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A study of heat transfer enhancement in a new solar air heater having circular type turbulators
Accepted Manuscript A study of heat transfer enhancement in a new solar air heater having circular type turbulators Adem Acır, İsmail Ata PII:
S1743-9671(14)20429-6
DOI:
10.1016/j.joei.2015.05.008
Reference:
JOEI 164
To appear in:
Journal of the Energy Institute
Received Date: 3 December 2014 Revised Date:
25 May 2015
Accepted Date: 27 May 2015
Please cite this article as: A. Acır, İ. Ata, A study of heat transfer enhancement in a new solar air heater having circular type turbulators, Journal of the Energy Institute (2015), doi: 10.1016/j.joei.2015.05.008. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ACCEPTED MANUSCRIPT
A STUDY OF HEAT TRANSFER ENHANCEMENT IN A NEW SOLAR AIR
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Adem Acıra*), İsmail Atab)
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HEATER HAVING CIRCULAR TYPE TURBULATORS
a,*)
Gazi University, Faculty of Technology, Department of Energy Systems Engineering Teknikokullar - Ankara - TURKEY
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Telephone: +90-312-202 86 05 Fax: +90-312-202 89 47
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E-mail: [email protected]
b)
Gazi University
Institute of Science and Technology, Ankara, Turkey
a,*)
Corresponding Author
ACCEPTED MANUSCRIPT ABSTRACT In this study, the heat transfer, friction factor and thermal performance factor characteristics of a new solar air heaters (SAHs) with circular type turbulators having different relief angles and distances were performed. Effect of the pitch ratio (PR) and angle ratio (AR) were investigated to
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improve in heat transfer in SAHs in a range of between 3000 and 7500 Reynolds number under solar radiation heat flux. The experimental results obtained using various turbulators were compared with conventional plain tube. Nusselt numbers were increased with increasing Reynolds
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number whereas; it was decreased with increasing PR and AR. The friction factor and thermal
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performance factor was decreased with increasing Reynolds number, PR and AR. It was found that the turbulators arrangement with PR = 2 and AR=0.125 shown the maximum heat transfer enhancement of 416 %, friction factor of 511 % than conventional plain tube and the thermal performance factor was obtained as ~2.9. In addition, the empirical equations for heat transfer, friction factor and thermal performance factor were derived and compared with the experimental
error.
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results. The predicted results indicated a good agreement with experimental results within 10 % of
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factor
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Keywords: Heat transfer enhancement, Turbulator, Solar air heater, Friction factor, Enhancement
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ACCEPTED MANUSCRIPT 1. Introduction In solar energy applications, the solar air heaters (SAHs) were commonly used as heat exchanger. The incoming solar radiations into thermal energy were converted with SAHs. The converting
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thermal energy was extracted by air flowing under the absorbing surface. However, the heat transfer to the flowing air was obstructed due to generation laminar sublayer over the heattransferring surface. Therefore, the thermal performance of SAHs was negatively affected by the formation laminar sublayer. The artificial surface roughness was used to eliminate negative effect
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in SAHs [1]. Namely, the artificial roughness having form of repeated ribs was used to improve
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heat transfer in SAHs. The heat transfer enhancement characteristics of solar air heater with roughness elements on the surface were investigated by some researchers. Hans et. al., [1] investigated out to improve of heat transfer coefficient and friction factor with multiple V-rib roughness in solar air heater duct. Hans et. al., carried out the effect relative roughness height (e/D), pitch (P/e), width (W/w) and angle of attack (a) in a range of Reynolds number (Re) from
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2000 to 20000. The heat transfer and friction factor in the convection flow air in SAHs having rectangular duct were investigated by Bhagoria et. al., [2] for Reynolds number range from 3000
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to 18000. Kumar et. al.,[3]., performed the improvement of the heat transfer in the SAHs with artificial roughness in the form 60° inclined discrete rib for relative roughness height (e/D), pitch
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(P/e) and gap posistion (d/W) for Reynolds number (Re) varied from 4105 to 20526. The enhancement of heat transfer characteristics a rectangular duct roughened SAH in the relative roughness height (e/D), width (W/w), pitch (p/e) and arc angle (a) were investigated by Singh et al., [4] for Reynolds number (Re) in the range of 2200–22,000. The effect of relative roughness height 0.0181–0.0363; relative roughness pitch 4.5–10.0, and groove position to pitch ratio 0.3– 0.7 to improve of heat transfer was performed by Ref. [5] and the obtained results were compared for SAHs having ribbed and smooth duct. Sethi et al. [6] conducted experimental study to enhancement of heat transfer in SAHs having dimple shaped roughness. Also, empirical equations for Nusselt number and friction factor were developed and compared with experimental results. 3
ACCEPTED MANUSCRIPT The thermal hydraulic performance analysis in SAHs with 60 V-down discrete rib roughnesses was performed by Karwa and Chitoshiya [7]. The heat transfer enhancement characteristics in SAHs with 60 V-down discrete rib roughness performances were investigated. The thermal performance analysis of a flat-plate solar collector was investigated by Ref. [8] with twisted strip,
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coil-spring wire and conical ridges. Skullong et. al.,[9] presented an experimental study on heat transfer characteristics in a SAH channel fitted with combined wavy-rib and groove turbulators for different rib-pitch to channel-height ratios (PR) with a single rib-to-channel height ratio (BR) for
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Reynolds numbers in the range of 4000 to 21.000. The heat transfer characteristics with the form of specially prepared inverted U-shaped turbulators on the absorber surface of an SAHs was
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performed by Ref. [10]. The heat transfer characteristics in the SAHs having corrugated trapeze, reverse corrugated, reverse trapeze, and a base flat-plate collector was investigated by Benli [11].
