temperatures of about 100, 90, 80, and 70  C. The following conclusions are drawn from the energetic and exergetic performance investigation carried out in the current paper: 1) The specific enthalpy drop of the HCFC-123 during expansion and the turbine rotation speed varied in a similar trend in the dynamic tests of the ORC. A more sufficient expansion resulted from a higher turbine rotation speed. The isentropic turbine efficiency was 0.68, with a rotation speed of around 40,000 rpm. The ORC successfully converted the low-grade heat of less than 80  C into power. A thermal efficiency of 7.4% was obtained, with a hot side temperature of about 80  C and a cold side temperature of about 5  C. 2) The thermodynamic irreversibility taking place in the evaporator, condenser, turbine, pump, and separator was appreciable, and the system exergy efficiency was around 40%. The exergy loss in the evaporator was largest, followed by that in the condenser. The exergy loss associated with the heat exchangers amounted to 74% of the total system exergy loss in the tests. Two-stage heat exchangers could be a good solution to reduce the heat-transfer irreversibility in the evaporator, and a regenerator could help reduce the exergy loss in the condenser. Acknowledgments

Fig. 14. Percentage of the exergy loss in each component.

The current study was sponsored by the National Science Foundation of China (50974150, 50978241), the National Basic Research Program of China (2011CB211703), the Excellence Youth Science Foundation of Anhui Province of China (10040606Y20), and the Fundamental Research Funds for the Central Universities of China.

J. Li et al. / Energy 38 (2012) 85e95

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Nomenclature E: exergy flow, W h: enthalpy, J/Kg m: mass flow rate, Kg/s p: pressure, Pa Q: heat flow, W s: entropy, kJ/kg/ C T: temperature,  C W: power, W 3 : machine efficiency h: energy/exergy efficiency x: rate of exergy loss, W Subscripts/superscripts 1e10: state points b: gearbox c: condenser cond: conduction conv: convection e: evaporator f: fluid g: generator i: element in: input loss: heat loss n: net out: out p: pump rad: radiation s: separator t: turbine u: used