Performance comparison of two-tank direct and thermocline thermal energy storage systems for 1 MWe class concentrating solar power plants

Energy 81 (2015) 526e536

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Performance comparison of two-tank direct and thermocline thermal energy storage systems for 1 MWe class concentrating solar power plants Daniele Cocco a, *, Fabio Serra b a b

University of Cagliari, Department of Chemical, Mechanical and Materials Engineering, Via Marengo, 2, 09123 Cagliari, Italy Solar Concentration Technologies and Hydrogen from RES Laboratory, Sardegna Ricerche, Z.I. Macchiareddu, 09010 Uta, CA, Italy

a r t i c l e i n f o

a b s t r a c t

Article history: Received 9 April 2014 Received in revised form 23 December 2014 Accepted 27 December 2014 Available online 20 January 2015

This paper compares the performance of medium size CSP (Concentrating Solar Power) plants based on an ORC (Organic Rankine Cycle) power generation unit and using linear Fresnel collectors, thermal oil as heat transfer fluid and two-tank direct and thermocline energy storage systems. The comparative performance analysis was carried out by means of specifically developed simulation models and considering different values of solar multiple and thermal energy storage capacity. The results of the performance assessment demonstrate that if the CSP plant has to be optimized for the highest specific energy production, that is, for the highest solar energy conversion efficiency, two-tank energy storage systems show slightly higher performance than thermocline storage systems. However, the study also demonstrates that thermocline storage systems can be an interesting option to reduce the energy production cost of CSP plants. © 2015 Elsevier Ltd. All rights reserved.

Keywords: Thermal energy storage Thermocline storage CSP plants Linear Fresnel collector

1. Introduction Concentrating Solar Power (CSP) plants are today one of the most interesting options in the field of solar energy technologies. CSP plants use solar collectors to increase the temperature of a Heat Transfer Fluid (HTF) and the high temperature thermal energy produced is converted into mechanical work by a suitable power generation section. Moreover, to offset the intermittence of solar energy and increase power plant dispatchability, CSP plants are usually coupled with a Thermal Energy Storage (TES) section. Today, the current CSP world generating capacity is around 3000 MW and is rapidly increasing. More than 1300 MW of additional CSP capacity is currently under construction and about 10 GW is expected before 2015. Spain is the country with the highest CSP production, thanks to the operation of more than 50 power plants with an installed capacity of more than 2300 MW [1,2]. The preferred choice for current CSP plants is based on largesize power generation units (often in the range of 20e50 MW) mainly due to their higher conversion efficiency and lower specific * Corresponding author. Tel.: þ39 070 6755720; fax: þ39 070 6755717. E-mail address: [email protected] (D. Cocco). http://dx.doi.org/10.1016/j.energy.2014.12.067 0360-5442/© 2015 Elsevier Ltd. All rights reserved.

capital costs. However, the construction of large-size CSP units requires the availability of large areas and noteworthy capital investments (a typical 50 MWe CSP plant requires a total capital investment of about 250e300 MV and the availability of about 150e250 ha of land). For this reason, the construction of mediumsize CSP plants (around 1-5 MWe) may be a more suitable option for countries where large areas can be hard to find (as in Italy, for example). During the design of CSP plants, different options are available for solar field (parabolic trough, linear Fresnel, solar tower and solar dish systems), heat transfer fluid (thermal oil, molten salts, steam, etc.), power generation section (steam Rankine and organic Rankine cycles, Stirling engines, combined cycles, etc.) and thermal energy storage (active, passive, two-tank, thermocline, etc.) [3e5]. For large-size CSP plants, the preferred configurations rely on solar fields based on Parabolic Trough Collectors (PTC) or solar tower systems, thermal oil as HTF and two-tank indirect systems using molten salts as storage medium for the TES section. The power generation section is almost always represented by steam Rankine cycles with superheated steam produced at about 370e380
C and 80e100 bar, high-pressure and low-pressure steam turbines with reheating and 4-6 steam extractions for feed-water heating [6,7].

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Nomenclature A C Cp d DNI E F G h HB HST i I k L m N Nu p Pr q Q R

area (m2) annual operating cost (V/yr) specific heat (J/kg K) particle diameter (m) direct Normal Irradiation (W/m2) energy (J) focal length (m) mass velocity (kg/m2 s) volumetric heat transfer coefficient (W/m3 K) bed height (m) hours of energy storage capacity (hr) annual interest ratio () investment cost (V) thermal conductivity (W/m K) collector length (m) mass flow (kg/s) operating lifetime (yr) Nusselt number () pressure (Pa) Prandtl number () specific thermal losses (W/m2) thermal power (W) lines distance (m)