In previous SAHs studies, various designs with different rib roughness geometry in flat plate type
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solar air collectors were developed. In this study, a new type of SAHs having semi circular absorber plate including various type turbulators was designed and fabricated as differently from previous SAHs studies. In this experimental study, three different copper tubes inside of the SAHs
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were arranged and turbulators was inserted inside the tube. Effect of the pitch ratio (PR) and angle ratio (AR) were investigated to improve in heat transfer in SAHs in a range of between 3000 and
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7500 Reynolds number. Nusselt number, friction and thermal performance factor results obtained in various PR and AR were compared with obtained results in conventional plain tube in a range of between 3000 and 7500 Reynolds number. Also, the empirical equations for Nusselt number, friction factor and thermal performance was developed and compared with experimental results. 2. Experimental set-up The SAHs experimental setup was shown in Fig. 1. The experimental set-up, diagram with parameters (L, D, α) of circular type turbulators and details of the SAHs having semi circular
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ACCEPTED MANUSCRIPT type’s absorber plates were given in Figs. 1-3. The collector dimensions were 800 mm × 450 mm × 200 mm. The absorbing surfaces were produced with a dull black painted galvanized copper with a thickness of 0.5 mm. In this study, a new approach in SAHs as different from literature was applied. The tube and turbulator diameters were 50 and 50 mm, respectively. Turbulators having
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various obstacle relief angles and distances were inserted inside the tube for working fluid (air) circulations. The ideas/criteria for designing various type turbulators was to increase in heat transfer and pressure drop. In this study, the circular type turbulators were considered. In the next
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studies, the other type’s turbulators which has conical-ring, conical-nozzle, V-nozzle, twisted-tape and screw-tape will applied to compare of enhancement of the heat transfer. The circular type
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turbulators were designed and fabricated as various obstacle relief angles (α= 45°, 90°, 135° and plain tube) and obstacle distances (L= 100 mm, 140 mm, 175 mm) to create of the turbulence flow and enhance of heat transfer rates. Turbulators dimensions and ratios using experimental tests in the SAHs were given in Table 1. The Reynolds number used by experimental tests was in a range
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of between 3000 and 7500. The sides and back of the collector were isolated to prevent in thermal losses. The collector was placed facing south with a slope angle of 30°. The solar radiation incident on the surface was measured with a Solorimeter (KIMO-SL100) and was recorded at
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noon. The ambient temperature, collector input, output and surface temperatures were measured with K type thermocouple. The inlet and outlet temperatures in the collectors, ambient and
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absorber plate temperatures were measured with digital data logger (AMATEK ASM 802-B) and recorded from several selected locations. The pressure loss and velocity of the air were measured by a digital manometer (BEAMEX-EXT200MCS) and the anemometer (LUTRON AM-4202), respectively. 3. Calculation Method SAHs are irradiative heat exchangers. Solar radiation energy converted to heat in the collector. The heat obtained solar radiation energy was transferred by convection from the absorber to the
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ACCEPTED MANUSCRIPT working fluid. In this study, the air as the working fluid (air) was used. At noon, the maximum instantaneous radiation energy was measured as 1100 W/m2 and used in the heat and flow friction characteristic calculations in SAHs. In this experimental study, the copper tubes inside the solar air collector were arranged as semi insulated tube. In the test section in SAHs, the useful heat gain
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( Qu ) of the air is calculated as [12-18]:
Qu = m& .cp .(Tout − Tin )
(1)
Qu Ap .(Ts − Tb )
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h=
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The heat transfer coefficient for the test section is:
(2)
where the bulk temperature of the working fluid (air), Tb, can be found by averaging the inlet and outlet temperatures of the air [12-18]. The physical properties of the air can be determined by Tb: (3)
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Tb = (Tout + Tin ) / 2
and Ts is the average surface temperature of the SAHs absorber surface. The heat transfer
h.D k air
(4)
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Nu =
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coefficient and Nusselt number were calculated as below [12-18]:
The Reynolds number was written by Re =
U.D
(5)
ν
Friction factor, f can be written as
f =
∆P 2 L U ρ D 2
(6)
in which U is mean air velocity in the copper tube inserted in SAHs. 6
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For a constant pumping power [12-18], (V∆P ) p = (V∆P ) t
(7)
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and the relationship between friction and Reynolds number were given as below [12-18]:
( f Re3 ) p = ( f Re3 ) t
(8)
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At the same pumping power, the thermal enhancement efficiency (η), were given as below by Ref [12-18]:
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Nut Nu p η= 1 ft 3 f p
(9)
Also, the uncertainties of measurements in this study were determined using Kline and
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McClintock uncertainity method [19]. The uncertainty during the measurements of the parameters was given in Table 2. Also, the maximum uncertainties of non-dimensional parameters were
conditions.
