Thermo-economic analysis of solar-biomass organic Rankine cycle powered cascaded vapor compression-absorption system

Thermo-economic analysis of solar-biomass organic Rankine cycle powered cascaded vapor compression-absorption system

Solar Energy 157 (2017) 920–933 Contents lists available at ScienceDirect Solar Energy journal homepage: www.elsevier.com/locate/solener Thermo-eco...

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Solar Energy 157 (2017) 920–933

Contents lists available at ScienceDirect

Solar Energy journal homepage: www.elsevier.com/locate/solener

Thermo-economic analysis of solar-biomass organic Rankine cycle powered cascaded vapor compression-absorption system

MARK



Bhavesh Patel, Nishith B. Desai , Surendra Singh Kachhwaha Department of Mechanical Engineering, School of Technology, Pandit Deendayal Petroleum University, Raisan, Gandhinagar 382007, Gujarat, India

A R T I C L E I N F O

A B S T R A C T

Keywords: Biomass Concentrated solar power Cascade refrigeration system Hybrid system Organic Rankine cycle Selection diagram

In this paper, a novel solar-biomass organic Rankine cycle (ORC) powered cascaded vapor compression-absorption system is proposed for low temperature cooling applications. The proposed system achieves clean and efficient low temperature cooling and heating with zero dependency on fossil fuels. Thermo-economic analysis is reported to assess the performance and commercial viability of the system. The solar fraction and break-even point (BEP), considering paraboloid dish, n-pentane organic fluid, straw type biomass, and Jodhpur location, are calculated as 0.254 and 7.71 years, respectively. Due to lower annual efficiency, the solar fraction for the linear Fresnel reflector (LFR) based system is 0.179; however, the lower cost of LFR field and lower cost of energy generation from biomass leads to the lower BEP (7.43 years). Thermo-economic performance of the system is also affected by the ORC working fluid and the calculated break-even values are 7.85 years for Toluene and 8.16 years for R245fa. In comparison with the solar-biomass powered system, the fully biomass powered system achieves 39% lower capital cost and 30% lower BEP. The selection of the biomass is influenced by the calorific value and cost. The decision of selection between the proposed system and the equivalent stand-alone cooling and heating is influenced by the characteristics of solar field and biomass, ORC working fluid, electricity and process heat cost, location of installation, discount rate, cost of process heat and electricity.

1. Introduction The rising trends of fuel prices and global warming with increasing energy demand urge the researchers to design energy efficient, environmental friendly, and commercially viable systems (Islam et al., 2016). Hybrid technologies based on cogeneration and trigeneration systems achieve higher efficiency and fuel utilization factor (Karellas and Braimakis, 2016). Renewable energy sources based hybrid systems can reduce the fossil fuel consumption and carbon footprint as well as suitable for the decentralized applications (Sahoo et al., 2015). Typically, the hybrid system is a traditional cogeneration system with an absorption and/or a vapor compression system (Baghernejad et al., 2016). Hybrid systems with prime movers, like, steam turbine or gas turbine or internal combustion engine are commercially available (Jradi and Riffat, 2014a). For small-medium scale applications, organic Rankine cycle (ORC) is a promising option compared to the steam Rankine cycle due to its superior thermodynamic performance and potential to operate with low temperature energy sources (Desai and Bandyopadhyay, 2016a). ORC can utilize different energy source, like, solar thermal (Rodríguez et al., 2016), biomass (Al-Sulaiman et al., 2012), geothermal (Coskun et al., 2012), combined solar thermal and



geothermal (Bicer and Dincer, 2016), ocean thermal energy (Yang and Yeh, 2014) and waste heat (Liu et al., 2016), etc. Several studies on solar thermal and biomass energy powered ORC based hybrid systems are summarized in Table 1. Al-Sulaiman et al. (2011a) studied the parabolic trough collector (PTC) operated ORC integrated to absorption chiller and reported comparisons between cogeneration and trigeneration systems. Al-Sulaiman et al. (2011 b) analyzed hybrid systems using different sources, like, solid oxide fuel cell (SOFC), biomass and solar thermal and reported biomass and solar thermal energy powered hybrid system achieves the highest trigeneration efficiency (90%). Al-Sulaiman et al. (2012) performed energetic and exergetic analyses of biomass-ORC powered absorption chiller and reported trigeneration efficiency of 89% for trigeneration mode, 87% for heating cogeneration, 17% for cooling cogeneration, and 14% for power mode. Huang et al. (2013) performed thermo-economic analysis of biomass-ORC driven absorption system and reported payback period of 17 years. Buonomano et al. (2015) carried out thermo-economic analysis of solar-geothermal energy powered system and reported 69.4% system efficiency with 7.6 years of payback period. Karellas and Braimakis (2016) performed thermo-economic analysis of biomass–solar driven ORC integrated vapor compression system for

Corresponding author. E-mail address: [email protected] (N.B. Desai).

http://dx.doi.org/10.1016/j.solener.2017.09.020 Received 9 March 2017; Received in revised form 5 September 2017; Accepted 8 September 2017 0038-092X/ © 2017 Elsevier Ltd. All rights reserved.

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Nomenclature

PTC SOFC VARS VCACRS VCRS

Notations used in formulation A AC BEP C cp COP CRF CV d DNI E h I IAM k m P Q S T Ul1 Ul2 W x

area (m2) Annualized cost (USD/y) break-even point cost (USD) specific heat at constant pressure (kJ/kg-K) coefficient of performance capital recovery factor (y−1) calorific value (kJ/kg) discount rate (%) direct normal irradiation (W/m2) annual energy (kWh/y) specific enthalpy (kJ/kg) aperture effective direct normal irradiance (W/m2) incidence angle modifier lifetime (year) mass flow rate (kg/s) pressure (kPa) heat duty (kW) solar fraction temperature (°C) first order heat loss coefficient of collector (W/m2-K) second order heat loss coefficient of collector (W/m2-K) power (kW) concentration fraction (%)

Subscripts 0 a BB bio c cc CL comp D d e ele-g EV exp h HE k m min o O&M p PH PRV s shx UPH sys UE sa

Greek symbols Δ η ρ

difference efficiency density

Abbreviations LFR ORC

parabolic trough collector solid oxide fuel cell vapor absorption refrigeration system vapor compression absorption cascade refrigeration vapor compression refrigeration system

linear Fresnel reflector organic Rankine cycle

ambient absorber biomass boiler biomass condenser cascade condenser solar collector field compressor design desorber evaporator electric generator expansion valve expander heater heat exchanger component mean minimum optical operation and maintenance pump process heat pressure reducing valve isentropic solution heat exchanger unit process heat system unit electricity stand-alone

(Kowalski and Zenouzi, 2006). The electricity consumption of VCACRS is 30–50% lower compared to equivalent VCRS (Cimsit and Tekin, 2012).. Extensive investigations on VCACRS for space cooling and/or refrigeration applications powered by gas engine (Jeong et al., 2011), waste heat of gas turbine (Garimella et al., 2011), industrial waste heat (Colorado and Rivera, 2015), and solar thermal using flat plate solar collector (Boyaghchi et al., 2016) are summarized in Table 2. The range of COP for the cascade system with series and parallel configurations are 0.27–0.67 and 1.24–1.84, respectively. The payback period for the waste heat based cascade system is reported between 4.5 to 8.7 years. As comprehended from the literature, the ORC integrated VARS based hybrid systems are typically limited to space cooling. On the other hand, ORC integrated VCRS based hybrid systems (for refrigeration applications) are having higher compressor size/capacity and higher electricity consumption. Recently, Patel et al. (2017) proposed an integration of waste heat ORC with cascaded refrigeration system, which fulfills the electrical energy requirement of vapor compression section and heat duty requirement of vapor absorption section. An ORC integrated VCACRS combines the advantages of both systems and achieves low temperature (up to −20 °C) cooling efficiently. Detailed energy, exergy and preliminary economic analyses of the waste heat ORC powered cascaded refrigeration system is presented by Patel et al. (2017). In the present study, extensive thermo-economic analysis of the solar–biomass ORC integrated cascaded refrigeration system is

cogeneration and trigeneration applications and reported 7 year payback period for trigeneration system. Amirante et al. (2016) conducted energetic–economic investigations of biomass–ORC integrated absorption system and reported efficiency and payback period as 71.8% and 6 year, respectively. Applications of hybrid systems at a commercial level are limited to space cooling using lithium bromide-water (LiBr–H2O) absorption refrigeration systems (Tassou et al., 2010). Hybrid system with ammonia–water (NH3–H2O) absorption system is less advisable due to toxicity, flammability, less boiling point temperature difference of refrigerant and absorbent, incompatibility with materials and lower coefficient of performance (COP) (Deng et al., 2011). Hybrid systems for low temperature (below 0 °C) applications, like, food processing, storage, retail market applications etc., are having limited commercial availability (Tassou et al., 2010). Liquid desiccant cooling, adsorption cooling, and multi effect absorption cooling technologies based smallscale integrated systems are still at research and development phase (Jradi and Riffat, 2014a). In recent years, numerous studies have been reported on vapor compression–absorption cascade refrigeration system (VCACRS) because of its potential to achieve low temperature cooling with lower electricity consumption (Jain et al., 2015a). The VCACRS system combines the advantages of stand-alone vapor compression refrigeration system (VCRS) and vapor absorption refrigeration system (VARS) 921

