Selection of working fluids for a novel low-temperature geothermally-powered ORC based cogeneration system

Energy Conversion and Management 52 (2011) 2384–2391

Contents lists available at ScienceDirect

Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman

Selection of working fluids for a novel low-temperature geothermally-powered ORC based cogeneration system T. Guo ⇑, H.X. Wang, S.J. Zhang Department of Thermal Energy and Refrigeration Engineering, School of Mechanical Engineering, Tianjin University, No. 92 Weijin South Road, Nankai District, Tianjin 300072, PR China

a r t i c l e

i n f o

Article history: Available online 4 March 2011 Keywords: Organic Rankine cycle (ORC) Low-temperature geothermal Working fluids Cogeneration Heat pump Disturbance condition

a b s t r a c t A novel cogeneration system driven by low-temperature geothermal sources was investigated in this study. This system consists of a low-temperature geothermally-powered organic Rankine cycle (ORC) subsystem, an intermediate heat exchanger and a commercial R134a-based heat pump subsystem. The main purpose is to identify appropriate fluids which may yield high PPR (the ratio of power produced by the power generation subsystem to power consumed by the heat pump subsystem) value and QQR (the ratio of heat supplied to the user to heat produced by the geothermal source) value. Performances of the novel cogeneration system under disturbance conditions have also been studied. Results indicate that fluids group presenting higher normal boiling point values shows averagely 7.7% higher PPR values and R236ea and R245ca outstand among the group. DTP (pinch temperature difference in heat exchangers) and gt (turbine efficiency) values play more important roles on the variation of PPR values. QQR values change slightly with various DTP, gt and grp (refrigerant pump efficiency) values while the variation range is larger under various geothermal source and heating supply parameters. Smaller DTP value, higher gt value, higher geothermal source parameters and lower heating supply parameters lead to higher PPR values but lower QQR values. Ó 2011 Elsevier Ltd. All rights reserved.

1. Introduction Low-temperature (<100 °C) geothermal, being an environmentally friendly heat source, has been widely used for heating in many places in China. Tianjin is the largest user of geothermal resources in China. Direct uses have been extended from industry and agriculture to direct heating, swimming pools, green house and medical therapy, etc. [1]. But for many cases where there exists a large temperature difference between the geothermal temperature and the heating supply temperature, a large irreversibility rate loss may appear. However, this issue would be overcome if a power generation system is incorporated, in order to make the geothermal source temperature decreased to a certain value which is a little higher than the heating supply temperature. Another issue is that the exhaust temperature of geothermal source is generally high which can lead to two disadvantages. On the one hand, the thermal energy in the geothermal source is not fully used. On the other hand, it will cause thermal pollution. However, this issue would be overcome if a heat pump system is incorporated, in order to make the exhaust temperature of geothermal source decreased to an appropriate value. So an innovative cogeneration system [2]

⇑ Corresponding author. Tel.: +86 22 2740 5049. E-mail address: [email protected] (T. Guo). 0196-8904/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2010.12.038

driven by low-temperature geothermal sources consisting of a power generation subsystem and heat pump subsystem is proposed in this study. The power generation subsystem employed in the novel cogeneration system is a subcritical organic Rankine cycle (ORC) system which has become a promising technology for conversion of lowtemperature heat to electricity and offers an advantageous efficiency in small-scale applications in recent years. The heat pump subsystem employed in the innovative cogeneration system is a commercial R134a-based heat pump system which is commonly used, technologically matured and of high efficiency. Many investigations concerning low-temperature geothermally-powered ORC have been carried out [3–13]. Hettiarachchi et al. [3] investigated the optimum working fluid and cycle parameters of a low-temperature geothermal source (70–90 °C) powered ORC, choosing the ratio of total heat transfer area to the net power out as the objective function. The working fluids considered were PF5050, R123, Ammonia and n-pentane. Subbiah and Natarajan [4] investigated the performance of a binary-fluid cycle based geothermal power plant using R113. Parameters under investigated were the first and second law cycle efficiencies. And it showed that the second law approach helped to identify optimal working conditions for maximizing the work output. Saleh et al. [5] investigated 31 different pure working fluids for a low-temperature geothermal (100 °C or higher) powered ORC system.