For medium-size CSP plants the technology options differ from those of large-size plants. In fact, for power outputs in the range of 1 MWe Organic Rankine Cycles (ORC) can be a more viable alternative to steam Rankine cycles for the power generation section. ORC systems use organic fluids with high molar weight instead of steam and require thermal energy inputs with temperature levels starting from 80 to 100
C (low temperature ORC cycles) up to 300e400
C (high temperature ORC cycles). Low temperature ORC cycles often use refrigerants as working fluid while high temperature ORC cycles often use siloxanes, even if different working fluids can be used in both cases [8e10]. Moreover, with such temperature levels, the most suitable option for the heat transfer fluid is today represented by thermal oils (whose maximum bulk temperature is 390e400
C), since molten salts (60% NaNO3 and 40% KNO3) and the direct production of steam in the solar field (Direct Steam Generation, DSG) are specifically developed to raise the maximum HTF temperature, especially for large-size units [11,12]. For the solar field of medium-size CSP plants, linear concentrating collectors appear to be the most tailored choice. Moreover, Linear Fresnel Collectors (LFC) may be a viable alternative to PTC, especially if the land requirement is a key feature. In comparison to PTC, LFCs have a simpler design, use less expensive mirrors and tracking systems, show lower land requirements and lower capital costs. On the other hand, the optical efficiency of LFC is lower than that of PTC [13e16]. With reference to the TES section, thermal energy can be stored as sensible heat, latent heat or chemical energy. Latent heat and chemical energy are considered the most promising technologies but sensible heat storage is the simplest method to store thermal energy and some technological and economic aspects make it superior, especially in the case of medium-size CSP systems [17e23]. In particular, the most mature solution for the TES section is today represented by two-tank direct systems, composed of a lowtemperature and a high-temperature tank, where thermal oils are used as storage medium. To reduce the cost of the TES section, the two-tank configuration can be substituted by a thermocline system,

Ra Re SM t T U V W

a

ε

h q r

527

Rayleigh number () Reynolds number () solar multiple () time (s) temperature (K) global heat transfer coefficient (W/m2 K) volume (m3) collector width (m) convective heat transfer coefficient (W/m2 K) bed void fraction () efficiency () incidence angle (
) density (kg/m3)

Acronyms CSP concentrating solar power DSG direct steam generation HTF heat transfer fluid IAM incidence angle modifier LCOE levelized cost of energy LFC linear fresnel collector ORC organic Rankine cycle PTC parabolic trough collector TCI total capital investment TES thermal energy storage

based on a single-tank packed bed containing a low-cost filling material [24e28]. In thermocline systems the hot HTF is pumped into the top of the tank, flows downwards and gradually heats the filling material. During the charging phase the high temperature zone is separated from the low temperature zone by a thermal gradient that moves downwards in the tank. During the discharging phase the cold HTF is pumped into the bottom of the tank so that the thermal gradient moves upwards. The use of low-cost filling materials in a single-tank system reduces the cost of the TES section and the volume of thermal oil required [24]. However, owing to the presence of the temperature gradient inside the tank, thermocline systems are less efficient than two tank systems because the useful thermal energy recovered during the discharging phase is lower than that supplied during the charging phase [28]. Recent studies on thermocline TES systems mainly focus on large-size CSP plants and on the use of molten salts [29e32]. However, for medium-size CSP units, the use of thermal oil can be a more interesting choice due to the lower operating temperatures. A thermal oil thermocline TES system was used from 1982 to 1987 by the Solar One CSP plant [33]. More recently, a similar TES systems was proposed for the 1 MW Saguaro CSP plant [34]. Studies on thermocline systems using thermal oil were also presented in Refs. [35e38]. In the field of medium-size units, a comparative performance analysis of CSP plants based on parabolic trough and linear Fresnel collectors was carried out in a previous paper [39]. The performance of the CSP plants were evaluated on the basis of a 1 MW ORC unit and with the use of thermal oil as heat transfer fluid and as storage medium in a two-tank direct thermal storage system. The results of the performance assessment demonstrate that CSP plants based on linear Fresnel collectors lead to higher values of electrical energy production per unit area of occupied land while parabolic troughs gives better values of energy production per unit area of solar collector owing to their better optical efficiency.

