ENHANCING HEAT TRANSFER IN AIR TUBULAR ABSORBERS FOR CONCENTRATED SOLAR THERMAL APPLICATIONS

Transcription

1 he final efinitive version of this manuscript as publishe in Applie hermal Engineering. ENHANCING HEA RANSFER IN AIR UBULAR ABSORBERS FOR CONCENRAED SOLAR HERMAL APPLICAIONS Yen Chean Soo oo a an Regano Benito b a CSIRO Energy echnology, 10 Murray Dyer Cct, Steel River Estate, Mayfiel West NSW 304 Australia b CSIRO Energy echnology, 11 Julius Avenue, North Rye NSW 113 Australia ABSRAC he evelopment of pressurise air tubular absorbers has been a major challenge for solar thermal applications involving high temperatures. he lo internal heat transfer coefficients that are expecte hen air is use as the heat transfer flui can be possibly alleviate through the use of enhancement methos such as helically coil/ire insert, tiste-tape insert an impling. A parametric stuy on the heat transfer an pressure rop performance of solar tubular air absorbers ith an ithout these heat transfer enhancements as conucte using a simplifie steay-state heat transfer moel ith bounary conitions representing typical solar applications. CO an helium as orking gases ere also consiere. his stuy shoe that tubes ith eeper protrusions on the inner all offer superior overall performance in terms of heat transfer an pressure rop compare to a plain tube an tube ith ifferent enhancement methos. he ifferential temperature beteen the tube all an bulk air coul be minimise. Consequently, tube lengths that can meet practical solar receiver sizes are possible for gas temperatures of over 800 C. he use of CO an Helium as orking gases can further reuce the total pressure rop in a solar tubular gas receiver. 1. INRODUCION o pressurize air receiver concepts for solar toers have been propose an investigate, namely, the irectly-irraiate volumetric receivers SOLGAE (Heller at. el, 006) an DIAPR (Kribus at. el, 001) an the cavity-type receivers ith metallic tubular absorbers SOLGAE an SOLHYCO (Amsbeck at. el, 008). he latter concept has remaine an attractive option ue to the simplicity in its construction an flexibility for scale-up, an to the likely loer overall cost as compare to the former option. Hoever, the esign an construction of pressurise tubular absorbers has been a major challenge in the evelopment of air-base solar technologies ue to the severe operating conitions involving high temperatures an pressures as ell as lo internal heat transfer coefficients using air as the heat transfer flui. When air temperatures above 800 C are targete in solar tubular absorbers, augmentation of tubular absorber performance becomes necessary to meet practical temperature limits of the tube material. In this paper, a brief revie an assessment of selecte enhancement techniques for the reuction of the air-sie heat transfer resistance in tubular absorbers are presente. hese techniques inclue internal enhancements through tube insertions of coil/ire, tiste tape an porous foams as ell as enhancements through impling of the tubes. A steay-state solar tubular receiver moel that takes into account the heat transfer across tube all an the air-sie