T. Guo et al. / Energy Conversion and Management 52 (2011) 2384–2391

2385

Nomenclature ALT cp EXPr GWP h m M ODP p P,W Q T Tr DT DHH V

g

atmosphere life time (yr) isobaric specific heat (kJ/(kgK)) expansion ratio in the turbine global warming potential enthalpy (kJ/kg) mass flow rate (kg/s) molecular weight (g/mol) ozone depletion potential pressure (MPa) power produced or consumed (kW) heat rate injected or rejected (kW) temperature (°C) reduced temperature temperature difference (K) enthalpy of vaporization (kJ/kg) volume flow rate (m3/s) efficiency

Subscripts bp normal boiling point comp compressor

Performance parameters considered mainly included thermal efficiency, volume flow rate at turbine inlet and expansion ratio. Working fluid R152a was recommended. Desideri and Bidini [6] investigated three configurations of Rankine cycle driven by geothermal sources and compared them with conventional single and dual flash stream power plants. The regenerated Rankine cycle with a closed heat exchanger was recommended. Iqbal et al. [7] carried out a sensitivity analysis of an isobutane-based binary ORC system driven by low-temperature geothermal sources, to determine the optimum operating conditions with minimum cost. It showed that a subcritical pressure cycle was lower in cost than a supercritical pressure cycle which had a higher thermodynamic efficiency. Gu and Sato [8,9] investigated a transcritical ORC system using geothermal sources (>190 °C). Performance parameters considered included thermal efficiency and work output. R134a, R290 and R125 were considered and R134a was recommended. Guo et al. [10] investigated pure working fluids for the transcritical ORC driven by low-temperature geothermal sources (80–120 °C). And R125 were recommended as the supercritical fluid. Heberle et al. [11] carried out a second law analysis for the series and parallel circuits of an organic Rankine cycle (ORC) and an additional heat generation for combined heat and power generation. The working fluids investigated were R227ea, R245fa, pentane and isopentane. Results showed that a series circuit with isopentane was most efficient. For parallel circuits and for power generation, R227ea was preferred. A model of an ammonia-water Rankine power cycle was developed and examined by Wagar et al. [12]. Kontoleontos et al. [13] presented an ORC machine that generated both heat and power by heat recovery from the cooling water circuit, corresponding to geothermal source of 120–150 °C and cooling water supplying a district heating system at 60/80 °C. Working fluids of R134a and R600a were compared based on a criterion of minimum system cost which was represented as minimum heat transfer areas given a constant net power output. Results showed that R134a was preferred. CO2 which has a lower critical temperature and good environmental characteristics have been most commonly used as the working fluid for a transcritical Rankine cycle, called CO2-based transcritical Rankine cycle instead of transcritical ORC. Zhang et al. [14–19] investigated a solar Rankine system using supercritical CO2 theoretically and experimentally with hot water temperature of above 150 °C. Performances of the system for combined

cri e, H geo gt i max min net o p r_ORC r_R134a rp s sup t th 1–4 1–8

critical evaporator or evaporating geothermal source generator in maximum minimum net out pinch working fluid in power generation subsystem R134a in heat pump subsystem refrigerant pump isentropic heating supply water turbine thermal geothermal temperature nodes states points in the cycle

