Student: A. Mwesigye. Supervisor: Prof. T. Bello-Ochende. Co-supervisor : Prof. J. P. Meyer. Date: February 2015

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1 THERMAL PERFORMANCE AND HEAT TRANSFER ENHANCEMENT OF PARABOLIC TROUGH RECEIVERS NUMERICAL INVESTIGATION, THERMODYNAMIC AND MULTI-OBJECTIVE OPTIMISATION Date: February 2015 Student: A. Mwesigye Supervisor: Prof. T. Bello-Ochende Co-supervisor : Prof. J. P. Meyer

2 Presentation outline Introduction Motivation and objectives of the research Research methodology Results and discussions Parabolic trough receiver thermal models Parabolic trough thermodynamic performance Heat transfer enhancement Conclusion and recommendations 2

3 INTRODUCTION

4 Introduction Why renewable energy? Clean sources of energy with minimum environmental impacts (less or no pollution) They are not finite like conventional sources Potential for distributed generation of electricity Posses significant potential to meet increasing energy demand and also provide for those without access Potential to improve security of energy supply Why solar energy Widely available Has lots of potential to supply a significant portion of world s energy needs Energy received on earth s surface is about 1.7x10 14 kw 84 minutes of solar radiation gives 900 EJ (World s energy demand in 2009) 4

5 Introduction Solar resource For CSP, DNI > kwh/m 2 /yr is suitable 5 Worldwide annual direct normal irradiation Source: (Breyer and Knies, 2009)

6 Introduction Pathways to electricity Photovoltaic systems Concentrated solar thermal systems Solar thermal is preferred It is easier to store thermal energy than electricity Several concentrated solar thermal technologies exist Parabolic trough systems Parabolic dish systems Solar tower Linear Fresnel systems Parabolic trough systems are the most technically and commercially developed Due to successful operation of SEGS plants in Mojave desert, California 6

7 Introduction Solar thermal technologies (a) Parabolic trough collector (Maria Trimarchi, 2012) (b) Solar pond (Hignett et al., 2009) (c) Solar dish (C. Rubens, 2008) 6 (d) Solar tower (Ariel Schwartz, 2012) Different solar thermal technologies (e) Solar chimney(unenergy blog)

8 Introduction Parabolic trough geometrical relations Equation of the parabola y 2 = 4fx Focal length is given by f = W a / 4tan r ( ϕ / 2) W f/w a Parabolic trough collector geometry 8

9 MOTIVATION AND RESEARCH OBJECTIVES

10 Motivation Some of the challenges with parabolic trough collector system include: - High capital costs - Higher thermal losses at elevated temperatures - HTF decomposition at higher temperatures (> 400 o C) - Circumferential temperature gradients in the absorber tube can cause glass envelop breakages and lead to receiver vacuum loss Most of these challenges are subject of current R&D efforts Using high concentration ratios is one of the ways with potential for further cost reductions - Possible with availability of light weight materials and advances in mirror manufacture - Reduces number of drives and controls and thus reduced capital costs 10

11 Motivation Why this study? Influence of high concentration ratios on receiver thermal and thermodynamic performance has not been investigated Limited studies on possible improvement in receiver thermal and thermodynamic performance with heat transfer enhancement Most studies use constant heat flux boundary conditions With constant heat flux, receiver circumferential temperature differences are not accounted for Thermodynamic performance of the receiver obtained with constant heat flux boundary conditions is not accurate In this study, realistic non-uniform heat flux boundary conditions on the receiver s absorber tube are determined and used 11

12 Motivation The receiver Central component of the system Accounts for about 30% of total plant cost Costly to replace All thermal losses take place here Parabolic trough receiver 12

13 Objectives of the research Overall objective was: to develop thermal and thermodynamic models for investigating performance of a parabolic trough receiver as well potential for improved performance with heat transfer enhancement The specific objectives were Develop and validate numerical model to predict receiver performance with realistic non-uniform heat flux profile Investigate entropy generation from heat transfer irreversibility and fluid friction irreversibility at different concentration ratios and rim angles Investigate potential for improved thermal and thermodynamic performance with heat transfer enhancement For twisted tape inserts, perforated plate inserts and perforated conical inserts Determine optimal configurations and conditions for each of the heat transfer enhancement techniques using multi-objective optimisation and thermodynamic optimisation 13

