National Renewable Energy Laboratory 

Innovation for Our Energy Future 

A national laboratory of the U.S. Department of Energy

Office of Energy Efficiency & Renewable Energy

NREL is operated by Midwest Research Institute ● Battelle Contract No. DE-AC36-99-GO10337 

Thermal Storage Commercial 

Plant Design Study for a 2-Tank 

Indirect Molten Salt System 

Final Report 

May 13, 2002 — December 31, 2004 

B. Kelly 

Nexant, Inc. 

San Francisco, California 

D. Kearney 

Kearney & Associates 

Vashon, Washington 

Subcontract Report 

NREL/SR-550-40166 

July 2006 
Thermal Storage Commercial 

Plant Design Study for a 2-Tank 

Indirect Molten Salt System 

Final Report 

May 13, 2002 — December 31, 2004 

B. Kelly 

Nexant, Inc. 

San Francisco, California 

D. Kearney 

Kearney & Associates 

Vashon, Washington 

NREL Technical Monitor: H. Price 

Prepared under Subcontract No. AAA-2-32432-01 

Subcontract Report 

NREL/SR-550-40166 

July 2006

National Renewable Energy Laboratory

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303-275-3000 • www.nrel.gov 

Operated for the U.S. Department of Energy 

Office of Energy Efficiency and Renewable Energy 

by Midwest Research Institute • Battelle 

Contract No. DE-AC36-99-GO10337
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Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Thermal Storage Commercial Plant Design Study 

for a 2-Tank Indirect Molten Salt System 

Final Report 

Table of Contents 

A. Introduction and Background.............................................................................................. 1

B. Reference Conditions......................................................................................................... 2

C. Parametric Studies............................................................................................................. 2

D. Plant Design Conditions ..................................................................................................... 4

E. Part Load Analysis ............................................................................................................10

F. Capital Cost Estimates ......................................................................................................17

G. Annual Performance and Electricity Cost ..........................................................................22

Contributors 

Bruce Kelly, Nexant 

Henry Price, NREL 

David Kearney, K&A 

- i - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

[ This page intentionally blank. ] 

- ii - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

A. Introduction and Background 

The development of a near-term thermal energy storage (TES) option is a major breakthrough for 

parabolic trough solar power plant technology. The 2-tank indirect TES concept proposed by Nexant 

and SunLab is being implemented in the AndaSol 50 MWe (net) trough plant currently under 

development in Spaini

. The project will have between 6-12 hours of TES depending on the economic 

optimization. In the U.S., Duke Solar is seriously considering this option where operational and 

economic factors favor its use. The economic optimum depends on the cost and performance of the 

TES system. The cost and performance of the indirect TES system are not independent because the 

size (cost) of the heat exchanger affects the physical size of the storage system and the performance 

of the turbine. 

The objectives of this Task were to develop a set of conceptual designs for 2-tank molten salt thermal 

energy storage (TES) systems of varying thermal capacities for trough plants, and to estimate 

levelized electricity costs for optimized plant configurations with storage. The solar plant assumed in 

the analysis is the Solar Electric Generating System (SEGS)-type plant using VP-1 for the solar field 

heat transport fluid. 

The availability of low cost and efficient thermal storage is one of the attributes of large-scale solar 

thermal technologies. Thermal storage allows electric energy to be dispatched at the times when it is 

needed most, and allows parabolic trough projects to achieve favorable capacity factors in excess of 

50 percent. During the past two years, a significant effort has been undertaken to identify thermal 

storage technologies for parabolic trough plantsii,iii. The initial goal was to develop a thermal storage 

option that would be perceived as low risk, and could be used with confidence in near-term projects. 

An indirect 2-tank thermal storage system has been proposed which meets these requirementsiv v

. 

The system uses the following equipment: a cold storage tank, operating a nominal temperature of 

290 °C; a hot storage tank, operating at a nominal temperature of 385 °C; a storage inventory of a 

mixture of binary nitrate salts; an oil-to-salt heat exchanger; and nitrate salt circulation pumps. 

During storage charging, heat is transferred from the collector field heat transport fluid to the nitrate 

salt through an oil-to-salt heat exchanger. During storage discharging, the fluid flows are reversed, 

and heat is transferred from the nitrate salt to the oil through the same heat exchanger. 

The overall goal of this effort was to help transfer this TES technology from a prototype 

development status to a commercial status. The approach taken was to develop an industry/lab team 

to evaluate the potential TES designs for a near-term commercial plant. The team included a 

parabolic trough developer (Duke Solar), TES system designer (Nexant), and SunLab. Duke Solar 

defined the overall system requirements. 

- 1 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

B. Reference Conditions 

The reference solar plant assumed for the optimization includes: 

• 55 MWe gross (or 50 MWe net) steam Rankine cycle power plant (100 and 66 bar) 

• LS-2+ solar field using VP-1 HTF @ 391C 

• Thermal storage of 2-12 hours equivalent full load capacity 

The storage system consists of an oil-to-salt heat exchanger, a cold storage tank, a hot storage tank, 

and two circulation pumps. The storage medium is a mixture of 60 percent by weight sodium nitrate 

and 40 percent by weight potassium nitrate. The salt mixture, which was used successfully at the 10 

MWe Solar Two central receiver project in southern California, offers a favorable combination of 

high density, low vapor pressure, moderate specific heat, low chemical reactivity, and low cost. 

During a thermal storage charge cycle, a portion of the oil from the collector field is directed to the 

oil-to-salt heat exchanger, where the oil cools from a nominal inlet temperature of 391 °C to an outlet 

temperature of about 298 °C. Nitrate salt from the cold storage tank flows in a countercurrent 

arrangement through the heat exchanger. The salt is heated from an inlet temperature of 291 °C to an 

outlet temperature of 384 °C, and then stored in the hot storage tank. During a discharge cycle, the 

oil and salt flow paths are reversed in the oil-to-salt heat exchanger. Heat is then transferred from the 

salt to the oil to provide the thermal energy for the steam generator. 

An optimization of the thermal storage system involves the assessment of numerous parameters, 

including the following: 

 • The inverse relationship between 1) the surface area, and cost, of the oil-to-salt heat exchanger, 

and 2) the quantity and cost of the storage inventory 

 • The inverse relationship between 1) the surface area of the oil-to-salt heat exchanger, and 2) the 

part load performance penalty of the Rankine cycle when operating from thermal storage. 

C. Parametric Studies 

A series of parametric studies was undertaken to determine the preferred design parameters for a 

thermal storage system with several thermal storage capacities, varying from 2 to 12 hours. Nexant 

conducted a series of parametric thermodynamic and capital cost studies on the Rankine cycle, the 

storage system, and the steam generator for the representative 50 MWe projectvi. System designs and 

capital cost estimates were developed for 140 combinations of the following parameters: 

• Rankine cycle live steam pressures of 101 bar and 66 bar 

• Thermal storage capacities of 2, 4, 6, 9, and 12 hours of turbine operation at full load 

• Oil-to-salt heat exchanger log mean temperature differences of 2 °C to 15 °C. 