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For this reason, this paper reports on a comparative performance analysis of 1 MWe class CSP plants based on linear Fresnel collectors, thermal oil as heat transfer fluid, an ORC module and using two-tank direct and thermocline TES systems. The comparative study was carried out by considering different values for solar multiple and thermal energy storage capacity, which are two of the most important design parameters of CSP power plants. In particular, the study aims to evaluate the influence of the latter design parameters on specific energy production and energy production cost. 2. CSP power plant configuration Fig. 1 shows a simplified diagram of the CSP plant considered in this paper. The CSP plant includes three main sections: the solar field, the power block and the thermal energy storage section. The solar field is based on several lines of linear Fresnel collectors connected in parallel to achieve the required thermal oil mass flow and therefore the required thermal power output. As shown by Fig. 2, each collector line includes several collector modules connected in series and each collector module is composed of several rows of flat mirrors whose slope continuously changes to follow the sun's position. The mirrors are mounted on a fixed steel structure placed near the ground (about 0.5e1.0 m above the ground) and concentrate solar radiation onto a fixed receiver installed several meters above the mirror plane. The receiver includes a secondary reflector that redirects the incoming solar rays towards the evacuated receiver tube. The solar collector lines are aligned along the NortheSouth direction and are equipped with a single-axis tracking system to follow the sun's path. The power block is based on an ORC unit, where thermal energy is converted to electrical energy by using an organic fluid (a siliconic oil in this case) that follows a regenerated Rankine cycle. As shown in Fig. 1, the thermal energy produced by the solar field is used in the ORC unit to heat and vaporize the organic fluid. The produced organic vapor expands in the turbine, is cooled in the regenerator and condensed. After the condenser, the organic fluid is compressed by the feeding pump and then preheated in the regenerator. The condensing heat is removed by a cooling tower. Fig. 3 shows a simplified scheme of the two TES solutions compared here. In the two-tank direct system (left side of Fig. 3), the thermal oil from the low-temperature storage tank (the cold tank) flows through the solar field, where it is heated from TC to TH, sent to the high-temperature storage tank (the hot tank) and subsequently pumped to the power block, where it is cooled and sent back to the cold storage tank. In general, the mass flow of thermal oil produced by the solar field differs from that required by the power block. For this reason, the mass of thermal oil stored in the

Fig. 2. Simplified scheme of the linear Fresnel collectors.

hot tank increases (and therefore the mass stored in the cold tank decreases) during periods of high solar energy availability, where the thermal power required by the ORC unit is lower than that produced by the solar field. The stored oil is subsequently used to operate the ORC unit during periods of low or no solar energy availability. The thermocline system (right side of Fig. 3) is based on a singletank packed bed containing a filling material (pebble rock, in this case). During periods of high solar availability, the excess mass flow of thermal oil at temperature TH produced by the solar field is pumped into the top of the tank and gradually heats the filling material. The high temperature zone of the packed bed is separated by the low temperature zone by a thermal gradient that moves downwards the tank during the charging phase. The cold oil exiting from the thermocline tank mixes with the oil exiting from the ORC unit and is pumped to the solar field. The stored thermal energy is subsequently recovered by pumping the cold oil from the ORC unit into the bottom of the tank. Therefore, during the discharging phase the thermal oil flows upwards, gradually cools the filling material and the thermal gradient moves upwards. For both CSP configurations, the piping system includes the two circulating thermal oil pumps, the cold and hot header pipe (one for distributing the cold thermal oil throughout the collector loops and the other for collecting the hot oil), the main pipes, valves, fittings and pressure, temperature and flow meters.

Fig. 1. Process flow diagram of the CSP power plant.

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Fig. 3. Simplified schemes of the two-tank direct and thermocline TES systems.

3. System modeling and assumptions The performance of the two CSP solutions was evaluated on a yearly basis by means of a specifically developed simulation model under the Matlab environment. In particular, the simulation model evaluates the performance of the CSP plant as a function of solar radiation and solar position for given values of the main geometrical and technical characteristics of solar field, thermal energy storage system and power generation section. 3.1. Meteorological data The comparative study was carried out using a data set for a typical meteorological year obtained from the Meteonorm software [40] for the site of Cagliari (39
130 2500 N, 9
070 2000 E), in the south of Sardinia (Italy). The meteorological data set includes DNI (Direct Normal Irradiation), solar azimuth and elevation, ambient temperature (dry and wet bulb), relative humidity and wind velocity. Fig. 4 gives the annual frequency distribution of the DNI for the site of Cagliari and Table 1 summarizes the most important meteorological data and the design conditions assumed for the analysis. 3.2. Solar field The performance of the solar field was calculated on an hourly basis as a function of solar radiation and solar position for given

values of the main geometrical and technical characteristics of the solar collectors, as well as for assigned thermodynamic properties of the heat transfer fluid. In particular, the thermal power QFLD transferred to the thermal oil is given by the difference between receiver available power QRCV and thermal losses QTHR:

QFLD ¼ mO $CpO $ðTH  TC Þ ¼ QRCV  QTHR

(1)

where mO and CpO are the mass flow and the specific heat of thermal oil and TC and TH the inlet and outlet oil temperatures. Thermal power QRCV is evaluated here by means of the following equation:

QRCV ¼ AC $DNI$hOPT;R $IAML $IAMT $hEND $hCLN

(2)

where AC is the collecting area, hOPT,R is the reference optical efficiency, IAML and IAMT are the longitudinal and transversal components of the Incidence Angle Modifier (IAM), hEND is the end-loss optical efficiency and hCLN is the mirror and glass tube surface cleanliness factor. Reference optical efficiency hOPT,R depends on mirror reflectivity, glass tube transmissivity, absorptivity of the selective coating of the receiver tube, imperfections in the collector mirrors, tracking errors, shading of receiver supports on mirrors, etc., and is commonly evaluated for an incidence angle q (that is, the angle between solar rays and the direction normal to the collector surface) equal to 0
. The reference optical efficiency is multiplied by the Incidence Angle Modifier to take into account the effect of incidence angle q on the optical properties of the different materials (mirrors, glass tube, selective coating) and the relative shading of mirror surfaces. Fig. 5 shows the two IAM components considered in this study for linear Fresnel collectors in function of the longitudinal and transversal components qL and qT of the incidence angle. These two components can be calculated in function of the solar position, which is completely defined by the azimuth and elevation angles [39].