Solar thermal and geothermal

Solar (evacuated tube collector)

Biomass and Solar (using PTC)

Biomass

Waste heat

Buonomano et al. (2015)

Boyaghchi and Heidarnejad (2015)

Karellas and Braimakis (2016)

Amirante et al. (2016)

Patel et al. (2017)

Biomass

Huang et al. (2013)

Biomass

Solar (using flat–plate collector)

Wang et al. (2012)

Jradi and Riffat (2014b)

Biomass

Al-Sulaiman et al. (2012)

Biomass

(R245fa)

SOFC–Biomass–Solar thermal

Al-Sulaiman et al. (2011 b)

Maraver et al. (2013)

500 kW (n-octane)

Solar (using PTC)

Al-Sulaiman et al. (2011a)

922 9.18 kWe (n-pentane)

281 kW (OMTS)

1.42 kWe (R134a, R152a, R245fa)

2.7 kW (R123)

6 kWe (R245fa)

0.5 KWe (HFE7100)

R245fa, R134a, R152a, npentane, Toluene, Siloxanes

222 kW (R245fa)

500 kW (n-octane)

500 kW (n-octane)

Source of energy

Author

ORC capacity (working fluid)

Table 1 Summary of literature review on Solar–Biomass driven ORC based hybrid system.

Trigeneration exergy efficiency: 20% for solar mode, 8% for solar and storage mode, and 7% for storage mode Cooling cogeneration exergy efficiency: 6.1% for solar mode, 3.5% for solar and storage mode, and 2.5% for storage mode. Heating cogeneration exergy efficiency: 19% for solar mode, 7.5% for solar and storage mode, 6% for storage mode Trigeneration energy efficiency: 90% for biomass mode; 90% for solar mode 76% for SOFC mode. 46% for solar and storage mode; 41% for storage mode System efficiency: 89% for trigeneration mode, 87% for heating cogeneration; 17% for cooling cogeneration, 14% for power mode Exergy efficiency: 28% for trigeneration mode, 27% for heating cogeneration mode, 13.5% for cooling cogeneration mode, 13% for power mode System efficiency: 27% for cooling cogeneration, 19% for heating cogeneration mode, and 10% for power mode Cooling capacity: 226 kW; Heating capacity: 889 kW. System efficiency: 71% for trigeneration mode, 85% for heating cogeneration mode, 11% for power mode. Payback period: 17 years R245fa, R134a, R152a are promising working fluids for 20 to 35 °C condensing temperature and small scale applications. Toluene, n-pentane, siloxanes are promising working fluids for 60 to 80 °C condensing temperature and medium and large scale applications Cooling capacity: 6.5 kW, Heating capacity: 9.6 kW. System efficiency: 85% for trigeneration mode, 83% for heating cogeneration, 4.2% for power mode Cooling capacity: 30 kW, Heating capacity: 87 kW. System efficiency: 69.4% for trigeneration mode, 6.4% for power mode Payback period: 2.5–7.6 years. Cooling capacity: 4.5 kW, Heating capacity: 11 kW Thermal efficiency: 23.66% (summer mode), 48.45% (winter mode) Exergy efficiency: 9.51% (summer mode), 13.76% (winter mode) Cooling capacity: 5 kW, Heating capacity: 53.5 kW Solar–ORC thermal, Electrical and Exergy efficiency: 6%, 3%, and 7% COP of VCRS is 3.88, Payback period is 7 years Cooling capacity: 500 kW, Heating capacity 1516 kW System efficiency: 71.8% for heating cogeneration, 31.2% for cooling cogeneration, 11.2% for power mode Payback period: 6 years Cooling capacity: 30.7 kW; Heating capacity: 77.9 kW; System efficiency: 79.02%; Rational (exergy) efficiency: 46.7%; Break-even point: 4.9 years

Single effect absorption chiller (LiBr–H2O)

Cascaded VCRS-VARS. (R410 A & LiBr–H2O)

Absorption chiller (LiBr–H2O)

Vapor compression system (VCS) (R134a, R152a, R245fa)

Ejector refrigeration cycle (R123)

Single stage absorption chiller (LiBr–H2O)

Liquid desiccant cooling unit (Potassium formate)

Absorption chiller (NH3–H2O and LiBr–H2O) and Adsorption chiller (Silica gel–water Zeolite–water)

Absorption chiller (LiBr–H2O)

Ejector cooling unit (R245fa)

Single effect absorption chiller (LiBr–H2O)

Single effect absorption chiller (LiBr–H2O)

Key findings

Cooling system (Working Fluid)

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50 kW, Refrigeration (–17 °C)

170 kW, Water chilling (4 °C)

51–56 kW, Refrigeration (4 °C)

200 W, Ultra–low temperature cooling (–170 °C)

100 kW, Refrigeration (0 °C)

Colorado and Rivera (2015)

Jain et al. (2015 b)

Boyaghchi et al. (2016)

Xu et al. (2016a)

Xu et al. (2016 b)

100 kW, Air conditioning (17 °C)

Mohammadi and Ameri (2014)

60.70 kW, Refrigeration (−55 °C)

66.67 kW, Refrigeration (−4 °C)

Jain et al. (2013)

Chen et al. (2015)

50 kW, Low temperature cooling (−10 °C)

Cimsit and Tekin (2012)

50 kW, Low temperature cooling (−10 °C)

Gas turbine waste heat

51 MW–82 MW, High heat flux electronics applications (−40 °C)

Garimella et al. (2011)

Cimsit et al. (2014)

Gas engine

100 kW, air conditioning (7 °C)

Jeong et al. (2011)

923 Waste heat

R410A

Mixture (R728, R50, R1150, R290, R600a)

R134a, R407 C, R22, R1234ze, and R1234yf

Solar thermal (using flat plate solar collector) Low–grade heat

R410a

CO2 and R134a

CO2

R134a

R–22

R22, R341a, R410 A, R407 C

R134a, R410a and NH3

CO2

NH3

R22

Working fluid VCRS

Industrial waste heat

Industrial waste heat

Thermal Energy

Thermal energy

Micro gas turbine

Waste heat

Thermal energy

Natural gas based engine

596 kW, Air conditioning

Sun (2008)

Source of energy

Cooling capacity and application

Author

Table 2 Summary of literature review on vapor compression–absorption cascade refrigeration system.