- 2 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Based on the Nexant designs, NREL used the Excelergy model to calculate annual performance of 

each combinationvii. For each design a parametric analysis was conducted to determine the optimum 

solar field size required to minimize the cost of energy. For this analysis, each of the 140 designs, 

plus the two no-storage cases, were modeled with 7 different solar field sizes. Care was taken to 

assure that the optimum solar field size was in the range of sizes selected. The optimum log-mean 

temperature difference (or heat exchanger size) that produced the lowest electric energy cost was 

found for each storage size for both power cycles. 

The studies involved the following steps: 

1) A representative computer model of the Rankine cycle was developed using the GateCycle 

programviii. 

2) The duties, overall heat transfer coefficients, surface areas, and pressure losses for the 

superheater, reheater, evaporator, and preheater of the steam generator were estimated using an 

in-house program 

3) The duty, overall heat transfer coefficient, surface area, and pressure losses for the oil-to-salt heat 

exchanger were calculated for log mean temperature differences of 2 °C to 15 °C. The solar 

multiple of the collector field was specified by NREL. For a solar multiple > 1, the design 

condition for the oil-to-salt heat exchanger was during thermal storage charging, with a duty 

equal to the solar multiple value times the design duty of the steam generator. 

4) Part load models were developed for the steam generator, the oil-to-salt heat exchanger, and the 

Rankine cycle to estimate the Therminol, nitrate salt, water, and steam temperature distributions 

during thermal storage discharging for heat exchanger approach temperatures of 2 °C to 15 °C. 

5) Parametric cost estimates for the thermal storage tanks, the nitrate salt inventory, the oil-to-salt 

heat exchangers, and the nitrate salt pumps were developed for each of the heat exchanger 

approach temperatures. 

6) The oil-to-salt heat exchanger approach temperature which provided the lowest unit storage cost, 

in $/kWht, was identified. 

7) For each heat exchanger approach temperature, estimates of the plant capital cost, annual energy 

production, and levelized energy cost were developed to determine the preferred thermal storage 

design parameters. 

8) Utilizing the NREL performance/cost model Excelergy, carry out performance analyses and 

electricity cost estimates for each heat exchanger approach temperature case to identify the 

optimum parameters. 

- 3 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

D. Plant Design Conditions 

The procedures for calculating the design point performance of the Rankine cycle, and estimating the 

surface areas and pressure losses in the steam generator and the oil-to-salt heat exchanger, are 

discussed in the following sections. 

Rankine Cycle 

The Rankine cycle design is a conventional, single reheat design with 5 closed and 1 open extraction 

feedwater heaters. The live steam pressure and temperature are 100 bar and 373 °C, respectively, 

and the reheat steam temperature is 373 °C. The principal design parameters are shown in Table 1, 

and the GateCycle flow diagram is shown in Figure 1. 

Cold and hot reheat steam pressures, feedwater heater extraction pressures, feedwater heater terminal 

temperature differences, feedwater heater drain cooler approach temperatures, condenser pressure, 

circulating water temperature range, and condenser approach temperature are typical values. 

Turbine expansion efficiencies, and the required live and reheat steam flow rates to achieve a gross 

output of 55.0 MWe, were calculated by GateCycle. Simultaneously, the low pressure turbine 

exhaust loss was adjusted manually to yield the desired gross cycle efficiency of 0.375. 

Steam Generator 

The surface areas and pressure losses of the preheater, evaporator, superheater, and reheater for the 

Therminol VP-1 steam generator were calculated using the following procedure in an Excel 

spreadsheet: 

 • Curve fits, as functions of temperature, were developed for the thermodynamic properties of the 

Therminol HTF, including density, viscosity, thermal conductivity, and specific heat. 

 • A Visual Basic version of the ASME Steam Tables was attached to the spreadsheet. 

 • Duties were calculated for the water/steam side of the heat exchangers from the pressure, 

temperature, and flow rate information provided by GateCycle. The required Therminol flow 

rate and inter-heat exchanger temperatures were calculated, by iteration, using the following 

constraints: a) the live and reheat steam temperatures were 20 oC less than the collector field 

outlet temperature; b) the evaporator pinch point was set to 7.2 °C to achieve the specified 

preheater oil outlet temperature of 301.7 °C; c) the temperature of the feedwater at the outlet 

from the preheater was set to 0.5 °C less than the saturation temperature to improve the accuracy 

of the log mean temperature difference calculations for the evaporator; and d) the reheater oil 

flow rate was set to achieve the specified oil outlet temperature of 225.2 °C. 

- 4 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

- 5 - October 25, 2002

G ross generator output, M W e 55.04

H igh pressure turbine

 - L ive steam pressure, bar 100.01

 - L ive steam tem perature, C 373

 - Live steam flow rate, kg/sec 60.231

 - Expansion efficiency 0.852

Interm ediate pressure turbine

 - R eheat steam pressure, bar 16.50

 - R eheat steam tem perature, C 373

 - R eheat steam flow rate, kg/sec 54.044

 - Expansion efficiency 0.900

Low pressure turbine

 - Expansion efficiency 0.850

 - E xhaust loss, kJ/kg 22.9

Final feedw ater tem perature, C 235

Feedw ater heater 1

 - E xtraction line inlet pressure, bar 33.5

 - E xtraction line pressure loss, percent 0.0

 - T erm inal tem perature difference, C 5.0

 - D rain cooler approach, C 5.0

Feedw ater heater 2

 - E xtraction line inlet pressure, bar 14.0

 - E xtraction line pressure loss, percent 0.0

 - Surface area, m

 - T erm inal tem perature difference, C 5.0

 - D rain cooler approach, C 5.0

D eaerator

 - E xtraction line inlet pressure, bar 6.18

 - E xtraction line pressure loss, percent 0.0

 - Feedw ater outlet tem perature, C 160.0

Feedw ater heater 4

 - E xtraction line inlet pressure, bar 3.04

 - E xtraction line pressure loss, percent 0.0

 - T erm inal tem perature difference, C 4.0

 - D rain cooler approach, C 5.0

Feedw ater heater 5

 - E xtraction line inlet pressure, bar 1.17

 - E xtraction line pressure loss, percent 0.0

 - T erm inal tem perature difference, C 4.0

 - D rain cooler approach, C 5.0

Feedw ater heater 6

 - E xtraction line inlet pressure, bar 0.37

 - E xtraction line pressure loss, percent 0.0

 - T erm inal tem perature difference, C 4.0

 - D rain cooler approach, C 5.0

Condenser

 - Inlet steam pressure, bar 0.080 2 5,659

C ooling tow ers

 - D ry bulb tem perature, C 4 0

 - R elative hum idity, percent 2 0

 - A pproach, C 6.8

 - R ange, C 9.7

Table 1 Rankine Cycle Design Parameters 
mal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

L Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

- 6 - October 25, 2002

ST1 

 0.08

 41.34

P 

W 

CND1

CNDPMP

FWH4

FDWPMP

ST2

FWH5

 33.48

 6.19

242.782811.9

P T

W H

 13.99

 2.73

351.613154.2

P T

W H

DA1

FWH0

FWH1

 0.37 2.24

 74.032515.2

P T

W H

IPST

 18.50 54.04

208.482712.4

P T

W H

 16.50 54.04

373.003195.8

P T

W H

 6.18 2.69

254.632966.7

P T

W H

 5.17

 48.63

235.082929.1

P T

W H

FWH2

 3.04 2.59

183.342830.7

P T

W H

 1.17 2.46

104.022675.3

P T

W H

 1.72

374.81

815.56859.76

P T

W H

 1.72

374.81

572.94584.45

P T

W H

VRHTR

VSUHTR

STMGEN

100.01

 60.23

373.003011.7

P T

W H

M1

SP1

V2

 6.18 2.69

254.632966.7

P T

W H

 6.18

 60.29

160.01675.49

P T

W H

 5.03 0.06

152.112726.6

P T

W H

V3

100.67

 0.06

311.462726.6

P T

W H 22.918

100.67

 60.29

235.011015.0

P T

W H

 9.31

 48.75

 70.03293.80

P T

W H

101.36

 60.29

190.02811.75

P T

W H

S2

S4

S5

S7

S8

S9

S11

S12

S15

S16

S17

S26

S27

S28

S32

S33

S35

S1

S18

S19

S22

S24

S30

S6

S13

S14

S3

S31

S20

S29

S23

S39

S40

S41

Figure 1 GateCycle diagram for 50 MWe Rankine Cycle Model 

TherNRE
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

 • Conventional shell and tube heat exchangers were selected for the preheater, the superheater, 

and the reheater. The high pressure water or steam was placed on the tube side, and the low 

pressure Therminol was placed on the shell side. 

 • A kettle evaporator, rather than a forced recirculation evaporator, was used. In general, kettle 

evaporators are normally less expensive than recirculation designs. Also, placing the heat 

transport fluid on the tube side posed only a minor potential for damage due to freeze-thaw 

cycles because the fluid has a favorable freezing point of only 11 °C. 

 • The Therminol flow through the reheater was assumed to be parallel to, and independent of, the 

Therminol flow through the superheater - evaporator - preheater combination. The Therminol 

flow from the reheater combined with the Therminol flow from the superheater - evaporator - 

preheater combination prior to returning to the collector field. 

 • Representative tube inside diameters, tube wall thicknesses, and water or steam velocities were 

selected for each heat exchanger, from which tube side heat transfer coefficients and pressure 

losses were calculated from published formulas. 

 • A tube baffle plate spacing was selected for each heat exchanger, from which the shell side heat 

transfer coefficients and pressure losses were calculated from published formulas. 

 • Representative tube side and shell side fouling coefficients were selected, and then combined 

with the convection coefficients and thermal resistance of the tube wall, to calculate overall heat 

transfer coefficients for each heat exchanger. 

 • Counterflow heat exchanger layouts, with two tube passes and two shell passes, were selected 

for the preheater, the superheater, and the reheater, and the surface areas were calculated as 

follows: 

C) o C)(Log Temperatur Mean ,Difference e o - 2 kJ/m , overall (U

2 kJ/sec Q, m Area, =

 • The overall heat transfer coefficient and surface area for the evaporator were calculated from 

published formulas for a kettle design. 

The procedure was also subject to the following constraints: 

 • The maximum length of tube which could be fabricated was assumed to be 25 meters. If the 

calculated tube length exceeded 25 meters, two heat exchangers were used in series. 

- 7 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

 • The desired pressure loss on the shell side of the preheater, the superheater, and the reheater was 

in the range of 1.4 to 1.7 bar. If the pressure loss was less than or more than the desired value, 

the tube baffle spacing was decreased or increased, respectively. Once the desired pressure loss 

was obtained, the final heat exchange areas were calculated. 

The steam generator configuration, showing fluid flow paths, fluid flow rates and temperatures, 

duties, heat exchanger areas, and pressure losses is illustrated in Figure 2 for a typical case. 

Figure 2 Typical 50 MWe Plant Steam Generator Configuration 

- 8 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Oil-to-Salt Heat Exchanger Design 

The preferred oil-to-salt heat exchanger is a conventional shell and tube designix. The surface area 

and pressure loss of the heat exchanger were calculated using the following procedure in an Excel 

spreadsheet: 

 • Curve fits, as functions of temperature, were developed for the thermodynamic properties of the 

Therminol and the nitrate salt, including density, viscosity, thermal conductivity, and specific 

heat. 

 • A duty was calculated from the heat balance information provided by GateCycle. The required 

fluid flow rates were calculated using the following constraints: a) the approach temperature at 

the hot end of the heat exchanger was the same as the approach temperature at the cold end; and 

b) the approach temperature was an independent variable in the optimization studies. 

 • The high pressure Therminol was placed on the tube side, and the low pressure nitrate salt was 

placed on the shell side. 

 • Representative values for the tube inside diameter, the tube wall thickness, and the Therminol 

velocity were selected, from which a tube side heat transfer coefficient and a pressure loss were 

calculated from published formulas. 

 • A tube baffle plate spacing was selected, from which the shell side salt velocity, heat transfer 

coefficient, and pressure loss were calculated from published formulas. 

 • Representative tube side and shell side fouling coefficients were selected, and then combined 

with the convection coefficients and thermal resistance of the tube wall, to calculate the overall 

heat transfer coefficient. 

 • A counterflow heat exchanger layout, with two tube passes and two shell passes, was assumed, 

and the surface area was calculated as follows: 

C) o C)(Log Temperatur Mean ,Difference e o - 2 kJ/m , overall (U

2 kJ/sec Q, m Area, =

The procedure was also subject to the following constraints: 

 • The maximum length of tube which could be fabricated was assumed to be 25 meters. If the 

calculated tube length exceeded 25 meters, multiple heat exchangers in series were used. 

 • For one heat exchanger, or for multiple heat exchangers in series, the desired total pressure loss 

on the shell side was in the range of 4.5 to 5.0 bar. If the pressure loss was less than or more 

than the desired value, the tube baffle spacing was decreased or increased, respectively. Once 

the desired pressure loss was obtained, the final heat exchange area was calculated. 

- 9 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

E. Part Load Analysis 

Steam Generator, Rankine Cycle, and Oil-to-Salt Heat Exchanger 

The oil-to-salt heat exchanger imposes a temperature drop when energy is transferred into storage, 

and a temperature drop when energy is transferred out of storage. Thus, when the steam generator is 

operating from thermal storage, the temperature of the Therminol is less than the design value of 

393 °C. The extent of the temperature decays is dependent on the log mean temperature difference 

selected for the heat exchanger. As expected, the decay in the steam generator source temperature 

leads to decays in both the live and reheat steam temperatures. Furthermore, if the flow rate of 

Therminol through the steam generator is fixed at the design flow rate, the thermal input to the steam 

generator is less than the design value. As a result, the flow rates of the live and reheat steam are 

lower than the design values, and the Rankine cycle operates under part load conditions. 