Table 1 Meteorological data for the site of Cagliari and design assumptions.

Fig. 4. Frequency distribution of the DNI for the site of Cagliari.

Annual direct normal solar radiation Average ambient temperature Average wind velocity Design DNI Design elevation/azimuth angles Design dry/wet bulb temperatures

1720 kWh/m2y 17.2
C 3.96 m/s 800 W/m2 74.2
/0.0
32.0/22.0
C

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design performance of the ORC module was evaluated with reference to data available for commercial units [41]. In particular, Table 3 gives the most important operating parameters of the ORC unit while Fig. 6 shows its off-design efficiency as a function of thermal load (cooling water at 25
C) and wet bulb temperature (thermal power input of 4043 kW and 3
C of temperature difference between cooling water and wet bulb temperature). In this study, due to the presence of the thermal energy storage section, the oil mass flow at the inlet of the ORC unit is constant and equal to the design value. Therefore, the inlet thermal load of the ORC unit varies only if the TES supplies oil at lower temperatures (in particular with thermocline systems). As better explained in the following, in this study, the lowest oil temperature supplied by the TES section is about 285
C while the design inlet temperature of the ORC unit is 305
C. The resulting lowest value of the inlet/outlet oil temperature difference of the ORC unit (81
C) is about 80% of the corresponding design value (101
C) and therefore the lowest ORC input load (3234 kW) is about 80% of the corresponding design value (4043 kW). Fig. 6 demonstrates that with an input thermal load ranging from 80% to 100% the decrease of the ORC efficiency is lower than 1 percentage point.

Fig. 5. Longitudinal and transversal IAM components.

With longitudinal components qL higher than 0
, the useful mirror area of linear collectors is reduced by the geometrical endlosses, which depend on collector length L and focal length F. Overall, these losses are taken into account by means of the endloss optical efficiency:

F hEND ¼ 1  $tan qL L

(3)

Thermal losses were evaluated by the sum of receiver thermal losses and piping thermal losses:

i  i h h QTHR ¼ qrec þ qpip $AC ¼ a1 $DT þ a2 $DT 2 þ qpip $AC

(4)

where DT is the difference between the average oil temperature in the receiver tube and ambient temperature. Table 2 shows the main geometrical and performance parameters of the solar field used in the following comparative study [13e17,39]. The main power consumption of the solar field is due to the collector tracking system and oil circulating pumps. The first term was assumed equal to 1.5 W/m2 of collecting area while the second was evaluated by imposing a fluid velocity of about 1 m/s and a pump efficiency of 75%. 3.3. Power generation section The power block considered in this comparative performance analysis is based on an ORC unit with a power output of 1 MWe integrated with a closed circuit cooling tower. Design and off-

3.4. Thermal energy storage section For each operating hour of the year, Eqs. (1)e(4) allow calculation of the mass flow of thermal oil produced by the solar field. As mentioned, the thermal energy produced by the solar field can be directly used by the ORC unit or stored in the TES section. The energy management strategy considered in this paper assumes that the ORC unit is operated only at its reference load. Moreover, to avoid and excessive number of starts and stops of the ORC unit and therefore to favor its longer operation at full power the energy management strategy assumes that the hot oil is pumped to the power block only if the thermal energy stored in the hot tank is sufficient to operate the ORC unit for at least 2 h. Therefore, at the beginning of the day the oil is generally sent to the TES section. Subsequently, if the thermal energy produced by the solar field is higher than that required by the ORC unit, the latter starts and the excess of thermal energy is stored in the TES section. Obviously, if the TES section is completely full, when the thermal oil produced by the solar field exceeds that required by the ORC module a share of the available solar energy is lost (unused solar energy). As already mentioned, the two TES systems considered here are based on a two-tank direct system and a single-tank packed bed containing a low-cost filling material (thermocline system). The size of each storage tank, the mass and volume of thermal oil and the filling material were evaluated in function of the required energy storage capacity EST and the thermodynamic properties of both thermal oil and filling material. The energy storage capacity can also be expressed in terms of equivalent hours of thermal energy supply HST (the storage energy capacity is therefore the product of ORC thermal power input and equivalent hours of storage).