power and heat generation were also analyzed. Chen et al. [20] analyzed performances of a CO2-based transcritical Rankine cycle theoretically for automobile waste heat recovery with turbine or expander inlet temperature of 200 °C. According to the previous references with respect to low-temperature geothermal powered ORC systems discussed in the above section, it shows that more attention has been paid to the screening of appropriate working fluids. Moreover, different screening criteria may lead to different working fluid recommendations [3,5,8–11]. Few Refs. [3,11,13] investigated low-temperature geothermal-powered ORC based combined heat and power generation system, just using the heat of cooling water to provide the heat supply. The exhaust temperature of geothermal source was generally high and the energy in geothermal sources has not been fully used. In addition, no reference has been found discussing such a system that combines an ORC system and a heat pump system for the combined heat and power generation. In this study, an innovative cogeneration system [2] powered by low-temperature geothermal sources is proposed and thermodynamically analyzed. The suitable working fluids have been screened and recommended for the ORC-based power generation subsystem. Performances of the novel cogeneration system under disturbance conditions have also been studied. The operation conditions are: geothermal source temperature ranging from 80 to 100 °C, geothermal exhaust temperature from 10 to 30 °C, heating supply temperature from 30 to 50 °C and heating return water temperature from 25 to 40 °C, pinch temperature difference in heat exchangers (DTP) from 1 to 6 K, turbine efficiency (gt) and refrigerant pump efficiency (grp) both ranging from 40% to 80%. The considered criteria are high PPR (the ratio of power produced by the power generation subsystem to power consumed by the heat pump subsystem) and QQR (the ratio of heat supplied to the user to heat produced by the geothermal source) values.

2. Choice of working fluids There are several general criteria while screening the proper working fluids. Stability, non-fouling, non-corrosiveness, non-toxicity and non-flammability are a few preferable physical and chemical characteristics. However, in a cycle design, not all the desired

2386

T. Guo et al. / Energy Conversion and Management 52 (2011) 2384–2391

Table 1 Physical, safety and environmental data of working fluids [21]. Substances

1 2 3 4 5 6 7 8

Physical data

R717 R152a R134a R600a R236ea R600 R245fa R245ca

Environmental data

M (g/mol)

Tbp (°C)

Tcri (°C)

pcri (MPa)

ALT (yr)

ODP (–)

GWP (100 yr)

17.03 66.05 102.03 58.12 152.04 58.12 134.05 134.05

33.33 24.02 26.07 11.67 6.19 0.55 14.90 25.13

132.25 113.26 101.06 134.67 139.29 151.98 154.05 174.42

11.333 4.517 4.059 3.640 3.502 3.796 3.640 3.925

0.01 1.4 14 0.02 8 0.02 7.6 6.2

0 0 0 0 0 0 0 0

<1 124 1430 20 710 20 1030 693

Results of Eq. (1)

Type of fluids

11.30 1.97 0.86 0.27 0.03 0.42 0.40 0.14

Wet Wet Isentropic Isentropic Isentropic Isentropic Isentropic Isentropic

Fig. 1. Schematic diagram of the novel cogeneration system.

where n (=dS/dTH) is the slope of saturated vapor curve on T–s diagram. Types of working fluids can be predicted by Eq. (1). That is, n > 0: a dry fluid, n  0: an isentropic fluid, and n < 0: a wet fluid [22]. Table 1 shows physical properties and environmental data of fluids considered in this study and the results predicted by Eq. (1).

3. System description and modeling

Fig. 2. T–s diagram of the novel cogeneration system.

general requirements can be satisfied. As a result of a first screening, 8 fluids presented in Table 1 emerged as potential candidates. The criteria considered here are as follows: zero-ODP value, GWP value below 1500, appropriate critical temperature values and they can be commercially purchased. And they can be categorized into two groups as wet and isentropic fluids according to Eq. (1) presented in Ref. [22].