To estimate the effects of decays in the Therminol source temperature on the performance of the 

Rankine cycle and the steam generator, the following iterative process was adopted: 

 • A trial Therminol, water, and steam temperature distribution through the superheater - 

evaporator - preheater combination was selected, from which the log mean temperature 

difference for each shell was calculated. 

 • A trial Therminol outlet temperature from the reheater was selected. 

 • Thermal duties on the water and steam sides of the four heat exchangers were calculated. 

 • Therminol flow rates through the reheater and the superheater - evaporator - preheater 

combination were calculated based on the trial temperature distribution and the thermal duties. 

 • The Therminol flow rate through the reheater was assumed to be fixed at the design flow rate. 

 • From the water, steam, and Therminol flow rates, tube side and shell side heat transfer 

coefficients were calculated. From these, overall heat transfer coefficients were calculated. 

 • The required log mean temperature difference for each shell was calculated, using the thermal 

duties on the water and steam sides and the overall heat transfer coefficients, as follows: 

) 2 m C)(Area, o- 2 kJ/m , overall (U

kJ/sec Q, Co

 ,difference etemperatur mean log Required =

- 10 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

 • The calculated log mean temperature difference was developed using the standard formula: 

C = o ,difference etemperatur mean log Calculated

⎥

⎥

⎦

⎤

⎢

⎢

⎣

⎡

Co ,difference etemperatur Lesser

Co ,difference etemperatur Greater ln

Co ,difference etemperatur Lesser - Co ,difference etemperatur Greater

 • The required and calculated values were then compared. If the values differed, a new trial 

Therminol temperature distribution was selected, and the process was repeated until the log 

mean temperature differences converged. 

 • The resulting values for the live steam temperature, the live steam flow rate, the reheat steam 

temperature, and the reheat steam flow rate were exported from the part load steam generator 

model to the GateCycle program. 

 • In GateCycle, the turbine was assumed to operate without throttling; i.e., the live steam pressure 

was determined by the steam flow rate through the turbine. The GateCycle program calculated 

the live steam pressure, the condenser pressure, the final feedwater temperature, and the gross 

generator output, and then exported these values back to the part load steam generator model. 

 • The process was repeated until constant values were obtained for the live steam pressure, live 

steam temperature, reheat steam pressure, reheat steam temperature, and final feedwater 

temperature. 

The part load steam generator model was coupled with the oil-to-salt heat exchanger and the thermal 

storage system, as illustrated in Figure 3. The fluid temperature distribution in the oil-to-salt heat 

exchanger and the thermal storage system was calculated as follows: 

 • The hot salt tank temperature was set equal to the collector field outlet temperature minus the 

approach temperature of the oil-to-salt heat exchanger under full load conditions. 

 • The Therminol source temperature for the steam generator was set equal to the hot salt tank 

temperature minus the approach temperature of the oil-to-salt heat exchanger under part load 

conditions, with the duty of the heat exchanger equal to the duty of the steam generator. 

 • The cold salt tank temperature was set equal to the steam generator outlet temperature plus the 

approach temperature of the oil-to-salt heat exchanger under part load conditions. 

 • The collector field return temperature was set equal to the cold salt tank temperature plus the 

approach temperature of the oil-to-salt heat exchanger under full load conditions. 

- 11 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Figure 3 Part Load Steam Generator Operation 

In Conjunction with the Oil-to-Salt Heat Exchanger and the Thermal Storage System 

The calculation of the fluid temperature distribution within the thermal storage system was clearly an 

iterative process. First, a trial hot salt tank temperature was selected, from which the thermal input to 

the steam generator was calculated based on the following: a trial log mean temperature difference 

for the oil-to-salt heat exchanger; and an assumption the Therminol flow rate through the steam 

- 12 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

generator was equal to the flow rate under design point conditions. Using the procedures described 

above, a set of iterative calculations between the steam generator model and the GateCycle program 

was performed to reach convergence on the working fluid duties, temperatures, and pressures, and to 

calculate the Therminol exit temperature from the preheater. A trial log mean temperature difference 

for the oil-to-salt heat exchanger was selected, from which the cold salt tank temperature was 

calculated. The full load performance of the oil-to-salt heat exchanger was then analyzed to 

determine the 1) the required Therminol inlet temperature during thermal storage charging, and 2) the 

hot salt tank temperature. If hot salt tank temperature did not equal the desired value of [(391 °C) – 

(Log mean temperature difference for the oil-to-salt heat exchanger)], the trial hot salt temperature 

was revised, and the process repeated until all of the system part load requirements were satisfied. 

Preheater Feedwater Exit Conditions 

The fluid temperature distributions within the steam generator and the oil-to-salt heat exchanger 

generally converged in a smooth manner. However, for oil-to-salt heat exchanger approach 

temperatures greater than 4 °C, an enthalpy balance on the preheater showed the exit feedwater to be 

a mixture of saturated water and saturated steam. 

Strictly speaking, a log mean temperature analysis of a heat exchanger is valid only if the specific 

heats of both the tube and the shell fluids are constant. For two phase flow, this is not the case. To 

improve the accuracy of the temperature distribution calculations, the part load model for the 

preheater was modified, as follows: 

 • The heat exchanger was divided into 100 sections along the length of the tubes. 

 • The temperature distribution on the shell side of the heat exchanger was estimated by assuming 

the enthalpy change of the Therminol in each section of the tube length was equal. The 

temperature distribution along the length of the tube was calculated by solving for the local fluid 

temperature which yielded the desired enthalpy. Since the specific heat of the Therminol does 

not change very much over the temperature range of interest, the temperature distribution 

between the entrance and the exit was essentially linear. 

 • A similar approach was taken on the water side to determine the point along the tube length at 

which the water reached saturation conditions. The temperature of the Therminol at this 

location was calculated from the distribution of Therminol temperatures. The method for 

estimating part load heat exchanger performance (i.e., comparing the required and the calculated 

log mean temperature differences) was then applied to the sensible heat transfer portion of the 

preheater. For the oil-to-salt heat exchanger log mean temperature differences of interest, 

sensible heat transfer occurred along 90 to 100 percent of the preheater tube length. 

 • The enthalpy of the water-steam mixture leaving the preheater was estimated from the fraction 

of latent heat transfer. This enthalpy, together with the saturation temperature of the feedwater, 

were used in the part load performance calculation for the evaporator. 