LiBr–H2O

Refrigerant: R23/R134a; Absorbent: DMF

LiBr–H2O

LiBr–H2O

LiBr–H2O

NH3–H2O

LiBr–H2O

LiBr–H2O

LiBr–H2O

LiBr–H2O and NH3–H2O

LiBr–H2O

LiBr–H2O

LiBr–H2O

VARS COP: 1.84; 40% annual energy saving compared to stand–alone VCRS COP: 1.87 for cascade system, 5.5 for compression section, 0.78 for absorption section. 14% performance improvement at 50% partial load operation COP: 0.594 for cascade system, 0.78 for absorption section, 2.17 for compression section. 31% annual energy saving compared to stand–alone VCRS. 48–51% annual energy saving compared to stand–alone VCRS using LiBr–H2O for VARS and R134a for VCRS. 33% annual energy saving using LiBr–H2O/R134a compared to NH3–H2O/R134a fluid pair 61% reduction in electric power consumption. 155% improved COP of the compression section. 4.5 years payback period and 3.5 years break-even point 50% increase in efficiency compared to stand–alone VCRS. 50% lower water consumption and 10% less second law efficiency for air–cooled absorption and water cooled compression section compared to both water cooled. COP: 0.61 for cascade system. 50% annual energy saving compared to stand–alone VCRS. Thermo–economically optimized system and improved COP and exergetic efficiency about 7% and 3.1%. COP: 0.277 and Exergy efficiency: 21.03%. 49.73% and 6.78% higher COP and exergy efficiency compared to two stage NH3–H2O absorption system COP of cascade system: 0.58 using single effect and 0.91 using double effect absorption system 45% annual energy saving compared to stand–alone VCRS. COP for different cascade configurations: 0.67 for series, 1.24 for parallel, 0.7 for combined series parallel. Payback period: 8.7 years for series, 11.5 years for parallel, 9.2 years for combined series parallel. Energy coefficient: 9.34 and exergy coefficient: 58%. R134a is better fluid based on energy and exergy analysis and R1234ze is better fluids based on exergo–economic analysis. COP of the compression section increased by 28.6%. Compressor power consumption decreased by 5.2%. 6.4 year payback period. COP and exergy efficiency of absorption-compression (ejector based double evaporator) system: 5.43 and 20.5%. Maximum COP is found at 5 °C intermediate cascade temperature.

Key findings

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produces electricity for compression section and thermal energy available at expander exhaust is utilized as a process heat and in the absorption section (see Fig. 1). Due to diurnal and seasonal variations in the solar radiation, the collector heat gain (state 29 to 19′) varies and therefore, the biomass boiler is used to supply continuous and fixed heat duty (state 30 to 19″) to ORC evaporator (state 19 to 20). The biomass consumption rate supplements according to the variation in heat duty provided by concentrating solar collectors. Thermal energy generated by the solar-biomass source is transferred to heat up the pressurized water. The heated water supplies energy to the organic working fluid through the ORC evaporator (state 19 to 20) and the condition of organic fluid at the ORC evaporator outlet is saturation vapor (state 18). Further, the ORC working fluid expands through the expander (state 18 to 15) and produces useful energy to run the compressor of the VCRS. It may be noted that the condition of the organic working fluid at the outlet of the ORC expander (state 15) is always superheated for dry working fluids. The energy available at the exhaust of expander is utilized for process heating (state 15 to 15a) and in the desorber (from state 15a to 16) to run the VARS. The ORC pump (state 16 to 17) drives the saturated organic liquid to the ORC evaporator and recirculate in the cycle. In the present study, the design capacity of the ORC is selected to meet the requirements of the compressor and pumps. The selection of higher design capacity of ORC results in extra electrical output from the system; however, the solar collector field area and annual biomass consumption increases accordingly. In the vapor absorption system, the weak LiBr-H2O solution enters into the desorber (state 8). In the desorber, LiBr-H2O is heated through ORC expander exhaust (state 15a to 16) where water vapor is separated (state 12) and strong LiBr solution leaves the desorber (state 9) and enters the absorber (state 11). The water vapor condenses into the VARS condenser by rejecting heat (state 12 to 13). That condensed liquid is expanded (state 13 to 14) and flows into the cascade condenser,

2. System description The combined solar-biomass energy assisted organic Rankine cycle integrated cascaded vapor compression and absorption system is shown in Fig. 1. The ORC is used as a power generating cycle as well as to supply thermal energy for process (water) heating and to drive the VARS. The VCACRS is used to produce low temperature cooling. The working fluids used for ORC, compression section and absorption section are n-pentane, R410A and LiBr–H2O, respectively. Pressurized water is considered as a working fluid for the hybrid solar-biomass sources (Karellas and Braimakis, 2016). On a typical sunny day, concentrated solar energy powered ORC

Desorber 8

EV 1

15a 9

28 7

PRV 11

Organic Rankine cycle

19

18 Expander

20

20' Spliter

BB=Biomass boiler CL=Concentrating solar collector field

Motor

Vapor compression cycle

Compressor 3

Evaporator 21

19''

19'

2930

23 24

4

EV 2

G

17

Absorber

1

2

6

5

Cascade condenser

15

10

Pump 14

27

Shx

Vapor absorption cycle

Mixer

CL

12

Condenser

Pump

16

ORC evaporator

13

26

Heater

25

BB

presented. Integration of solar-biomass based heat input makes the proposed system independent of the fossil fuel and grid connectivity; therefore, this system is ideal for remote and decentralized requirements of cooling and heating. The proposed system enables clean and efficient low temperature cooling as well as heating suitable for off-grid applications, like, process industries, shopping malls, hospitals, etc. It may be noted that the waste heat ORC powered VCACRS will be always cheaper than the fully biomass or solar-biomass ORC powered VCACRS. Detailed thermo-economic analyses of the proposed system is carried out considering different solar collector fields, organic working fluids, biomass type, cost of electricity and heating, discount rate, and location of installation. Detailed thermo-economic analyses and comparisons with the equivalent stand-alone cooling system (VCRS) and heating (with a given cost of heating, USD/MWhth) are also reported. The small capacity cooling system (30.7 kW) is used in the present analysis due to size limitations of VARS. Higher cooling capacity of the system can avail the advantages of lower break-even point and better thermoeconomic performance compared to the stand-alone cooling and heating.

22 Fig. 1. Solar-biomass ORC powered cascaded vapor compression-absorption system.

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(Desai et al., 2014),

where the water–vapor mixture evaporates (state 14 to 5). In the absorber, vapor gets mixed with LiBr solution and converts into weak LiBr-H2O solution (state 6) and finally recirculates in the desorber through the solution heat exchanger. The vapor compression and absorption sections are thermally coupled by means of the cascade condenser and forms the VCACRS. In the cascade condenser, superheated vapor from the compressor rejects the thermal load (state 4 to 1) to the absorption section and condenses at lower condensing temperature (T1) maintained by the VARS. The condensed liquid refrigerant is expanded into the expansion valve (state 1 to 2). In evaporator, low pressure and temperature refrigerant mixture evaporates (state 2 to 3) and the low temperature cooling is achieved (state 21 to 22). In the compressor, saturated vapor is compressed and converted into superheated vapor (state 3 to 4). Due to the lower condensing temperature of VCRS, the pressure ratio across the compressor reduces which causes reduction in electricity consumption of compression section and size of compressor as well as expander. The assumptions used while modeling the organic Rankine cycle (Tchanche et al., 2010) and the vapor compression–absorption cascade system (Jain et al., 2014) are,

Ap,CL =

QCL,D ηCL,D ·ID

(3)

where ηCL,D is solar collector field efficiency at design condition and QCL,D is solar collector heat gain at design condition (in this case QCL,D = Qe,ORC). The efficiency of the solar collector field (ηCL), at aperture effective design DNI (ID) and at any aperture effective DNI (I), can be calculated by,

ηCL,D = ηo,CL−Ul1·⎛ ⎝ ⎜

ηCL = ηo,CL−Ul1·⎛ ⎝

(Tm,CL−T0)2 ⎞ Tm,CL−T0 ⎞ −Ul2·⎛ ID ID ⎠ ⎝ ⎠ ⎜





(4a)

(T −T )2 Tm,CL−T0 ⎞−Ul2·⎛ m,CL 0 ⎞ I I ⎠ ⎠ ⎝ ⎜



(4b)

It may be noted that the Eq. (4a) and (4b) are generalized equations for finding the solar collector efficiency at design and off-design conditions, respectively (IIT Bombay, 2014). Moreover, these equations can be used for analysis of any type of collectors (e.g., PTC, LFR or paraboloid dish). The solar collector field heat gain, QCL, at any aperture effective DNI (I) is calculated by,

• The system is assumed to perform under steady state. • Pressure drops and heat losses in the pipe lines and heat exchangers are neglected. • The lithium bromide-water solution in absorber and desorber are assumed to be in equilibrium at a given pressure and temperature. • The fluids at states 1, 13, 16 are considered as saturated liquids and

QCL = ηCL ·I ·AP,CL

(5)

In case of low or negligible solar radiations, heat requirement of the ORC evaporator is generated by the biomass boiler (back up energy source) for stable and continuous system operation,

at states 3, 5, 18 are considered as saturated vapor.