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obtained as ± 10 % for Reynolds number, ±7 % for Nusselt number and ± 10% for friction for all
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4. Results and Discussions
4.1.Validation test of plain tube In this study, the effects of the circular type turbulators inducing swirl flows on the heat transfer enhancement (Nusselt number, Nu), friction factor (f) and thermal performance factor (η) in SAHs were investigated. In the calculations, angle ratio (AR) and pitch ratio (PR) were taken into consideration as shown in Fig. 4 and 5. The Reynolds number was changed between 3000 and 7500. The pitch ratios (PR) was 2, 2.8, 3.5 whereas, the angle ratios (AR) was 0.125, 0.25 and 0.375. The experimental results for the heat transfer rate and the friction factor in a circular plain tube are obtained and confirmed with the Eqs. (10)-(13) correlations [12]. Nusselt number with the 7
ACCEPTED MANUSCRIPT experimental data’s obtained from the plain tube in SAHs was compared using Gnielinski [12,16] and Kays&Crowford equations [11,17]. Also, the friction factor with the experimental data’s obtained from the plain tube was compared using Petukhov [12,16] and Blasius [12,16] for the experimental system validations. The correlation equations used to provide experimental system
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validations were given as below: Gnielinski correlation for Nusselt number,
( f 8)(Re− 1000) Pr 2 1 1 + 12.7( f 8) 2 (Pr 3 − 1)
3000 < Re < 5x106
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Nu =
Nu = 0.0158 Re 0.8 Petukhov correlation for friction factor,
f = (0.79 ln Re− 1.64) −2
f = 0.316 Re −0.25
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Blasius correlation for friction factor,
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Kays&Crowford correlation for Nusselt number,
(10)
(11)
3000 < Re < 5x106
(12)
3000 < Re < 20,000
(13)
The correlation Eqs.(10-13) results used to provide experimental system validations were
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compared with experimental plain tube data’s. The Nusselt number deviations between present
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study and Gnielinski and Kays&Crowford were obtained within ±15% whereas; the friction factor deviations according to Petukhov and Blasius correlations were computed as ±11%. The experimental plain tube data’s were found to be in good agreement with the previous correlations given of Eqs. (10) – (13) for both the Nusselt number and friction factor as shown in Fig. 4a and 5a.
4.2.Heat transfer The circular type turbulators having obstacle relief angles and distances in SAHs were used to improve in the heat transfer. The circular type turbulators having obstacle relief angles and 8
ACCEPTED MANUSCRIPT distances created boundary layer of the air flow and increase the turbulence in the air flow. Effects of the pitch ratio (PR), angle ratio (AR) and various Reynolds numbers on the Nusselt number and Nusselt number ratio (Nut/Nup) were illustrated in Fig. 4a and 4b. As shown in Fig. 4a, the Nusselt numbers were increased with increasing Reynolds number whereas; it was decreasing with
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increasing PR and AR. It can be seen that the enhancement of the heat transfer in the SAHs having turbulators were significantly higher than obtained results in plain tube. The increase of enhancement of the heat transfer with decreasing PR and AR was due to the increased heat
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transfer surface area and prolonged residence time of the air flow in the tube in SAHs. As investigated in Fig 4., the maximum heat transfer enhancement was obtained with the use of the
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circular turbulator at the smallest pitch ratio (PR= 2.0) and angle ratio (AR= 0.125) for between 3000 and 7500 Reynolds number. For the PR=2 and AR=0.125, the increase in the heat transfer was between 416 % and 258 % in a range of between 3000 and 7500 Reynolds number, respectively, according to conventional plain tube. At the same Reynolds number, the increase in
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the heat transfer was between 339 % and 190 % for the PR=3.5 and AR=0.125 according to conventional plain tube. It was seen that the use of the circular type turbulator was shown a significant increase in Nusselt number according to the plain tube. In the PR=2.0, the Nusselt
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number for AR=0.125 were improved ~19 % and ~33 % according to AR= 0.375 for 3000 and 7500 Reynolds numbers, respectively. At the same Reynolds numbers, the Nusselt number
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obtained with AR=0.125 was higher than AR=0.375 at about ~10% and ~37% for PR=3.5, respectively. Evaluations from PR=2 to PR=3.5 for AR=0.125, the Nusselt numbers were decreased as ~23% and ~18% for 3000 and 7500 Reynolds numbers, respectively. These results were shown that the heat enhancement factor of AR ratios was higher than PR ratios. In addition, Nusselt number ratio (Nut/Nup) against the Reynolds number was illustrated in Fig. 4b. It can observe that the Nusselt number ratio was decreased with increasing Reynolds number from 3000 to 7500 for all conditions. The highest Nusselt number ratio obtained with low Reynolds number. Also, Nusselt number ratio decreased with the rise of PR and AR as shown Fig. 4b. The best 9
ACCEPTED MANUSCRIPT Nusselt number ratio was obtained as 5,16 and 3,58 with at the smallest pitch ratio (PR= 2.0) and angle ratio (AR= 0.125) for between 3000 and 7500 Reynolds number, respectively. The Nusselt number ratios of all conditions were above unity. This situation provided an advantage the use of
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circular type turbulators according to the plain tube.
4.3.Friction factor
The friction factor (f) and the friction factor ratio (ft/fp) plotted against the Reynolds number
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values were shown in Figs. 5a and 5b. The friction factor was decreased with the increasing Reynolds number, PR and AR as shown in Fig. 5a. However, the friction factor was increased
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with decreasing PR and AR depends on circular type turbulators having obstacle relief angles and distances. The fiction factor was higher than obtained values in plain tube as illustrated in Fig. 5a. The increase in friction factor was resulted by the increased contact surface area and prolonged residence time of the flow in the tube in SAHs. Because, the turbulators can be created the
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dynamic pressure of the fluid with increasing contact area and number of obstacle depends on PR and AR. As shown in Fig. 5a, the maximum friction factor was obtained as 511 % and 295 % in a range of between 3000 and 7500 Reynolds number, respectively, according to conventional
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plain tube for the PR=2 and AR=0.125. The average friction factor obtained by the circular turbulator with the lowest pitch ratio (PR= 2.0) was higher than the circular rings having PR= 2.8
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and 3.5 at around 6% and 17%, respectively, for AR=0.125. The average friction factor generated by the circular turbulator having the lowest angle ratio (AR= 0.125) was higher than the circular rings with AR = 0.25 and 0.375 at around 19% and 60%, respectively, for PR= 2. It was seen that the friction factor values of AR ratios were higher than PR ratios.