Table 2 Main geometrical and performance parameters of the solar field. Collector length L Collector width W Collector area AC Focal length F Lines distance R Reference optical efficiency hOPT,R Cleanliness efficiency hCLN Oil inlet (TC)/outlet (TH) temperature a1 coefficient a2 coefficient Specific piping losses qpip

150 m 16.56 m 1712.0 m2 7.40 m 4.00 m 0.67 0.98 204/305
C 0.056 W/m2 K 0.000213 W/m2 K2 5 W/m2

Table 3 Design parameters of the ORC unit. Gross power output Thermal power input Oil inlet (TH)/outlet (TC) temperature Oil mass flow Condenser power output Cooling water inlet/outlet temperature Gross electrical efficiency ORC internal consumption Cooling tower electrical consumption

1000 kW 4043 kW 305/204
C 16.3 kg/s 3040 kW 25/35
C 24.7% 3.6% of gross power 0.6% of cond. power

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between the layer and the environment. Inside each layer the model assumes constant characteristics of both solid and oil along the radial direction. The properties of the oil varies along the vertical axis in function of the corresponding temperature [45], while the physical properties of the solid material are kept constant (the density is 2750 kg/m3 and the specific heat is 900 J/kg K). The main heat transfer mechanism considered by the model is forced convection. Thermal conduction is considered only for the oil because the contact surface of solid particles is very small, while thermal radiation is neglected. The convective heat transfer coefficient hV between solid particles and oil is calculated by means of the following equation:

hV ¼

6ð1  εÞ a d

(8)

The heat transfer coefficient a can be calculated according to the relation:

a¼

Fig. 6. Off-design efficiency of the ORC unit.

The required storage volume of each tank was calculated by means of the following equation:

V¼

EST ½rS CpS ð1  εÞ þ rO CpO εðTH  TC Þ

vTS hV ¼ ðT  TS Þ vt rS CpS ð1  εÞ O

(9)

where the Nusselt number is a function of Reynolds (based on particle diameter) and Prandtl numbers [46]:

  0:33 Nu ¼ 2 þ 1:1 Re0:6 d Pr

(10)

(5)

where rS and rO are the densities of the solid filling material and thermal oil, CpS and CpO are the specific heats of solid material and thermal oil and ε is the bed void fraction. Obviously, for the twotank direct system the bed void fraction to be inserted in Eq. (5) is equal to 1. The mass of thermal oil and filling material was calculated starting from V and from their thermodynamic and physical properties while the design volume of each storage tank was assumed 10% higher than the storage tank volume V. For the two-tank direct system, the storage inefficiencies are caused only by thermal energy losses. The latter were evaluated on an hourly basis by assuming a temperature of the metal surface equal to the oil temperature, heat conduction through an insulation layer of 0.3 m of mineral wool (thermal conductivity equal to 0.036 W/m K) and natural convection between the tank external surface and ambient air. The operation of the thermocline storage system requires the availability of a proper simulation model [28]. In particular, a twophase one-dimensional model based on the work originally developed by Schumann [42] allows prediction of the temperature time evolution along the tank and the corresponding stored energy. The temperatures of both solid material (TS) and thermal oil (TO) are evaluated along the flow direction z within the layers resulting from the chosen spatial step of the tank, according to the following equations [42e44]:

vTO mO vTO hV kO v2 TO þ ¼ ðTS  TO Þ þ vt AT rO ε vz rO CpO ε rO CpO ε vz2 hw ðT  Tenv Þ  rO CpO ε O

Nu$k d

The pressure drop along the bed is evaluated by using the Ergun equation [47,48]:



Dp rO G2



d HB



QW ¼ U$AW $DTW

where AT is the cross sectional area of the tank, kO the effective thermal conductivity of the oil, hV the volumetric convective heat transfer coefficient and hw the overall heat transfer coefficient

¼ 1:75 þ 150

ð1  εÞ Red

(11)

(12)

where U is the global heat transfer coefficient between oil and external air, AW is the walls area and DT W is the difference between the internal oil temperature and air temperature. For the thermocline tank the global heat transfer coefficient is calculated by considering the conductive heat resistance of the wall insulation and the internal and external convective heat transfer coefficients. For the conventional oil tanks the global heat transfer coefficient is calculated by considering the conductive heat resistance of the wall insulation and only the external convective heat transfer coefficient. In both cases, the conductive heat resistance of the metal walls is neglected. In particular, the internal heat transfer coefficient for the thermocline tank is assumed equal to 0.8 times the heat transfer coefficient a [44]. The external heat transfer coefficient is calculated in function of the Rayleigh number Ra by means of the following relation: 1=3

(7)



Finally, the thermal energy losses QW through the tank walls are evaluated with the following equation:

NuHB ¼ 0:1RaHB (6)

ε3 1ε

(13)

According to the energy management strategy considered in this paper, during hours of DNI availability the portion of the hot oil produced by the solar field exceeding that required by the ORC module is pumped to the thermocline TES section. The thermal energy stored in the thermocline tank is recovered and sent to the ORC unit during hours of low DNI availability. However, for thermocline systems, in addition to the thermal energy losses, a storage energy inefficiency is introduced by the partial utilization of the available storage volume. In fact, in thermocline tanks the useful thermal energy recovered during the discharging phase is lower