3. Thermo-economic analyses

QBB = Qe,orc−QCL

The penetration of any technology to commercial market depends on its thermo-economic viability. In this regards, thermo-economic model of the proposed system is formulated and presented in this section. Moreover, the comparative thermo-economic model of the proposed system and stand-alone VCRS and heating (with a given cost of heating, USD/MWhth) is presented.

where QBB and Qe,orc is heat duty of biomass boiler and ORC evaporator, respectively. The biomass fuel consumption can be calculated as:

mbio =

The fundamental equations, used to formulate mathematical model of basic ORC and VCACRS, are based on Desai and Bandyopadhyay (2009) and Jain et al. (2013), respectively (see Table 3). It may be noted that the proposed system is sized and modeled considering commercially available minimum cooling capacity of the vapor absorption refrigeration system. The collector useful energy gain at design condition (QCL,D) can be calculated as:

Component/ system Evaporator Compressor Expansion Valve 2 Cascade Condenser Desorber Absorber

QCL,D = ηo,CL ·DNI ·IAM·Ap,CL −Ul1·(Tm,CL−T0)·Ap,CL −Ul2·(Tm,CL−T0)2 ·Ap,CL (1)

QCL,D =

QBB ηBB ·CVbio

(7)

Table 3 Thermodynamic equations of the ORC integrated VCACRS.

3.1. Thermodynamic modeling of the proposed system

ηo,CL ·ID·Ap,CL −Ul1·(Tm,CL−T0 )·Ap,CL −Ul2·(Tm,CL−T0)2 ·Ap,CL

(6)

(2)

where ηo,CL is collector field’s optical efficiency (considering cleanliness factor), Ul1 and Ul2 are heat loss co-efficient based on aperture area of collector field, Tm,CL is mean temperature of collector field ((T29 + T19′)/2), T0 is ambient temperature, and ID is aperture effective design direct normal irradiance (product of direct normal irradiance (DNI) and incident angle modifier (IAM)) (Desai and Bandyopadhyay, 2015). It may be noted that the incidence angle modifier for PTC field is the ratio of optical efficiency at any incidence angle and the efficiency at incidence angle equals to zero. IAM for linear Fresnel reflector (LFR) field gives the reduction in optical efficiency due to incidence as well as transversal angle (Schenk et al., 2014). It may be noted that the shadow losses and end-losses are neglected in the present analysis. The aperture area of the solar collector field, AP,CL is calculated by

Equations based on first law of thermodynamics

Qe = m3 (h3−h2) = (m21·cp)e (T21−T22) Ws,comp = m3 (hs,4−h3) ; Wele,comp =

Ws,comp ηs,comp·η ·η m ele − motor

h1 = h2 Qcc = m5 (h5−h14 ) = m 4 (h4−h1) ; T1−T14 = ΔTmin,cc m8 = m9 + m12 ; Qd = m12 h12 + m9 h9−m8 h8 m6 = m5 + m11 ; x 9 ·m11 = x 6 ·m6 Qa = m5 h5 + m11 h11−m6 h6 = (m23·cp)a (T24−T23)

Condenser

Qc = m13 (h12−h13) = (m25·cp)c (T26−T25)

Solution heat exchanger Expansion valve 1 Pressure reducing valve VARS pump

m9 h9−m10 h10 = m8 h8−m7 h7 ; εshx =

T9 − T10 T9 − T 7

h13 = h14 h10 = h11 Wp =

(P12 − P5)·m6 ; ρ·ηs,p

m7 h7 = m6 h6 + WP

ORC evaporator

Qe,orc = m18 (h18−h17) = (m20 ·cp)e,orc (T19−T20) ;

Expander

Ws,exp = m18 (h18−hs,15) ; Wele − exp = Ws,exp. ηs,exp . ηele − m

Qe,orc = QCL + QBB ORC Pump

Ws,P,orc =

m16 (P17−P16) ; ρ

WP,orc =

Ws,P ,orc ; ηs,P . orc

WP,orc = m17 h17−m16 h16 Heater

Qh = m15 (h15−h15a) = (m27·cp)h (T28−T27)

The states point in Table 3 corresponds to Fig. 1.

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electric output of generator, and ORC pump work, respectively. The cost of heat exchangers used in the ORC power block are calculated by (Xu et al., 2017),

where mbio is consumption of biomass (kg/s), and CVbio is the lower calorific value of biomass (kJ/kg). The design point overall efficiency of the system is given by,

ηoverall,D =

Qe + Qh ID·AP,CL

where CHE is the cost of heat exchangers, namely, ORC evaporator (Ce,orc), heater (Ch), desorber (Cd) which is a function of heat transfer area of aforementioned heat exchanger (AHE), a1, b1, c1, a2, a3, a4, and b4 are the regression cost coefficients. The area of heat exchanger (AHE) can be calculated following the procedure given in Appendix A.

where Qe is the VCRS evaporator heat duty and Qh is the process heat generation in the heater. The economic model for the proposed hybrid system is developed to evaluate the thermo-economic performance of the system. The economic model is used to estimate the break-even point compared to the equivalent stand-alone cooling system (VCRS) and heating (with a given cost of heating, USD/MWhth). The capital cost of the present system, which comprises the capital costs of solar collector field (CCL), biomass boiler (CBB), VARS (CVARS), VCRS (CVCRS) and ORC (CORC), is calculated by,

Csys = CCL·AP,CL + CBB + CVARS + CVCRS + CORC

3.2. Economic modeling of the proposed system The break-even point (BEP) is the period after which the additional cost of investment for the proposed system is recurred.

(9)

BEP =

The capital cost of the ORC power block (CORC) is given by,

CORC = ( ∑ CK )

(14)

CHE = a4 ·AHE + b4 (8)

Csys−CVCRS,sa ((CUE ·EVCRS + CO & M ) + (CUPH ·EPH ))sa−(Cbio·mbio + CO & M )sys (15)

(10)

where, ∑ CK is summation of costs of expander (Cexp), electric generator (Cele-g), ORC pump (Cp,orc), ORC evaporator (Ce,orc), heater (Cheater), desorber (Cd), and integration and installation cost of the ORC system (Cmisc). It may be noted that the condenser of the power block is replaced by the heater and desorber in the proposed system. The cost of ORC expander is calculated by (Boyaghchi and Heidarnejad, 2015),

where Cbio (USD/kg) is cost of biomass, mbio is biomass fuel consumption (kg/y); CVCRS,sa is cost of equivalent stand-alone vapor compression refrigeration system; CUPH is cost of unit process heat (USD/kWh); EPH is annual heating requirement (kWh/y); EVCRS,sa is annual energy consumption of stand-alone VCRS (kWh/y); CUE (USD/kWh) is cost of unit electricity. The annualized cost (AC) for the proposed system can be calculated by,

log10 (Cexp) = a1 + (b1·log10 (Wexp))−(c1·(log10 (Wexp))2 )

ACsys = (Csys·CRF + CO & M + Cbio·Mbio)

(11)

The cost of ORC electrical generator is calculated by (Boyaghchi and Heidarnejad, 2015), 0.95 Cele − g = a2 ·Wele −g

CRF = d·

(12)

(1 + d ) k ((1 + d ) k )−1

(17)

where Csys is capital cost of system (USD), CO & M is annual operation and maintenance cost (USD/y), CRF is capital recovery factor (y−1), d is discount rate and k is lifetime (year). Using Eq. (9) as well as operation and maintenance cost as a fraction (f) of capital cost, the annualized cost equation (Eq. (16)) can be expressed as:

The cost of ORC pump is calculated by (Boyaghchi and Heidarnejad, 2015),

Cp,orc = a3·W p0.71 ,orc

(16)

(13)

where Wexp, Wele-g and Wp,orc are work output of expander (shaft work), Table 4 Input data for design condition of ORC integrated VCACRS. Parameter