The changes of the friction factor ratio (ft/fp) with the Reynolds number values were shown in Fig. 5b. The friction factor ratios with increasing Reynolds number, PR and AR were decreased. As investigated in Fig. 5b, the average friction factor ratio obtained by using turbulators with at the
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ACCEPTED MANUSCRIPT smallest pitch ratio (PR= 2.0) and angle ratios (AR= 0.125 and 0.375) were found to be about 5,1 and 3.2 times from the plain tube. It was observed that the use of the circular type turbulators provided an increase in the friction factor ratio than plain tube under the same conditions.
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4.4.Thermal Performance Factor
The thermal performance factor was shown in Fig. 6 with the rise of Reynolds number from 3000 to 7500. The thermal performance factor was evaluated for the same pumping power criteria. As
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shown in Fig. 6., the thermal performance factor decreased with increasing Reynolds number, PR and AR. The thermal performance factors above ~2 were obtained for in all the cases. The
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maximum thermal performance factor was computed as ~2.9 and ~2.3 in the PR=2 and AR= 0.125 for between 3000 and 7500 Reynolds number, respectively. The thermal performance factor for PR=2 and AR = 0.125 and 0.375 were obtained as ~2.9 and 2.6 for 3000 Reynolds number. The thermal performance factor was raised as ~ 8% with decreasing AR. The thermal performance
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factor for PR=3.5 was computed as ~2.5 and ~1.9 with AR= 0.125 for between 3000 and 7500 Reynolds number, respectively. On the other hand; the thermal performance factor for AR=0.125 and PR=2 and 3.5 were computed as ~2.9 and 2.5 for 3000 Reynolds number, respectively. The
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thermal performance factor was increased as ~ 13% with decreasing PR. The average thermal performance factor obtained with the lowest angle ratio (AR= 0.125) was higher than AR= 0.25
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and 0.375 at around 1% and 7%, respectively, for PR=2. The average thermal performance factor with lowest pitch ratio (PR= 2) was higher than PR = 2.8 and 3.5 at around 13% and 14%, respectively, for AR= 0.125.
4.5. Correlations Nusselt number, friction number and thermal performance factor are strong function of a system and operation parameter. Functional relationship for Nusselt number, friction number and thermal performance factor can be written as below: 11
ACCEPTED MANUSCRIPT Nu = f1 (Re, PR, AR)
(15)
f = f 2 (Re, PR, AR)
(16)
η = f 3 (Re, PR, AR)
(17)
Nusselt number, friction factor and thermal performance factor correlations was developed for
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range of system and operation parameter with regression analysis. The correlations obtained from experimental results in a range of Reynolds number between 3000 and 7500 was shown in Eqs. (18)–(20), respectively. The predicted results from Eqs (18)–(20) were compared with
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experimental results and the obtained comparatively results were illustrated in Figs. 7–9. The
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predicted values obtained empricial equations have indicated in good agreement with experimental results. The maximum deviation of the between predicted and experimental results were within ± 10% for Nusselt number, friction factor and thermal performance factor, respectively.
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Nu = 0.127 Re 0.764 ( PR ) −0.356 ( AR ) −0.160 f = 50 .089 Re −0.713 ( PR ) −0.297 ( AR ) −0.409
(19) (20)
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η = 10 .006 Re −0.155 ( PR ) −0.230 ( AR ) −0.05
(18)
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5. Comparison with previous work In this section compared the thermal performance characteristics of a SAHs fitted with circular type turbulators with previous SAHs studies. In previous SAHs studies, various designs with different rib roughness geometry (transverse, V-shape, V shaped inclined discrete W-shape and inclined rib with gap viz.) and dimensions of the air flow passage in flat plate type solar air collectors were developed in order to increase the thermal performance of SAHs. In this study, a new type of SAHs having various type turbulators was designed as differently from previous SAHs studies. Turbulators as swirl generators which are conical-ring, conical-nozzle, V-nozzle, twisted-tape and screw-tape were used to improve of heat transfer with constant heat flux electric 12
ACCEPTED MANUSCRIPT heater on heat exchangers without solar radiation in earlier investigations [14-16, 23]. In order to see the advantage of turbulators of SAHs having semi circular type’s absorber plate, the present work was performed and compared with previous SAHs works as shown in Table 3. As investigated in Table 3, the thermo-hydraulic performance parameters of the present study have
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indicated in a good performance compared to previous studies. Thus, the use of the turbulator as well as the different rib roughness geometry used in SAHs to heat transfer enhancement was
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proposed and investigated.
6. Conclusions
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In this study, the effects of circular type turbulators having various obstacle relief angles (AR) and pitch ratio (PR) on heat transfer, flow friction and thermal performance factor characteristics in SAHs were investigated. The main results of this study are summarized below:
The heat transfer enhancement in the SAHs having turbulators was significantly higher
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than obtained in conventional plain tube. The maximum heat transfer enhancement obtained with decreasing pitch and angle ratios in the PR=2 and AR= 0.125. The maximum increase in the heat transfer was in a range of 416 % and 258 % for about 3000
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and 7500 Reynolds numbers, respectively, according to plain tube.
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The friction factor decreased with the increasing Reynolds number, PR and AR. The maximum friction factor was obtained as 511 % and 295 % in a range of between 3000 and 7500 Reynolds number, in the PR=2 and AR= 0.125 according to conventional plain tube.
