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Energy Procedia 75 (2015) 394 - 402

The 7th International Conference on Applied Energy - ICAE2015

Thermal and thermodynamic performance of a parabolic trough receiver with Syltherm800-Al203 nanofluid as the heat

transfer fluid

Aggrey Mwesigyea?*9 Zhongjie Huana

a Department of Mechanical Engineering, Tshwane University of Technology, Private Bag X680, Pretoria, 0001, South Africa

Abstract

With the recent advances in technology, the use of higher concentration ratios has become feasible and shown potential to reduce the cost of electricity from parabolic trough collector systems. With this, efficient and improved heat transfer performance of parabolic trough receivers is becoming essential. Nanofluids are known to increase the heat transfer capabilities to levels higher than ordinary heat transfer fluids. In this work, the thermal and thermodynamic performance of a parabolic trough receiver using Syltherm800-Al203 nanofluid as the heat transfer fluid is presented. Fluid inlet temperatures in the range 350 K to 650 K, nanoparticle volume fractions in the range 0 - 8% and Reynolds numbers in the range 3 560 to 1 334 000 depending on the inlet temperature were considered. Results from this study show improved thermal and thermodynamic performance of a parabolic trough receiver with the use of nanofluids. The thermal efficiency increases by up to 8% for the range of parameters considered. Results further show that there is a Reynolds number beyond which the use of nanofluids becomes thermodynamically undesirable at a given inlet temperature.

© 2015 The Authors. Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license

(http://creativecommons.Org/licenses/by-nc-nd/4.0/).

Peer-review under responsibility of Applied Energy Innovation Institute

Keywords: Heat transfer irreversibilities; Nanofluids; Parabolic trough receiver; Temperature gradients; Volume fraction

1. Introduction

The parabolic trough technology is the most commercially and technically developed concentrated solar power technology. Several concentrated solar power plants have been built based on this technology, including the first solar electricity generation systems (SEGS) in California's Mojave desert [1]. The

* Corresponding author: Tel.: +27 12 382-4469; fax: +27 12 382-5602 E-mail address: mwesigvea@tut.ac.za

1876-6102 © 2015 The Authors. Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license

(http://creativecommons.org/licenses/by-nc-nd/4.0/).

Peer-review under responsibility of Applied Energy Innovation Institute

doi: 10. 1016/j. egy pro .2015.07.402

successful operation of the SEGS plants is one of the reasons for the technological and commercial development of the parabolic trough technology. With several research and development initiatives [1,2], the cost of electricity from these systems has continued to reduce.

Nomenclature

a Aperture width, m

Ac Collector area, m2

cp Specific heat capacity, J/kg K

CR Concentration ratio

dri Absorber tube inner diameter, m

dgi Receiver glass cover outer diameter, m

d.ro Absorber tube outer diameter, m

f Collector focal length, m

h Heat transfer coefficient, W/m2 K

lb Direct normal irradiance, W/m2

L Length, m

m Mass flow rate, kg/s

Nu Nusselt number

P Pressure, Pa

Ç loss Receiver thermal loss per unit meter, W/m

q« Useful heat gain, W

S" •J gen Entropy generation per unit meter, W/m K

T Temperature, K

V Volumetric flow rate, m3/s

w p Pumping power, W

x,y,z Spatial coordinates AP Pressure drop, Pa Greek symbols

a Absorber tube absorptivity

Çr Collector rim angle, degrees

* Nanoparticle volume fraction, %

X Thermal conductivity, W/m K

t]th Collector thermal efficiency, %

tfel Power block electrical efficiency, %

H Viscosity, Pa s

P Density, kg/m3

P Collector reflectivity

0 Receiver circumferential angle, degrees

% Glass cover transimissivity

Subscripts

amb Ambient state

b Base fluid

inlet Absorber tube inlet

nf Nanofluid

p Nanoparticles

outlet Absorber tube outlet

Increasing concentration ratios is one of the areas deemed to have large potential for further cost reductions [2,3]. With higher concentration ratios, larger collector apertures are utilized reducing the required length for the same aperture area. This means reduced connections and drives, thereby reducing the cost of installation, operation and maintenance of these systems. An example is the Ultimate Trough® with an aperture of 7.5 m and a solar collector assembly length of 247 m making it the world's largest parabolic trough collector [3]. This collector is expected to reduce the solar field cost by 20 to 25 % [3].