2 flui film an along the length of a tubular absorber is also evelope. he moel allos analytical comparison of a plain tube an selecte heat transfer enhance tubes in terms of heat transfer an pressure rop performance. CO an helium as orking gases are also consiere for comparison ith air.. HEA RANSFER ENHANCEMEN MEHODS he heat transfer process beteen the tube all an the orking gas insie the tubular absorber can be enhance by inucing higher egree of turbulence through the interruption of the evelopment of the bounary layer, enlargement of effective heat transfer area an generation of seconary/sirl flo. Most of the thermal enhancement techniques combine more than one mechanism that can lea to an increase in pressure loss. herefore, the relationship beteen the thermal an hyraulic performance of the enhance tube is an important consieration. ube inserts such as helical coil/ire (Figure 1a) an tiste tape (Figure 1b) ith various pitch ratios are simple an lo-cost passive techniques that improve heat transfer by isturbing the flui bounary layer at the all. In general, these techniques force separation an reattachment of the flui layer at the all causing higher egree of turbulence in the bulk flo. Helical coil inserts mainly isturb the bounary layer near the tube all hile tiste tapes isrupt the bounary layer at the tube all introucing superimpose vortex motion (sirl) into the bulk flo. Moifications that provie three-imensional surface roughness by knurling the inner tube surface to prouce internal fins an by using special inenting machines to prouce surface imples or protrusions offer promising heat transfer enhancement methos in the turbulent flo regions. Hoever, except for the latter metho, these are not suitable for high-temperature an high pressure applications ue to ifficulties in removing inner materials from high strength alloy metal tubes. Commercially available imple tubes (Figure 1c) are manufacture from stanar tubes in hich an externally applie inenting tool provies imples an protrusions on both sies of the all ithout removing any surface material. Each imple or protrusion acts as a passive vortex generator that effects tangential an raial mixing of the flui layer near the all an the bulk or core flui. In terms of heat transfer an pressure rop performance, Chen at el. (001) claime that the performance of imple tubes are superior to other conventional heat transfer augmentation esigns. Dimple tubes also offer improvement in mechanical strength an resistance to thermal expansion. he porous matrix of foam (Figure 1) typically has high surface-area ensity ith extensive contact surface beteen the soli an flui. he tortuous irregularly-shape flo passages continuously isrupts the bounary layers inucing a high egree of three-imensional flo mixing. he main isavantage of the tubular packe-be concept for solar air receiver systems is the high pressure rop across the porous be. he associate pressure rop can be still significantly high even for pores structure ith porosity higher than 90% compare to other enhancement techniques. ubular packe-bes therefore are not consiere in this stuy. 4

3 Figure 1 (a) Helical coil/ire insert, (b) tiste-tape insert, (c) imple tube, () porous foam. 3. SEADY-SAE HEA RANSFER MODEL Figure a shos schematically solar tubular absorbers vertically installe opposite the aperture plane in a cavity-type receiver. he absorbers consist of a bank of metal-alloy tubes hich connects in parallel to an inlet an an outlet manifols. he aperture of the cavity is tilte an facing onars alloing solar energy to be incient on only the front all of the absorbers. he heat transfer to the gas in tubular absorbers is governe by the conuctive heat transfer from the soli all in the raial, circumferential an axial irections an convective heat transfer from the all to the orking gas. A complete accounting of the thermal loss from the tube involves convection, conuction an raiation heat transfer from the absorber tubes to its environment such as the air insie the receiver, receiver alls an the ambient air.

4 Figure (a) A cavity-type solar receiver, (b) iscretisation of an axial segment of the absorber all. As shon in Figure b, the all of a tubular absorber is ivie into elemental noes ith thermal continuity at the surface bounaries of each element. he steay-state energy balance equations are then given as follos: For the absorber pipe, z r r r r k (1) here bounary conitions at the inner an outer surfaces of the absorber tube are: g inner h r k loss sol outer q q r k For the orking gas, g i g,g p h r z c m () here i g k Nu h 6

5 In all cases, a constant value for the thermal conuctivity of the pipe all ( k ) is assume. For each increment z along the length of the tube, the flo an bulk flui temperature are also assume uniform ith an inner tube circumferential average gas-sie heat transfer coefficient ( h ). he Nusselt number ( Nu ) correlations for pipe ith or ithout heat transfer enhancement techniques are use to preict the gas-sie heat transfer coefficient as given in able 1. he pressure rop across each tubular absorber is etermine as: P z g i f m Acs (3) here friction factor ( f ) is given in able 1. he pressure rop across each manifol is estimate as the sum of a series of -junctions. he pressure loss coefficients for these - junctions are given by Bassett et al. (001). able 1: Heat transfer correlations an pressure rop for ifferent tube configurations. Configuration Nusselt Number Friction Factor Reference Smooth tube f Re 1000Pr f log Re Gnielinski g Prg (1976) Nu f Pr. Pr g Re 300 Coil insert 716 e Nu 53 Re 15, 000 Re 100, p e p e p f 1. 38log Re Zhang at el. (1991) iste tape m 1 3 Nu C Re Pr p C.. e 31 m 8 375e p 5, 000 Re 5, p. 81 m f C Re 1. 81p C e 1. 3 p m 08 94e Chang at el. (007) Dimple tube m 1 3 Nu C Re Pr e e e C. N p e e e m. N p 7, 566 Re 51805, 031 e e p e 5 3 N 6 m f C Re e e e 013 C 067 N p e e e 6 m 098 N p Chen at el. (001)