14 NUMERICAL MODELLING AND OPTIMISATION PROCEDURE

15 Numerical modelling procedure Numerical modelling and optimisation with ANSYS 13, 14, 14.5 Important aspects of numerical modelling Non uniform heat flux boundary conditions obtained with SolTrace Turbulent flow conditions Evacuated receiver s annulus space Absorber tube is selectively coated with cermet coating (variable emissivity) Mixed boundary condition (radiation and convection) for receiver s glass cover Realizable k-ε model for turbulent closure Low values of y + ~ 1 to fully resolve the viscous sublayer and accurately predict heat transfer and fluid friction Enhanced wall treatment for near wall modelling SIMPLE algorithm for pressure and velocity coupling Temperature dependent fluid properties Mesh dependence studies are carried out for each receiver model 15

16 Optimisation procedure For a multi-objective optimisation problem Maximise/minimise f m (x), Subject to g j (x) 0 h k (x) = 0 x (L) i x i x (U) i m = 1, 2,..,M j = 1,2,.,J k = 1,2,..,K i = 1,2,,n - f m (x) is the objective function and Mis the number of functions - xis the vector of n decision variables such that x= (x 1, x 1,.., x n ) T - g j (x) and h k (x) are constraint functions (J are inequality constraints and K equality constraints) - x i (L) and x i (U) are lower and upper limits respectively 16

17 Optimisation procedure Multi-objective optimisation was implemented in ANSYS design explorer Optimise heat transfer and fluid friction performance The two objectives are conflicting Pareto optimal solutions are sought solutions not dominated with respect to another Entire optimisation procedure is shown in the adjacent figure 17 Multi-objective optimisation procedure

18 Optimisation procedure Goodness of fit one of the metrics for assessing accuracy of response surfaces (b) (a) Goodness of fit (a) using standard response surfaces and (b) goodness of fit after refinement 18

19 Optimisation procedure Thermodynamic optimisation based on the entropy generation minimisation method configurations with lower entropy generation rates are better entropy generation rates of interest are the ones due to Fluid flow irreversibility S gen, F = µ u T x i j + u x i j u x i j + ρε T And heat transfer irreversibility S α λ α T 2 t 2 gen, H = ( T) + ( T) 2 2 T λ Total entropy is obtained from Sgen = S V gen dv With, S gen = Sgen F + Sgen, H 19

20 RESULTS Receiver thermal models

21 Results -Receiver thermal models Simulation parameters Reflector Receiver Aperture width, W a 4 10 m Absorber tube inner diameter, d ri m Collector length, L c 5 m Absorber tube outer diameter, d ro 0.07 m Reflectivity, ρ c 0.96 Glass cover inner diameter, d gi 0.11 m Rim angle, φ r o Glass transmissivity, τ g 0.97 Concentration ratio, C R =A c /A r Environmental conditions Direct normal irradiance, DNI 1000 W/m 2 Absorber tube absorptivity, α abs 0.96 Glass cover emissivity, ξ g Ambient temperature, T amb Wind velocity, V w Inlet temperature, T inlet 300 K 2 m/s K

22 Results -Receiver thermal models LCR Yang et al.[12] He et al.[15] Jeter[18] This work 10 Ray tracing results for a parabolic trough collector with an aperture width of 10 m and rim angle, φ r = 80 o Angle (θ o ) Comparison of the present study prediction of local concentration ratio (LCR) as a function of circumferential angle (θ) with literature 22

23 Results -Receiver thermal models Heat flux (W/m 2 ) ϕ r (degrees) Heat flux (W/m 2 ) C R Αngle (θ o ) (a) Angle (θ ο ) Heat flux on the absorber tube as a function of circumferential angle (a) at different rim angles for a concentration ratio of 86 (b) at different concentration ratios and a rim angle of 100 o (b) 23

24 Results -Receiver thermal models 2-D computational domain and discretized domain of the receiver tube (c) 24 2-D computation domain (a) Longitudinal view (b) cross-sectional view Discretized receiver domain Mesh

25 Results -Receiver thermal models Validation of receiver thermal performance with data from Sandia National laboratories Wind Air DNI Flow rate speed temperature (W/m 2 ) (L/min) (m/s) ( o C) T inlet ΔT ( o C) ( o C) (Experimental) ΔT ( o C) % Efficiency Efficiency % (Present error (Experimental) (present error study) ΔT study)