With these advances in collector development, improved designs of receiver tubes are essential for improved thermal performance. At a given flow rate, increasing concentration ratios will result in high absorber tube circumferential temperature gradients and high absorber tube temperatures leading to increased receiver thermal loss [4]. Moreover, with these high temperature gradients, entropy generation rates in the parabolic trough receiver will increase, degrading the thermodynamic performance of the receiver [5,6].

Even though, improved designs of receiver tubes have been developed such as the Schott PTR® 70 receiver tube with low emissivity at high temperatures and a durable vacuum [7], enhancement of convection heat transfer in the receiver's absorber tube has potential to further improve the performance of the receiver. As such, several investigations on heat transfer enhancement in parabolic trough receivers have been carried out in the recent past [8-11]. Most of these studies have investigated the use of passive heat transfer enhancement techniques to improve the performance of parabolic trough receivers.

Recently the use of nanofluids for heat transfer enhancement in heat transfer devices has gained significant attention. Nanofluids are simply engineered dilute colloidal suspensions of particles with sizes in the nano-scale range (less than 100 nm) in a base fluid [12]. The heat transfer performance obtainable with nanofluids is known to surpass the performance of heat transfer liquids available today [13,14].

Some studies on the use of nanofluids for solar energy harvesting are available in literature. Javadi et al. [15] presented an extensive review on the performance of solar collectors using nanofluids. Taylor et al.[16] used a conservative simplified analysis to compare a nano-based concentrating solar thermal system with a conventional one. They showed that nanofluids have excellent potential for power tower applications with 5%-10% improvement in efficiency. Waghole et al.[17] experimentally investigated the heat transfer and friction factor of silver nanofluids in absorber/receiver of a parabolic trough receiver with twisted tape inserts. They showed increase in heat transfer performance of the receiver with the use of nanofluids. Recently, Sokhansefat et al.[18] investigated the heat transfer enhancement in a parabolic trough collector tube using Al203/synthetic oil nanofluid for the nanoparticle concentration less than 5% and operating temperatures of 300 K, 400 K and 500 K. They showed substantial increase in heat transfer performance of the receiver with the use of nanofluids.

From the above literature review, it is evident that there is potential for improved heat transfer performance in parabolic trough receivers using nanofluids. However, studies on heat transfer enhancement in parabolic trough receivers using nanofluids are still limited. In the few available studies, only the thermal performance of parabolic trough receivers with nanofluids is presented with no studies investigating the thermodynamic performance of the receiver with nanofluids. Therefore, the main purpose of this paper is to present an analysis of both the thermal and thermodynamic performance of a parabolic trough receiver with syltherm800-Al203 nanofluid.

2. Physical model

The parabolic trough collector system consists of a mirror bent in a parabolic shape with a heat collection element at its focus as shown in Fig.l. The receiver of the parabolic trough system consists of a steel absorber tube that is enclosed in a glass envelope. The space between the absorber tube and the glass cover is evacuated to very low pressures about 0.0103 Pa [1] to suppress the convective heat loss. The absorber tube is coated with a selective coating that has a high absorptivity for the incoming solar radiation and a low emissivity for infrared radiation [1,19]. The geometry of the parabola forming the collector is given by [19]

Where/is the focal length, determining the placement of the receiver tube for optimal interception of the incident rays. For a given aperture width a, and rim angle the focal length is given by [19]

The receiver tube used in this study is similar to the SEGS LS-2 receiver with an absorber tube outer diameter of 70 mm and inner diameter of 66 mm and a glass cover of outer diameter 120 mm and inner diameter of 115 mm [20]. The computational domain of the receiver used in this study is shown in Fig. 2.

y2 = 4 fx

f = a MtanO,./2)

Glass cover

Absorber tube

Parabolic concentrator

Fig.l. 2-D Cross section view of a parabolic trough system showing incident and reflected rays

The heat loss from the receiver can be represented on a thermal resistance network. The heat losses making up this resistance network depend on the state of the annulus space [19]. In this study, the annulus space between the absorber tube and the glass cover is considered evacuated, therefore, only the radiation heat loss is present between the absorber tube and the glass cover. The representative equations for the receiver thermal losses are presented in Refs. [19,21,22].

Fig.2. 2-D representation of the computational domain for a parabolic trough receiver (a) longitudinal view (b) cross-section view

3. Heat transfer fluid properties

In this work, AI2O3, a commonly used and inexpensive nanoparticle was used with Syltherm800 as the base fluid. In the Syltherm-A^Os nanofluid mixture, the nanoparticles were taken tobe 38 nm in diameter similar to previous investigations [23,24]. The density was obtained using the classical formula for conventional solid-liquid mixtures and the specific heat capacity was obtained by assuming thermal equilibrium between particles and the surrounding liquid [23,24].

