Simulation of the parabolic trough solar energy generation system with Organic Rankine Cycle

Applied Energy 97 (2012) 630–641

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Simulation of the parabolic trough solar energy generation system with Organic Rankine Cycle Ya-Ling He ⇑, Dan-Hua Mei, Wen-Quan Tao, Wei-Wei Yang, Huai-Liang Liu Key Laboratory of Thermo-Fluid Science and Engineering of MOE, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an, Shaanxi 710049, China

a r t i c l e

i n f o

Article history: Received 3 October 2011 Received in revised form 15 February 2012 Accepted 19 February 2012 Available online 26 April 2012 Keywords: Solar thermal power generation Organic Rankine Cycle (ORC) Organic working fluid

a b s t r a c t A model for a typical parabolic trough solar thermal power generation system with Organic Rankine Cycle (PT-SEGS–ORC) was built within the transient energy simulation package TRNSYS, which is formed by integrating several submodels for the trough collector system, the single-tank thermal storage system, the auxiliary power system and the heat-electricity conversion system. With this model, the effects of several key parameters, including the interlayer pressure between the absorber tube and the glass tube (pinter), the flow rate of high temperature oil in the absorber tube (v), solar radiation intensity (Idn) and incidence angle (h), on the performance of the parabolic trough collector field based on the meteorological data of Xi’an city were examined. The study shows that the heat loss of the solar collector (qloss) increases sharply with the increase in pinter at beginning and then reaches to an approximately constant value. The variation of heat collecting efficiency (ghc) with v is quite similar to the variation of qloss with pinter. However, Idn and h exhibit opposite effect on ghc. In addition, it is found that the optimal volume of the thermal storage system is sensitively dependent on the solar radiation intensity. The optimal volumes are 100, 150, 50, and 0 m3 for spring equinox, summer solstice, autumnal equinox and winter solstice, respectively. Ó 2012 Elsevier Ltd. All rights reserved.

1. Introduction Among all the solar thermal power generation technologies, the parabolic trough solar thermal power generation systems (PTSEGSs) have attracted great attentions and achieved commercial applications. From 1985 to 1991, the US Luz company has built 9 PT-SEGS in the California Mojave Desert with the total installed capacity reaching about 354 MW and the generation efficiency reaching about 15% [1,2]. And, a number of new plants are currently in the planning process. To date, great efforts had been devoted for further advancing this technology for power generation [3–11]. Lippke [3] simulated a typical 30 MW PT-SEGS. The results indicated that the solar radiation intensity greatly influences the optimum temperatures of the steam and heat transfer oil (HTO). Thomas [4], Kalogirous et al. [5] and Zarza et al. [6] carried out investigations on the parabolic trough collector systems for steam generation. Their studies demonstrated that the calculation error was within 1.2% and only 48.6% of the solar radiation energy falling on the collector was utilized for steam generation, others was dissipated to the environment in different forms: collection losses (41.5%), thermal losses (6.9%), energy losses due to raising the water temperature from environment temperature to 100 °C (2.2%) and for the rig (0.5%). A direct ⇑ Corresponding author. Tel./fax: +86 29 82665930. E-mail address: [email protected] (Y.-L. He). 0306-2619/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2012.02.047

steam generation PT-SEGS bench was also built by Almanza and Lentz [7] using the recirculation process concept. In the bench, four modules were connected in series and the first three were adapted with copper pipe absorbers covered with black chrome while the last one used steel pipe covered with black chrome. The results indicated that the copper pipes could eliminate the bend due to thermal stress and the measured system efficiency of this system reached about 3%. In addition, the basic energy and exergy analysis on the PT-SEGS was also performed by Singh et al. to evaluate the respective losses as well as exergetic efficiency of the whole system [8]. The study indicated that within the system, most of the heat loss occurred in the condenser of the heat engine part while most of the exergy loss is located in the PT collector–receiver assembly. Besides, several approaches have been proposed for further improving the efficiency of PT-SEGS, among which the PT-SEGS with Organic Rankine Cycle (ORC) gets more and more attentions. A 150-kWe trough-ORC solar power plant was demonstrated in the Coolidge Solar Irrigation Project [9]. Although the plant operated successfully for several years, it suffered from many problems like low collector performance, high operation and maintenance costs and low annual output, hindering its wide applications. Since then, the performance of PT-SEGS–ORC has been being investigated by many researchers. Price and Hassani [10] built the system models of PT-SEGS–ORC with Aspen Plus, in which Pentane and a combination of mixed working fluids were used as the working fluid and three ORC cycles were analyzed including

Y.-L. He et al. / Applied Energy 97 (2012) 630–641

631

Nomenclature a A b c cp cp d DE f h

Dh HTO i, j I K l m _ m MEP OWF p Pe PT q,Q R Re S SEGS T T DT dT U v; V W Z

control function area, m2 control function or modified coefficient control function constant pressure specific heat capacity, kJ kg1 K1 average constant pressure specific heat capacity, kJ kg1 K1 diameter, m thermodynamic energy change, W frictional resistance coefficient enthalpy, kJ kg1, or convection heat transfer coefficient, W m2 K1, pressure loss coefficient enthalpy rise, kJ kg1 heat transfer oil serial number solar radiation intensity, W m2 solar incidence angle modifier length of finite unit, m mass, kg mass flow rate, kg s1 gas mean molecular free path organic working fluid condensation pressure, Pa electric energy production, kJ parabolic trough thermal energy, W thermal resistance, m KW1 Reynolds number numbers solar energy generation system temperature, °C average temperature, °C temperature rise, °C temperature difference, °C overall heat transfer coefficient, W m2 K1 volume, m3; HTO flow rate, ms1 work, kJ heat recovery series