26 Results -Receiver thermal models Heat flux distribution on the receiver s absorber tube at a concentration ratio of 86 (a) ϕ r = 40 o (b) ϕ r = 120 o 26

27 Results -Receiver thermal models (a) Receiver temperature distribution at a concentration ratio of 86, inlet temperature of 400K, ϕ r =70 o andflowrate4.9m 3 /h(a)insidetheabsorbertube (c)annulusspace (b) 27

28 Results -Receiver thermal models From the thermal performance evaluation Simple, easy to use and accurate correlations were developed Nusselt number is given by Nu =0.0104Pr Re Friction factor is given by f = 0.173Re Obtained in the range Re ; 9.29 Pr 33.7 and 400 K T 650 K Valid within ±5% 28

29 Results -Receiver thermal models Comparison of present study developed heat transfer and fluid friction correlations with Gnielinski and Petukhov correlations respectively [84] on a scatter plot. 29

30 Results -Receiver thermal models Heat loss (W/m) ϕ r (degrees) Heat loss (W/m) C R T inlet (K) T inlet (K) Receiver heat loss as a function of fluid temperature and collector rim angle at a flow rate of 4.9 m 3 /h and concentration ratio of 86 Receiver heat loss as a function of fluid temperature and concentration ratio at a flow rate of 4.9 m 3 /h and rim angle

31 RESULTS Thermodynamic optimisation

32 Results -Thermodynamic optimisation Nu C R f C R Re [ x 10 3 ] Re [ x 10 3 ] Nusselt number and friction factor respectively as functions of Reynolds number and concentration ratio 32

33 Results -Thermodynamic optimisation Thermodynamic analysis and optimisationare based on the entropy generation minimisation method (EGM) S gen (W/K) (S ) : Direct method gen F (S ) : Direct method gen H S : Direct method gen S : Indirect method [61] gen Re Validation of the entropy generation model with Bejan s analytical correlation 33

34 Results -Thermodynamic optimisation S ' gen (W/m K) (a) S' gen ; C R = 71 (S' gen ) H ; C R = 71 (S' gen ) F ; C R = 71 S' gen ; C R = 129 (S' gen ) H ; C R = 129 (S' gen ) F ; C R = Re [ x 10 3 ] T (K) inlet Entropy generation in a parabolic trough receiver (a) irreversibilities in the receiver (b) Entropy generation with inlet temperature at different rim angles and flow rate of 9.25 m 3 /h S' gen (W/m K) C R = 86 (b) ϕ r (degrees)

35 Results -Thermodynamic optimisation x/l = 0.5 x/l = 0.75 x/l = S''' gen (Wm -3 K -1 ) Be y/r = y/r = y/r = 0 y/r = 0.61 y/r = y/r x/l Entropy generation in the receiver s absorber tube as a function of absorber tube s radial position (y/r) and different positions along the tube s streamwise direction (x/l). Bejan number as a function of absorber tube streamwise position (x/l) and absorber tube radial position (y/r) 35

36 Results -Thermodynamic optimisation 2.0 ϕ r (degrees) 10 C R S' gen (W/m K) S'gen (Wm -1 K -1 ) Re [ x 10 3 ] (a) Re [ x 10 3 ] (b) Entropy generation in a parabolic trough receiver (a) as a function of Reynolds number at differentrimangles,concentrationratioof86andinlettemperatureof600k(b)asafunction ofreynoldsnumberatdifferentconcentrationratios,rimangleof80 0 and400k 36

37 Results -Thermodynamic optimisation ϕ r (degrees) Be C R Re [ x 10 3 ] Be Re [ x 10 3 ] Bejan number as a function of Reynolds number and concentration ratio (C R ) for an inlet temperature of 500 K and a rim angle (φ r )of80 o. Bejan number as a function of Reynolds number and rim angle (φ r ) for a concentration ratio (C R ) of 86 inlet temperature of 600 K. 37

38 Results -Thermodynamic optimisation Optimal flow rates at different concentration ratios Concentration ratio (C R ) Optimal flow rate (m 3 /h)

39 RESULTS Heat transfer enhancement

40 Results -Heat transfer enhancement Data reduction Friction factor f = P ρ u 1 2 L 2 m d ri Nusselt number Nu q dri = T T λ ri b Absorber tube circumferential temperature difference φ =T r, max Tr,min Entropy generation rate S gen = S V gen Bejan number Be = ( S ) gen S gen dv Entropy generation ratio N s, en = S S H gen, en gen, p Thermal efficiency η th qɺ Wɺ = I A / η u p el b c 40