The nanofluid density is determined according to [23,24]

Pnf = (1 Pb (3)

The specific heat capacity is determined from [23,24]

% = K1 - ^ P* + *cPp Pp ] ' K1 - ftp* + fop ] (4)

The dynamic viscosity is obtained from [24,25]

^ =Mi (123^2 + 7.3^+1) (5)

For thermal conductivity, the Bruggeman model [24,26] which considers interaction among spherical particles with various concentrations of inclusion was used. The thermal conductivity is given as [24,26]

Xtf = 0.25 [(30 - \)Xp + (2 - 30)Ab + VÂ] (6)

A = [(30 - \)Xp + (2 - 30)^ ]2 +

The properties of the base fluid are temperature dependent. For this study, curve fitted polynomials obtained using regression analysis from manufacturer data sheets [27] were used. The specific heat capacity cpb, the density pb and the thermal conductivity Xp are given by the polynomials given by Eqs. (8) - (10) respectively.

For 233.15 < T< 673 K

Cpb = 1.10787 +1.70736 xl0~3 T (kJ / kgK )

pb = 1.2691 x 103 -1.52115^ +1.79133 x 10^T2 -1.67145 x 10^6T3

■ / m3)

(8) (9)

xb = 1.90134 X 10_1 -1.88053 x \0TAT (W / mK) The viscosity is given by piecewise polynomials given by Eqs. (11) and (12) For233.15 <T< 343 K

/ub = 5.14887 x 104 - 9.61656 x 102T + 7.50207^2 -3.12468x 10~2T3 + 7.32194 x 10~5T4 - 9.14636 x 10~8 T5 + 4.75624 x lO^Y6 (mPa.s) For 343 < T< 673.15 K

jub = 9.88562 x 101 - 7.30924 x lO^T + 2.21917 xl0~3 T2 - 3.42377 x 10~6 T3 + 2.66836 x 10~9 T4 - 8.37194 x 10~13 T5 (mPa.s)

4. Numerical analysis

4.1 Boundary conditions

For this study, the boundary conditions used include: (1) a non-uniform heat flux on the absorber tube's outer wall. The heat flux is obtained using a Monte-Carlo ray tracing procedure using SolTrace [28]. The resulting heat flux profiles used are similar to those used in Mwesigye et al. [5]. The rim angle used is 80° and the aperture width is 6 m or a geometric concentration ratio (Cr) of 86. A direct normal irradiance (DNI) of 1 000 W/m2 was assumed throughout this work. The collector is assumed to be of perfect alignment and perfect shape thus, negligible specular and slope errors. (2) Velocity inlet and pressure outlet boundary conditions were used for the absorber tube's inlet and outlet respectively as shown in Fig. 2(a). (3) No-slip and no-penetration boundary conditions were specified for the absorber tube and glass cover walls. (4) There is no flow in the receiver's annulus space; therefore, for the inlet and outlet of the receiver's annulus space, a symmetry boundary condition was used such that the normal gradients of all flow variables are zero. (5) A mixed boundary condition was used on the outer wall of the glass cover to account for both radiation and convection heat transfer. Stefan Boltzmann's law gives radiation between the glass cover and the sky. Convection heat transfer from the receiver's glass was modelled by specifying a convection heat transfer coefficient and free stream temperature. The detailed equations for the convection heat transfer coefficient and sky temperature are presented in previous studies [5,22]. The ambient temperature used was maintained at 300 K and the wind speed was fixed at 2 m/s. Table 1, shows the summary of the simulation parameters used in this study.

Table 1. Simulation parameters used in this study

Parameter Value Parameter Value

a 6 m dH 0.066 m

L 5.0 m dro 0.07 m

P 0.96 tg 0.97

<Pr 80° a 0.96

cR 86 t 0 - 8%

Tinlet 350-600 K Tomb 300 K

4.2 Numerical procedure

The flow inside the receiver's absorber tube is turbulent. Therefore, the solution was obtained numerically by solving the Reynolds averaged Navier-Stokes equations together with the Realisable k-s model [29] for turbulence closure. The discrete ordinates model was used to model the radiation heat transfer in the receiver's annulus space.