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than that supplied during the charging phase due to the presence of the temperature gradient inside the tank. Fig. 7 shows a typical temperature profile along the vertical axis of the thermocline tank considered here. In particular, the dashed lines describe the temperature profile at the end of the charging phase (right) and discharging phase (left) if the extreme temperatures are kept constant (during the charging phase the oil outlet temperature cannot be higher than TC and during the discharging phase the oil outlet temperature cannot be lower than TH). In this case, Fig. 7 demonstrates that the available storage volume is only partially used (the area under each curve is proportional to the stored energy). The continuous lines give the temperature profiles achieved by introducing a temperature threshold between oil outlet temperature and TC (charging phase) and between TH and oil outlet temperature (discharging phase). Fig. 7 demonstrates that the stored energy increases with increasing values of the admitted temperature threshold (20% of the overall temperature difference in Fig. 7). Obviously, the overall available storage volume can be completely exploited only if the oil outlet temperature equals TH at the end of the charging phase. Otherwise, the stored thermal energy can be completely recovered only if the oil outlet temperature equals TC at the end of the discharging phase. The foregoing considerations show that while a two-tank direct system is able to fully exploit its energy storage volume and completely recover the stored thermal energy at the higher temperature level, a thermocline system is not able to do so. In particular, a thermocline TES system uses only a portion of its energy storage volume and a portion of the stored thermal energy can be recovered only at a temperature lower than TH. Therefore, as shown in Fig. 8, for a thermocline TES system the thermal energy recovered during the discharging phase can be compared to the storage energy capacity to give the charge/discharge cycle efficiency. Fig. 8 refers to a single charge/discharge cycle and demonstrates that the cycle efficiency increases with the admitted temperature threshold. However, an increase in the temperature threshold decreases the temperature of the oil produced during the last stage of the discharging phase with a corresponding decrease in ORC efficiency (the decrease of the oil temperature leads to a lower ORC thermal input and therefore, as shown by Fig. 6, a lower efficiency). For the temperature threshold considered in this study (20% of the

Fig. 7. Typical temperature profile inside the thermocline system.

Fig. 8. Charge/discharge cycle efficiency for a thermocline storage system.

maximum temperature range) the cycle efficiency of the thermocline tank is about 75%. As mentioned, except for the heat losses, a two-tank direct system is able to fully exploit its energy storage capacity and therefore its cycle efficiency is virtually one (the calculated annual heat losses are less than 1% of the energy stored). On the average, the oil temperature decrease caused by heat losses is generally less than 5e10
C, so that the ORC efficiency is almost constant.

4. Energy performance of CSP plants For a given nominal power output of the ORC unit, the extension of the solar field and the capacity of the TES section mainly depend on two important design parameters: the solar multiple and energy storage capacity. The solar multiple (SM) is the ratio between the thermal power produced by the solar field at design conditions and the thermal power required by the power block at nominal conditions. CSP plants are usually designed for SM values above 1 to allow operation of the power block at nominal conditions even during low DNI periods. With SM values higher than 1 the annual energy production of the CSP plant benefits from the adoption of a suitable TES section. In particular, the latter is usually adopted by CSP plants with SM values in the range of 1.5e2.5. By increasing the solar multiple the collecting area increases and therefore the land area required by the solar field must increase. Land area depends on the number, width and distance of the collector lines, the free space around the collectors (10 m in this study) and the area required by the power block and the TES section (2500 m2 in this study). Fig. 9 shows the collecting area and the land requirement for the CSP plant studied here in function of SM. Plot labels also give the corresponding number of collector rows and the thermal power output of the solar field at design conditions. As mentioned, the energy storage capacity of a CSP plant is usually expressed in terms of equivalent hours of thermal energy supply HST. In particular, for the ORC unit considered here (4043 kW of thermal power input), 1 h of storage capacity (which is 4043 kWh of stored energy) requires, for the two-tank TES system, about 60 t of thermal oil and two tanks with a net volume of about 75 m3 (3.6 m in diameter and 7.2 m in height each, for example). For the thermocline TES system, the same energy storage capacity requires about 16 t of thermal oil, 120 t of pebble rock and a single tank with a net volume of about 62 m3 (3.4 m in diameter and 6.8 m in height, for example).

D. Cocco, F. Serra / Energy 81 (2015) 526e536

Fig. 9. Collecting and land area for the CSP plant in function of the solar multiple.

Fig. 10 shows the net electrical production on a yearly basis for the two CSP plants based on the two-tank and thermocline TES systems, in function of both solar multiple and equivalent hours of thermal energy storage. For both solutions, annual energy production increases with SM and HST even though the presence of the TES section improves net energy production only for SM values above 1.4e1.6. For high values of SM, Fig. 10 shows that the use of the two-tank solution leads to higher energy yields in comparison to the thermocline one owing to its better efficiency. In fact, due to the presence of the thermal gradient, the thermocline system does not fully exploit its energy storage volume and recovers a portion of the stored thermal energy at temperatures lower than TH. However, with low SM values and high storage capacities (HST ¼ 12 h), the energy production of the CSP plants based on the thermocline TES section is only slightly lower than that based on the two-tank system owing to the smaller external surface and therefore the lower thermal energy losses.

Fig. 10. Annual energy production in function of solar multiple and energy storage capacity.