Greek symbols a absorptivity or extraction coefficient D surface roughness of the absorber tube, m r Stefan–Boltzmann constant; molecular diameter, m e emissivity or performance coefficient g efficiency h incidence angle of beam radiation

a simple Rankine cycle, a Rankine cycle with recuperation and a simple Rankine cycle with reheat and recuperation. The results indicated that the efficiency were 12.5%, 20.1%, 20.4%, 20.6% and 20.7%, respectively, for basic ORC using pure working fluid, ORC with recuperation using pure working fluid, ORC with using mixed working fluid, ORC with recuperation and reheat using pure working fluid and ORC with recuperation and reheat using mixed working fluid. Prabhu [11] conducted simulation investigations on PTSEGS–ORC, and compared the difference between the Organic Rankine Cycle and steam Rankine cycle. The results indicated that the steam cycle outperformed the ORC by 15–25% at the average summer high temperature conditions even though the ORC systems were carefully designed and the working fluids were carefully selected. Pei et al. [12] developed a PT-SEGS–ORC with heat regeneration, in which the CPC collectors were made up of compound parabolic concentrator and the organic working fluid (OWF) was

k

q s U

conductivity, W m1 K1 reflectivity transmissivity, or enthalpy, kJ kg1 energy, kJ

Subscripts a absorber tube aux auxiliary energy subsystem c collecting con condenser dn direct solar radiation en environment ev evaporation ext extraction f heat transfer oil (HTO) fe feed g glass tube ge generator h high temperature hc heat collecting he heat exchange i,i inner, serial number in inlet inter interlayer j serial number lr,LR last stage surface-type regenerator loss heat loss L low temperature or liquid o outer or OHT op optical out outlet p pump pr pooled regenerative heater pt parabolic trough collector r radiation s single tank sat saturation set control set point sh superheating sky outer space sys system t total turb steam turbine V vapor w, w water, work

HCFC-123. The study indicated that the regenerative cycle has negative effects on the collector efficiency due to increment of the average working temperature of the first-stage collectors while positive ones on the ORC efficiency and the system electricity efficiency with regenerative ORC is about 8.6% when the solar radiation is 750 W m2. Our literature review indicated that most of the models for simulating trough collector field in the PT-SEGS are empirical models; the accuracy of those models is limited. Also, the systems considered in those models are relatively simple. In this work, a detailed procedure for modeling the PT-SEGS–ORC through the energy simulation package TRNSYS [13] is presented, in which the onedimensional model [14] is adopted to simulate the trough collector field to improve the modeling accuracy. The effect of heat recovery system on the system efficiency is also considered in the model by introducing the last stage surface-type regenerator. With the

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model, the influences of several design and operation parameters on the performance of both the collector field and the whole system are examined. 2. System description and working principle A schematic diagram of the parabolic trough solar thermal power generation systems is shown in Fig. 1, which consists of the trough collector field, thermal storage system, auxiliary energy system, heat exchange system, heat recovery system, cooling system and power system, etc. The trough collector field includes several lines of the trough collectors, and each collector comprises of a receiver and a concentrator. With respect to the receiver, it is made of a stainless steel tube covered by a glass envelope for reducing heat loss due to radiation and convection, which is placed along the focal line of the concentrator. As the heat transfer oil flows through the receiver, it is heated up to a higher temperature. During the operation, a single thermal storage tank is employed to store the excessive thermal energy when the solar intensity is higher than that needed in the heat exchange system, whereas the auxiliary energy system will operate when the solar radiation intensity is inadequate to drive the subsequent procedures and/or the stored thermal energy is exhausted. As for the heat exchanger, it consists of a preheater, an evaporator and an overheater, through which the organic working fluid (OWF) is turned into steam by the high temperature HTO from the trough collector, thermal storage system or the auxiliary energy system. The OWF steam is then used as the working fluid for driving the steam turbine. 3. Mathematical model 3.1. Model formulations for each subsystem In this section, formulations for mathematical model of each subsystem are given. Also, the validation of the model of trough collector field with the available experimental results is carried on to make sure the correctness and reliability of the simulation of PT-SEGS–ORC. For the other subsystems, as they are more advanced ones, no relative test data are available in the open literatures, hence their validations are not conducted for the time being. 3.1.1. Trough collector field 3.1.1.1. Model formulations. Fig. 2 illustrates the sketch of the crosssection of the collector tube in the trough collectors. During the

Fig. 2. Cross-section view of the collector tube.

operation, the heat reflected by the parabolic trough is mainly absorbed by the HTO in the absorber tube and the rest dissipates to the ambient air. The amount of the solar radiation concentrated on the absorber tube (Qa) is decided by the geometric and optical parameters of the parabolic trough collector, which represents the superposition of the thermal energy absorbed by HTO (Qf) and the heat loss of the collector (Qloss). Mathematically, it can be written as:

Q a ¼ gop  K  Idn  Apt

ð1Þ

Q a ¼ Q f þ Q loss

ð2Þ

where Apt is the orifice area of the parabolic trough collector and Idn is the direct solar radiation intensity. The terms gop and K are the optical efficiency of collector and solar incidence angle modifier, which are, respectively, given by [3]:

gop ¼ qpt sg aa

ð3Þ

K ¼ cosðhÞ  0:003512h  0:00003137h2

ð4Þ

with qpt, sg, aa and h, respectively, denoting the reflectivity of parabolic trough, the transmissivity of the glass tube, the absorptivity of the absorber tube and the incidence angle of beam radiation. A one-dimensional steady-state model [14] is employed herein to simulate the trough collector. In the model, the trough collector is divided into a certain number of finite units along the flowing

Fig. 1. Schematic diagram of parabolic trough SEGS system with Organic Rankine Cycle.