6 For the purposes of this stuy, the receiver an aperture sizes are only selecte on the basis of overall requirements of the absorber an are not optimise for any solar fiel arrangement. he cavity is assume to be perfectly thermally insulate. Only the emission ( Q loss,ra ) an convection ( Q loss,conv ) losses term are therefore consiere in the moel. Raiation exchange is estimate beteen the irectly heate front surface of the absorber ( abs =9) an the aperture (assuming ap =5 C, ap =1, tilte angle of 45 from horizontal, r ap =5m). he corresponing raiation-netork equation is then given by (Holman, 001): 4 Aabs ap Q loss,ra (4) Aabs Aap Aabs Fabsap 1 A abs A A ap abs 4 F abs A abs ap ap ap here area-eighte average temperature of the heate all is use. With minimal clearance beteen ajacent tubes (1.3 o ) an beteen the tubes an the rear receiver all, the raiation exchange beteen ajacent tubes an beteen the tubes an the rear alls are assume negligible. he back alls of the tubes are mainly heate via circumferential thermal conuction from the heate front alls. he natural convection losses ( Q loss,conv ) through the aperture ithout in effects is estimate using the implicit metho from Leibfrie an Ortjohann (1995). Amsbeck at el. (009) reporte that their CFD preictions agree reasonably ell ith the Leibfrie s correlations. In this 8 stuy, natural convection insie the cavity receiver is in the laminar region ( Ra ). o heating patterns ere consiere in setting the temperatures of the solar heate front all of the tubular absorbers (, front ) as shon in Figure b: - Case 1: emperatures are increasing along the length of the tube from 800 C to 950 C (maximum alloable temperature for the selecte tube material). his represents a typical heating pattern for solar absorbers here the all temperatures in the inlet region are loer than those in the outlet region. - Case : emperatures are constant equal to 950 C along the length of the tube. his heating pattern represents a limiting case for a selecte tube material. Both energy balance equations for the soli (Eqn 1) an gas phases (Eqn ) are iscretise using the finite ifference metho. hese equations together ith gas-sie convective heat transfer an heat losses from the outer all of the absorbers are solve iteratively to achieve relative resiuals less than

7 4. RESULS AND DISCUSSION 4.1 Base case Plain tubular absorber A plain tubular absorber ith an inlet air mass flo rate of 01 kg/s an temperature of 400 C an being heate up to 800 C is consiere for the base case stuy. A nominal tube size of ¾ (Sch5) is also selecte. For all cases, the inlet gas pressure of 5 barg is use. Figure 3a shos the bulk air an all temperature istribution along smooth tubes for the to heating patterns. he absorbe energy to achieve the to temperature profiles are shon in Figure 3b. As may be expecte, tube all an air temperatures are higher ith case heating pattern. In orer to achieve an outlet air temperature of 800 C, reasonable length (about 3 m) of plain absorber tubes are neee. Hoever, higher outlet air temperatures may require much longer tubes involving heating patterns that are impractical for any solar fiel layout. It may also be note from Figure 3b that approximately 18% more heat is neee near the inlet of the tube in orer to achieve a constant temperature of 950 C along the length of the tube. Case represents a more severe operating conition for tubular absorbers. herefore, case heating pattern is applie in all later comparison stuies.

8 Figure 3 Base case analysis of solar tubular air absorbers for ¾ smooth tubes, ith one sie subjecte to incient solar flux to achieve bulk air of 400K at flo rates of 01 kg/s, (a) bulk air an all temperatures, (b) total absorbe heat along the length of absorbers. As shon in able, oubling the base case flo rate to 0kg/s (case A) may only require half the require number of absorber tubes. Hoever an increment of 33% in length is neee to achieve the same outlet air temperature. Furthermore, a substantial increase in total pressure rop (98%) can be expecte. his is mainly ue to higher pressure rop across the manifols. In able, it is also seen that increasing the tube size from ¾ (case A) to 1 (case B) for a flo rate of 0kg/s may not further reuce the pressure rop but longer absorber length is require. his makes it ifficult for plain tubes to be use in practical solar applications since long tubes may not be able to fit limite receiver space. It may be note that there is a slight increase in 10