The thermal performance factor decreased with the increasing Reynolds number, PR and AR. The maximum thermal performance factors computed as ~2.9 and ~2.3 in the PR=2 and AR= 0.125 for between 3000 and 7500 Reynolds number, respectively.
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ACCEPTED MANUSCRIPT Nusselt number, friction factor and thermal performance factor were correlated and developed the empirical equations. The predicted values obtained empirical equations have indicated in good agreement with experimental results within ±10 %.
As a result, the maximum heat transfer enhancement obtained using turbulators in SAHs
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was found at higher Reynolds number, lower pitch and angle ratios. Also, the effect of the
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heat enhancement factor of AR was higher than PR in the general evalutions.
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ACCEPTED MANUSCRIPT Nomenclature
Ap
Absorber plate area of collector, m2
AR
Angle ratio
cp
Specific heat of air, j/kg°C
Greek letters
D
Diameter, mm
α
Obstacle clearance angle
f
Friction factor
ν
Kinematic viscosity, m2/s
h
Heat transfer coefficient, W/m2°C
ρ
Fluid density, kg/m3
I
Solar radiation W/m2
k
Thermal conductivity, W/m2°C
L
Lenght, mm
m&
Mass flow rate, kg/s
Nu
Nusselt number
P
Fluid pressure, Pa
Pr
Prandtl number
PR
Pitch ratio
Re
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Air velocity, m/s
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U
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Subscripts
Air
b
Bulk
e
Environment
in
Inlet
out
Outlet
p
Plain tube
Reynolds number
s
Surface
Qu
Useful heat gain, W
t
Turbulator
T
Temperature, °C
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EP
TE D
a
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ACCEPTED MANUSCRIPT Table 1. Turbulator dimensions using experimental tests
45°
0,125
90°
0,25
135°
0,375
Length (L),mm 100 140 175 100 140 175 100 140 175
Diameter (D),mm 50
50
Pitch Ratio (PR=L/D) 2 2,8 3,5 2 2,8 3,5 2 2,8 3,5
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Angle Ratio (AR=α/360)
50
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Obstacle relief angle (α)
Table 2. The uncertainty during the measurements of the parameters Unit
Collector inlet temperature Collector outlet temperature Absorber surface temperature Ambient temperature
Time Solar radiation
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Pressure loss
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Air flow velocity
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Parameter
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Uncertainty (δ)
0
C
± 0.10
0
C
± 0.10
0
C
± 0.10
0
C
± 0.10
m/s
± 0.02
min
± 0.10
W/m2
± 1,26
Pa
± 0,10
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Table 3. Comparison of optimum values of performance parameter for various enhancement geometries in SAHs. Enhancement Element
Nu
f
η
Present study
Circular turbulators and semi-circular absorber plate
5,16
6,11
2,82
Bhagoria et al. [2]
Transverse wedge shaped ribs
2.4
5.3
1.38
Singh and Saini [21]
W down rib
2.7
2.86
1.94
Jaurker et al. [5]
Rib groove
2.75
3.61
1.76
Prasad and Saini [20]
Transverse wire
2.38
4.25
1.78
Bopche and Tandale [10]
Inverted U shape rib
2.82
3.72
1.82
Momin et al. [22]
V sahaped rib
2.30
2.83
1.78
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Investigation
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FIGURE CAPTIONS
Fig. 1. Schematic view of experimental set-up Fig. 2. Sectional view of absorber plate in experimental setup
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Fig. 3. Schematic diagram with parameters of circular type turbulators
Fig.4. Changes of (a) Nusselt number and (b) Nut/Nup with Reynolds number. Fig.5. Changes of (a) Friction factor and (b) ft/fp with Reynolds number.
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Fig.6. Changes of thermal performance factor with Reynolds number Fig. 7. Comparison of experimental and predicted Nusselt number
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Fig. 8. Comparison of experimental and predicted friction factor
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Fig. 9. Comparison of experimental and predicted thermal performance factor
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Fig. 1. Schematic view of experimental set-up.
Fig. 2. Sectional view of absorber plate in experimental setup
Fig. 3. Schematic diagram with parameters of circular type turbulators
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Fig.4. Changes of (a) Nusselt number and (b) Nut/Nup with Reynolds number.
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Fig.5. Changes of (a) Friction factor and (b) ft/fp with Reynolds number.
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Fig.6. Changes of thermal performance factor with Reynolds number
Fig. 7. Comparison of experimental and predicted Nusselt number
25
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Fig. 8. Comparison of experimental and predicted friction factor
Fig. 9. Comparison of experimental and predicted thermal performance factor
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ACCEPTED MANUSCRIPT Highlights
-
The heat transfer enhancement in a new Solar Air Heaters (SAHs) having various picth and angle ratios were investigated. The highest heat transfer enhancement was provided with decreasing picth and angle
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ratios.
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The effect of the heat enhancement of pitch ratio was higher than angle ratio.