Y.-L. He et al. / Applied Energy 97 (2012) 630–641

direction of the working fluid. To indicate the heat transport process in the trough collector, the thermal resistance network for each finite unit is presented in Fig. 3. Clearly, the heat energy absorbed by HTO (Qf) is determined by the thermal resistance (R1) due to the convective heat transfer between HTO and absorber tube surface and the thermal resistance (R2) due to the heat conduction through the tube walls. Mathematically, it can be given by:

Qf ¼

plðT a;o  T f Þ

ð5Þ

R1 þ R2

where l is the length of finite unit. Besides, R3, R4 represent the thermal resistance due convective, radiation heat transfer between absorber tube outer wall and glass tube inner wall. R5 stands for the thermal resistance due to heat conduction between the glass tube inner and outer wall. And R6, R7 means the thermal resistance due to convective, radiation heat transfer between glass tube outer wall and the out environment. With respect to the heat loss (Qloss), it is actually equal to the heat transfer rate between the outer wall of the absorber tube and the inner wall of the glass tube (Qa-g), and the heat transfer rate between the outer wall of the glass tube and the environment (Qasky). Mathematically, it gives:

Q loss ¼ Q ag ¼ Q gsky

ð6Þ

In Eq. (6), Qa-g is the summation of the convective heat transfer rate [15] and the radiative heat transfer rate between the glass tube and the absorber tube, which can be given by:

Q ag ¼ Q ag;c þ Q ag;r Q ag;c ¼

ð7Þ

2phag;c ½T a;o  T g;i  lnðdg;i =da;o Þ

ð8Þ

Q ag;r ¼ 2pea rlda;o ½T 4a;o  T 4g;i 

ð9Þ

where dg,i is the inner diameter of the glass tube, da,o is the outer diameter of the absorber tube, ea is the emissivity of the absorber tube outer wall, and ha-g,c is the convection heat transfer coefficient in the interlayer, given by Eq. (10) [15]:

hag;c ¼

kag da;o 2

d

g;i ln da;o þ b  MEP



dg;i da;o

 þ1

ð10Þ

with ka-g representing gas conductivity in the interlayer. b and MEP are the modified coefficient and gas mean molecular free path, which are, respectively given by:

b¼

9c  5 2ðc þ 1Þ

MFP ¼

ð11Þ

1:748  1020 T pr2

ð12Þ

where c, r, p, T are the specific heat capacity ratio, the molecular diameter, the pressure, and the average temperature of the gas in the interlayer given by T ¼ ðT a;o þ T g;i Þ=2 . As for the heat transfer rate between the outer wall of the glass tube and the environment (Qa-sky), it is the summation of convection and radiation heat transfer rate between the glass tube and the environment, which is given by: R3

R6 Ten

Tf

R1

Ta,i

R2

R5 Ta,o

Tg,i R4

633

Q gsky ¼ Q gsky;c þ Q gsky;r

ð13Þ

Q gsky;c ¼ 2pldg;o hg;c ½T g;o  T en 

ð14Þ

h i Q gsky;r ¼ 2pldg;o eg r T 4g;o  T 4sky

ð15Þ

where dg,o is the outer diameter of the glass tube, hg,c is the convection heat transfer coefficient outside the glass tube, eg is the emissivity of the glass tube outer wall, r is the Stefan–Boltzmann constant and Tsky is the effective sky temperature. With respect to the pressure drop (i.e., the frictional head loss) of HTO in the absorber tube, it is related to the HTO flow rate (v), the inner diameter (da,i) and the surface roughness of the absorber tube (D) [16]. The equivalent pressure loss coefficient can be expressed by:

h¼f

l m2 da;i 2g

ð16Þ

where f is the frictional resistance coefficient, given by:

pffiffiffi pffiffiffi 1= f ¼ 2:0 logðD=da;i Þ=3:7 þ 2:51=ðRef f ÞÞ

ð17Þ

with Ref denoting the Reynolds number of HTO in the tube. Based on above equations, the general heat transfer and pressure loss in the trough collector can be calculated according to the processes presented in Fig. 4. 3.1.1.2. Model Validation. In order to validate the theoretical model of the one-dimensional steady-state model for the trough collector, simulations is carried out in LS-2 collector under the operation condition shown in Table 1. The simulation results are compared with the experimental results presented in Ref. [17], as indicated in Fig. 5. It can be seen that the simulation collector efficiencies are accordant with the experimental values [17]. The good agreement shows that the theoretical model for trough collector in the present paper is correct and reliable. 3.1.2. Single-tank thermal storage subsystem As mentioned earlier, a single tank is used to storage the heat energy in the system, and the heat transfer oil (HTO) is adopted as the heat storage medium. Simply, the thermocline model [18], which is an existing model in TRNSYS [13], is employed to simulate the heat transfer process inside the thermal storage system. In the model, the tank is divided into n units with equal volume from the _ f;h are, respectop to the bottom as illustrated in Fig. 6. Tf,h and m tively, the temperature and mass flow rate of HTO from collector _ f;L are those field to the thermal storage system, while Tf,L and m from heat exchange system to the thermal storage system, respectively. To further simplify the model, assumptions are also made as listed below: (1) Uniform temperature distribution of HTO in each unit. (2) Ignoring the heat conduction between adjacent units. (3) Neglecting the flow rate of HTO due to the large volume of the tank. Generally, the conservation of heat energy in the ith unit can be expressed as:

mf;i cpf;i

dT f;i _ f;h cpf;i ðT f;h  T f;i Þ þ bi m _ f;L cpf;i ðT f;L  T f;i Þ ¼ ai m dt  ci ðT i1  T i Þcpf;i ci > 0 þ U sen As;i ðT f;i  T en Þ þ ci ðT i  T iþ1 Þcpf;i ci < 0 ð18Þ

Tg,o R7 Tsky

Fig. 3. Thermal resistance network of trough collector.

where Tf,i, mf,i and cpf,i are the temperature, the mass and the constant pressure specific heat of HTO in the ith unit, respectively. As,i is the surface area of the ith element, Ten is the environment

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Y.-L. He et al. / Applied Energy 97 (2012) 630–641

Fig. 4. The calculation process for the trough collector.