9 thermal efficiency of the receivers ( rec ) as less amount of tubes are use ue to a reuction in reraiate heat losses. able Comparison of total pressure rop an efficiency of a 500 kw t receiver for ifferent mass flo rates per absorber tube Case m (kg/m 3 ) i (mm) L P (m) N pipe P total (kpa) rec (%) Base A B ubular absorber ith internal heat transfer enhancements A comparative stuy of heat transfer an pressure rop in ¾ solar absorber tubes incorporating either inserts or imple surface is carrie out for mass flo rates beteen 005 an 03kg/s (7,083 < Re < 4,499). Uner base case conitions, Nusselt numbers (Nu HE /Nu smooth ) an friction factors (f HE /f smooth ) relative to plain tubes for helically-coile ires (Coil 1 an Coil ) an smooth tiste tapes (ape 1 an ape ) are shon in able 3 for pitch ratios of.48 an 1.56 ith helix angles of 3 an 45 respectively. Both tiste-tape inserts offer better heat transfer enhancement than coil inserts. Hoever, the tape insert incurs higher pressure rop since it isturbs the entire flo fiel hile the coil insert only isturbs the flo near the all. able 3 also shos the Nu HE /Nu smooth an f HE /f smooth ratios for to imple tubes in hich Dimple 1 has less number of shalloer ents (N=3, e=6mm) an Dimple has more number of eeper ents (N=6, e=1.3mm). Dimple 1 can achieve heat transfer enhancement of 140% ith insignificant increase in pressure rop relative to the smooth tubes. As the imple epth an the number of imples increases, the gas-sie heat transfer coefficient increases ue to stronger effect of vortex interaction an mixing of bulk gas. able 3 shos that 50% improvement in heat transfer enhancement relative to plain tubes is achieve ith Dimple. Hoever the pressure rop is corresponingly ouble. able 3 Comparison of Nu HE /Nu smooth an f HE /f smooth ratios of a 500 kw t receiver for absorber tube ith an ithout heat transfer enhancement Case Configuration Nu HE /Nu smooth f HE /f smooth Base Smooth 1 1 Coil 1 p/=.48, e/= Coil p/=1.56, e/= ape 1 p/= ape p/= Dimple 1 e/=036, e/p=06, e/phi=1091, N= Dimple e/=0783, e/p=13, e/phi=3714, N=6.5 In Figure 4, the effect of mass flo rate on the heat transfer enhancement an friction factors in thermally enhance tubes are carrie out an compare to that of the smooth tube uner base case conitions. In both figures, the lines represent ata estimate from the reporte correlations

10 (able 1). he increase of mass flo rate or helix angle reuces the Nu HE /Nu smooth ratio an increases the f HE /f smooth ratio reflecting the iminishe sirling flo generate by both type of inserts ue to high momentum in the bulk gas. It can also be seen in Figure 4 that favourable heat transfer enhancement can be achieve in imple tubes at lo as ell as high flo rate conitions. For the investigate imple configurations, the value of Reynols number exponent in the Nusselt number correlation (able 1) is consistently higher (m =88-93) than those for smooth tubes (m =8). his suggests that, on the basis of heat transfer correlations from Chen s ork, imple tubes have better heat transfer performance compare to plain tubes. Figure 4 Comparison of tubes ith or ithout heat transfer enhancement, (a) ratio of heat transfer coefficients, (b) ratio of friction factors. A simple parameter calle performance ratio can be use to evaluate the overall performance of a thermally enhance tube relative to plain tube for the same compression poer. his is 1