Value

References

Refrigeration capacity of compression chiller (Qe) Brine (PG-Water) temperature at exit of evaporator (T22) Temperature difference between inlet and exit of evaporator external fluid (PG-water) (T21-T22) Concentration of Propylene-Glycol Desorber temperature (TD or T12) ORC fluid temperature at exit of desorber (T16) Water temperature at inlet of condenser (T25) Water temperature at exit of condenser (T26) Absorber temperature (T6) Water temperature at inlet of absorber (T23) Water temperature at exit of absorber (T24) Water temperature at exit of cascade condenser (T5) Temperature driving force for cascade condenser (ΔTmin,cc) Water temperature at inlet of heater (T27) Temperature driving force for heater (ΔTmin,heater) Temperature driving force for ORC evaporator (ΔTmin,e,orc) Temperature difference between heat source and expander inlet (T19 – T18) Compressor isentropic efficiency (ηs,comp) Compressor mechanical efficiency (ηm) Compressor electro-motor efficiency (ηele-motor) Pump isentropic efficiency (ηs,p) Solution heat exchanger effectiveness (Ɛshx) Expander pressure ratio (PR) Expander isentropic efficiency (ηs,exp) Efficiency of ORC electrical generator-motor (ηele-m)

30.7 kW -15 °C 10 °C 40% 80 °C TD + 5 °C 27 °C 32 °C 37 °C 27 °C 32 °C 6 °C 8 °C 65 °C 5 °C 10 °C 15 °C 0.85 – (0.0467 Pressure ratio) 0.90 0.90 0.90 0.70 3 0.70 0.90

Refcon (2015) Refcon (2015) Cimsit et al. (2014) Refcon (2015) Thermax Limited (2015) Jain et al. (2015 b) Jain et al. (2015 b) Jain et al. (2015 b) Jain et al. (2015 b) Jain et al. (2015 b) Jain et al. (2015 b) Jain et al. (2015 b) Jain et al. (2013) Rentizelas et al. (2009) Desai and Bandyopadhyay (2016 b) Lecompte et al. (2014) Tchanche et al. (2009) Mohammadi and Ameri (2014) Nikolaidis and Probert (1998) Colorado and Velazquez (2013) Jain et al. (2013) Colorado and Velazquez (2013) Tchanche et al. (2010) Tchanche et al. (2010) Al-Sulaiman et al. (2011a)

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annually. Therefore, the economic performance of the proposed plant will be always lower than the data reported in this study. As a base case, the indigenous paraboloid dish, ORC with n-pentane, straw type biomass, and location at Jodhpur (Longitude 26.28°N and Latitude 73.02°E) are considered. The flow parameters of the system (at each state points) and the results of thermo-economic analysis obtained for the base case are summarized in Tables 7 and 8, respectively. The indegeneious paraboloid dish (Arun 160) with two-axis tracking mechanism use inverted cavity type MS receiver which collects concentrated solar energy and transmits to the receiver fluid (Kedare et al., 2016). Various heat transfer fluids, like, pressurized water, steam or thermic fluid can be used in the receiver (Kedare et al., 2016). Hourly DNI data and ambient temperature data for Jodhpur, India are taken from the work of Ramaswamy et al. (2013) and IIT Bombay (2014). The system requires the 137.6 kW heat duty to generate required cooling (30.7 kW) along with the process heat generation of 77.6 kW. The temperature of process heat ranges between 60 to 84 °C, which can be further used for domestic water heating, preheating, cleaning, drying, pasteurizing in food and beverage industries, washing/bleaching in textile industries, air-conditioning etc. The overall efficiency of the proposed hybrid solar-biomass based ORC powered cascade system (base case) is calculated 47.1%. The cascading of VARS and VCRS improves the COP of the compression subsystem by 122% compared to the equivalent stand-alone VCRS (COP = 1.98). As a result, the lower size of the compressor is required in the proposed system compared to the stand-alone system. The required area of the paraboloid dish, for generating 137.6 kW of ORC evaporator heat duty, is 288 m2 (see Table 8). The solar fraction (S), the ratio of annual energy generated through solar collectors to the total annual energy required by the system, is 0.254 (total annual energy requirement is 1205 MWh). The remaining energy requirement, 899 MWh out of 1205 MWh, is fulfilled by biomass having annual biomass consumption of 236 tons. The total investment cost and break-even point of the system are estimated about USD 183,657 and 7.71 years, respectively. It may be noted that the higher cooling capacity of the proposed system can avail advantages of lower break-even point and better thermo-economic performance compared to the equivalent stand-alone systems as demonstrated by Patel et al. (2017). The effects of various parameters important from thermo-economic point of view are discussed below:

ACsys = (CCL·AP,CL + CVCRS + CVARS + CORC + CBB )·(CRF + f ) (18)

+ (Cbio·Mbio)

3.3. Economic modeling of the stand-alone cooling system (VCRS) and heating The proposed system is compared with the equivalent stand-alone cooling system and heating to see the feasibility. The annualized cost for the equivalent stand-alone VCRS and heating (with a given cost of heating, USD/MW hth) can be given as,

ACsa = ((C·CRF + CO & M + CUE ·E )VCRS + (CUPH ·EPH ))sa

(19)

ACsa = (CVCRS ·(CRF + f ) + CUE ·EVCRS + CUPH ·EPH )sa

(20)

4. Results and discussions Thermo-economic analyses have been carried out, for a typical year on hourly basis, using the Engineering Equation Solver (EES) (Klein, 2016). Thermophysical properties of all fluids are taken from the inbuilt functions of EES. A commercially available compression chiller of 30.7 kW capacity (with R410A) to achieve cooling at −20 °C (Refcon, 2015) and hot water driven single effect LiBr-H2O absorption chiller of 37.7 kW (Thermax Limited, 2015) are considered. The small size cooling application is considered as per the minimum commercial size available for the vapor absorption refrigeration system. In the ORC, npentane is used as a working fluid. Pressurized water is considered as a working fluid for the hybrid solar-biomass heat source (Karellas and Braimakis, 2016). The evaporator approach temperature, which is the difference between temperature of brine leaving the evaporator and the evaporator temperature, is taken as 5 °C (Hojjat Mohammadi and Ameri, 2014). The condenser approach temperature, which is the difference between the condenser temperature and temperature of water leaving the condenser, is taken as 3 °C. The pressure and temperature of the water (as external fluid for cooling) are taken as 101.325 kPa and 27 °C, respectively. The design and operating parameters used for the analysis of the present system are given in Table 4 and Table 5. The cost data used for the economic analysis are given in Table 6. In the present case, the constant cooling and heating demands for 8760 h/y are taken. If the system is installed for a specific heating and cooling load profile (provided both are completely known for 8760 h) then these variations can be match by higher or lower firing of the biomass. This may result in excess process heat or electrical output from the ORC. Based on the availability of the solar radiation, the consumption of the biomass changes and it is important to know, from the load profiles, which time plant is non-operational. It may be also noted that the real systems, with variable cooling-heating loads, typically operate lower than the 8760 h

4.1. Effect of solar collector type Comparisons of PTC and LFR is done against paraboloid dish (base case) and reported in the Table 8. Based on the assumptions of the optical efficiency and heat loss coefficients for solar collector fields, the calculated aperture areas of the PTC and LFR are 251 m2 and 271 m2. It may be noted that the required solar field area for the indigenous dish is

Table 5 Input design data for solar concentrating collector-biomass boiler circuit. Parameters of concentrating Solar collector

DISH

LFR

PTC

Optical efficiency of the solar field (ηo) Heat loss co-efficient based on aperture area of solar field (Ul1 in W/ m2-K) Heat loss co-efficient based on aperture area of solar field (Ul2 in W/ m2-K2) Collector tracking mode

0.65 0.35

0.65 0.1

0.7 0.1

0.00002

0

0

Two-axis tracking

Focal axis N-S horizontal and E-W tracking 800 Novatech design (Schenk et al., 2014)

Focal axis N-S horizontal and E-W tracking 800 Euro Trough design (Schenk et al., 2014)

2

Aperture effective design DNI (ID in W/m ) Incidence angle modifier (IAM) Parameters of biomass boiler Capacity of the biomass boiler (kW) Efficiency of biomass boiler (ηBB) Calorific value of Straw type biomass Calorific value of Rice husk biomass Calorific value of Willow chips biomass