Table 1 Working parameters for trough collector validation. Parameter

Value

Unit

Parameter

Value

Unit

Direct solar radiation intensity Optical efficiency of collector Environment temperature Environment wind speed

940 0.731 22 0

W m2 – °C m s1

Incidence angle HTO flow rate in collector field Effective sky temperature HTO temperature

11.6 0.5 14 100–350

° m s1 °C °C

73 Experimental Results Simulation Results

72 71

ηhc

70 69 68 67 66

100

150

200

250

300

350

Tf ( ) Fig. 5. Validation for model of the trough collector.

temperature, and Us-en is the overall heat transfer coefficient between the storage tank and the environment. The constants ai, bi and ci are the control functions, which are, respectively, given by:

ai ¼



1; i ¼ Sh 0; i–Sh

ð19Þ

Fig. 6. Diagram of the single-tank thermocline thermal storage subsystem.

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Y.-L. He et al. / Applied Energy 97 (2012) 630–641

bi ¼



1; i ¼ SL

j1 X

_ f;h ci ¼ m

T/

ð20Þ

0; i–SL _ f;L aj  m

n X

bj

T2

with Sh and SL denoting the numbers of control volume in the thermal storage system, into which the HTO flows from collector field and the heat exchange system, respectively. Totally, the heat loss in the thermal storage tank (Qloss,s), the heat transferred by HTO from the trough collector field to the thermal storage tank (Qpt-s), the heat transferred by HTO from the thermal storage tank to the heat exchange system (Qs-he), and the thermodynamic energy change in the tank (DE) can be, respectively, determined by:

Q loss;s ¼

n X

ΔT

ð21Þ

j¼iþ1

j¼1

T4

Temperature change process for HTO T3

U sen As;i ðT f;i  T en Þ

Tsh

T1 Tev

Temperature change process for OWF

Tfe preheating process

1

evaporating overheating process process

2

3

4

Fig. 7. Illustration of heat transfer process.

ð22Þ

i¼1

Pn

i¼1 c pf;i

_ f;h Q pts ¼ m

n

ðT f;h  T f;n Þ

_ o;he cpo ;he ðT ev  T gs Þ ¼ ghe m _ f;he cpf;he ðT 2  T 1 Þ Q 12 ¼ m

ð28Þ

_ o;he hlatent ¼ ghe m _ f;he cpf;he ðT 3  T 2 Þ Q 23 ¼ m

ð29Þ

ð24Þ

_ o;he cpo;he ; m _ f;he cpf;he ÞðT 4  T ev Þ Q 34 ¼ ehe minðm

ð30Þ

ð25Þ

_ f;he and cpf,he are the mass flow rate and the specific heat where m _ O;he and cpo,he are those of the OWF. hlacapacity of the HTO, while m tent is the latent heat of OWF. ghe and ehe are the heat exchange efficiency and performance coefficient of the heat exchange system.

ð23Þ

Pn

i¼1 c pf;i

_ f;L Q she ¼ m

DE ¼

n X

n

ðT f;1  T f;L Þ  Pn

Pn mf;i 

i¼1

i¼1 c pf;i

n



i¼1 T i



Pn

i¼1 T i jt¼0



n

3.1.3. Auxiliary energy subsystem To maintain the continuous and stable operation of the whole power system, the auxiliary energy system has to be operated when the solar radiation is inadequate and the heat energy in the thermal storage system is exhausted. In the auxiliary energy system, Tset is the auxiliary energy control temperature. By comparing the temperature of HTO that flows into the auxiliary energy subsystem (Tf,aux) and Tset, it can be decided whether the auxiliary energy subsystem should be on or not. When Tf,aux < Tset, the auxiliary energy subsystem is on and the heat provided by the auxiliary energy system (Q aux ) and heat loss in the auxiliary energy system (Q loss;aux ) are, respectively, determined by:

(

_ f;aux cpf;aux ðT set T f;aux ÞþU auxen Aaux ðT f;aux T en Þ m

gaux

Q aux ¼

0 (

Q loss;aux ¼

;

T f;aux < T set

ð26Þ

; T f;aux  T set

U auxen Aaux ðT f;aux  T en Þ þ ð1  gaux ÞQ_ aux ;

T f;aux < T set

0

; T f;aux  T set

3.1.4.2. Heat recovery system. 3.1.4.2.1. The last stage surface-type regenerator. The flow and heat transfer process in the last stage surface-type regenerator is illustrated in Fig. 8. The temperature of the overheated OWF at the outlet of the steam turbine is Tturb,out, and it decreases to To,lr after the heat releasing process in the regenerator, then the cooled OWF condenses to liquid in the condenser. The temperature of supercooled liquid OWF at the outlet of the condenser is Tcon,out, and it increases to To,LR after absorbing the heat in the regenerator. Simply, the heat transfer rate in the last stage surface-type regenerator (Qlr) can be calculated by:

_ o;lr ðcpV;lr ; cpL:lr Þmin ðT turb;out  T con;out Þ Q lr ¼ elr m

_ o;lr is the OWF mass flow rate, elr is the regenerator perforwhere m mance coefficient. cpV,lr and cpL,lr are the specific heat capacities of the vapor and liquid OWF, respectively. And, To,LR and To,lr are, respectively, given by:

T o;LR ¼

Q lr þ T con;out _ o;lr cpL;lr m

ð27Þ _ f;aux , T f;aux and cpf,aux are the mass flow rate, the average where m temperature and the constant pressure specific heat of HTO in the auxiliary energy system, respectively. Uaux-en is the overall heat transfer coefficient between the auxiliary energy subsystem and the environment. Aaux and gaux are the surface area and the efficiency of the auxiliary energy subsystem, respectively. 3.1.4. Heat-electricity conversion subsystem 3.1.4.1. Heat exchange system. Fig. 7 illustrates the heat transport processes in the heat exchange system. In the figure, Process 1–2, process 2–3 and process 3–4 are, respectively, the preheating process, evaporating process and overheating process. And, T4–T3–T2– T1 and Tfe–Tev–Tsh are the temperature change process for HTO and OWF, respectively. The thermal energy absorbed by the OWF in these three processes can be determined by:

ð31Þ

T o;lr ¼ T turb;out 

ð32Þ

Q lr

ð33Þ

glr m_ o cpV;lr

where glr is the regenerator efficiency. Tcon,out

To,lr

To,LR

Tturb,out

Fig. 8. Fluid flow in the last stage surface-type heat regenerator.