41 Results -Receiver with twisted tape inserts Twisted tape inserts are widely used Give high improvements in heat transfer performance with moderate fluid friction Several modifications have been studied in other applications Low twist ratio, wall detached twisted tapes are considered Physical model of a receiver with twisted tape inserts 41 Computational domain

42 Results -Receiver with twisted tape inserts Low twist ratio, wall detached twisted tapes are considered (their use in parabolic trough receivers has not been studied before) Twist ratio is given by ~ y = H d ri (in the range ) The width ratio W w ~ = d ri (in the range ) Physical model of receiver s absorber tube with twisted tape inserts 42

43 Results -Receiver with twisted tape inserts High heat transfer rates are achieved due to a longer helical path taken by the fluid and resulting fluid mixing 4 4 Re = , ~ y = 0.5 andw~ = 0.91 Re = , ~ y = 0.8 andw~ = Streamlines in the absorber tube of a receiver with twisted tape inserts 43

44 Results -Receiver with twisted tape inserts Thermal Performance Nu yɶ Nu p Re p (a) [ x 10 3 ] Nu yɶ Nu p wɶ (b) Heat transfer performance (a) as a function of Reynolds number and twist ratio for a width ratio of 0.76 and inlet temperature of 400 K (b) as a function of width ratio and twistratiofor600kandre p =8.13x

45 Results -Receiver with twisted tape inserts Fluid Friction f yɶ f p f/f p yɶ Re p [ x 10 3 ] Re p [ x 10 3 ] (a) (a) Friction factor as afunction of Reynolds numberand twist ratio for awidthratioof 0.76 and inlet temperature of 400 K (b) friction factor ratio as a function of Reynolds numberandtwistratiofor600kandwidthratioof0.76 (b) 45

46 Results -Receiver with twisted tape inserts Thermal Performance Lower than a given Reynolds number, thermal efficiency improves with the use of twisted tape inserts η th yɶ (η th ) p Re [ x 10 3 ] p (a) (η ) th p Re p [ x 10 3 ] Thermal efficiency as a function of Reynolds number and twist ratio (a) for 500 K and width ratio of 0.91 (b) for 600 K and width ratio of 0.76 η th yɶ (b) 46

47 Results -Receiver with twisted tape inserts Thermal Performance φ ( o C) yɶ φ p φ ( o C) yɶ φ p Re p (a) [ x 10 3 ] Absorber tube circumferential temperature difference as a function of Reynolds number and twist ratio (a) for 400 K and width ratio of 0.76 (b) for 500 K and width ratio of 0.83 Re p (b) [ x 10 3 ] 47

48 Results -Receiver with twisted tape inserts Heat transfer correlation Nu = Re Pr yɶ wɶ p Valid for Re p and 10.7 Pr K T 600 K 0.53 wɶ yɶ 2.0 Nu (Observed) % -15% Nu (Predicted) Comparison of the predicted heat transfer performance with the observed heat transfer performance for a receiver with twisted tape inserts. 48

49 Results -Receiver with twisted tape inserts Fluid friction correlation 0.6 f = yɶ wɶ Reen Valid for Re p and 10.7 Pr K T 600 K 0.53 wɶ yɶ 2.0 f (Observed) % -16% f (Predicted) Comparison of the predicted friction factor with the observed friction factor for a receiver with twisted tape inserts. 49

50 Results -Receiver with twisted tape inserts Thermodynamic Performance Be yɶ Be p Be wɶ Be p Re p [ x 10 3 ] (a) Bejan number for a receiver with twisted tape inserts for a temperature of 600 K as a function of Reynolds number (a) and twist ratio for a width ratio of 0.61 (b) and width ratio for a twist ratio of 0.5 Re p (b) [ x 10 3 ] 50

51 Results -Receiver with twisted tape inserts Thermodynamic Performance S' gen (W/m K) wɶ (S' gen ) p S' gen (W/m K) yɶ (S' gen ) p Re p [ x 10 3 ] Re p [ x 10 3 ] (a) (b) Entropy generation for a receiver with twisted tape inserts for an inlet temperature of 600 K as a function of Reynolds number (a) and width ratio for a twist ratio of 1.0 (b) and twist ratio for a width ratio of