- 13 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

- 14 - October 25, 2002

A calculation of the water-steam mixture velocity at the exit from the preheater was also made to 

ensure the velocity was not excessive. Interestingly, the mixture velocities were typically below the 

design point feedwater velocity due to 1) lower feedwater flow rates under part load conditions, and 

2) vapor fractions of only 1 to 9 percent. 

The results of the part load steam generator and Rankine cycle analyses for oil-to-salt heat exchanger 

approach temperatures in the range of 2 °C to 15 °C are shown, for 100 bar and 66 bar steam 

conditions respectively, in Table 2 and Table 3. As expected, the hot salt tank temperature decreases 

uniformly, and the collector field supply temperature increases uniformly, with an increase in the oilto-salt heat exchanger approach temperature. However, the temperature of the cold salt tank is, to a 

large degree, independent of the approach temperature. The effect can be traced to the live steam 

pressure. As the approach temperature increases, the thermal input to the steam generator and the 

live steam production rate both decrease. Since the turbine operates under sliding pressure, the live 

steam pressure also decreases. A reduction in live steam pressure is accompanied by reduction in the 

following: the saturation temperature in the evaporator; the Therminol temperature at the exit from 

the preheater; and the Therminol inlet temperature to the oil-to-salt heat exchanger during a charge 

cycle. 

The analysis shows the impact of each heat exchanger size on power plant electric output and 

power cycle efficiency. The heat exchanger size is shown in the table but each size refers to a 

specific log-mean-temperature-difference (LMTD) between the HTF and the salt. Key to note is 

that the electric output and efficiency both drop when operating from storage. The magnitude of 

the drop depends on the size of the heat exchanger. The larger the heat exchanger, the smaller the 

impact on turbine performance. Note the heat exchanger with a LMTD of 2°C is 7.5 times the 

size of the heat exchanger with an LMTD of 15°C. The electric output only drops to 53.4 MWe 

and the steam cycle efficiency drops to 37.5% for TES with a LMTD of 2°C compared to 55 

MWe and 37.9% when operating directly from the solar field (case 1). The electric output drops 

to 43.6 MWe and efficiency to 37.0% for the LMTD of 15°C case. 

Also evaluated was the impact of TES on a lower pressure (66 bar) steam turbine. An earlier 

study had concluded that a lower pressure steam cycle would help reduce the cost of TES and 

result in a lower overall cost of energy. The lower pressure steam cycle allows a lower HTF 

return temperature to the solar field. This increases the temperature difference across the TES 

system, reducing the required storage volumes (and cost) for the same amount of thermal energy 

storage. This cycle also a 55 MWe plant (gross) when operating directly from the solar field and 

had a steam cycle efficiency of 35.5%. Similar reductions in efficiency and electric output occur 

when this cycle operated from TES, again depending on the size of the heat exchanger.
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Table 2 Rankine Cycle, Steam Generator, and Oil-to-Salt Heat Exchanger Part Load Performanc

for 100 bar steam pressure 

Case 1 Case 2 Case 3 Case 4 Case 5 Case 6 Case 7 Case 8 Case 9 Case 10 Case 11 Case 

Oil-to-salt heat exchanger - LMTD, oC N/A 2 3 4 5 6 7 8 9 10 11 12 

 - Heat exchanger area, m2 N/A 80,860 49,625 34,755 26,408 21,201 17,693 15,116 13,252 11,789 10,559 9,62

 - Salt side pressure loss, bar N/A 5.14 5.20 5.09 5.03 5.10 5.05 5.06 4.99 4.97 5.03 4.90  - Oil side pressure loss, bar N/A 4.05 3.36 3.42 3.45 3.18 2.90 2.70 2.38 2.14 1.93 1.77 System temperatures, oC

 - Cold tank N/A 292 293 293 293 293 294 294 294 294 294 295  - Hot tank N/A 389 388 387 387 386 385 384 383 382 381 380  - Collector field supply 294 294 295 296 298 299 300 301 302 303 304 305  - Preheater Therminol exit 294 295 294 293 292 291 290 290 289 288 287 286 Rankine cycle - Live steam pressure, bar 101.0 97.9 96.6 95.2 93.8 92.5 91.2 89.9 88.4 87.0 85.9 84.4  - Live steam temperature, oC 374 372 370 369 368 366 365 364 362 361 359 358 

 - Reheat steam temperature, oC 374 369 368 367 366 365 364 362 361 360 359 358 

 - Gross output, MWe 55.0 53.1 52.4 51.7 50.9 50.2 49.5 48.7 47.9 47.2 46.5 45.7  - Gross efficiency 0.379 0.375 0.375 0.374 0.374 0.374 0.373 0.373 0.372 0.372 0.372 0.37Steam generator preheater - Exit steam quality 0.00 0.00 0.01 0.01 0.02 0.03 0.04 0.05 0.05 0.06 0.07 0.0 - Mixture exit velocity, m/sec 1.19 1.15 1.14 1.12 1.11 1.10 1.10 1.09 1.07 1.06 1.05 1.04 Unit storage cost, $/kW ht - 2 hours N/A $80.31 $61.72 $52.86 $48.06 $45.29 $43.48 $42.28 $41.44 $40.89 $40.59 $40.2 - 4 hours N/A $51.99 $42.84 $38.52 $36.26 $35.01 $34.26 $33.81 $33.54 $33.42 $33.43 $33.4 - 6 hours N/A $42.43 $36.42 $33.62 $32.20 $31.46 $31.06 $30.86 $30.78 $30.79 $30.91 $31. - 9 hours N/A $39.61 $34.39 $31.97 $30.77 $30.14 $29.83 $29.69 $29.65 $29.70 $29.84 $29. - 12 hours N/A $38.69 $33.89 $31.68 $30.59 $30.03 $29.77 $29.66 $29.66 $29.73 $29.88 $30.

- 15 - October 25, 2002
hermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

EL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

- 16 - October 25, 2002

Table 3 Rankine Cycle, Steam Generator, and Oil-to-Salt Heat Exchanger Part Load Performance 

for 66 bar steam pressure 

TNR

Case 1 Case 2 Case 3 Case 4 Case 5 Case 6 Case 7 Case 8 Case 9 Case 10 Case 11 C

Oil-to-salt heat exchanger - LMTD, oC N/A 2 3 4 5 6 7 8 9 10 11 

 - Heat exchanger area, m2 N/A 94,401 56,463 39,633 30,107 23,959 19,878 16,819 14,663 13,052 11,697 1

 - Salt side pressure loss, bar N/A 5.00 5.09 4.94 4.95 4.96 4.99 4.95 5.04 4.98 5.02  - Oil side pressure loss, bar N/A 3.43 3.52 3.49 3.44 3.47 3.30 3.44 3.02 2.70 2.43 System temperatures, oC