n ¼ cp =T H  DHH  ð0:38  T rH =ð1  T rH Þ þ 1Þ=T 2H

ð1Þ

The innovative cogeneration system layout and the corresponding T–s diagram are shown in Figs. 1 and 2, respectively. The system consists of a low-temperature geothermally-powered ORC subsystem, an intermediate heat exchanger and a commercial R134a-based heat pump subsystem. The geothermally-powered ORC subsystem can be identified as 1 ? 2 ? 3 ? 4 ?1. The process proceeding in the intermediate heat exchanger can be presented by the temperature decrease of geothermal source (T2 ? T3) and the temperature increase of heating supply water (T5A ? T6A) simultaneously. The R134a-based heat pump subsystem process is 5 ? 6 ? 7 ? 8 ? 5. The novel system also includes three types of fluid circuits, that is, the geothermal water circuit, the heating supply water circuit and the working fluid circuit, respectively. During the geothermal water circuit, three temperature decrease processes occur in the evaporator of power generation subsystem, intermediate heat exchanger and evaporator of heat pump subsystem presented by T1 ? T2, T2 ? T3 and T3 ? T4, respectively. The circuit of heating supply water can be presented by a temperature increase (T5 ? T6) occurs in the condenser of power generation subsystem (B), intermediate heat exchanger (A) and condenser of heat pump subsystem (C). The working fluid circuit occurs in the power

2387

T. Guo et al. / Energy Conversion and Management 52 (2011) 2384–2391

generation and heat pump subsystem, respectively and is discussed in the following section. The low-temperature geothermally-powered ORC subsystem analyzed here is a subcritical ORC, consisting of an evaporator, a turbine/expander, a condenser, a refrigerant pump and the heat transfer fluid systems (geothermal water and heating supply water cycling systems). The schematic diagram of the subcritical ORC as a subsystem of the novel cogeneration system is shown in Fig. 1. The generated high pressure vapor in the evaporator flows into the turbine/expander and its enthalpy is converted into work. The low pressure vapor exits the turbine/expander and is led to the condenser where it is liquefied by heating supply water. The liquid available at the condenser outlet is pumped into high pressure liquid, and then flows into the evaporator where it is heated and vaporized by the geothermal water and a new cycle begins. The above described processes are presented in the T–s diagram shown in the right part of Fig. 2. The heat pump subsystem presented here is a commercial R134a-based heat pump system, consisting of an evaporator, a compressor, a condenser, an expansion valve and the heat transfer fluid systems (geothermal water and heating supply water cycling systems). The schematic diagram of the heat pump system as a subsystem of the novel combined power and heat generation system is shown in Fig. 1. The high pressure vapor at the outlet of the compressor flows into the condenser where it is liquefied by heating supply water. The high pressure liquid available at the condenser outlet is converted into low pressure two phase fluid through the expansion valve. The low pressure two phase working fluid flows into the evaporator where it is heated and vaporized by the geothermal water, and then it is pumped into high pressure vapor through the compressor and a new cycle begins. The above described processes are presented in the T–s diagram shown in the left part of Fig. 2. The advantages of the novel cogeneration system are as follows, on the one hand, the heating return water is employed as the cooling circling water for the power generation subsystem avoiding the use of an additional cooling system, leading to a lower system investment. On the other hand, an intermediate heat exchanger and a commercial R134a-based heat pump subsystem are incorporated into the innovative combined power and heat generation system leading to a lower geothermal source exhaust temperature, maximizing the usage of thermal energy in the low-temperature geothermal source and avoiding the thermal pollution. The following general assumptions are formulated for this study: Each component is considered as a steady-state steady-flow system, the kinetic and potential energies as well as the heat and friction losses are neglected. The following sections introduce the equations used to perform the analysis. 3.1. Power generation subsystem Evaporator

Q e ¼ mr

ORC ðh1

 h4 Þ:

ð2Þ

Turbine

W t ¼ mr

ORC ðh1

 h2s Þgt ¼ mr

ORC ðh1

 h2 Þ:

ð3Þ

Condenser

Q sup1 ¼ mr

 h3 Þ:

ORC ðh2

ð4Þ

Refrigerant Pump

W rp ¼ mr

ORC ðh4s

 h3 Þ=grp ¼ mr

ORC ðh4

 h3 Þ:

ð5Þ

Thermal efficiency

gth ¼ ðW t  W rp Þ=Q e :

ð6Þ

Net power output

Pnet ¼ ggt W t  W rp :

ð7Þ

3.2. Intermediate heat exchanged

Q sup2 ¼ cp mgeo ðT 2  T 3 Þ:

ð8Þ

3.3. R134a-based heat pump subsystem Compressor

Pcomp ¼ mr

R134a ðh5

 h8 Þ:

ð9Þ

 h6 Þ:

ð10Þ

Condenser

Q sup3 ¼ mr

R134a ðh5

Coefficient of performance (COP)

COP ¼ ðh5  h6 Þ=ðh5  h8 Þ:

ð11Þ

System parameter

Q sup ¼ Q sup1 þ Q sup2 þ Q sup3 :

ð12Þ

Q geo ¼ cp mgeo ðT 1  T 4 Þ: PPR ¼ P net =Pcomp :

ð13Þ ð14Þ

QQR ¼ Q sup =Q geo :

ð15Þ

4. Validation Numerical solution is validated with the results of Saleh et al. [5] for various working fluids-based ORC without regenerator (internal heat exchanger (IHX)) and for the same operating conditions listed as follows. The cycle in Ref. [5] was using geothermal heat source with 120 °C available temperature and the mass flow rate refer to a net power output of 1 MW. The condensation temperature was 30 °C, the isentropic turbine efficiency was 0.85, and the isentropic pump efficiency was 0.65. The turbine inlet temperature was 20 °C lower than the geothermal source temperature, i.e. 100 °C. The pinch point temperature difference was 10 °C. The comparison shows very good agreement between present solution and the results of Saleh et al. [5] as shown in Table 2. The discrepancies of different parameters showed in Table 2 mainly derive from the choosing of equation of state (EOS) that the BACKONE

Table 2 Validation of the numerical model with the previously published data [5] for various fluids-based ORC. Substance

Tc (°C)

pcri (MPa)

Tt,i (°C)

Tt,o (°C)

pmax (MPa)

pmin (MPa)

mr_ORC (kg/s)

Vt,i (m3/s)

Expr (–)

gth (%)

Source

-IHX R245ca R245ca R245fa R245fa

174.42 174.42 154.05 154.05

3.925 3.925 3.640 3.640

100.0 100.0 100.0 100.0

53.75 51.56 50.70 49.68

0.934 0.928 1.267 1.261

0.123 0.122 0.180 0.179

30.548 32.124 33.424 35.466

0.619 0.638 0.468 0.493

7.88 8.01 7.61 7.65

12.79 12.44 12.52 12.03

[5] Present [5] Present

2388

T. Guo et al. / Energy Conversion and Management 52 (2011) 2384–2391

Table 3 Specifications of the cycle parameters considered. Cycle parameters

Labels

Value range

Typical value

Unit

Geothermal source temperature Geothermal exhaust temperature Mass flow rate of geothermal Heating supply temperature Heating return temperature Pinch temperature difference Turbine efficiency Generator efficiency Refrigerant pump efficiency Compressor efficiency

T1

80–100

90

°C

T4

10–30

25

°C

mgeo T6 T5 DTP

1 30–50 25–40 1–6 40–80 96 40–80 60

1 45 35 5 60 96 70 60

kg/s °C °C K % % % %

gt ggt grp gcomp

Fig. 4. Thermal efficiency (gth) and COP values of fluids versus pinch value in heat exchangers (DTP).

Fig. 3. PPR and QQR values of fluids versus pinch value in heat exchangers (DTP).

EOS was selected in Ref. [5] while the standard EOS was chosen in the present study.