800 – Values 20% higher than the heat duty of ORC evaporator 0.90 (Boyaghchi and Heidarnejad, 2015) 15330 kJ/kg (Huang et al., 2013) 12340 kJ/kg (Huang et al., 2013) 11450 kJ/kg (Huang et al., 2013)

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Table 6 Input cost data used for economic analysis. Component

Cost correlation variable of cost correlation

References

PTC field (CCL) LFR collector field (CCL) Paraboloid Dish collector field (CCL) Biomass boiler (CBB) Vapor absorption refrigeration system (CVARS) Vapor compression refrigeration system (CVCRS) Stand-alone compression cooling system (CVCRS)sa Integration and installation cost (Cmisc) Expander (Cexp)

370 USD/m2 210 USD/m2 250 USD/m2 255 USD/kW 26,978 USD 10,492 USD 14,998 USD

Karellas and Braimakis (2016) Cocco and Cau (2015) Kedare et al. (2016) Karellas and Braimakis (2016) Thermax Limited (2015) Refcon (2015) Refcon (2015)

20% of the total ORC component cost log10(Cexp) = a1 + (b1·log10(Wexp))-c1·log10(wexp)2); where, a1 = 2.6259, b1 = 1.4398, c1 = 0.1776 Cele-generator = a2·(Wele-exp)0.95; where, a2 = 60 Cp,orc = a3·(Wp,orc)0.71; where, a3 = 3540 CHE = a4·AHE + b4; where, a4 = 516.621, b4 = 268.45 69 USD/ton 58 USD/ton 72 USD/ton 2% of the capital cost 24 USD/MWh 0.15 USD/kWh

Lemmens (2016) Boyaghchi and Heidarnejad (2015)

Electric generator (Cele-g) ORC pump (Cp,orc) Heat exchanger (CHE) Straw biomass Rice husk biomass Willow chip biomass Fraction of O & M cost (f) Cost of unit process heat (CUPH) Cost of unit electricity (CUE)

Boyaghchi and Heidarnejad (2015) Boyaghchi and Heidarnejad (2015) Xu et al. (2017) Huang et al. (2013) Huang et al. (2013) Huang et al. (2013) Baral et al. (2015) Huang et al. (2013) Karellas and Braimakis (2016)

The cost data are updated to year 2015 using Chemical Engineering Plant Cost Index (CEPCI) index (CEPCI, 2015).

trough collector. It is worth to mention that the annual biomass consumption is higher for solar collector having lower solar fraction. LFR based system is having the lowest break-even point (7.43 years) due to the lower cost of both the LFR field (USD/kW) and heat generation through biomass. Based on the Eq. (18), annualized costs of the proposed system using Dish, LFR, and PTC, are calculated and reported in Table 8. The capital recovery factor is calculated using 10% discount rate and 25 years lifetime (Desai and Bandyopadhyay, 2016 b). The results of annualized cost reveal that the proposed system with LFR (39,658 USD/y) and paraboloid dish (USD 39,965 USD/y) are better compared to PTC (43,301 USD/y).

Table 7 Flow parameters of the proposed system. State point

Fluid

T (°C)

P (kPa)

m (kg/s)

h (kJ/kg)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 15a 16 17 18 19 19′ 19″ 20 20′ 20″ 21 22 23 24 25 26 27 28 29 30

R410a R410a R410a R410a Water LiBr–H2O LiBr–H2O LiBr–H2O LiBr–H2O LiBr–H2O LiBr–H2O Water Water Water n-pentane n-pentane n-pentane n-pentane n-pentane Pressurized Pressurized Pressurized Pressurized Pressurized Pressurized PG-Water PG-Water Water Water Water Water Water Water Pressurized Pressurized

14.0 −20.0 −20.0 43.3 6.0 37.0 37.0 63.3 80.0 49.9 49.9 80.0 35.0 6.0 109.1 85.0 85.0 85.5 136.5 151.5 151.5 151.5 143.6 143 143 −5.0 −15.0 27.0 32.0 27.0 32.0 65.0 84.4 143 143

1223 400 400 1223 0.94 0.94 5.63 5.63 5.63 5.63 0.94 5.63 5.63 0.94 415 415 415 1247 1247 495 495 495 495 495 495 101 101 101 101 101 101 101 101 545 545

0.159 0.159 0.159 0.159 0.016 0.211 0.211 0.211 0.195 0.195 0.195 0.016 0.016 0.016 0.352 0.352 0.352 0.352 0.352 4.024 – – 4.024 – – 0.851 0.851 2.283 2.283 1.906 1.906 0.957 0.957 – –

221.8 221.8 414.2 457.8 2512 90.9 90.9 144.2 195.8 138.2 138.2 2650 146.6 146.6 510.2 288.4 146.3 148.0 539.2 639.0 639.0 639.0 604.8 602.2 602.2 −40.2 −76.3 113.2 134.1 113.2 134.1 272.1 353.6 602.2 602.2

water water water water water water

water water

4.2. Effect of ORC working fluid The comparative results of n-pentane (base case), Toluene, and R245fa as organic working fluids are reported in Table 8. The required heat duty in the ORC evaporator about 137.6 kW for n-pentane is lowest compared to other fluids. Based on the ORC evaporator heat duty, the minimum solar collector field area is 288 m2 in case of npentane. The solar collector field area for R245fa and toluene based ORCs are 8% and 15% higher compared to n-pentane. Therefore, the annual biomass consumption is 236 tons for n-pentane, 254 tons for R245fa, and 273 tons for Toluene. The lower biomass consumption in case of n-pentane is due to the lower heat requirement of system as discussed earlier. It is worth to mention that the process heat generation is different for n-pentane (77.9 kW), R245fa (89.0 kW), and toluene (100.8 kW). The break-even point for n-pentane based system is the lowest (7.71 years) as a result of lower solar field area requirement and lower power block cost. Toluene based system achieves 12% higher annual process heat generation and 24% lower ORC power block cost compared to the R245fa based system; however, the solar collector field cost is higher by 7%. The overall system cost for Toluene working fluid is about 1% lower compared to R245fa based system. As a result of higher incentives by process heat, toluene based system achieves 7.85 years of break-even point, which is lower compared to R245fa based system. The annualized cost for the proposed system using npentane (39,965 USD/y) working fluid is lower compared to R245fa (USD 42,683 USD/y) and Toluene (USD 43,384 USD/y).

higher due to its lower optical efficiency and higher heat loss coefficient. For the annual energy requirement of the system (1205.4 MWh), the paraboloid dish (solar fraction 0.254) generates 12% and 42% higher energy output compared to the PTC (solar fraction 0.226) and LFR (solar fraction 0.179). The higher solar fraction for the paraboloid dish is achieved due its two-axis tracking mechanism. In case of the LFR based system, the lower value of solar fraction is due to the lower annual optical efficiency compared to the paraboloid dish and parabolic

4.3. Effect of biomass type Different types of biomass, namely, straw type (base case), rice 928

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Table 8 Summary of simulation results (Cooling capacity of 30.7 kW). Parameters

Base Case

Effect of Solar Collector

Effect of ORC fluids

Effect of biomass type

Effect of places

Type of collector

Dish

LFR

PTC

Dish

Dish

Dish

Dish

Dish

Dish

ORC working fluid Biomass type Place QCL,D (kW) ACL (m2) Auxiliary Boiler capacity (kWth) Process heat generation from heater (kW) Design point overall efficiency (%) Total annual energy demand (MWh/y) Net annual energy generation from solar (MWh/y) Net annual energy generation from biomass (MWh/y) Solar fraction Annual biomass consumption (tons/y) Total system cost (Csys in USD) Break-even point (y) Annualized system cost (ACsys in USD/y)

n-pentane Straw Jodhpur 137.6 288 165 77.9 47.1 1205.4 306.2 899.2 0.254 236 183,657 7.71 39,965

n-pentane Straw Jodhpur 137.6 271 165 77.9 50.1 1205.4 215.8 989.6 0.179 258 168,508 7.43 39,658

n-pentane Straw Jodhpur 137.6 251 165 77.9 54.1 1205.4 272.4 932.9 0.226 243 204,537 9.03 43,301

Toluene Straw Jodhpur 160.3 332 192 100.8 49.5 1404.2 356.7 1047.5 0.254 273 201,666 7.85 43,384