 - Cold tank N/A 271 272 272 272 272 273 273 273 274 274  - Hot tank N/A 389 388 387 387 386 385 384 383 382 381  - Collector field supply 272 273 274 275 277 278 279 280 282 283 284  - Preheater Therminol exit 272 271 271 270 269 268 268 267 266 265 265 Rankine cycle - Live steam pressure, bar 66.5 64.7 64.0 63.1 62.5 61.7 60.8 60.1 59.3 58.5 57.7  - Live steam temperature, oC 374 372 370 369 367 366 364 363 361 360 358 

 - Reheat steam temperature, oC 374 368 367 366 365 364 362 361 360 358 357 

 - Gross output, MWe 55.0 53.3 52.8 52.0 51.4 50.8 50.1 49.4 48.7 48.1 47.4  - Gross efficiency 0.355 0.352 0.351 0.351 0.351 0.350 0.350 0.349 0.349 0.349 0.348 0Steam generator preheater - Exit steam quality 0.00 0.00 0.00 0.00 0.02 0.03 0.03 0.04 0.05 0.06 0.06  - Mixture exit velocity, m/sec 1.13 1.09 1.08 1.06 1.07 1.07 1.05 1.05 1.04 1.04 1.02 Unit storage cost, $/kW ht - 2 hours N/A $79.55 $58.18 $48.68 $43.53 $40.27 $38.20 $36.69 $35.76 $35.07 $34.58 $ - 4 hours N/A $49.52 $38.93 $34.26 $31.80 $30.27 $29.33 $28.68 $28.32 $28.09 $27.95 $ - 6 hours N/A $39.40 $32.40 $29.35 $27.78 $26.82 $26.26 $25.90 $25.73 $25.65 $25.63 $ - 9 hours N/A $36.50 $30.41 $27.76 $26.41 $25.60 $25.14 $24.84 $24.72 $24.67 $24.68 $ - 12 hours N/A $35.18 $29.57 $27.15 $26.20 $25.46 $25.05 $24.80 $24.69 $24.67 $24.69 $
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

F. Capital Cost Estimates 

A series of parametric capital cost estimates were developed for storage systems with approach temperatures 

of 2 °C to 15 °C for the oil-to-salt heat exchanger. The parametric studies identified the approach 

temperature which yielded the lowest unit storage capital cost, in $/kWht. 

Thermal Storage Tanks 

The capital cost of the thermal storage tanks was developed as follows: 

1) The cost of the tank shell was estimated by calculating the weight of the tank, and then applying a unit 

price. The unit price for carbon steel tanks, including material, shop fabrication, shipping, and field 

fabrication, was estimated to be $4.40/kg. Unit costs were derived from earlier central receiver cost studiesx

and Bechtel historical cost information. The tank weight was calculated from the following: 

 - The thickness of the wall at the bottom of the tank was calculated using the height of the tank, the 

density of the inventory fluid, the allowable material stress at the tank operating temperature, and 

conventional formulas for stresses in a hoop. The minimum wall thickness at the top was assumed to be 

6 mm, and the wall thickness was assumed to vary linearly from the bottom of the tank to the top. 

 - The thickness of the floor was assumed to be 8 mm. 

 - The roof was assumed to be a self supporting dome, with a thickness of 6 mm. 

2) The walls and roof were insulated with calcium silicate block insulation, which in turn was covered with 

a corrugated aluminum jacket for weather protection. The thickness of the insulation was assumed to vary 

linearly with the tank design temperature, increasing from a minimum value of 300 mm at a temperature of 

290 oC to a maximum value of 500 mm at a temperature of 565 oC. The unit insulation costs were assumed 

to vary linearly with the thickness, increasing from a minimum value of $160/m2

 at a thickness of 300 mm to 

a maximum value of $235/m2

 at a thickness of 500 mm. 

3) Starting from the bottom and moving up, the tank foundation consisted of a concrete slab, an insulating 

concrete slab, foam glass insulation, insulating firebricks, and a steel slip plate. The perimeter of the 

foundation was somewhat different, consisting of a ring wall of firebricks to support the additional weight of 

the walls and roof. An illustration of the foundation is shown in Figure 5. Costs for each of the elements 

were calculated from the following: 

 - The thickness of the concrete slab was assumed to be 610 mm, and 73 kg of reinforcing steel were 

provided in each cubic meter of concrete. Unit prices for the concrete were estimated to be $85/m3

 and 

1.3 installation hour/m3

, and unit prices for the reinforcing steel were estimated to be $0.80/kg and 

0.022 installation hour/kg. 

- 17 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Concrete Foundation

Thermal Foundation

Foamglass Insulation

Insulating Firebricks

Perimeter

Ringwall

Tank Shell

Cooling

Pipes

Figure 4 Thermal Storage Tank Foundation 

 - The thickness of the insulating concrete slab was assumed to vary linearly with the tank design 

temperature, increasing from a minimum value of 0 mm at a temperature of 290 oC to a maximum value 

of 230 mm at a temperature of 565 oC. Unit prices for the insulating concrete were estimated to be 

$100/m3

 and 1.3 installation hour/m3. 

 - The thickness of the foam glass insulation was assumed to vary inversely with the tank design 

temperature, decreasing from a maximum value of 400 mm at a temperature of 290 oC to a minimum 

value of 300 mm at a temperature of 565 oC. Subcontract unit prices for the insulation were estimated 

to be $356/m3

. 

 - The thickness of the insulating firebrick was assumed to vary linearly with the tank design temperature, 

increasing from a minimum value of 0 mm at a temperature of 290 oC to a maximum value of 165 mm 

at a temperature of 565 oC. Unit prices for the firebricks were estimated to be $1 each. Installation 

costs, without mortar, were estimated to be 0.10 labor hours for each brick. 

 - The height of the perimeter ring wall was assumed to be the sum of the foam glass thickness and the 

insulating firebrick thickness. Unit prices for the firebricks were estimated to be $1 each. Installation 

costs, with mortar, were estimated to be 0.33 labor hours for each brick. 

 - The thickness of the steel slip plate was assumed to be 6 mm. Unit prices for the plate were estimated 

to be $1.30/kg, with 0.022 installation hour/kg. 

 - The foundation cooling pipes were assumed to be 200 mm, Schedule 20 carbon steel pipes located 

1,200 mm on center. Unit prices for the pipe were estimated to be $2.20/kg, with 1.15 installation 

hour/m. 

- 18 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Storage Media 

A delivered unit price for binary nitrate salt was estimated to be $0.43/kg, based on budgetary information 

from the Solar Tres central receiver project. To the material price were added $0.02/kg for the fuel to melt 

the salt, and $0.05/kg for the labor to handle the salt once at the site, for a total installed cost of $0.50/kg. 

Oil-to-Salt Heat Exchangers 

The unit price for the oil-to-salt heat exchanger was estimated to be $146/m2 (see xi). Installation labor costs 

were estimated to be 2.2 hours per metric ton, based on a unit weight of 200 kg/m2

. 