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Y.-L. He et al. / Applied Energy 97 (2012) 630–641

3.1.4.2.2. Pooled regenerative heater. The fluid flow in the pooled regenerative heater is shown in Fig. 9. In the figure, hj is the extraction enthalpy in the jth steam turbine, aj is the jth extraction coefficient and sj is the jth feeding fluid enthalpy at the outlet of the regenerative heater. The enthalpy conservation in the pooled regenerative heater can be written as:

"

z X

gpr aj hj þ 1 

!

#

ai sj1 ¼ 1 

i¼j

z X

!

3.2. Performance analysis

ai sj

ð34Þ

i¼j1

where gpr is the efficiency of the heater. 3.1.4.3. Power system. The power system contains a steam turbine with pentane being the OWF and a generator. In the system, the work done by the ORC (Wturb), the efficiency of the steam turbine (gturb) and the electric energy production (Pege) are, respectively, given by:

" _ o;turb W turb ¼ m

z X 



ai hturb;in  hi þ 1 

i¼1

z X

!

ai



hturb;in  hturb;out



#

i¼1

ð35Þ W

turb gturb ¼ _ mo;turb ðhturb;out  ho;fe Þ

Pege ¼

ð36Þ

gge W turb

ð37Þ

3600

where hturb,in and hturb,out are, respectively, the specific enthalpies of OWF at the inlet and outlet of the steam turbine, ho,fe is the specific enthalpy of the feeding OWF and gge is the efficiency of the generator. 3.1.4.4. Cooling system. The overheated OWF from the last stage surface-type regenerator condenses into the saturated fluid in the condenser. The saturation temperature (Tsat) corresponding to a certain pressure in the condenser (pcon) is decided by the condenser temperature difference. Simply, the saturation temperature (Tsat), the temperature rise of cooling water (DTw,con) and the condenser temperature difference (dTcon), are, respectively, given by:

T sat ¼ T w;con þ DT w;con þ dT con

ð38Þ

_ o;con qo;con m _ w;con cpw m

ð39Þ

DT w;con ¼

 hcon Acon dT con ¼ ðT sat  T w;con Þ exp  _ w;con cpw m

αj

1−

i =j -1

3.2.1. Simulation methods In this work, the energy simulation package TRNSYS [13] is adopted. TRNSYS is a transient systems simulation program package with modular structures, in which the user can specify the components that constitute the unevaluated system and the manner in which they are connected. As the modular package of the single-tank thermal storage subsystem already exists in TRNSYS, the modular packages of the submodels for the remaining subsystem presented above are established and added to the DLL files in the TRNSYS database and the PT-SEGS–ORC is then constructed by calling the DLL files. With the selected basic parameters, the performance of the trough collector field and the whole system is evaluated based on the meteorological data of city Xi’an on the specific day of spring equinox, summer solstice, autumnal equinox and winter solstice, respectively. 3.2.2. Basic parameters 3.2.2.1. Selection of organic fluids and cycle parameters. In this work, R113, R123 and pentane are used as the OWF. Table 2 shows their thermodynamic parameters. Due to the high critical temperature and low ODP and GWP, three OWF are all suitable for high temperature ORC. Fig. 10 shows the ORC efficiencies (go) and the unit mass OWF output powers (wo) for three different working fluids (i.e., R113, R123 and Pentane) at different evaporation pressures (pev). As indicated in the figure, for the same turbine inlet temperature (Tturb,i), wo of pentane is far greater than those of R112 and R123. Also, it is seen that when Tturb,i is less than 185 °C, go of R123 is the highest, which is followed by that of pentane and R113. However, it is shown that when Tturb,i is greater than 185 °C, go and wo of pentane are both the greatest. As a result, pentane is selected as the organic working fluid in the following studies. And, the rated operating conditions of pev and Tturb,i are, respectively, set to be 2.0 MPa and 190 °C. 3.2.2.2. Optimal feeding fluid temperature for ORC system. The feeding fluid temperature (Tfe) of the heat recovery system depends on the first stage heat recovery extraction pressure (pext). Fig. 11 shows the effect of the first stage heat recovery pressure (pext) on go under different heat recovery series (Z). Clearly, with the increase in Z, go increases gradually. For each Z, there exists an optimal pext. As indicated in the figure, the optimal first stage heat recovery pressures are, respectively, 0.9, 1.2, 1.4 and 1.6 MPa when Z is 1, 2, 3 and 4.

hj 3.2.2.3. Suitable number of heat recovery series. Fig. 12 shows the effect of Z on go. It is seen that go increases sharply with Z at the beginning and then reaches to an approximately constant value. However, it is noticed that too high Z results in the large cost of the generation system and serious security risks. In this work, three stages heat recovery system is proposed.

z

z

∑α

ð40Þ

τ j -1

τj

_ o;con and m _ w;con are, respectively, the mass flow rates of where m OWF into the condenser and the cooling water, qo,con is the condensation heat release of the OWF in the condenser, cpw is the specific heat capacity of cooling water, hcon is the condensation heat transfer coefficient and Acon is the heat exchange area of condenser.

i

1 − ∑αi i =j

Fig. 9. Fluid flow in the pooled regenerative heater.