R245fa Straw Jodhpur 148.7 311 178 89.0 48.1 1302.6 330.9 971.7 0.254 254 199,752 8.16 42,683

n-pentane Rice husk Jodhpur 137.6 288 165 77.9 47.1 1205.4 306.2 899.2 0.254 291 183,657 7.92 40,650

n-pentane Willow chips Jodhpur 137.6 288 165 77.9 47.1 1205.4 306.2 899.2 0.254 314 183,657 10.84 46,352

n-pentane Straw type Ahmedabad 137.6 288 165 77.9 47.1 1205.4 254.3 951.0 0.211 248 183,657 8.01 40,931

n-pentane Straw type Daggett 137.6 288 165 77.9 47.1 1205.4 424.3 781.1 0.352 204 183,657 7.00 37,826

4.5. Comparative analysis between solar-biomass powered and fully biomass powered systems

husk, and willow chips are considered to evaluate the thermo-economic performance of the system. Paraboloid dish collector, n-pentane organic fluid, and Jodhpur location are selected for the analysis. The annual biomass consumption in case of straw type biomass (about 236 tons) is lower about 24% and 34% compared to the rise husk and willow chips type due to their different calorific values. It may be noted that the costs of different types of biomass are also different as mentioned in Table 6. The calculated break-even point is minimum about 7.71 years for straw type biomass which is 3% higher for rice husk type, and 40% higher for willow chip type. The higher value of break-even point for the willow chips are due to the lower calorific value and higher cost compared to the straw type and rice husk. It may be noted that the biomass price varies with the place and season; however, in the present study these have been assumed constant. The annualized cost for Rice husk type biomass based system is 1.7% higher and for willow chips biomass based system is 16% higher, compared to the straw type biomass based system (base case).

A system powered by fully biomass is also compared with the solarbiomass powered system and the result shows that the fully biomass powered system achieves lower BEP, as expected. However, the use of solar energy reduces the dependency on biomass, a carbon-neutral source with higher water and land foot-prints. The annual biomass fuel consumption is calculated as 315 tones, 391 tones, and 421 tones for straw, rice husk and willow chips types biomass, respectively. The capital cost of the biomass powered system (about USD 111,615) is about 42% lower compared to the solar-biomass powered system (base case). The calculated BEP of the biomass powered system is 5.4 for straw type, 5.7 for rice husk, and 10.4 for willow chips. It may be noted that the BEP and annualized cost term will be always lower for the waste heat power system compared to the fully biomass or solar-biomass powered system. The annualized cost of the fully biomass powered system (base case comparison) is about 9% lower compared to the proposed hybrid solar-biomass powered system (see Table 9).

4.4. Effect of installation location 4.6. Comparative analysis between hybrid solar-biomass powered proposed system and equivalent stand-alone cooling system (VCRS) and heating

The effect of location of installation have been studied using paraboloid dish, straw type biomass, and n-pentane organic fluid. Hourly DNI data for Jodhpur, India (Longitude 26.28°N and Latitude 73.02°E) and Ahmedabad, India (23.0225°N, 72.5714° E) are taken from the work of Ramaswamy et al. (2013) and for Daggett, USA (40.9054°N, 109.5211°W) from US DOE (2014). As the diurnal and seasonal values of DNI varies with the place, the solar fraction is highest (0.352) as compared to the other locations (see Table 8). The higher value of solar fraction causes decrease in the annual biomass consumption. Therefore, in case of the location of Daggett (USA), the minimum break-even point (7 years) is obtained. For Daggett (USA), the solar fraction is highest (0.35) and biomass consumption (204 t/y) is lowest compared to Jodhpur (India) and Ahmedabad (India). Therefore, the annualized cost of the system (37,826 USD/y) is lowest for Daggett (USA).

The proposed system generates the cooling and heating simultaneously and thus it is compared with the stand-alone equivalent capacity cooling system (VCRS) and heating (with a given cost of heating, USD/MWhth). Based on the annualized cost, the proposed system with dish (39,965 USD/y), LFR (39,658 USD/y), and PTC (43,301 USD/y) are better compared to the stand-alone VCRS and heating (43,491 USD/ y). The proposed system with straw type and rice husk biomass is preferred over equivalent stand-alone cooling system and heating. However, for willow chips based biomass system, the annualized cost is about 6.2% higher compared to the stand-alone cooling system and heating. The annualized costs for proposed hybrid system at all selected locations (Jodhpur, Ahmedabad, and Daggett) are low compared to the stand-alone VCRS and heating. The proposed system can avail the advantages of greenhouse gas emissions compared to the standalone

Table 9 Comparison with stand-alone cooling system and heating for different discount rates (Cooling Capacity: 30.7 kW; Process Heat: 77.9 kWth).

Discount rate (d)

Hybrid solar-biomass based proposed system (USD/y) 5% 10% (Base case) 15%

Biomass based proposed system (USD/y) 5% 10% (Base case) 15%

Stand-alone cooling and heating (USD/y) 5% 10% (Base case) 15%

Annualized cost (USD/y)

32,768

31,881

42,903

39,965

48,154

929

36,256

41,234

43,491

44,159

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alone cooling system and heating for the process heating prices of 20 USD/kWh (40,761 USD/y) and 24 USD/kWh (43,491 USD/y). For the lower unit process heat price (15 USD/kWh), the annualized cost of stand-alone VCRS and heating decreases (37,349 USD/y) and makes it favorable. Due to the increasing trends of the electricity and heating prices, break-even point for the proposed system is expected to be lower in future.

cooling system (30.7 kW) and heating (77.9 kWth). The calculations of the carbon dioxide (CO2) emissions are based on the assumptions that the coal based electricity and heating emits 940 g/kWhe (IEA, 2016)) and 500 g/kWhth (POST, 2016). As a result, the proposed system reduces CO2 emission by 549 t/y. In case of natural gas based electricity (405 g of CO2 per kWhe (IEA, 2016)) and heating (210 g of CO2 per kWhth (POST, 2016)), the proposed system reduces CO2 emission by 233 t/y. Annualized cost for n-pentane based system (39,965 USD/y) is lower compared to the stand-alone cooling system and heating (43,491 USD/y). It may be noted that for all the variations the cooling capacity (30.7 kW) is taken as constant. The process heat generation (77.9 kWth) is also constant with variations in solar collector field type, biomass type, place of installation, and stand-alone system, except with variation in the ORC working fluid (Toluene: 100.8 kWth; R245fa: 89.0 kWth). These variations in the generation of process heat are due the change in ORC thermal efficiencies with working fluids. It may be noted that the solar field area requirement also increases in case of Toluene and R245fa working fluid. An equivalent stand-alone cooling system and heating are compared with the Toluene and R245fa based system. It is demonstrated that the proposed system with Toluene (43,384 USD/ y) and R245fa (42,683 USD/y) are better than the stand-alone cooling system (30.7 kW) and heating with 100.8 kWth capacity (48,307 USD/ y) and 89.0 kWth capacity (45,826 USD/y).

5. Conclusions In the present study, solar-biomass organic Rankine cycle powered cascaded vapor compression-absorption system is proposed and analyzed. Thermo-economic performance of the proposed system is influenced by the type of solar collector field, working fluid of organic Rankine cycle, biomass type, electricity and heating cost, discount rate, and location of installation. Thermo-economic performance of the proposed system is also compared with the equivalent stand-alone cooling system and heating (with a given cost of heating, USD/MWhth). As a base case, indigenous paraboloid dish, Jodhpur location, npentane organic fluid, and straw type biomass are taken. Compared to the stand-alone VCRS and heating, the proposed system has the breakeven point of 7.71 years as well as the lower annualized cost about 8%. A system powered by fully biomass (BEP = 5.4 years) is thermo-economically better than the combined solar-biomass system (BEP = 7.71 years). However, the use of solar energy reduces the dependency on biomass, a carbon-neutral source with higher water and land foot-prints. Comparative analysis between different collectors shows that the value of solar fraction is highest for the paraboloid dish (0.254) compared to the PTC (0.226) and LFR (0.179). However, the lower cost of LFR field and lower cost of energy generation from biomass lead to lower BEP (7.43 years), compared to the paraboloid dish based system. The requirement of ORC evaporator heat duty and power block cost changes with the working fluid. The required ORC evaporator heat duty is higher about 8% for R245fa and 16% for Toluene compared to the npentane (about 136.5 kW). Thermo-economic performance of the system is also affected by the annual process heat generation and ORC power block cost; therefore, the break-even point values are 7.85 years for toluene and 8.16 years for R245fa. The selection of the biomass is influenced by the calorific value and cost of the biomass. As a result, the minimum BEP of 7.71 years is calculated for straw type biomass. The higher value of solar fraction reduces the annual biomass consumption and the annualized cost of the system. Therefore, the location Daggett (USA) is having the lowest BEP of 7 years. In general, higher values of solar fraction, calorific value of biomass, heating and electricity cost are favorable for the selection of the proposed system; however, lower values of cost of biomass and discount rate are preferable. It is worth to mention that the small capacity cooling system (30.7 kW) is used in the present analysis. Higher cooling capacity of the system can avail the advantages of lower break-even point and better thermo-economic performance compared to the standalone cooling and heating. The present study includes steady-state simulation of the proposed system with fixed cooling and heating demands and its comparison with the stand-alone cooling and heating. Present research is directed towards detailed dynamic simulation of the system with variable heating and cooling load profile in future.