All the numerical modelling was implemented in ANSYS® 14.5. The procedure includes, the modelling of the geometry of the receiver in ANSYS design modeller, the discretisation of the computational domain using ANSYS meshing and solution of the governing equations together with the

(11) (12)

boundary conditions using ANSYS Fluent [29]. Second order upwind schemes were employed for integrating the governing equations together with the boundary conditions over the computational domain. For pressure and velocity coupling, the SIMPLE algorithm was used. To capture the near wall gradients, the dimensionless wall coordinate y+ of about 1 was ensured in all simulations. For these low values of y+, the enhanced wall treatment option was used for modelling the near-wall regions. Convergence was obtained with scaled residuals of mass, momentum, turbulent kinetic energy (k) and turbulence dissipation rate (e) less than 10"5 while the energy residuals were less than 10"7. Solution was taken to be fully converged when the convergence history of the heat transfer rate, receiver heat loss and entropy generation ceased changing for more than 150 successive iterations.

5. Results and discussions

5.1 Validation of the numerical models

The numerical results obtained in this study were validated against experimental data and analytical solutions available in literature. The validation of the receiver model, the ray trace results are presented in an earlier study [5] and the entropy generation model was also validated in previous studies [5,6]. In all cases there was excellent agreement with available data.

5.2 Thermal performance

The thermal performance of the receiver can be presented in terms of the receiver thermal loss, the collector thermal efficiency and also the heat transfer coefficient. In determining the thermal efficiency, pumping power is included in the thermal efficiency equation to cater for increase in pumping power with

inclusion of nanoparticles. As such, the thermal efficiency is given by = (qu - w / )l lbAc . The

electrical efficiency of the power block, was taken as 32.7% [30], the useful heat gain is given by

q^ = mc^ - TM) and the pumping power is given by = VAP .

Fig. 3 (a) shows the heat transfer coefficient as a function of Reynolds number and nanoparticle volume fraction at a temperature of 600 K. At a given Reynolds number, the heat transfer coefficient is shown to increase as nanoparticle volume fraction increases as expected. For the range of parameters considered, the heat transfer performance increases by up to 76 % as the volume fraction increases from 0 - 8%. Fig. 3(b) shows the receiver thermal loss as a function of Reynolds number and volume fraction at a temperature of 450 K. As shown in this figure, the receiver thermal loss decreases as the volume fraction increases. This is mostly at low values of Reynolds numbers or flow rates. At low values of Reynolds numbers, the absorber tube temperatures are higher and heat transfer performance is low, thus the use of nanofluids leads to improved heat transfer performance and reduced absorber tube temperatures which leads to reduced radiation loss from the receiver's absorber tube.

20 ........................

0 5 10 15 20 25 Re [ x 104 ]

(b) Receiver thermal loss as a function of Reynolds number and nanoparticle volume fraction at an inlet temperature of 450 K

-•-0 %

— 2 % — 4 %

— 6 %

60 80 100

[xlO4]

Fig. 3. (a) Heat transfer performance as a function of Reynolds number and nanoparticle volume fraction at an inlet temperature of600 K

Fig. 4 (a) and (b) shows the variation of the thermal efficiency as a function of Reynolds number and nanoparticle volume fraction at inlet temperatures of 450 K and 550 K respectively. As shown in the figures, the variation of the thermal efficiency depends on the inlet temperature of the fluid. At low temperatures, the thermal efficiency continuously reduces as the Reynolds number increases. As temperatures increase, the trend of the Reynolds number changes, it increases and attains a maximum value at some Reynolds number and decreases again. This is probably because, at low temperatures, the absorber tube temperatures are low and, therefore, receiver thermal loss is low, so that increasing the flow rates does not significantly affect the receiver thermal loss and thermal performance. However, since the fluid is denser and more viscous at low temperatures, increasing the flow rates decreases the thermal efficiency due to increased pumping power. As temperatures increase, absorber tube temperatures, as well as receiver radiation losses, increase. As such, increasing flow rates reduce the absorber tube temperatures and thus radiation losses, thereby increasing receiver thermal efficiency. A further increase in flow rates leads to much higher pumping power requirements and the efficiency decreases.

90 86 82 78 74 70

0 2 4 6 8 10 12 14 16

Re [ X 104 ]

Fig. 4. (a) Thermal efficiency as a function of Reynolds number and nanoparticle volume fraction at an inlet temperature of 400 K

-•-0 %

Ti 1 %

-*-2 % -•-4 %

■—6 % -o-8 %

70 68 -66 -64 62 -60

* ■ — 0 % '

4 1 \ \ V >, \ \ \\ -»-2% '■ -»-4 % ■ "-6 % : ■0-8 % .