Nitrate Salt Pumps 

Budgetary prices for the nitrate salt pumps heat exchangers were obtained from the Solar Two project and 

the Solar Tres project. A regression analysis of the data provided the following equation for the unit cost of 

cold salt pumps: 

Unit cost, $/kWe = $14,720 (Pump motor power, kWe) – 0.4488

A regression analysis of the data for the hot salt pumps provided the following: 

Unit cost, $/kWe = $5,512 (Pump motor power, kWe) – 0.1845

The installation labor rate was estimated to be 100 hours per pump for motor ratings below 75 kWe, 

300 hours for motor ratings between 75 and 750 kWe, and 500 hours for motor ratings between 750 and 

1,500 kWe. 

Nitrate Salt Piping 

Unit costs for the nitrate salt piping costs were developed on the following basis: 

 • Pipe material costs were estimated to be $2.20/kg for carbon steel, $4.40/kg for ferritic steel, and 

$6.60/kg for stainless steel lines 

 • Unit pipe weights were calculated from the design system pressures, maximum allowable material 

stresses, and conventional hoop stress formulas. 

 • Installation labor rates for material handling, lineup, tack welds, and production welds were developed 

from Bechtel standards 

 • Allowances for hangers, supports, inspection, and testing were developed as factors of the material and 

installation costs 

 • Unit electric heat trace power rating were developed from a differential equation which calculated 

temperature rise rates for the pipe material and thermal insulation. A preheat time of 2.0 hours was 

assumed, with initial and final pipe temperatures of 0 oC and 175 oC, respectively. The number of heat 

- 19 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

trace cables was calculated from a maximum unit power rating of 165 W/m per cable, and 2 spare 

cables were installed with the active cables. Unit heat trace costs were estimated to be $50/m-cable, 

including the cables, control thermocouples, electric power supply to the cables, and installation. 

 • Pipe thermal insulation costs were developed from budgetary price information provided to the Solar 

Two project. Insulation thicknesses were a function of the pipe diameter and fluid temperature, and 

ranged from a minimum of 150 mm to a maximum of 200 mm. 

Summary Investment Cost Plots 

The total storage system unit cost results of the parametric cost analyses are summarized in Table 2 and 

Table 3, and illustrated in Figure 5 and Figure 6. Table 4 shows the thermal storage system cost breakdown 

for storage capacities of 2 to 12 hours for the 100 bar case at an LMTD of 7°C. 

Table 4 Thermal Storage System Costs 

101 Bar Rankine Cycle and 7°C Oil-To-Salt Heat Exchanger LMTD 

- 20 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

$20

$30

$40

$50

$60

$70

2 712

Oil-to-Salt Heat Exchanger LMTD, o

C

Unit Storage Cost, $/kWht

2 Hour

4 Hour

6 Hour

9 Hour

12 Hour

Figure 5 Effect of LMTD on unit cost for several storage capacities, at 100 bar 

$20

$30

$40

$50

$60

$70

2 712

Oil-to-Salt Heat Exchanger LMTD, o

C

Unit Storage Cost, $/kWht

2 Hour

4 Hour

6 Hour

9 Hour

12 Hour

Figure 6 Effect of LMTD on unit cost for several storage capacities, at 66 bar 

- 21 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

G. Annual Performance and Electricity Cost 

NREL adapted the Excelergy model for this study to more accurately model the performance impact of 

operation from thermal storage. For the analysis, the levelized cost of energy was used as the figure of 

merit to compare each design. Figure 7 below shows the results for the 6-hour storage case where seven 

solar field sizes were evaluated with each of the TES heat exchanger sizes. The optimum 6 hour TES 

case has a solar field size of about 460,000 m2

 of solar field with an LMTD of about 7°C. 

6 Hours of TES

0.10

0.11

0.11

0.12

0.12

0.13

0.13

0.14

0 4 812 16

LMTD

LEC $/kWh

 323,360 

 368,480 

 413,600 

 458,720 

 503,840 

 552,720 

 597,840 

Solar Field

Area (m2)

Figure 7 LMTD Optimization for 7 hours of TES, 55 MWe (gross), 101 bar steam cycle 

Figure 8 shows the LEC for each storage volume and each heat exchanger size. The LMTDs that provide the 

minimum LEC are indicated on the graph for each TES capacity. Larger storage sizes are optimized with 

larger heat exchangers (lower LMTDs). Note that the curve is fairly flat and the minimum LEC occurs over a 

fairly wide range of LMTDs. 

- 22 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

100 Bar Turbine Cycle

0.100

0.105

0.110

0.115

0.120

0.125

0.130

NA 2 3 4 5 6 7 8 9 10 11 12 13 14 15

LMTC (C)

LEC ($/kWh)

0

2

4

6

9

12

Hrs TES

Figure 8 TES Optimization for 55 MWe Trough Plant, 101 bar steam cycle 

Also important to note is that the LEC for the plant decreases with capacity. Table 5 compares the cost of 

energy for the plant with the minimum TES cost with the plant with the minimum LEC. The optimum LEC 

usually occurs with a larger TES heat exchanger (lower LMTD). However, the LEC for the optimized design 

is not substantially lower than the lowest cost TES design (0.1% - 0.6%). 

Table 5 TES Design Comparison – Minimum TES Cost Verses Minimum LEC 

Thermal Energy Storage (hours) 2 4 6 9 12 

System with minimum TES cost 

Heat Exchanger LMTD (°C) 14 10 9 9 8 

TES System Cost ($/kWht) 40.20 33.42 30.78 30.78 29.66 

LEC ($/kWh) 0.1183 0.1134 0.1097 0.1079 0.1067

System with minimum LEC 

Heat Exchanger LMTD (°C) 12 8 7 6 6 

TES System Cost ($/kWht) 40.26 33.81 31.06 30.14 30.03 

LEC ($/kWh) 0.1182 0.1133 0.1095 0.1074 0.1061 

Figure 9 and Table 6 show the comparison of cost of energy for 66 and 100 bar turbine steam cycles. In this 

study the 100 bar turbine cycle appears to result in a lower cost of energy than the 66 bar system. This is a 

different result than was observed in the earlier K&A study 3

. Additional comparisons with the earlier work 

is needed to understand this difference. 

- 23 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Steam Cycle Pressure

0.105

0.110

0.115

0.120

0.125

0.130

0 2 4 6 810 12

Hours of TES

LCOE ($/kWh)

66 Bar

100 Bar

Figure 9 Comparison of LEC for 55 MWe plants with 66 and 100 bar turbine steam cycles 

Table 6 TES Design Comparison – Minimum TES Cost Verses Minimum LEC 

Thermal Energy Storage (hours) 2 4 6 9 12 

100 Bar Turbine Cycle 

TES System Cost ($/kWht) 40.26 33.81 31.06 30.14 30.03 

LEC ($/kWh) 0.1182 0.1133 0.1095 0.1074 0.1061 

66 Bar Turbine Cycle 

TES System Cost ($/kWht) 40.26 33.81 31.06 30.14 30.03 

LEC ($/kWh) 0.1203 0.1150 0.1108 0.1086 0.1069 

The SEGS plants allow the turbine to operate up to 115% of design turbine output. This analysis also looked 

at the impact on LEC of allowing the turbine to operate up to 115% of design output, assuming no change in 

power cycle equipment cost. Figure 10 shows the impact on plant LEC as compared to the plant that is 

constrained to a maximum of 100% turbine output. The overall LEC is lower in part because the plant is 

optimized to a larger solar field sizes. 

