THERMODYNAMIC OPTIMISATION AND EXPERIMENTAL COLLECTOR OF A DISH- MOUNTED SMALL-SCALE SOLAR THERMAL BRAYTON CYCLE WG LE ROUX Study-leaders: Prof. T. Bello-Ochende Prof. J.P. Meyer Department of Mechanical and Aeronautical Engineering, University of Pretoria , South Africa February, 2015 Submitted in partial fulfilment of the requirements for the degree PhD (Mechanical Engineering) 1
Presentation Outline 1. Introduction 2. Background 3. Literature Study 4. Modelling and Optimisation 5. Analytical Results 6. Experimental Study 7. Conclusion 8. Recommendations 2
1. Introduction Parabolic dish concentrator for a Stirling engine (Image extracted from Pitz-Paal, 2007) Long-term average of direct normal solar irradiance on a world map showing the potential of solar power generation in southern Africa (GeoModel Solar, 2014) A typical micro-turbine (the GT1241) as available from Honeywell, Garrett proposed for the small-scale solar thermal Brayton cycle 3 (Image extracted from Garrett, 2014)
1. Introduction Problem Solar-to-electricity technologies are required which are • more competitive • • more efficient • cost-effective Purpose of the study Small-scale dish-mounted open solar thermal Brayton cycle • optimise solar receiver and recuperator - method of total entropy generation minimisation test optimised receiver • Objectives • Second law of thermodynamics Entropy generation minimisation • Ray-tracing software • Geometry optimisation • • Experimental receiver test 4
2. Background Scope of Research – Thermodynamic Optimisation Open and direct solar thermal Brayton cycle • Second Law of Thermodynamics • Entropy Generation Minimisation • • Maximise net power output • Optimise geometry of recuperator and receiver Heat Transfer & Fluid Flow Irreversibilities • Experimental setup •
Solar resource – South Africa Why Solar?
Solar resource - World According to DLR •
Solar resource – South Africa Why Solar?
The Department of Minerals and Energy places South Africa’s annual direct normal irradiation (DNI) between 2 500kWh/m2 and 2 900 kWh/m2 with an average of almost 300 days of sunshine per year.
Solar resource – South Africa, Pretoria Meteonorm 1400 1200 1000 Irradiance (W/m^2) 800 Irradiance of beam 600 Mean irradiance of global 400 radiation, tracked Mean irradiance of global radiation horizontal 200 0 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 Time (h)
CSP - Concentrating methods Dish Trough Tower
Background Brayton cycle Small-scale solar power Photovoltaic cells • mobility, cost benefits • Solar water heaters • • micro-turbines • CSP (Concentrated solar power) hybrid system • – Trough • storage systems • Rankine Cycle water heating • – Dish-mounted efficient and highly competitive. • • Stirling Engine Maximum net power output • Brayton cycle • combined effort of • heat transfer, Air out • fluid mechanics and 11 • thermodynamics 10 9 Recuperator 3 4 W W W 8 net t c 2 Compressor Turbine Load Receiver 5 7 6 12 Q * 1 Air in
Solar tracking - Elevation • SunEarthtools
Solar tracking - Azimuth Measured angle of tracking system versus real azimuth angle of the sun 150 Morning measurements 100 Noon measurements Afternoon measurements SunEarthTools 50 Azimuth angle 0 -50 -100 -150 6 7 8 9 10 11 12 13 14 15 16 17 18 Time (h)
Two-axis solar tracking required for dish Solar Active tracking Passive Micro- Date and time processor based or a and combination Fluid Bi- electro- of sensor and Auxiliary metallic optical date/time bifacial strips sensor based solar cell based based Mousazadeh et al. (2004), Poulek and Libra (2000)
3. Literature Study Air out 11 10 9 Recuperator 3 4 W W W 8 net t c 2 Compressor Turbine Load Receiver 5 7 6 Q * 1 Air in The open and direct solar thermal Brayton cycle 16
3. Literature Study Test set-up of a solar thermal Brayton cycle (Image extracted from Heller et al., 2006) 17
Small-scale open and direct solar thermal Brayton cycle with recuperator Advantages • – High recommendation – Air as working fluid – Hot air exhaust • Water heating • Space heating • Absorpsion refrigeration – Recuperator • high efficiency and • low compressor pressure ratios Disadvantages • – recuperator and receiver pressure losses – turbo-machine efficiencies – recuperator effectiveness irreversibilities – Heat losses
Solar thermal Brayton - Recuperator Air out 11 10 9 Recuperator 3 4 W W W 8 net t c 2 Turbine Compressor Load Receiver 5 7 6 * Q 1 Air in
Solar thermal Brayton - Recuperator Image extracted from: Stine, B.S., Harrigan, R.W., 1985, Solar energy fundamentals and design. New York: John Wiley & Sons, Inc.