5. Results and discussion The operation conditions of the novel cogeneration system are given in Table 3. In the following paragraphs we present the results of a comparative study for different working fluids-based cogeneration system based on the model previously described. Performances of the cogeneration system under disturbance conditions have also been studied. The thermodynamic properties of fluids are evaluated with NIST RFPROP 7.0 [23], which has a sufficient accuracy. Given system parameters listed in Table 3, the evaporating temperature of the ORC-based power generation subsystem for each working fluid and each operation parameter combination is optimized. And the evaluation process can be carried out equally. Performances of different fluids under disturbance conditions for the novel cogeneration system are shown in Figs. 3–16 given other cycle parameters as typical values. The results shows that the capacity of the commercial R134a heat pump ranges from 1.9 to 42.2 kW with various parameters listed in Table 3. Figs. 3–8 show the variation of PPR and QQR values of 8 fluids considered and thermal efficiency and COP values under various pinch temperature difference values in heat exchangers (DTP), turbine efficiency (gt) and refrigerant pump efficiency(grp) listed in Table 3. With the increase of DTP value, the condensing temperature and the geothermal outlet temperature of the ORC subsystem increase leading to decreasing thermal efficiency (as shown in Fig. 4) and heat injected values. As a result, the net power output of ORC subsystem decreases. From Fig. 4, we can see that the COP value for the

Fig. 5. PPR and QQR values of fluids versus turbine efficiency (gt).

Fig. 6. Thermal efficiency (gth) and COP values of fluids versus turbine efficiency (gt).

heat pump also decreases with the increase of DTP value and more power will be consumed. Therefore, PPR values decrease with the increase of DTP value as shown in Fig. 3. But QQR values show slightly increasing tendency (1.01–1.05). The eight fluids can be divided into two groups according to PPR values. The first group includes three fluids, i.e. NH3, R134a and R152a presenting lower

T. Guo et al. / Energy Conversion and Management 52 (2011) 2384–2391

Fig. 7. PPR and QQR values of fluids versus refrigerant pump efficiency (grp).

Fig. 8. Thermal efficiency (gth) and COP values of fluids versus refrigerant pump efficiency (grp).

2389

Fig. 10. QQR values of R245ca versus DTP values with various turbine efficiencies (gt).

Fig. 11. Thermal efficiency (gth) and COP values of R245ca versus DTP values with various turbine efficiencies (gt).

Fig. 9. PPR values of R245ca versus DTP values with various turbine efficiencies (gt).

normal boiling point values, showing lower PPR values. The second group presents averagely 7.7% higher PPR values and R236ea and R245ca show a little higher PPR values among the group. All the fluids considered show nearly the same QQR values which change from 1.010 to 1.045 with various DTP values ranging from 1 to 6 K. The variation of PPR values are larger with DTP increasing from 1 to 6 K presenting 57.18–14.46% for R245ca.

Fig. 12. PPR and QQR values of R245ca versus geothermal source and exhaust temperatures.

As shown in Figs. 5–8, PPR and thermal efficiency values increase with the increase of efficiencies of turbine and refrigerant pump. It is apparent that turbine efficiency has more influence on PPR and thermal efficiency values than refrigerant pump efficiency. The COP value for the heat pump is independent of turbine

2390

T. Guo et al. / Energy Conversion and Management 52 (2011) 2384–2391

Fig. 13. Thermal efficiency (gth) and COP values of R245ca versus geothermal source and exhaust temperatures.

Fig. 14. PPR values of R245ca versus various heating supply temperature with various heating return water temperature.

Fig. 15. QQR values of R245ca versus heating supply temperature with various heating return water temperature.

and refrigerant efficiency as shown in Figs. 6 and 8. As shown in Fig. 5, average PPR values change from 10.88% to 24.64% with turbine efficiency increasing from 40% to 80%. The PPR values are nearly free of influence caused by the increase of refrigerant pump efficiency as shown in Fig. 7. And the variation of R134a presents the largest increasing from 14.34% to 17.80%.

Fig. 16. Thermal efficiency (gth) and COP values of R245ca versus heating supply temperature with various heating return water temperature.