52 Results -Receiver with twisted tape inserts Multi-objective optimisation Sample response surfaces y~ w ~ y ~ w ~ 3-D response surfaces for Nusselt number and friction factor for Re p = 8.28x10 4 and an inlet temperature of 600 K 52

53 Results-Receiver with twisted tape inserts Multi-objective optimisation Sample Pareto optimal front In multi-objective optimisation all solutions on the Pareto front are equally important f Nu Designer can chose any of them according to his/her needs Also a decision support tool can be used to select an appropriate solution Pareto optimal solutions for a receiver with twisted tape inserts for an inlet temperature of 600 K and Re p = 1.64x

54 Results -Receiver with twisted tape inserts With twisted tape inserts Significant improvement of receiver heat transfer performance Improvement of receiver thermal efficiency between 5-10% Reduction of absorber tube temperature gradients between 4-64% Reduction of entropy generation rates up to 59% Significant improvements in thermal efficiency and reduction in entropy generation obtainable for flow rates lower than 43 m 3 /h 54

55 Results -Receiver with perforated plate inserts Perforated inserts - light weight -lower pressure drop compared to solid inserts (c) Computational domain 55 Receiver with perforated plate inserts

56 Results -Receiver with perforated plate inserts Dimensionless parameters - Dimensionless insert spacing p pɶ = L (in the range: ) Computational domain of a receiver with perforated plate inserts - Dimensionless insert size ɶ β = β β max (in the range: ) - Dimensionless insert orientation angle dɶ = d d ri (in the range: 1 β 1) ~ 56

57 Results -Receiver with perforated plate inserts Thermal Performance Nu pɶ pɶ pɶ pɶ pɶ Nu p = 0.04 = 0.08 = 0.12 = 0.16 = Re [ x 10 3 ] Heat transfer performance increases - With increase in insert size - With reduction in insert spacing With improvement in heat transfer performance - Reduction in absorber tube circumferential temperature difference Nusselt number as a function of Reynolds ~ ~ number and insert spacing for β =1, d =0. 61 and inlet temperature of 400 K 57

58 Results -Receiver with perforated plate inserts Friction factor pɶ pɶ pɶ = 0.04 = 0.08 = 0.12 f p pɶ pɶ = 0.16 = 0.20 Fluid friction also increases - With increase in insert size - With reduction in insert spacing f Re [ x 10 3 ] Friction factor as a function of Reynolds ~ ~ number and insert spacing for β =1, d =0. 61 and inlet temperature of 400 K 58

59 Results -Receiver with perforated plate inserts - Nusselt number can be predicted from Nu = ɶ Re Pr p d ( tan ) 1000 ɶ β Valid for Re and 10.7 Pr β 30 o 400 K T 600 K [ x 10 3 ] Nu (observed) % -15% Nu (predicted) [ x 10 3 ] 59 Nusselt number parity plot

60 Results -Receiver with perforated plate inserts - Friction factor can be predicted from f = pɶ dɶ + β Re ( sin ) Valid for Re and 10.7 Pr β 30 o 400 K T 600 K f (observed) % -18% f (predicted) 60 Friction factor parity plot

61 Results -Receiver with perforated plate inserts η th pɶ pɶ pɶ = 0.04 pɶ pɶ = 0.16 = 0.08 = = 0.12 (η ) th o Re [ x10 3 ] Collector thermal efficiency for a receiver with perforated plate inserts as a function of Reynolds number - With perforated plate inserts - Improvement of collector thermal efficiency between 1.2 8% - Reduction of absorber tube temperature gradients up to 67% - Reduction of entropy generation rates up to 53% - Improvements in efficiency and reduction in entropy generation obtainable for flow rates lower than 36 m 3 /h and 44 m 3 /h respectively 61

62 Results -Receiver with perforated conical inserts Receiver with perforated conical inserts 62

63 Results -Receiver with perforated conical inserts Important dimensionless parameters - Dimensionless insert spacing pɶ = p/ L c (in the range: ) - Dimensionless insert size rɶ = 2 r / d c p ri Computational domain of a receiver with perforated conical inserts (in the range: ) - Dimensionless insert orientation angle ɶ β = β/ β c max (in the range: ) 63