3. Literature Study Efficiencies of different solar receivers – Pressurised volumetric Receiver Reference T out (K) T in (K) P (kPa) (kg/s) Working fluid ΔP (Pa) rec type number or m model Pressurised PLVCR-5 71% 1 323 - 420 - Air - volumetric (Ávila-Marín, 2011) PLVCR-500 57% 1 233 300 415 - Air - (Ávila-Marín, 2011) DIAPR 79% 1 477 308 1 800 0.0222 Air 25 000 (Karni et al., 1997), (Ávila-Marín, 2011) REFOS 67% 1 073 - 1 500 - Air 1 800 (Buck et al. 2002), (Ávila-Marín, 2011) Dickey, 2011 88% 871 542 273 0.409 Air 2 900 21
3. Literature Study Efficiencies of different solar receivers - Tubular Receiver Reference T out (K) T in (K) P (kPa) (kg/s) Working fluid ΔP (Pa) type number or m rec model Tubular Cameron et 51%* 1 089 865 370 0.73 He-Xe 7 000 al., 1972 Kribus et al., - 1 023 300 1 600 - 0.01 Air 40 000 1999 1 900 Heller et al., - 823 573 650 - Air 10 000 2006 Neber and 82% 1 500** - 760 0.0093 Air 40 Lee, 2012 Amsbeck et 43% 1 076 876 384 0.526 Air 7 330 al., 2010 Amsbeck et 39.7% 1 055 871 375 0.516 Air 7 400 al., 2010 Solugas - 873 598 850 5.6 Air (Quero et al., 2013) *calculated by author **proposed 22
3. Literature Study Particle receiver Open volumetric receiver – HiTRec (Image extracted from Miller and Koenigsdorff, (Image extracted from Ávila-Marín, 2011) 1991) 23
3. Literature Study Closed volumetric receiver, Longitudinal tubular receiver REFOS (Image extracted from Amsbeck et al., 2008) ( Image extracted from Buck et al., 2002) 24
3. Literature Study Ceramic counterflow plate-type recuperator Coiled tubular receiver (Image extracted from Pietsch and Brandes, 1989) (Image extracted from Kribus et al., 1999) 25
3. Literature Study Q 1 = 6.8 kW, T 1 = 1 308 K, Q 2 = 8.3 kW, T 2 = 1 179 K, Q 3 = 9.7 kW, T 3 = 1 054 K, Q 4 = 11.2 kW, T 4 = 904 K Q 5 = 12.7 kW, Q 6 = 14.1 kW, Q 7 = 15.9 kW Performance map 12 (in different weather conditions) Q 6 10 Q 7 small-scale open solar thermal • Brayton cycle 8 Q 5 fixed optimised geometries • T 1 (kW) T 2 6 T 3 4 T 4 2 Q 4 Q 2 Q 1 Q 3 0 1.4 1.6 1.8 2 2.2 2.4 26
4. Modelling and Optimisation Control volume for the open solar thermal Brayton cycle Q loss , j j m W W W m net t c Q * 27
4. Modelling and Optimisation Solar receiver - SolTrace Example of an analysis done for the solar dish and receiver 28
4. Modelling and Optimisation Solar receiver Rectangular open-cavity Heat loss from the solar receiver open-cavity receiver 29
4. Modelling Solar receiver air heating Rectangular open cavity tubular receiver • Stainless steel • Pressure drop (Colebrook equation) • Variables • Tube diameter, • Inlet temperature, • Mass flow rate 30
4. Modelling Solar receiver – conduction heat loss [1] Assumptions: Wind speed: 2.5 m/s • T 0 = 300 K • P 0 = 86.6 kPa • 100 mm insulation thickness • Conductivity of 0.061 W/mK at • 550 ° C average temperature [2] Elevation angle of 45 ° • ( 1 / h t / k ) 1 . 77 out ins ins A T T T T n s , n s , n Q loss , cond , n R 1 / h A t / k A cond out n ins ins n [1] Le Roux, W.G., Bello-Ochende, T. and Meyer, J.P., 2014, The efficiency of an open cavity solar receiver for a small-scale solar thermal Brayton cycle, Energy Conversion and Management 2014;84:457 – 70. 31 [2] Harris, J.A., Lenz, T.G., 1983, Thermal performance of solar concentrator/cavity receiver systems, Solar Energy 34 (2), pp. 135-142.
4. Modelling Solar receiver – radiation heat loss 4 4 Q A T T loss , n , rad ap s , n N 4 4 Q A F T T n n n j n s , n j s , j j 1 32
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