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S1743-9671(14)20429-6
DOI:
10.1016/j.joei.2015.05.008
Reference:
JOEI 164
To appear in:
Journal of the Energy Institute
Received Date: 3 December 2014 Revised Date:
25 May 2015
Accepted Date: 27 May 2015
Please cite this article as: A. Acır, İ. Ata, A study of heat transfer enhancement in a new solar air heater having circular type turbulators, Journal of the Energy Institute (2015), doi: 10.1016/j.joei.2015.05.008. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
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A STUDY OF HEAT TRANSFER ENHANCEMENT IN A NEW SOLAR AIR
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Adem Acıra*), İsmail Atab)
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HEATER HAVING CIRCULAR TYPE TURBULATORS
a,*)
Gazi University, Faculty of Technology, Department of Energy Systems Engineering Teknikokullar - Ankara - TURKEY
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Telephone: +90-312-202 86 05 Fax: +90-312-202 89 47
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E-mail: [email protected]
b)
Gazi University
Institute of Science and Technology, Ankara, Turkey
a,*)
Corresponding Author
ACCEPTED MANUSCRIPT ABSTRACT In this study, the heat transfer, friction factor and thermal performance factor characteristics of a new solar air heaters (SAHs) with circular type turbulators having different relief angles and distances were performed. Effect of the pitch ratio (PR) and angle ratio (AR) were investigated to
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improve in heat transfer in SAHs in a range of between 3000 and 7500 Reynolds number under solar radiation heat flux. The experimental results obtained using various turbulators were compared with conventional plain tube. Nusselt numbers were increased with increasing Reynolds
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number whereas; it was decreased with increasing PR and AR. The friction factor and thermal
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performance factor was decreased with increasing Reynolds number, PR and AR. It was found that the turbulators arrangement with PR = 2 and AR=0.125 shown the maximum heat transfer enhancement of 416 %, friction factor of 511 % than conventional plain tube and the thermal performance factor was obtained as ~2.9. In addition, the empirical equations for heat transfer, friction factor and thermal performance factor were derived and compared with the experimental
error.
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results. The predicted results indicated a good agreement with experimental results within 10 % of
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factor
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Keywords: Heat transfer enhancement, Turbulator, Solar air heater, Friction factor, Enhancement
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ACCEPTED MANUSCRIPT 1. Introduction In solar energy applications, the solar air heaters (SAHs) were commonly used as heat exchanger. The incoming solar radiations into thermal energy were converted with SAHs. The converting
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thermal energy was extracted by air flowing under the absorbing surface. However, the heat transfer to the flowing air was obstructed due to generation laminar sublayer over the heattransferring surface. Therefore, the thermal performance of SAHs was negatively affected by the formation laminar sublayer. The artificial surface roughness was used to eliminate negative effect
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in SAHs [1]. Namely, the artificial roughness having form of repeated ribs was used to improve
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heat transfer in SAHs. The heat transfer enhancement characteristics of solar air heater with roughness elements on the surface were investigated by some researchers. Hans et. al., [1] investigated out to improve of heat transfer coefficient and friction factor with multiple V-rib roughness in solar air heater duct. Hans et. al., carried out the effect relative roughness height (e/D), pitch (P/e), width (W/w) and angle of attack (a) in a range of Reynolds number (Re) from
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2000 to 20000. The heat transfer and friction factor in the convection flow air in SAHs having rectangular duct were investigated by Bhagoria et. al., [2] for Reynolds number range from 3000
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to 18000. Kumar et. al.,[3]., performed the improvement of the heat transfer in the SAHs with artificial roughness in the form 60° inclined discrete rib for relative roughness height (e/D), pitch
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(P/e) and gap posistion (d/W) for Reynolds number (Re) varied from 4105 to 20526. The enhancement of heat transfer characteristics a rectangular duct roughened SAH in the relative roughness height (e/D), width (W/w), pitch (p/e) and arc angle (a) were investigated by Singh et al., [4] for Reynolds number (Re) in the range of 2200–22,000. The effect of relative roughness height 0.0181–0.0363; relative roughness pitch 4.5–10.0, and groove position to pitch ratio 0.3– 0.7 to improve of heat transfer was performed by Ref. [5] and the obtained results were compared for SAHs having ribbed and smooth duct. Sethi et al. [6] conducted experimental study to enhancement of heat transfer in SAHs having dimple shaped roughness. Also, empirical equations for Nusselt number and friction factor were developed and compared with experimental results. 3
ACCEPTED MANUSCRIPT The thermal hydraulic performance analysis in SAHs with 60 V-down discrete rib roughnesses was performed by Karwa and Chitoshiya [7]. The heat transfer enhancement characteristics in SAHs with 60 V-down discrete rib roughness performances were investigated. The thermal performance analysis of a flat-plate solar collector was investigated by Ref. [8] with twisted strip,
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coil-spring wire and conical ridges. Skullong et. al.,[9] presented an experimental study on heat transfer characteristics in a SAH channel fitted with combined wavy-rib and groove turbulators for different rib-pitch to channel-height ratios (PR) with a single rib-to-channel height ratio (BR) for
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Reynolds numbers in the range of 4000 to 21.000. The heat transfer characteristics with the form of specially prepared inverted U-shaped turbulators on the absorber surface of an SAHs was
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performed by Ref. [10]. The heat transfer characteristics in the SAHs having corrugated trapeze, reverse corrugated, reverse trapeze, and a base flat-plate collector was investigated by Benli [11].
