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Research Papers

Experimental and Numerical Heat Transfer Analysis of an Air-Based Cavity-Receiver for Solar Trough Concentrators

[+] Author and Article Information
R. Bader

Department of Mechanical and Process Engineering, ETH Zurich, 8092 Zurich, Switzerland

A. Pedretti

Airlight Energy Holding SA, 6710 Biasca, Switzerland

A. Steinfeld1

Department of Mechanical and Process Engineering, ETH Zurich, 8092 Zurich, Switzerland;Solar Technology Laboratory,  Paul Scherrer Institute, 5232 Villigen PSI, Switzerlandaldo.steinfeld@ethz.ch

1

Corresponding author.

J. Sol. Energy Eng 134(2), 021002 (Feb 27, 2012) (8 pages) doi:10.1115/1.4005447 History: Received June 07, 2011; Revised November 03, 2011; Published February 27, 2012; Online February 27, 2012

We report on the field testing of a 42 m-long full-scale solar receiver prototype installed on a 9 m-aperture solar trough concentrator. The solar receiver consists of a cylindrical cavity containing a tubular absorber with air as the heat transfer fluid (HTF). Experimental results are used to validate a heat transfer model based on Monte Carlo ray-tracing and finite-volume techniques. Performance predictions obtained with the validated model yield the following results for the receiver. At summer solstice solar noon, with HTF inlet temperature of 120 °C and HTF outlet temperature in the range 250–450 °C, the receiver efficiency ranges from 45% to 29% for a solar power input of 280 kW. One third of the solar radiation incident on the receiver is lost by spillage at the aperture and reflection inside the cavity. Other heat losses are due to natural convection (9.9–9.7% of solar power input) and re-radiation (6.1–17.6%) through the cavity aperture and by natural convection from the cavity insulation (5.6–9.1%). The energy penalty associated with the HTF pumping work represents 0.6–24.4% of the power generated.

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Copyright © 2012 by American Society of Mechanical Engineers
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Figures

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Figure 1

(a) Solar receiver prototype cross-section; squares indicate thermocouple locations installed at the center between receiver inlet and outlet. (b) Enlarged cross-section of the right-wing secondary concentrator, consisting of an asymmetric CPC coupled to a circular segment.

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Figure 2

(a) Solar flux concentration distribution, and (b) angular flux distribution at the outlet of the right-wing secondary concentrator, for solar incidence angles θskew=0∘ (solid lines), θskew=30∘ (dashed lines), and θskew=60∘ (dashed-dotted lines)

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Figure 3

Convective heat transfer at surfaces 2 and 3, Q·convection,2, Q·convection,3, as functions of surface temperatures T2 and T3, for a downward-facing receiver aperture, obtained by CFD (+), by the empirical correlations (solid lines), and by applying Kuehn-Goldstein Nu-correlation [10] for convective heat transfer between nested cylinders (dashed lines)

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Figure 4

Experimental test facility in Biasca, Switzerland

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Figure 5

Measured and simulated pressure drops in the HTF flow between receiver inlet and outlet as a function of the HTF mass flow rate

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Figure 6

(a) Measured HTF inlet and outlet temperatures and simulated HTF outlet temperatures; and (b) ratios of simulated and measured receiver surface temperatures at the measurement locations indicated in Fig. 1 for 13 off-sun test runs

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Figure 7

Thermal heat losses from the receiver for 13 off-sun test runs

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Figure 8

(a) Measured DNI and HTF mass flow rate as a function of time and (b) measured HTF inlet and outlet temperatures and simulated HTF outlet temperature as a function of time for a representative transient on-sun test

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Figure 9

Efficiency of the solar receiver prototype as a function of the HTF outlet temperature, for the spring equinox (a) and the summer solstice (b), at 8:00, 10:00, and 12:00 solar time; HTF mass flow rate is varied from 0.1 to 1 kg/s; HTF inlet temperature is 120 °C; for comparison, the efficiency of the Schott PTR70 (2008) receiver at solar noon is shown

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Figure 10

Pressure drop ΔpHTF across the solar receiver prototype and corresponding isentropic HTF pumping power requirement W·pump,s for the spring equinox (a) and the summer solstice (b) at solar noon, as functions of the HTF mass flow rate

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Figure 11

Energy breakdown of heat losses by the receiver (gray areas) and heat gain by the HTF (white area), and the associated HTF pumping power requirement (black curve), normalized by the solar power incident from the primary concentrator, Q·solar, as functions of the HTF outlet temperature THTF,out in the range 250–450 °C for the spring equinox (a) and the summer solstice (b) at solar noon; THTF,in=120 °C

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