- 24 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Turbine Max Operation

0.102

0.104

0.106

0.108

0.110

0.112

0.114

0.116

0.118

0.120

0.122

0.124

0 2 4 6 810 12

Hours of TES

LCOE ($/kWh)

115%

100%

Figure 10 Impact on LEC of Turbine Over Design Operation 

Conclusions 

• The performance impact of the TES system must be considered in the overall cost optimization of the 

power plant. The lowest cost TES design does not correspond to the lowest cost of electricity. 

• Although the TES system cost is lower for a 66 bar turbine power cycle, the lower efficiency of the 

66 bar turbine results in a higher cost of solar energy. This is the opposite conclusion to an earlier 

study. 

• For the 50 MWe (net) trough plant, the indirect thermal storage system can help reduce the cost of 

electricity from the plant. 

• In order to develop an optimum solar plant design, it is important to conduct a detailed design 

analysis that accounts of costs, economics, and plant performance in order to determine the optimum 

TES configuration and solar field size. The analysis above was done to minimize the LEC, however 

other economic figures of merit (benefit/cost ratio, net present value or internal rate of return) would 

likely be used for a real project. These would likely result in different optimum TES configurations 

and solar field sizes. 

- 25 - October 25, 2002
Thermal Storage for Parabolic Trough Technology Commercial Plant Design for a 2-Tank 

NREL Contract No. AAA-2-32432-01 Indirect Molten Salt System -- Final Report 

Endnotes 

 

i

 Aringhoff, R., and M., Geyer, “AndaSol- 50MW Solar Plants with 9 Hour Storage for Southern Spain,” Proceeding of 

the 11th SolarPACES International Symposium on Concentrated Solar Power and Chemical Energy Technologies, 

September 4-6, 2002, Zurich, Switzerland. 

ii Hermann, U. “Survey of Thermal Storage for Parabolic Trough Power Plants,” NREL/SR-550-27925. NREL, Golden, 

CO, September 2000. 

iii Engineering Evaluation of a Molten Salt HTF in a Parabolic Trough Solar Field. Task 6 Report. Final System 

Performance and Cost Comparisons. Prepared by Kearney and Associates for NREL, August 2001. 

iv Nexant Inc., “Thermal Storage Oil-to-Salt Heat Exchanger Design and Safety Analysis”, Task Order Authorization 

Number KAF-9-29765-09, San Francisco, California, March 22, 2001. 

v

 Nexant Inc., “USA Trough Initiative: Thermal Storage Oil-to-Salt Heat Exchanger Design and Safety Analysis”, , 

National Renewable Energy Laboratory Task Order Authorization Number KAF-9-29765-09, San Francisco, California, 

March 22, 2001. 

vi Nexant Inc., “USA Trough Initiative: Rankine Cycle, Steam Generator, and Two-Tank Thermal Storage Parametric 

Studies”, , National Renewable Energy Laboratory Task Order Authorization Number KAF-9-29765-09, San Francisco, 

California, September 27, 2002. 

vii Price, H., “Heat exchanger Size Optimization for Indirect 2-Tank Molten-Salt Heat Thermal Storage System for 

Parabolic Trough Plants,” National Renewable Energy Laboratory, Golden, CO, September 27, 2002. 

viii GateCycle Program, Version 5.34, GE Enter Software, Inc. and the Electric Power Research Institute. 

ix “Thermal Storage Oil-to-Salt Heat Exchanger Design and Safety Analysis”, Task Order Authorization Number KAF-9-

29765-09, Nexant Inc. (San Francisco, California), March 22, 2001. 

x

 “Investigation of Thermal Storage and Steam Generator Issues”, Bechtel Corporation (San Francisco, California), Sandia 

National Laboratories Report SAND93-7084, August 1993. 

xi “USA Trough Initiative: Thermal Storage Oil-to-Salt Heat Exchanger Design and Safety Analysis”, Nexant Inc. (San 

Francisco, California), National Renewable Energy Laboratory Task Order Authorization Number KAF-9-29765-09, 

March 22, 2001. 

- 26 - October 25, 2002
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13 May 2002 - 31 December 2004 

5a. CONTRACT NUMBER

DE-AC36-99-GO10337 

5b. GRANT NUMBER

4. TITLE AND SUBTITLE

Thermal Storage Commercial Plant Design Study for a 2-Tank 

Indirect Molten Salt System: Final Report 

5c. PROGRAM ELEMENT NUMBER

5d. PROJECT NUMBER

NREL/SR-550-40166 

5e. TASK NUMBER

CP06.1301 

6. AUTHOR(S)

B. Kelly (Nexant, Inc.) and D. Kearney (Kearney & Associates) 

5f. WORK UNIT NUMBER

7. PERFORMING ORGANIZATION NAME(S) AND ADDRESS(ES)

Nexant, Inc. Kearney & Associates 

101 Second St., 11th Floor PO Box 2568 

San Francisco, CA 94105 Vashon, WA 98070 

8. PERFORMING ORGANIZATION 

REPORT NUMBER

AAA-2-32432-01 

10. SPONSOR/MONITOR'S ACRONYM(S)

NREL 

9. SPONSORING/MONITORING AGENCY NAME(S) AND ADDRESS(ES)

National Renewable Energy Laboratory 

1617 Cole Blvd. 

Golden, CO 80401-3393 11. SPONSORING/MONITORING 

AGENCY REPORT NUMBER

NREL/SR-550-40166 

12. DISTRIBUTION AVAILABILITY STATEMENT

National Technical Information Service 

U.S. Department of Commerce 

5285 Port Royal Road 

Springfield, VA 22161 

13. SUPPLEMENTARY NOTES

NREL Technical Monitor: H. Price 

14. ABSTRACT (Maximum 200 Words)

Subcontract report by Nexant, Inc., and Kearney and Associates regarding a study of a solar parabolic trough 

commercial plant design with 2-tank indirect molten salt thermal storage system. 

15. SUBJECT TERMS

concentrating solar power; solar parabolic trough; solar; solar thermal electricity 

16. SECURITY CLASSIFICATION OF: 19a. NAME OF RESPONSIBLE PERSON 

a. REPORT 

Unclassified 

b. ABSTRACT 

Unclassified

c. THIS PAGE 

Unclassified

17. LIMITATION 

OF ABSTRACT

UL 

18. NUMBER 

OF PAGES 

19b. TELEPONE NUMBER (Include area code)

Standard Form 298 (Rev. 8/98) 

Prescribed by ANSI Std. Z39.18