3.2.2.4. Site description and solar radiation parameter. In China, abundant solar energy resources are available, and its territory can be divided into five regions according to amount of the accepted global solar radiation [19]. Among the first-three regions, the annual sunshine duration is more than 2000 h and the annual

637

Y.-L. He et al. / Applied Energy 97 (2012) 630–641 Table 2 Thermodynamic performance parameters of organic working fluid. Working fluid

Mole Mass (g mol1)

Critical pressure (MPa)

Critical temperature (°C)

ODP

GWP

R113 R123 Pentane

187.38 152.93 72.15

3.39 3.66 3.37

214.05 183.67 196.54

0.9 0.02 0

1.55 29 11

120 25.6 100

o

21.5 80

24.8

60

140

150

160

170

180

190

200

Z=1 Z=2 Z=3 Z=4

24.4

20.5 20.0 130

25.2

-1

R113 Output Power R123 Output Power Pentane Output Power

21.0

wo /kJ·kg

/%

22.0

/%

22.5

26.0

R113 Efficiency R123 Efficiency Pentane Efficiency

o

23.0

40 210

24.0

0.4

0.6

0.8

Tturb,i /

1.0

1.2

1.4

1.6

1.8

2.0

pext /MPa

(a) pev = 1.5 MPa

Fig. 11. Effect of first stage heat recovery steam extraction pressure ORC system efficiency.

24.0 120 100

o

80 23.0 60

26.0

wo /kJ·kg

/%

23.5

25.5

22.0 150

/%

-1

40 R113 Efficiency R123 Efficiency Pentane Efficiency

160

170

180

R113 Output Power R123 Output Power Pentane Output Power 20

190

200

25.0

o

22.5

24.5

210

24.0

Tturb,i /

(b) pev = 2.0 MPa

23.5

0

1

120

3

4

Fig. 12. Effect of heat recovery series on ORC system efficiency.

100

o

R113 Efficiency R123 Efficiency Pentane Efficiency

80

-1

23.5

wo /kJ·kg

/%

24.0

60

23.0 160

2

Z

24.5

R113 Output Power R123 Output Power Pentane Output Power

170

180

190

200

40 210

Tturb,i /

(c) pev = 2.5 MPa Fig. 10. Effect of turbine inlet temperature on ORC system efficiency and unit mass output power.

where annual global solar radiation is about 5000–5850 MJ/m2, approximately equivalent to the amount of daily radiation at 3.8–4.5 kW h/m2. The direct solar radiation intensity on the specific day of spring equinox, summer solstice, autumnal equinox and winter solstice, shown in Fig. 13, is used as the basic input parameters for system simulation. 3.2.3. System performance In this section, the effects of the designing parameters on the performance of PT-SEGS–ORC are examined. Generally, the power generation efficiency and the power output are employed to indicate the performance of the PT-SEGS–ORC. The total system power generation efficiency (gsys) can be given by:

3600 

global solar radiation amount is larger than 5000 MJ/m2, which makes them very favorable regions for solar energy utilization. In the present work, Xi’an (latitude: 34.267°N, longitude: 108.9°E) of Shaanxi Province is selected as the site where the PT-SEGS– ORC is built in the simulation work. It is in the above third region,

gsys ¼ P24

P24

i¼1 Q c ðiÞ þ

i¼1 Pege ðiÞ P24 i¼1 Q aux ðiÞ

ð41Þ

where Qc(i), Qaux(i) and Pege(i) are the thermal energy collected by the field, the energy consumption by the auxiliary systems and the system output power, respectively.

638

Y.-L. He et al. / Applied Energy 97 (2012) 630–641

3000

120 Spring equinox Summer solstice Autumnal equinox Winter solstice

100

Tf=100 Tf=150 Tf=200

Idn/kJ·h

qloss,t /W·m

-1

-1

2000

1000

80

Tf=250

60 40 20

0

0

4

8

12

16

20

0 0.01

24

0.1

1

10

100

1000

10000

1000

10000

pinter /Pa

Time/h

(a) qloss,t

Fig. 13. Direct solar radiation on four typical days.

70

4. Results and discussion

60

The performances of the parabolic trough collector field and the whole system are evaluated under the basic operation conditions listed in Table 3. Details are as follows.

qloss,c /W·m

-1

50

4.1. Performance of solar collector field

40

Tf=100 Tf=150 Tf=200 Tf=250

30 20

In this section, the effects of the interlayer pressure between absorber tube and glass tube (pinter), HTO flow rate in the absorber tube (v), solar radiation intensity (Idn) and the incidence angle (h) on the performance of the collector field are examined. 4.1.1. Interlayer pressure between absorber tube and glass tube (pinter) As radiation heat loss (qloss,r) is the same at the same Tf, the difference of total heat loss (qloss,t) is thus mainly decided by the convection heat loss (qloss,c). Fig. 14 presents the effect of pinter on the unit length heat loss of the solar collector (qloss,t) at different HTO temperatures (Tf). As indicated in the figure, qloss,t and qloss,c exhibit similar variation trends with pinter. With the increase in Tf, both qloss,t and qloss,c increase. When pinter is smaller than 10 Pa, the variations of qloss,t and qloss,c are sharp. When pinter is higher than 10 Pa, however, both qloss,t and qloss,c are almost independent on pinter. This may be due to the fact that when the interlayer is highly rarefied the major thermal resistance of the overall heat transfer of the glass tube to the environment is in the interlayer conduction, and qloss,c is mainly dependent on the random collision between gas molecules in the interlayer. When pinter is smaller than 10 Pa, increase of pressure in the interlayer will lead to rapid decrease of the molecular mean free path and intense collisions between the molecules, resulting in the rapid increase of qloss,c; when pinter is higher than 10 Pa, the thermal resistance of the interlayer become trivial, hence the effect of further increase in pinter becomes much more mild.