4.6.1. Effect of discount rate The effects of discount rate for hybrid solar-biomass powered proposed system, biomass powered proposed system, and equivalent cooling system and heating is shown in Table 9. For the base case (d = 10%), the annualized cost is lower for only biomass powered system (36,256 USD/y) compared to the hybrid solar-biomass powered system (39,965 USD/y) and equivalent cooling system and heating (43,491 USD/y). For discount rate of 5%, the decrease in annualized cost of 18% for hybrid solar-biomass powered system, 12% for fully biomass powered system, and 1% for equivalent stand-alone cooling system and heating is observed. Higher discount rate of 15% leads to the selection of stand-alone cooling system and heating compared to the hybrid solar-biomass powered system. It may be noted that for all selected values of discount rates, the fully biomass powered system is found to be better compared to other two configurations. 4.6.2. Effect of unit electricity price Hybrid solar-biomass powered proposed system and equivalent cooling system and heating are compared for different unit electricity prices and results are reported in Table 10. Electricity price of 0.15 USD/kWh (base case) is favorable for the selection of hybrid solarbiomass powered system (39,965 USD/y) over stand-alone cooling system (VCRS) and heating (43,491 USD). For the lower electricity price (0.08 USD/kWh), the annualized cost of the proposed system remains same; however, the annualized cost of stand-alone cooling system (VCRS) and heating decreases (31,748 USD/y) and makes it favorable. 4.6.3. Effect of unit process heat price As compared in Table 10, the proposed system (Base case: 39,965 USD/y) achieves lower annualized cost compared to the stand

Table 10 Effect of unit electric cost, process heat cost and comparison with stand-alone cooling system and heating (Cooling Capacity: 30.7 kW; Process Heat: 77.9 kWth). Proposed system (Base case) Annualized cost (USD/ y)

39,965

Stand-alone cooling and heating CUPH = 15 USD/ CUPH = 20 USD/ kWh kWh 37,349 40,761

930

CUPH = 24 USD/kWh (Base case) 43,491

CUE = 0.08 USD/ kWh 31,748

CUE = 0.15 USD/kWh (Base case) 43,491

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Appendix A. Appendix In the present work, shell and tube type of counter flow heat exchangers (for ORC evaporator, desorber, and heater) are assumed (Buonomano et al., 2015). The outside tube diameter is taken as 14 mm with 1.2 mm thickness and the tube materials is considered as copper (Buonomano et al., 2015). The heater is designed assuming as a condensing unit. In the present study, the length of tube is taken as 3.5 m for the ORC evaporator, 4.5 m for the desorber, and 6 m for the heater. Moreover, the number of tube passes are taken as 2, 2, and 4 for the ORC evaporator, desorber and heater, respectively. It may be noted that the length of tube and number of tube passes are taken in such a way that the fluid velocity inside the tube can be maintained between 1 m/s to 2.5 m/s. In the present study, the design of heat exchangers are done considering that the refrigerant flows at shell side. The surface area of each heat exchanger (AHE) for a given heat load (QHE) can be calculated as (Xu et al., 2017): (A.1)

QHE = (U ·A·LMTD)HE

where QHE is the heat duty of particular heat exchanger, A is the area of heat exchanger to be calculated, LMTD is the logarithmic mean temperature difference (ORC evaporator, desorber, and heater). The overall heat transfer coefficient (UK) of heat exchanger can be expressed as (Xu et al., 2017):

1

UHE = ⎡ ⎣

( )·( ) + ( )·F + ( Do Di

1 hi

Do Di

i

Do 2·k copper

)·ln ( ) + F + ( ) ⎤⎦ Do Di

o

1 ho

(A.2)

where Do and Di are the outside and inside tube diameters, Fo and Fi are the fouling factors, ho and hi are the outside and inside heat transfer coefficients, kcopper is the thermal conductivity of the tube material. A.1. ORC evaporator The logarithmic mean temperature difference for the ORC evaporator is calculated as:

LMTD =

(T19−T18)−(T20−T18) ln

(

T19 − T18 T20 − T18

)

(A.3)

Internal heat transfer coefficient (hi) for the ORC evaporator is estimated using the correlation proposed by Gnielinski (1976) for water being cooled inside the pipe.

hi =

Nu·k ; Di

(A.4)

fr Pr ⎞⎟ Nu = ⎛ ⎞·(Re−1000)·⎛⎜ 1/2 2/3 ⎝8⎠ ⎝ 1 + 12.7·(fr /8) ·(Pr −1) ⎠

(A.5)

where Nu is the average Nusselt number, Re is the Reynold number, k is the thermal conductivity of the fluid, Di is the internal tube diameter, and Pr is the Prandtl number. The friction factor (fr), Reynolds number (Re), and velocity of fluid in tube (u) are estimated as:

fr = (0.79·ln (Re )−1.64)−2 ;Re =

ρ ·u·Di ṁ ; and u = D2 μ (ρ)·(3.14· 4i )·

(

Notube passtube

)

(A.6)

External heat transfer coefficient (ho) for the ORC evaporator is estimated using the correlation proposed by the Rohsenow et al. (1998) for vapor evaporation. 3

ho·(Tmean−T18) = μ·hfg ·((g ·

ρl −ρv 0.5 ⎛ cpl ·(Tmean−T18) ⎞ ) )·⎜ 1 ⎟ σ ⎝ 0.013·hfg ·Pr f ⎠

where Tmean is the average temperature

(

T19 + T20 2

(A.7)

)

A.2. Desorber The logarithmic mean temperature for the desorber is calculated as (Jain et al., 2016):

LMTD = (Tm−TD )

(A.8)

where Tm is the mean temperature (T15 + T16)/2 and TD is the desorber temperature. Internal heat transfer coefficient (hi) for the desorber is obtained from the Dabson and Chato correlation (Bergman and Incropera, 2011) for vapor condensation inside the tube. 1/4

3 ⎛ g ·ρ ·(ρ −ρ )·(ks )·hfg ⎞ ⎞ hi = 0.555·⎜ ⎛⎜ l l v ⎟ μl ·(Tm−TD )·Di ⎠ ⎟ ⎝⎝ ⎠

(A.9)

External heat transfer coefficient (ho) of the desorber can be estimated from the correlation developed by the Bakhtiari et al. (2011)

ho = 5554.3·(Γ)0.236

(A.10)

where Γ is the mass flow rate of LiBr solution (internal fluid) per unit length of wetted tube.

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A.3. Heater The logarithmic mean temperature for the heater is calculated as:

LMTD =

(T15a−T28)−(T15a−T27) ln

(

T15a − T28 T15a − T27

)

(A.11)

Internal heat transfer coefficient (hi) of the heater is obtained using Gnielinski correlation as previously discussed for calculating hi for the ORC evaporator. External heat transfer coefficient (ho) is calculated using the correlation proposed by the Holman (2002) for condensation at shell side.

g ·ρ ·(ρ −ρ )·kl3·hfg ⎞ ho = 0.725·⎜⎛ l l v ⎟ ⎝ N ·μl ·(Tsat −Ts )·Do ⎠

0.25

(A.12)

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