11 expresse in terms of Nusselt numbers an friction factors as shon in Equation 6. A goo thermally enhance tube shoul have performance ratios much higher than Nu HE f HE HE (6) Nu smooth f smooth Figure 5 shos the performance ratios for the enhance tubes that ha been consiere earlier. It can be seen that at lo Reynols number (Re <10,000), performance in terms of heat transfer an pressure rop for tubes ith tiste tape inserts an imple tubes are close. As the flo rate increases, the imple tubes sho a steay improvement relative to the other enhancement techniques. Figure 5 Comparison of overall performance ratio for tubes ith heat transfer enhancement. Figure 6 shos comparisons of bulk air temperatures an absorbe heat in imple tubes (Dimple 1 ith N=3, e=6mm an Dimple ith N=6, e=1.3mm) an smooth tubes uner base case operating conitions. Dimple 1 an Dimple provie heat transfer enhancement up to 1.4 an.5 times respectively. In this stuy, only internal enhancement of imple tube is consiere here the protrusions on the inner tube surface generate the vortex flo mixing. If the esire outlet air temperature is set to be 800 C, substantial tube length reuction of 4% an 49% for imple tubes 1 an respectively (Figure 6a) can be achieve compare to plain tube. he total pressure rop for these tube configurations is generally similar an is mainly ue to pressure rop in the inlet an outlet manifols. For the case here higher outlet air temperature (>800 C) is esire, the use of plain tube ithout heat transfer enhancement as solar absorber may require long tube length in orer to achieve the temperature ifferential ithout exceeing the alloable material temperature of 950 C. In this case, it might not be practical particularly here parts of the absorber all at the further en from the cavity aperture may not be irectly expose to solar raiation. herefore, enhancing heat transfer insie the absorbers tube can be beneficial uner these operating conitions. For the basis of same tube length an total absorbe heat as the base case, enhance absorbers oul result in an increase outlet temperature to 856 C an 914 C for imple tubes

12 1 an respectively. Incorporation of enhance absorbers oul also require less number of tubes ue to higher absorbe heat per tube. As shon in Figure 7a, the total pressure rop can be reuce by 9.6% an 18.9% hen absorber tube length of.75m is use for the imple tubes 1 an respectively. he corresponing absorbe heat along the absorbers for these pipe configurations is shon in Figure 7b. Figure 6 Comparison for tubes ith or ithout imples, (a) air temperatures, (b) total absorbe heat. CO an helium ere consiere as orking gas for comparison ith air as shon in able 4. Complete system analysis that inclues other key components such as compressor, heat exchangers an thermal energy storage ere not carrie out. he poer (W c ) associate to the pressure rop across the tube absorbers ere estimate. Compare to air, CO have relatively higher ensity an slightly higher specific heat capacity hereas Helium ith loer ensity 14

13 offers better thermal properties 1. All thermal properties for these gases ere evaluate at 5 barg. As shon in able 4, a ramatic reuction in total pressure rop of 46% an 64.9% can be achieve hen CO an Helium respectively are use as the absorber orking gas for plain tubes an tubes ith imples. As also note earlier, the thermal efficiency of the receivers ( rec ) are not significantly affecte by using tubes ith or ithout thermal enhancements since the tube all temperatures an aperture size are all the same in all cases. his is also true hen tubes ith ifferent gases are use. able 4 Comparison of total pressure rop an efficiency of a 500 kw t receiver for tubes ith or ithout imples ith air an CO. ube type Gas m pipe (kg/s) L P (m) out ( C) Ptotal (kpa) W c (W) rec (%) Smooth Air Smooth CO Smooth He Dimple 1 Air Dimple 1 CO Dimple 1 He Dimple 1 Air Dimple 1 CO Dimple 1 He CONCLUSION In orer to improve the compactness an performance of solar tubular air receiver esign, selecte heat transfer enhancement an augmentation methos on the gas-sie of the absorber tubes ere reviee an investigate. A comparative assessment of solar tubular absorbers ith an ithout thermal enhancement in terms of heat transfer an pressure rop as conucte analytically. For plain tubes, loer flo rates an smaller tube sizes are neee to achieve acceptable rise in bulk air temperature across the length of the tube hile avoiing material failure an not exceeing practical temperature limits of the tube material. he investigation also shoe that tubes ith eeper protrusions on the inner all as can be foun in imple tubes offer superior overall performance in terms of heat transfer an pressure rop in the higher flo region compare to that of investigate tubes ith or ithout insertions such as helical coils an tiste tapes. Improvement on the total pressure loss of the absorber can also be achieve by the use of CO an Helium as the receiver orking flui ithout significant increase in the receiver efficiency. Hoever, this is achieve at the expense of higher compression poer in the case of helium. NOMENCLAURE A abs outer surface area of the absorber tube (m ) A ap aperture area of the cavity-receiver (m ) 1 For temperatures range of 400 to 800 C at 5 barg, the ensity an specific heat capacity ratios of CO an Helium compare to air are an for CO, an 138 an for Helium.