10 0 0.01

0.1

1

10

100

pinter /Pa

(b) qloss,c Fig. 14. Effect of interlayer pressure on unit length heat loss.

4.1.2. HTO flow rate in the absorber tube (v) The effect of Tf on the heat collecting efficiency (ghc) at different HTO flow rates (v) is presented in Fig. 15. Clearly, as shown in the figure, ghc decreases sharply with the increase in Tf at a given v. This is because with the increase in the oil temperature Tf the radiation heat transfer between the outer surface of the absorber tube and the inner surface of the glass tube deceases, leading to a reduction of the absorbed solar energy. The variation of ghc with v at a constant Tf of 250 °C is also shown in Fig. 16. It is seen that with the increase in v, ghc increases gradually. Interestingly, it is found that further increasing v exhibits little effect on ghc when v reaches a certain value. This is due to the reason that when v is low, the thermal resistance of the oil convective heat transfer is an important part of the overall thermal resistance from glass inner surface to oil. Increasing v can significantly enhance convection heat transfer inside the absorber tube, hence increases the absorbed solar

Table 3 Basic working parameters. Parameter

Value

Unit

Parameter

Value

Unit

Length of trough collector Focal length of parabolic trough Width of trough collector Initial temperature of HTF in the heat exchange system Efficiency of the heat exchange system Initial temperature at the entrance of steam turbine Initial pressure at the entrance of steam turbine Line of trough collector Number of collector each line Outer diameter of absorber

47.1 1.49 5.0 200 95 190 2.0 5 3 0.070

m m m °C % °C MPa – – m

Inner diameter of absorber Outer diameter of glass tube Inner diameter of glass tube Relative internal efficiency of steam turbine Efficiency of the last stage surface-type regenerator Performance of the last stage surface-type regenerator Mass flow rate of cooling water Temperature of cooling water Water pump efficiency Regenerative heater efficiency

0.065 0.105 0.095 70 95 0.6 1.0  104 20 67 99

m m m % % – kg h1 °C % %

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Y.-L. He et al. / Applied Energy 97 (2012) 630–641

absorber tube outer wall, thus more heat loss will be caused. However, heat absorbed by the HTO is much higher than heat loss under higher Idn, therefore, resulting in the increase of ghc with Idn. Besides, the effect of h on ghc at different temperatures of HTO (Tf) is also presented in Fig. 18. As can be seen from the figure, ghc decreases sharply with the increase of h at a given Tf. This highlights the importance of improving the tracking accuracy of the collector such that the efficiency of the collector can be boosted.

73

71

hc

/%

72

70

v =0.5m/s v =1.0m/s v =2.0m/s

4.2. Performance of the whole system

69 68 100

150

200

250

300

Tf / Fig. 15. Heat collecting efficiency on different HTO flow rate.

71.2

hc

/%

71.0

70.8

70.6

70.4 0.5

1.0

1.5

2.0

2.5

-1

v/m·s

Fig. 16. Effect of HTO flow rate on heat collecting efficiency.

energy and consequently reduces the heat loss. However, when the v reaches a certain value, oil convective thermal resistance becomes not important, hence further increasing v exhibits little effect on the efficiency. 4.1.3. Solar radiation intensity (Idn) and incidence angle (h) Fig. 17 shows the effect of different temperatures of HTO (Tf) on the heat collecting efficiency (ghc) at different solar radiation intensities (Idn). Clearly, as indicated in the figure, ghc significantly decreases with Tf at a given Idn, and its variation is much severer at lower Idn. Also, ghc gradually increases with Idn at a given Tf and the increasing trend becomes less under higher Tf. This is mainly due to the fact that higher Idn leads to higher temperature on the

In this part, the influences of the mass flow rate of HTO in the _ o ), thermal storage capacity and cooling water collector field (m parameters on the performance of the whole system are explored. _ o) 4.2.1. Mass flow rate of HTO in the collector field (m _ o on the collected thermal enFig. 19 shows the influence of m ergy (Qc), the work transferred from the thermal energy (WQ) and the pump work due to pressure loss (Wp) in the collector field at different seasons. As shown in the figure, Qc sharply increases _ o at beginning. Then, the increase of Qc with m _ o becomes with m _ o is weak. In addition, it is seen that the variation of WQ with m quite similar to that of Qc. However, the variation of Wp is quite dif_ o and the variation of Wp is much ferent, it greatly increases with m _ o. severer at higher m Also, it is found that the difference between WQ and Wp, i.e., _ o of 2.0  104, maximal useful work, can be achieved at m 4 4 4 2.0  10 , 1.4  10 and 1.4  10 kg/h for spring equinox, summer solstice, autumnal equinox and winter solstice, respectively. The difference in the optimal mass flow rate of HTO is due to the fact that the different solar radiation intensities bring about different optimum mass flows on these typical days, and the larger the solar radiation intensity is, the more significant effect of mass flow rate on collector efficiency will be. 4.2.2. Thermal storage capacity The performance of the auxiliary energy system with different thermal storage capacities is analyzed. Four thermal storage systems with different volumes (V) of 0 m3, 50 m3, 100 m3 and 150 m3 are studied. Fig. 20 shows the time-dependent consumptions of the auxiliary energy (Uaux) at different V for spring equinox, summer solstice, autumnal equinox and winter solstice. Clearly, it can be seen from Fig. 20a and c that Uaux is closely dependent on V at the time slot between 16:00 to 20:00 at spring equinox and autumnal equinox. In that period, Uaux is greatly decreased when V is increased from 0 m3 to 50 m3, whereas it changes little with further increasing V from 50 m3 to 150 m3. As a result, a volume of 50 m3 is recommended for the thermal

72

80 70 60

/%

70

69

hc

hc

/%

71

2

67 100

Tf=100

40

2

Idn=800W/m

68

50

Idn=400W/m

Tf=200

2

Idn=1200W/m

30

150

200

250

20

Tf=300

0

10

20

30

40

50

60

Tf / Fig. 17. Heat collecting efficiency at different solar radiation intensity.