In previous SAHs studies, various designs with different rib roughness geometry in flat plate type
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solar air collectors were developed. In this study, a new type of SAHs having semi circular absorber plate including various type turbulators was designed and fabricated as differently from previous SAHs studies. In this experimental study, three different copper tubes inside of the SAHs
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were arranged and turbulators was inserted inside the tube. Effect of the pitch ratio (PR) and angle ratio (AR) were investigated to improve in heat transfer in SAHs in a range of between 3000 and
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7500 Reynolds number. Nusselt number, friction and thermal performance factor results obtained in various PR and AR were compared with obtained results in conventional plain tube in a range of between 3000 and 7500 Reynolds number. Also, the empirical equations for Nusselt number, friction factor and thermal performance was developed and compared with experimental results. 2. Experimental set-up The SAHs experimental setup was shown in Fig. 1. The experimental set-up, diagram with parameters (L, D, α) of circular type turbulators and details of the SAHs having semi circular
4
ACCEPTED MANUSCRIPT type’s absorber plates were given in Figs. 1-3. The collector dimensions were 800 mm × 450 mm × 200 mm. The absorbing surfaces were produced with a dull black painted galvanized copper with a thickness of 0.5 mm. In this study, a new approach in SAHs as different from literature was applied. The tube and turbulator diameters were 50 and 50 mm, respectively. Turbulators having
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various obstacle relief angles and distances were inserted inside the tube for working fluid (air) circulations. The ideas/criteria for designing various type turbulators was to increase in heat transfer and pressure drop. In this study, the circular type turbulators were considered. In the next
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studies, the other type’s turbulators which has conical-ring, conical-nozzle, V-nozzle, twisted-tape and screw-tape will applied to compare of enhancement of the heat transfer. The circular type
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turbulators were designed and fabricated as various obstacle relief angles (α= 45°, 90°, 135° and plain tube) and obstacle distances (L= 100 mm, 140 mm, 175 mm) to create of the turbulence flow and enhance of heat transfer rates. Turbulators dimensions and ratios using experimental tests in the SAHs were given in Table 1. The Reynolds number used by experimental tests was in a range
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of between 3000 and 7500. The sides and back of the collector were isolated to prevent in thermal losses. The collector was placed facing south with a slope angle of 30°. The solar radiation incident on the surface was measured with a Solorimeter (KIMO-SL100) and was recorded at
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noon. The ambient temperature, collector input, output and surface temperatures were measured with K type thermocouple. The inlet and outlet temperatures in the collectors, ambient and
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absorber plate temperatures were measured with digital data logger (AMATEK ASM 802-B) and recorded from several selected locations. The pressure loss and velocity of the air were measured by a digital manometer (BEAMEX-EXT200MCS) and the anemometer (LUTRON AM-4202), respectively. 3. Calculation Method SAHs are irradiative heat exchangers. Solar radiation energy converted to heat in the collector. The heat obtained solar radiation energy was transferred by convection from the absorber to the
5
ACCEPTED MANUSCRIPT working fluid. In this study, the air as the working fluid (air) was used. At noon, the maximum instantaneous radiation energy was measured as 1100 W/m2 and used in the heat and flow friction characteristic calculations in SAHs. In this experimental study, the copper tubes inside the solar air collector were arranged as semi insulated tube. In the test section in SAHs, the useful heat gain
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( Qu ) of the air is calculated as [12-18]:
Qu = m& .cp .(Tout − Tin )
(1)
Qu Ap .(Ts − Tb )
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h=
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The heat transfer coefficient for the test section is:
(2)
where the bulk temperature of the working fluid (air), Tb, can be found by averaging the inlet and outlet temperatures of the air [12-18]. The physical properties of the air can be determined by Tb: (3)
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Tb = (Tout + Tin ) / 2
and Ts is the average surface temperature of the SAHs absorber surface. The heat transfer
h.D k air
(4)
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Nu =
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coefficient and Nusselt number were calculated as below [12-18]:
The Reynolds number was written by Re =
U.D
(5)
ν
Friction factor, f can be written as
f =
∆P 2 L U ρ D 2
(6)
in which U is mean air velocity in the copper tube inserted in SAHs. 6
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For a constant pumping power [12-18], (V∆P ) p = (V∆P ) t
(7)
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and the relationship between friction and Reynolds number were given as below [12-18]:
( f Re3 ) p = ( f Re3 ) t
(8)
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At the same pumping power, the thermal enhancement efficiency (η), were given as below by Ref [12-18]:
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Nut Nu p η= 1 ft 3 f p
(9)
Also, the uncertainties of measurements in this study were determined using Kline and
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McClintock uncertainity method [19]. The uncertainty during the measurements of the parameters was given in Table 2. Also, the maximum uncertainties of non-dimensional parameters were
conditions.
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obtained as ± 10 % for Reynolds number, ±7 % for Nusselt number and ± 10% for friction for all
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4. Results and Discussions
4.1.Validation test of plain tube In this study, the effects of the circular type turbulators inducing swirl flows on the heat transfer enhancement (Nusselt number, Nu), friction factor (f) and thermal performance factor (η) in SAHs were investigated. In the calculations, angle ratio (AR) and pitch ratio (PR) were taken into consideration as shown in Fig. 4 and 5. The Reynolds number was changed between 3000 and 7500. The pitch ratios (PR) was 2, 2.8, 3.5 whereas, the angle ratios (AR) was 0.125, 0.25 and 0.375. The experimental results for the heat transfer rate and the friction factor in a circular plain tube are obtained and confirmed with the Eqs. (10)-(13) correlations [12]. Nusselt number with the 7
ACCEPTED MANUSCRIPT experimental data’s obtained from the plain tube in SAHs was compared using Gnielinski [12,16] and Kays&Crowford equations [11,17]. Also, the friction factor with the experimental data’s obtained from the plain tube was compared using Petukhov [12,16] and Blasius [12,16] for the experimental system validations. The correlation equations used to provide experimental system
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validations were given as below: Gnielinski correlation for Nusselt number,
( f 8)(Re− 1000) Pr 2 1 1 + 12.7( f 8) 2 (Pr 3 − 1)
3000 < Re < 5x106
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Nu =
Nu = 0.0158 Re 0.8 Petukhov correlation for friction factor,
f = (0.79 ln Re− 1.64) −2
f = 0.316 Re −0.25
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Blasius correlation for friction factor,
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Kays&Crowford correlation for Nusselt number,
(10)
(11)
3000 < Re < 5x106
(12)
3000 < Re < 20,000
(13)
The correlation Eqs.(10-13) results used to provide experimental system validations were
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compared with experimental plain tube data’s. The Nusselt number deviations between present
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study and Gnielinski and Kays&Crowford were obtained within ±15% whereas; the friction factor deviations according to Petukhov and Blasius correlations were computed as ±11%. The experimental plain tube data’s were found to be in good agreement with the previous correlations given of Eqs. (10) – (13) for both the Nusselt number and friction factor as shown in Fig. 4a and 5a.