' 4 \ \ « » \ ■ > 1 \ ■ 1 1 \ ■..............1. » 4 \ \ \ \ 1 \ . \ \ • . V- ....

10 20 30 40 50 60 70 80

Re [ X 104 ]

(b) Thermal efficiency as a function of Reynolds number and nanoparticle volume fraction at an inlet temperature of 550 K

Accordingly the use of nanofluids is shown to be viable at low values of fluid temperatures and Reynolds numbers where heat transfer performance is usually low. At the lowest Reynolds number and an inlet temperature of 400 K, the thermal efficiency increases by as much as 8% as the nanoparticle volume fraction increases from 0% to 8%. As the Reynolds numbers and volume fraction increase, the required pumping power also increases and reduces the thermal efficiency. Further increase in Reynolds number and volume fraction reduces the thermal efficiency to values lower than those of a receiver with only the base fluid. Accordingly, significant improvements in efficiency are achievable at flow rates lower than 24.6 m3/h for most of the temperatures considered. The highest increase in efficiency being at the lowest flow rate of 4.93 m3/h and lowest temperature considered in this study. The increase in efficiency is attributed to increased heat transfer performance and reduced absorber tube temperatures which lead to lower thermal losses.

5.3 Thermodynamic performance

Regarding thermodynamic performance, the use of nanofluids in the receiver should not result in higher entropy generation rates compared to when only the base fluid is used. Using the method for determining entropy generation described in earlier studies [6, 7], entropy generation due to heat transfer and fluid friction is obtained and presented. As shown in Fig. 5(a), at the flow rate of 18.5 m3/h and inlet temperatures ranging from 350 K - 600 K, the entropy generation rates in the receiver are reduced with the use of nanofluids. It should be noted that at a given flow rate, the Reynolds number increases with temperature since the fluid properties are temperature dependent. The entropy generation rate is also shown to reduce as the nanoparticle volume fraction increases due to improved heat transfer performance. At low flow rates, the heat transfer irreversibilities are dominant and using nanofluid improves receiver heat transfer performance thus reducing the heat transfer irreversibilities. Fig. 5 (b) shows the entropy

generation as a function of Reynolds number and nanoparticle volume fraction, as shown, the use of nanofluids improves the thermodynamic performance at low Reynolds numbers. As the Reynolds numbers increase and become larger than some value at a given volume fraction, the entropy generation becomes more than that in a receiver with only the base fluid and makes the use of nanofluids thermodynamically useless. This is because at high Reynolds numbers, the heat transfer irreversibilities are significantly reduced while the fluid friction irreversibilities are increased significantly and become the dominant source of irreversibility. The Reynolds number beyond which the use of nanofluids does not make thermodynamic sense is shown to reduce as the nanoparticle volume fraction increases.

Fig. 5. (a) Entropy generation in the receiver as a function of Reynolds number and volume fraction at a flow rate of 18.5 m3/h for 350 K - 600 K

(c) Entropy generation in the receiver as a function of Reynolds number and volume fraction at an inlet temperature of 450 K

6. Conclusion

In this work a numerical study on the thermal and thermodynamic performance of a parabolic trough receiver using Syltherm800-Al203 nanofluids is presented. From the study, it is shown that the use of nanofluids improves the thermal and thermodynamic performance of the receiver for some range of flow rates (or Reynolds numbers) at a given fluid temperature. The heat transfer performance increases by up to 76%, the collector thermal efficiency increases up to 8% for flow rates lower than 24.6 m3/h for most fluid temperatures considered. From the thermodynamic analysis, it was shown that there is a Reynolds number beyond which use of nanofluids deteriorates the thermodynamic performance of the receiver since the entropy generation rates become higher than those in receivers using only the base fluid.

7. Copyright

Authors keep full copyright over papers published in Energy Procedia

Acknowledgements

The authors acknowledge the support received from the Faculty of Engineering and the Built Environment at Tshwane University of Technology. The funding received from the National Research Foundation (NRF) is also duly acknowledged and appreciated.

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Biography

Aggrey Mwesigye is currently a PhD student at the University of Pretoria and a Lecturer in the department of Mechanical Engineering at Tshwane University of Technology. His areas of interest are renewable and sustainable energy, computational fluid dynamics, entropy generation minimization, exergy analysis, heat transfer enhancement and thermodynamics.