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As mentioned, in many countries (and in Italy in particular) the large areas required by CSP plants are difficult to find and therefore an important performance parameter of such power plants is given by the specific energy production per unit area of solar field. For this reason, Fig. 11 shows the annual energy production per unit area of collectors (a) and per unit area of land occupied by the solar field (b) of the two CSP configurations studied here in function of both SM and HST. Since the land area of the solar field is about twice that of collectors, the annual energy production per unit area of solar field land shows the same behavior with half values. Moreover, as annual solar energy available per m2 of collector area (1720 kWh/m2y for the site considered here) does not depend on SM and HST, energy production per unit area of collectors is proportional to the average solar conversion efficiency of the CSP plant. Fig. 11 demonstrates the importance of a proper trade-off between solar multiple and storage capacity to maximize the specific energy production and therefore the average conversion efficiency. In fact, the increase in SM increases the collecting area, the land area and the annual thermal energy produced by the solar field. However, for a given nominal power output of the ORC unit, the increase in SM increases the share of thermal energy that cannot be directly used for power generation. Therefore, the effective utilization of this surplus energy for power generation requires the increase in energy storage capacity. Fig. 11 also shows that with high SM and low HST values, the two-tank option leads to higher specific energy productions because the TES section is more frequently used along the year. On the contrary, if the TES system is used only for short time periods and with low states of charge (low SM and high HST values), the higher thermal energy losses of the two-tank system (two tanks with a corresponding higher external area) reduces the difference between the specific energy productions of the two TES solutions. Overall, Fig. 11(a) demonstrates that specific energy production (and therefore solar conversion efficiency) can be optimized for suitable values of both SM and HST. In particular, the highest specific energy production is about 140 kWh/y per m2 of aperture area for both the two-tank and the thermocline TES systems. The corresponding value of the solar conversion efficiency is about 8.14%. If land availability is the most limiting factor for the construction of the CSP power plant, solar multiple and energy storage capacity can be set to optimize the net energy production per unit area of occupied land. In this case, Fig. 11(b) shows that the highest specific energy production per land area is about 60 kWh/ ym2 for the two-tank system (achieved for SM ¼ 2.3 and HST ¼ 12 h), and about 59 kWh/ym2 for the thermocline system (achieved for SM ¼ 2.1 and HST ¼ 12 h). Therefore, optimization for the occupied land area requires higher SM values than optimization for the collecting area. To show the different utilization of the available storage volume of the two TES options, Fig. 12 illustrates the frequency distribution of their state-of-charge (that is, the ratio of stored energy to storage capacity) during the year for a CSP plant designed with 12 collector lines (SM ¼ 2.4) and 4 h of energy storage capacity. In Fig. 12, each bar gives the annual hours during which the TES system shows a state-of-charge between the corresponding labels of the horizontal axis. Fig. 12 clearly demonstrates the less efficient utilization of the available storage capacity of the thermocline system. As shown in Fig. 12, the state of charge of this system roughly fluctuates from 20% to about 85% while that of the two-tank system varies from 0 to 100%. The thermocline system is at its lowest state-of-charge (around 20%) for about 4000 h/yr and at its highest state-ofcharge (around 85%) for about 1100 h/yr. The two-tank system is almost empty (state-of-charge between 0 and 5%) for about 3880 h/ yr while it is almost full (state-of-charge between 95% and 100%)

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Fig. 11. Specific energy production per collecting area (a) and land area (b) of the two CSP power plants in function of solar multiple and energy storage capacity.

for about 780 h/yr. Similar behaviors of the TES state-of-charge can be found for different values of SM and HST. 5. Preliminary economic analysis Thermal energy storage with thermocline systems is expected to reduce the overall cost of the TES section due to the use of a single tank instead of two tanks and the partial substitution of the thermal oil with a low-cost filling material. For thermocline systems, Kolb [24] estimated a capital cost reduction of about 33% with respect to conventional two-tank direct systems. For this reason a preliminary economic analysis to evaluate the Levelized Cost of Electricity (LCOE) of CSP plants with two-tank direct and thermocline TES systems was carried out. The LCOE was calculated here according to the simplified methodology proposed by the International Energy Agency [49], by means of the following equation:

LCOE ¼

TCI þ

 PN  k k¼1 Ik þ CO&M;k $ð1 þ iÞ PN k k¼1 ðEk Þ$ð1 þ iÞ

(14)

where TCI is the Total Capital Investment, Ik is the additional investment cost in year “k”, CO&M,k are the operation and maintenance costs in year “k”, Ek is the net electricity production in year “k”, i is the annual interest rate and N the operating lifetime. For simplicity, TCI has been concentrated at the beginning of the operating lifetime period and the investment costs in the following