64 Results -Receiver with perforated conical inserts Thermal Performance Nu pɶ c = 0.06 pɶ c = pɶ c = 0.14 pɶ 6000 c = 0.18 Nu p Re [ x 10 3 ] - Heat transfer performance and fluid friction increase with - increase in insert size - reduction in insert spacing - and increase in insert cone angle - Absorber tube circumferential temperature gradients reduce as heat transfer performance increases Heat transfer performance at an inlet temperature of 600 K, insert cone angle, ~ β ~ = 0.70 and insert size, r= 0.91 c c 64

65 Results -Receiver with perforated conical inserts - Nusselt number can be predicted from ( ɶ ) ( ɶ ) β ( ) ( ) Nu = p r ( tan ) Re Pr c c c For the range of parameters considered Nu (observed) % -12% Nu (predicted) 65 Nusselt number parity plot

66 Results -Receiver with perforated conical inserts - Friction factor can be predicted from ( ɶ ) ( ɶ ) ( ) f = p r 1+ sin β Re c c For the range of parameters considered % f (observed) % f (predicted) 66 Friction factor parity plot

67 Results -Receiver with perforated conical inserts 1 - Perforated conical inserts η th pɶ c = 0.06 pɶ c = 0.10 pɶ c = 0.14 pɶ = 0.18 c (η ) th p Re [ x 10 3 ] - Improve heat transfer performance for some range of Reynolds numbers - Improve collector thermal efficiency between 3-8% - Reduce absorber tube temperature gradients between % - Reduce entropy generation rates up to 53% Thermal efficiency for a receiver with perforated conical inserts at 650 K 67

68 Conclusions: Receiver thermal performance Key findings from this chapter include: Heat flux distribution is non-uniform around the absorber tube s circumference Rim angles lower than 60 o give high heat flux peaks With low flow rates, significantly higher temperature gradients are observed as concentration ratios increase as rim angles reduce With low rim angles (lower than 60 o ) thermal efficiency reduces by about 7.2% compared to values at 120 o Rim angles above 80 o show no significant increase in receiver heat loss and temperature gradients In general, rim angles lower than 60 o should be avoided especially at low flow rates and high concentration ratios 68

69 Conclusions: Receiver thermodynamic analysis Key findings from this chapter include: Nusselt number and friction factor are shown not to vary with concentration ratios However, second law analysis shows that changing concentration ratios and rim angles affects receiver thermodynamic performance High entropy generation rates exist in the part of the receiver with concentrated heat flux Entropy generation rates increase with increasing concentration ratio and reducing rim angles For rim angles greater than 80 0, thermodynamic performance does not change significantly as rim angles increase Entropy generation rates decrease with increasing temperatures There exist an optimal Reynolds number that minimises entropy generation rates Optimal flow rates do not change with inlet temperatures considered 69

70 Conclusions: Heat transfer enhancement For twisted tape inserts Heat transfer enhancement improves receiver thermal and thermodynamic performance* Thermal efficiency increases in the range 5-10% for twist ratios greater than 1.0 for twist ratios greater than 1.0 and all with ratios if flow rates are lower than 43 m 3 /h Absorber tube circumferential temperature differences reduce in the range 4-68% Entropy generation rates reduce by 59% for flow rates lower than 49 m 3 /h With multi-objective optimisation, optimal twist ratios of 0.42 and optimal width ratios of 0.65 were obtained (All objectives equally important) Combined use of multi-objective optimisation and thermodynamic optimisation is also demonstrated 70

71 Conclusions: Heat transfer enhancement For perforated plate inserts Heat transfer enhancement improves receiver thermal and thermodynamic performance* Thermal efficiency increases in the range 3-8% for insert spacing ( p~ ) in the range and insert size ( d ~ ) in the range if flow rates are lower than 43 m 3 /h Absorber tube circumferential temperature differences reduce up to 67% Entropy generation rates reduce by 53% for flow rates lower than 44 m 3 /h With multi-objective optimisation, optimal geometrical parameters were obtained and presented. 71

72 Conclusions: Heat transfer enhancement For perforated conical inserts Heat transfer enhancement improves receiver thermal and thermodynamic performance* Thermal efficiency increases in the range 3-8% for insert spacing ( p~c ) in the range and insert size ( r~c ) in the range if flow rates are lower than 37 m 3 /h Absorber tube circumferential temperature differences reduce up to 56% Entropy generation rates reduce by 45% With multi-objective optimisation, optimal geometrical parameters were obtained and presented (increase in thermal efficiency between 4-7% for flow rates lower than 37 m 3 /h ). 72