14 A cs cross-sectional area of the inner tube (m ) c p specific heat capacity at constant pressure (J/kg-K) c iameter of coil insert (m) i e f inner iameter of the tube (m) epth of imple (m) friction factor F vie factor abs ap h average heat transfer coefficient (W/m K) k thermal conuctivity (W/mK) L cav cavity ith (m) L p absorber pipe length (m) m N N p mass flo rate (kg/s) number of longituinal imple columns number of absorber tubes Nu p average Nusselt number pitch of imples (m) P tot total pressure rop (kpa) Pr Pranlt number q heat flux (W/m ) Q r heat rate (W) raius of the tube (m) r ap raius of the cavity aperture (m) r o outer raius of the tube (m) Re Reynols number 16

15 Ra Rayleigh number temperature ( C) ap temperature at the cavity aperture ( C) b bulk air temperature in the cavity ( C) c air temperature from the top of the cavity aperture ( C) a ambient temperature ( C) all temperature ( C) W c z compressor poer (W) axial irection of the tube Greek symbols abs emissivity of the absorber surface ap emissivity of the cavity aperture iameter of imples (m) ensity of the flui (kg/m 3 ) circumferential angle of the tube ( ) rec efficiency of the receiver (%) Subscripts HE heat transfer enhancement in out loss sol inlet outlet heat losses incient solar pipe all

16 g ra conv gas raiation convection REFERENCES 1. Amsbeck, L., Helsch, G., Roger, M. an Uhlig, R., 009, Development of A Broaban Antireflection Coate ransparent Silica Wino for A Solar-Hybri Microturbine System, In Proceeings of SolarPACES, Berlin, Germany.. Amsbeck, L., Buck, R., Heller, P. an Uhlig, R., 008, Development of A ube Receiver for A Solar-Hybri Microturbine System, In Proceeings of 14 th Biennial CSP SolarPACES Symposium, Las Vegas, NV USA. 3. Bassett, M.D., Winterbone, D.E. an Pearson, R.J., 001, Calculation of Steay Flo Pressure Loss Coefficients for Pipe Junctions, Proc. Instn. Mech. Engr, vol. 15, pp Chang, S.W., Jan, Y.J. an Liou, J.S., 007, urbulent Heat ransfer an Pressure Drop in ube Fitte ith Serrate iste ape, Int. J. hermal Sciences, vol. 46, pp Chen, J., Muller-Steinhagen, H. an Duffy, G.G., 001, Heat ransfer Enhancement in Dimple ubes, Applie hermal Engineering, vol. 1, pp Gnielinski, V., 1976, Ne Equations for Heat an Mass ransfer in urbulent Pipe an Channel Flo, International Chemical Engineering, vol. 1, pp Heller, P., Pfaner, M., Denk,., ellez, F., Valvere, A., Fernanez, J. an Ring, A., 006, est an Evaluation of A Solar Poere Gas urbine System, Solar Energy, vol. 80, pp Holman, J.P., 001, Heat ransfer, 8 th SI Metric Eition, McGra-Hill, Ne York. 9. Kribus, A., Doron, P., Rubin, R., Reuven, R., aragan, E., Duchan, S. an Karni, J., 001, Performance of the Directly-Irraiate Annular Pressurise Receiver (DIAPR) Operating at 0 Bar an 100 C, ASME J. Solar Energy Eng, vol. 13, pp Leibfrie, U. an Ortjohann, J., 1995, Convective Heat Loss from Upar an Donarfacing Cavity Solar Receivers: Measurements an Calculations, ASME J. Solar Energy Eng, vol. 117, pp Zhang, Y.F., Li, F.Y an Liang, Z.M., 1991, Heat ransfer in Spiral-Coil-Inserte ubes an its Application, Avances in Heat ransfer Augmentation an Mixe Convection ASME, vol. 169, pp