Fig. 18. Effect of incidence angle on heat collecting efficiency.

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Y.-L. He et al. / Applied Energy 97 (2012) 630–641 5

1.2x10

Qu

Wp

5

-1

6.0x10

Q /kJ·h

4

5

6

4

4

4

2

1.2x10 8.0x10

W /kW

4

8

WQ

W /kW

4

8.0x10

Wp

1.6x10

6

WQ

5

1.0x10 -1

10

Qu

5

Q /kJ·h

5

2.0x10

8

1.4x10

4

4.0x10

2

4.0x10

4

2.0x10

0 0.0 4 4 4 4 4 4 1.6x10 1.8x10 2.0x10 2.2x10 2.4x10 2.6x10

0 0.0 4 4 4 4 4 4 1.6x10 1.8x10 2.0x10 2.2x10 2.4x10 2.6x10

-1

-1

mo /kg·h

mo /kg·h

(a) Spring equinox

(b) Summer solstice

4

4

3

5.0x10

4

Qc 4

2.5x10

Wp

4

2.0x10

1

2 W /kW

3.0x10

WQ

2.0x10

-1

2

4

Wp

4

W /kW

Q /kJ·h

-1

WQ

Q /kJ·h

4.0x10

3

3.0x10

Qc

4

1.5x10

4

1

1.0x10

4

1.0x10

3

5.0x10

0.0 4 1.2x10

4

1.4x10

4

4

1.6x10

1.8x10

0 4 2.0x10

0.0 4 1.2x10

4

4

1.4x10

-1

0 4 2.0x10

4

1.6x10

1.8x10

mo /kg·h

mo /kg·h-1

(c) Autumnal equinox

(d) Winter solstice

Fig. 19. Impact of mass flow rate of HTO on collected energy and pressure loss.

6

6

2.0x10

2.0x10

3

V=0m 3 V=50m 3 V=100m 3 V=150m

1.2x10

aux

/kJ aux

6

3

5

8.0x10

0

4

8

12

16

20

0.0

24

4

8

12

16

20

Time/h

(a) Spring equinox

(b) Summer solstice

24

6

2.0x10 3

V=0m 3 V=50m 3 V=100m 3 V=150m

6

6

1.6x10

/kJ

6

1.6x10 1.2x10

aux

/kJ

0

Time/h

6

aux

5

8.0x10 4.0x10

2.0x10

5

8.0x10

6

1.2x10

3

V=0m 3 V=50m 3 V=100m 3 V=150m

5

8.0x10

5

5

4.0x10

4.0x10 0.0

6

1.2x10

5

5

4.0x10 0.0

V=0m 3 V=50m 3 V=100m 3 V=150m

6

1.6x10

/kJ

6

1.6x10

0

4

8

12

16

20

24

0.0

0

4

8

12

16

Time/h

Time/h

(c) Autumnal equinox

(d) Winter solstice

20

24

Fig. 20. Auxiliary thermal energy consumption on different heat storage capacities.

storage system at spring equinox and autumn equinox. At summer solstice, Uaux decreases gradually when V is increased from 0 m3 to 150 m3. It is noticed that when V reaches 150 m3, only a little amount of auxiliary energy is required from 4:00 to 8:00 as indi-

cated in Fig. 20b. Interestingly, it is found in Fig. 20d that at winter solstice, the auxiliary energy system should be operated all day long and the amounts of the auxiliary energy required is same for different storage volumes because of the rather weak solar

Y.-L. He et al. / Applied Energy 97 (2012) 630–641

_ o on spring equinox, summer solstice, (3) The optimum m autumnal equinox and winter solstice are determined, which are, respectively, 2.0  104, 2.0  104, 1.4  104,1.4  104 kg/h. And, the recommended volumes of thermal storage system on spring equinox, summer solstice, autumnal equinox and winter solstice are 100 m3, 150 m3, 50 m3, 0 m3, respectively.

16 15

sys

/%

14 13 12

Spring equinox Summer solstice

Acknowledgments

Autumnal equinox Winter solstice

11 10

641

0

50

100

150

3

V/m

Fig. 21. Efficiency on different heat storage capacity.

radiation intensity in winter. This implies that the thermal storage system can be removed if the weather condition is always similar to that at winter solstice in Xi’an. Fig. 21 also presents the variation of total system efficiency (gsys) with thermal storage capacity for different seasons. Clearly, as indicated in the figure, the highest gsys can be achieved when the thermal storage volumes are 100 m3, 150 m3, 50 m3 and 0 m3, respectively for spring equinox, summer solstice, autumnal equinox and winter solstice. Taking into account the dependence of Uaux and gsys on V, the volumes of 100 m3, 150 m3, 50 m3 and 0 m3 for the thermal storage systems are, respectively, recommended for the solar thermal power generation system under the weather conditions like spring equinox, summer solstice, autumnal equinox and winter solstice in Xi’an. 5. Conclusions This paper presents an integrated model for the typical parabolic trough solar thermal power generation system with Organic Rankine Cycle. The simulation is model built within the transient energy simulation package TRNSYS. With the model, the influences of several designing and operating parameters on the performance of the collector field as well as the whole system are examined. The main findings and conclusions are as follows: (1) When pinter is smaller than 10 Pa, the unit length heat loss of the solar collector (qloss) increase sharply. However, with further increasing pinter, the variation of qloss is limited. (2) With the increase in v, the heat collecting efficiency (ghc) increases quickly at beginning. With further increasing v, ghc is almost independent on v when v reaches a certain value. In addition, it is shown that ghc increases with the increase in Idn but decreases with the increase in h.

This work was supported by the Key Project of National Natural Science Foundation of China (No. 50736005) and the National Natural Science Foundation of China (No. 51176155).

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