From Figs. 3–8, it can also be seen that fluid presenting a higher PPR value would show a lower QQR value. Based on above discussions, QQR values change slightly with various DTP, gt and grp values. Moreover, DTP and gt values play more important roles on the variation of PPR values. The combined influence of DTP and gt values on PPR, QQR, thermal efficiency and COP values are shown in Figs. 9–11 (taking R245ca as an example). If a PPR baseline is designated, such as 20% in Fig. 9, the minimum turbine efficiency limitation increases with the increase of pinch temperature values in heat exchangers. As shown in Fig. 9, the minimum turbine efficiency limitation is 40% with DTP value of 3 K. However, when the DTP value increases to 5 K the minimum turbine efficiency limitation would reach 80%. The variation range of QQR values is 1.005–1.048 under various DTP (1–5 K) and gt (40–80%) values as shown in Fig. 10. The increasing DTP values lead to less and larger temperature difference between evaporating and condensing temperatures for the ORC and heat pump subsystem, respectively. As a result, thermal efficiency and COP values both decrease with the increase of DTP values as shown in Fig. 11. The variation of PPR, QQR, thermal efficiency and COP values under various geothermal source temperatures (80–100 °C), geothermal exhaust temperatures (10–30 °C), heating supply temperatures (30–50 °C) and heating return water temperatures (25– 40 °C) can be seen from Figs. 12–16. Lower geothermal source temperature and geothermal exhaust temperature would lead to less net power output and more power consumed. As a result, PPR values decrease as shown in Fig. 12. PPR values decrease from 59% to 2.8% under the considered geothermal source and exhaust temperature range. We can see from Fig. 12 that geothermal source parameters which make QQR values decreased from 1.14 to 1.01 playing more important roles on QQR values than DTP, gt and grp values. The evaporating and condensing temperature of the ORC subsystem is free of influence caused by geothermal source parameters. So the thermal efficiency values keep constant as shown in Fig. 13. However, the evaporating temperature of the heat pump subsystem is increasing with the increase of geothermal source exhaust temperature, making COP values for the heat pump subsystem increased. The heating return water temperature and heating supply temperature will influence condensing temperatures of ORC and heat pump subsystems and affect thermal efficiency and COP values as shown in Fig. 16 which is reflected by variation of PPR and QQR values consequently. The smaller heating return water temperature would lead to higher PPR values at constant heating supply temperature. Given the constant heating return water temperature, PPR values decrease with the increase of heating

T. Guo et al. / Energy Conversion and Management 52 (2011) 2384–2391

supply temperature as shown in Fig. 14. The tendency of QQR values are opposite to that of PPR values as shown in Fig. 15. 6. Conclusions An innovative cogeneration system powered by low-temperature geothermal sources is proposed and thermodynamically analyzed. The suitable working fluids have been screened and recommended for the ORC-based power generation subsystem. Performances of the novel cogeneration system under disturbance conditions have also been studied. The operation conditions are listed in Table 3. The considered criteria are high PPR and QQR values. The main results are as follows: 1. Fluids group presenting higher normal boiling point values show higher PPR values. It presents averagely 7.7% higher PPR values and R236ea and R245ca show a little higher PPR values among the group. 2. DTP and gt values play more important roles on the variation of PPR values. 3. QQR values change slightly with various DTP, gt and grp values while the variation range is larger under various geothermal source and heating supply parameters. 4. Fluid presenting a higher PPR value would show a lower QQR value. 5. Smaller DTP value, higher gt value, higher geothermal source parameters and lower heating supply parameters lead to higher PPR values.

Acknowledgement The authors gratefully acknowledge financial support provided by the National Natural Science Foundation of China (Grant No. 50976079). References [1] Zhu JL, Zhang W. Optimization design of plate heat exchangers (PHE) for geothermal district heating systems. Geothermics 2004;33:337–47. [2] Guo T, Wang HX, Zhang SJ. CN Patent ZL 200910228029.4 (in Chinese).