73 Recommendations Investigation of year round performance of the parabolic trough system for both non-enhanced and enhanced receivers Consideration of other heat transfer enhancement techniques Take into account the effect of any optical errors present in the system on heat flux distribution and temperature distribution. 73

74 Acknowledgements My supervisors(prof. T. Bello-Ochende and Prof. J.P Meyer) All academic and administrative staff in the Department of Mechanical and Aeronautical Engineering, University of Pretoria My entire family my dear wife, Ashimwe Charity Mwesigye; my Children Michelle J Atukundaand JoelMArinda;MyParents,Mr&MrsKamukama;myin-laws,Dr&Mrs Mfitundida All research students and friends in Thermoflow research group The funding received from the NRF, TESP and Stellenbosch University/University of Pretoria, SANERI/SANEDI, CSIR, EEDSM Hub and NAC is gratefully acknowledged and appreciated. 74

75 List of publications Mwesigye A., Bello-Ochende T. and Meyer J. P. Numerical investigation of entropy generation in a parabolic trough receiver at different concentration ratios. Energy 53(2013), Mwesigye A., Bello-Ochende T. and Meyer J. P. Thermodynamic performance of a parabolic trough receiver with centrally placed perforated plate inserts. Appl Energy 136(2014), Mwesigye A., Bello-Ochende T. and Meyer J. P. Minimum entropy generation due to heat transfer and fluid friction in a parabolic trough receiver with non-uniform heat flux at different rim angles and concentration ratios. Energy 73(2014), Mwesigye A., Bello-Ochende T. and Meyer J. P. Multi-objective and thermodynamic optimisation of a parabolic trough receiver with perforated plate inserts. Appl Therm Eng 77 (2015), Mwesigye A., Bello-Ochende T. and Meyer J. P., Thermal performance of a parabolic trough receiver with perforated conical inserts for heat transfer enhancement. In Conference Proceedings of ASME 2014 International Mechanical Engineering Congress and Exposition, IMECE2014, Nov 14-20, 2014, Montreal, Quebec, Canada, Paper ID: IMECE

76 List of publications Mwesigye A., le Roux W.G., Bello-Ochende T. and Meyer J. P., Thermal and thermodynamic analysis of a parabolic trough receiver at different concentration ratios and rim angles. In Conference Proceedings of the 10 th International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics, HEFAT2014, July 2014, Orlando, Florida. Mwesigye A., Bello-Ochende T. and Meyer J. P., Heat transfer enhancement in a parabolic trough receiver using perforated conical inserts. In Conference Proceedings of the 15th International Heat Transfer Conference, IHTC-15, August 10-15, 2014, Kyoto, Japan. Paper I.D. IHTC Mwesigye A., Bello-Ochende T. and Meyer J. P., Determination of heat flux and temperature distribution in a parabolic trough receiver at different rim angles and concentration ratios, In Conference Proceedings of the 2 nd Southern African Solar Energy Conference, SASEC 2014, January 27-29, 2014, Pine Lodge Resort, Nelson Mandela Bay, South Africa. Paper ID. 27. Mwesigye A., Bello-Ochende T. and Meyer J. P. Heat transfer enhancement in a parabolic trough receiver using wall detached twisted tape inserts. In Conference Proceedings of the ASME 2013 International Mechanical Engineering Congress and Exposition, IMECE2013, Nov 15-21, 2013, San Diego CA, USA. Paper ID: IMECE

77 List of publications Mwesigye A., Bello-Ochende T. and Meyer J. P. Thermodynamic performance of a parabolic trough receiver with centrally placed perforated plate inserts. In Conference Proceedingsofthe5 th InternationalConferenceonAppliedEnergy,ICAE2013,Jul 1-4, 2013, Pretoria, South Africa. Paper ID: ICAE Mwesigye A., Bello-Ochende T. and Meyer J. P., Numerical analysis of thermal performance of an externally longitudinally finned receiver for parabolic trough solar collector, In Conference Proceedings of the 9 th International conference on Heat Transfer, Fluid Mechanics and Thermodynamics, HEFAT 2012, pp , Malta, th July

78 Thank you

Aggrey Mwesigye. Submitted in partial fulfilment of the requirements for the degree. PHILOSOPHIAE DOCTOR in Mechanical Engineering.

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