years have been neglected. Moreover, annual operation and maintenance costs and annual electricity production have been assumed here in as constant during the overall operating lifetime. Total capital investment was estimated on the basis of published information on purchased equipment costs of the main CSP plant components and by adding the balance of plant and engineering costs [13e16,19,50]. Table 4 shows the main assumptions used for the economic analysis. With reference to the TES section, data of Table 4 apply to the two-tank direct system. Due to the lack of reliable data, the cost for the thermocline TES section was assumed equal to 70% of the cost of the corresponding two-tank TES section. Fig. 13 gives the LCOE for both TES solutions in function of the energy storage capacity for a solar field based on 8 (SM ¼ 1.6), 12 (SM ¼ 2.4) and 16 collector lines (SM ¼ 3.2). Like Figs. 11 and 13 demonstrates the importance of a proper trade-off between solar multiple and thermal storage capacity to minimize energy production cost. In fact, TCI linearly increases with the extension of the solar field (and therefore SM) and with the increase in storage capacity. However, as shown in Fig. 10, the contribution to annual net energy production given by the increase in both solar field area and storage capacity becomes less and less important for increasing values of SM and HST. For this reason, Fig. 13 demonstrates that for a given solar multiple, the LCOE can be minimized for a proper value of thermal energy storage capacity. In particular, in the field of SM and HST considered here, the lowest LCOE (420 V/MWh) is achieved for the CSP plant based on the thermocline option by considering 8 collector lines and about 5 h of storage capacity or 12 collector lines and 8e12 h of storage capacity. The lowest LCOE for the two-tank solution (about 430 V/MWh) is achieved by considering 8 collector lines and 4 h of storage capacity or 12 collector lines and 8 h of storage capacity. Table 4 Assumptions for the economic analysis of CSP systems.

Fig. 12. State-of-charge of the two TES systems (12 collecting lines, HST 4 h).

Solar field specific cost ORC specific cost Tank specific cost Thermal oil cost Piping specific cost Land cost Balance of Plant cost Engineering cost Insurance annual cost O&M annual cost Annual interest rate Operating lifetime

200 V/m2 of collector area 1000 V/kWe of nominal power output 625 V/m3 of storage volume 2.5 V/kg 30 V/m2 of collector area 10 V/m2 250 V/kWe of nominal power output 20% of equipment and BoP cost 1.0% of TCI 1.5% of TCI 7% 25 years

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aperture area (achieved for a storage capacity of about 8 h and SM ¼ 1.6 or for a storage capacity of about 12 h and SM ¼ 2.0) and 60 kWh/y per m2 of occupied land area (achieved for SM ¼ 2.3 and HST ¼ 12 h). Optimized CSP solutions based on thermocline TES systems show only slightly lower values of specific energy production. However, the study also demonstrates that optimization of the CSP power plant to achieve the best conversion efficiency does not automatically lead to the lowest energy production cost. In fact, the preliminary economic analysis shows that thermocline TES systems allows to slightly reduce the energy production costs with respect to two-tank systems (420 V/MWh vs 430 V/MWh for optimized solutions). Obviously, the results of the energy and economic comparative analysis presented in this paper are influenced by the energy management strategy adopted for the TES section as well as on the availability of detailed cost data. For this reason, a more in-depth study of these issues is currently in progress. Fig. 13. LCOE of the two CSP power plants in function of storage capacity and number of collector lines.

Moreover, comparison of Figs. 11 and 13 shows that optimization of the CSP power plant to achieve the best conversion efficiency (that is, the higher specific energy production) does not automatically lead to the lowest energy production cost. In particular, Figs. 11 and 13 demonstrate that the SM and HST design parameters to achieve the lowest energy production cost are lower than those required to achieve the highest specific energy production. As shown by Fig. 11, the highest specific energy production can be achieved with a storage capacity of about 8 h and SM ¼ 1.6 or with a storage capacity of about 12 h and SM ¼ 2.0 for both TES options. With reference to these SM and HST values, Fig. 13 shows LCOE values higher than the lowest ones. Overall, two-tank direct storage systems allow to achieve a slightly higher specific energy production but thermocline TES systems allow to minimize energy production costs. It should be observed that a more accurate economic comparison between two-tank and thermocline storage systems requires more reliable data about their capital and operating costs. Moreover, higher energy productions and lower LCOE can be achieved with reference to sites with annual solar energy availability higher than that considered in this study. For example, the site of Daggett (California, US), often used as a reference site, is characterized by an annual DNI availability of about 2724 kWh/m2, which is about 58% higher than that of Cagliari. Therefore, the annual electrical production of a CSP plant located in Daggett should be about 50e60% higher than that of the same plant located in Cagliari. 6. Conclusions This paper compares the performance of concentrating solar power plants using an ORC power generation unit, linear Fresnel collectors, thermal oil as heat transfer fluid and two-tank direct and thermocline energy storage systems. The comparative study aims to evaluate the influence of TES configuration, solar multiple and energy storage capacity on specific energy production and the energy production cost of 1 MWe class CSP plants. The results of the performance assessment demonstrate that two-tank direct storage systems allow to achieve a slightly higher specific energy production and that thermocline TES systems can be an interesting option to reduce energy production costs. In particular, the highest specific energy production for CSP power plants with two-tank TES systems is about 140 kWh/y per m2 of

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