2391

[3] Hettiarachchi HDM, Golubovic M, Worek WM, Ikegami Y. Optimum design criteria for an Organic Rankine cycle using low-temperature geothermal heat sources. Energy 2007;32:1698–706. [4] Subbiah S, Natarajan R. Thermodynamic analysis of binary-fluid Rankine cycles for geothermal power plants. Energy Convers Manage 1988;28(1):47–52. [5] Saleh B, Koglbauer G, Wendland M, Fischer J. Working fluids for lowtemperature organic Rankine cycles. Energy 2007;32:1210–21. [6] Desideri U, Bidini G. Study of possible optimization criteria for geothermal power plants. Energy Convers Manage 1997;38(15–17):1681–91. [7] Iqbal KZ, Fish LW, Starling KE. Isobutane geothermal binary cycle sensitivity analysis. Proc Okla Acad Sci 1977;57:131–7. [8] Gu ZL, Sato H. Optimization of cyclic parameters of a supercritical cycle for geothermal power generation. Energy Convers Manage 2001;42:1409–16. [9] Gu ZL, Sato H. Performance of supercritical cycles for geothermal binary design. Energy Convers Manage 2002;43:961–71. [10] Guo T, Wang HX, Zhang SJ. Comparative analysis of natural and conventional working fluids for use in transcritical Rankine cycle using low-temperature geothermal source. Int J Energy. doi:10.1002/er.1710. [11] Heberle F, Brüggemann D. Exergy based fluid selection for a geothermal Organic Rankine Cycle for combined heat and power generation. Appl Therm Eng 2010;30:1326–32. [12] Wagar WR, Zamfirescu C, Dincer I. Thermodynamic performance assessment of an ammonia–water Rankine cycle for power and heat production. Energy Convers Manage 2010;51:2501–9. [13] Kontoleontos E, Karytsas C, Mendrinos D, Georgilakis PS. Optimized geothermal binary power cycles. J Optoelectron Adv Mater 2008;10(5): 1228–32. [14] Zhang XR, Yamaguchi H, Uneno D. Analysis of a novel solar energy powered Rankine cycle for combined power and heat generation using supercritical carbon dioxide. Renew Energy 2006;31:1839–54. [15] Zhang XR, Yamaguchi H, Fujima K. Study of solar energy powered transcritical cycle using supercritical carbon dioxide. Int J Energy Res 2006;30:1117–29. [16] Zhang XR, Yamaguchi H, Fujima K. Experimental performance analysis supercritical CO2 thermodynamic cycle powered by solar energy. AIP 2006;832:419–24. [17] Zhang XR, Yamaguchi H, Fujima K. Theoretical analysis of a thermodynamic cycle for power and heat generation using supercritical carbon dioxide. Energy 2007;32:591–9. [18] Zhang XR, Yamaguchi H, Uneno D. Thermodynamic analysis of the CO2-based Rankine cycle powered by solar energy. Int J Energy Res 2007;31:1414–24. [19] Zhang XR, Yamaguchi H, Uneno D. Experimental study on the performance of solar Rankine system using supercritical CO2. Renew Energy 2007;32: 2617–28. [20] Chen Y, Lundqvist P, Platell P. Theoretical research of carbon dioxide power cycle application in automobile industry to reduce vehicle’s fuel consumption. Appl Therm Eng 2005;25(14–15):2041–53. [21] Calm JM, Hourahan GC. Refrigerant data update. Heat/Pip/Air Cond Eng 2007;79(1):50–64. [22] Liu BT, Chien KH, Wang CC. Effect of working fluids on organic Rankine cycle for waste heat recovery. Energy 2004;29:1207–17. [23] REFPORP. Reference fluid thermodynamic and transport properties. NIST Standard Reference Database 23-Vision 7.0. Gaithersburg (MD): National Institute of Standards and Technology 2002.