4.2.Heat transfer The circular type turbulators having obstacle relief angles and distances in SAHs were used to improve in the heat transfer. The circular type turbulators having obstacle relief angles and 8
ACCEPTED MANUSCRIPT distances created boundary layer of the air flow and increase the turbulence in the air flow. Effects of the pitch ratio (PR), angle ratio (AR) and various Reynolds numbers on the Nusselt number and Nusselt number ratio (Nut/Nup) were illustrated in Fig. 4a and 4b. As shown in Fig. 4a, the Nusselt numbers were increased with increasing Reynolds number whereas; it was decreasing with
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increasing PR and AR. It can be seen that the enhancement of the heat transfer in the SAHs having turbulators were significantly higher than obtained results in plain tube. The increase of enhancement of the heat transfer with decreasing PR and AR was due to the increased heat
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transfer surface area and prolonged residence time of the air flow in the tube in SAHs. As investigated in Fig 4., the maximum heat transfer enhancement was obtained with the use of the
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circular turbulator at the smallest pitch ratio (PR= 2.0) and angle ratio (AR= 0.125) for between 3000 and 7500 Reynolds number. For the PR=2 and AR=0.125, the increase in the heat transfer was between 416 % and 258 % in a range of between 3000 and 7500 Reynolds number, respectively, according to conventional plain tube. At the same Reynolds number, the increase in
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the heat transfer was between 339 % and 190 % for the PR=3.5 and AR=0.125 according to conventional plain tube. It was seen that the use of the circular type turbulator was shown a significant increase in Nusselt number according to the plain tube. In the PR=2.0, the Nusselt
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number for AR=0.125 were improved ~19 % and ~33 % according to AR= 0.375 for 3000 and 7500 Reynolds numbers, respectively. At the same Reynolds numbers, the Nusselt number
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obtained with AR=0.125 was higher than AR=0.375 at about ~10% and ~37% for PR=3.5, respectively. Evaluations from PR=2 to PR=3.5 for AR=0.125, the Nusselt numbers were decreased as ~23% and ~18% for 3000 and 7500 Reynolds numbers, respectively. These results were shown that the heat enhancement factor of AR ratios was higher than PR ratios. In addition, Nusselt number ratio (Nut/Nup) against the Reynolds number was illustrated in Fig. 4b. It can observe that the Nusselt number ratio was decreased with increasing Reynolds number from 3000 to 7500 for all conditions. The highest Nusselt number ratio obtained with low Reynolds number. Also, Nusselt number ratio decreased with the rise of PR and AR as shown Fig. 4b. The best 9
ACCEPTED MANUSCRIPT Nusselt number ratio was obtained as 5,16 and 3,58 with at the smallest pitch ratio (PR= 2.0) and angle ratio (AR= 0.125) for between 3000 and 7500 Reynolds number, respectively. The Nusselt number ratios of all conditions were above unity. This situation provided an advantage the use of
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circular type turbulators according to the plain tube.
4.3.Friction factor
The friction factor (f) and the friction factor ratio (ft/fp) plotted against the Reynolds number
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values were shown in Figs. 5a and 5b. The friction factor was decreased with the increasing Reynolds number, PR and AR as shown in Fig. 5a. However, the friction factor was increased
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with decreasing PR and AR depends on circular type turbulators having obstacle relief angles and distances. The fiction factor was higher than obtained values in plain tube as illustrated in Fig. 5a. The increase in friction factor was resulted by the increased contact surface area and prolonged residence time of the flow in the tube in SAHs. Because, the turbulators can be created the
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dynamic pressure of the fluid with increasing contact area and number of obstacle depends on PR and AR. As shown in Fig. 5a, the maximum friction factor was obtained as 511 % and 295 % in a range of between 3000 and 7500 Reynolds number, respectively, according to conventional
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plain tube for the PR=2 and AR=0.125. The average friction factor obtained by the circular turbulator with the lowest pitch ratio (PR= 2.0) was higher than the circular rings having PR= 2.8
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and 3.5 at around 6% and 17%, respectively, for AR=0.125. The average friction factor generated by the circular turbulator having the lowest angle ratio (AR= 0.125) was higher than the circular rings with AR = 0.25 and 0.375 at around 19% and 60%, respectively, for PR= 2. It was seen that the friction factor values of AR ratios were higher than PR ratios.
The changes of the friction factor ratio (ft/fp) with the Reynolds number values were shown in Fig. 5b. The friction factor ratios with increasing Reynolds number, PR and AR were decreased. As investigated in Fig. 5b, the average friction factor ratio obtained by using turbulators with at the
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ACCEPTED MANUSCRIPT smallest pitch ratio (PR= 2.0) and angle ratios (AR= 0.125 and 0.375) were found to be about 5,1 and 3.2 times from the plain tube. It was observed that the use of the circular type turbulators provided an increase in the friction factor ratio than plain tube under the same conditions.
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4.4.Thermal Performance Factor
The thermal performance factor was shown in Fig. 6 with the rise of Reynolds number from 3000 to 7500. The thermal performance factor was evaluated for the same pumping power criteria. As
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shown in Fig. 6., the thermal performance factor decreased with increasing Reynolds number, PR and AR. The thermal performance factors above ~2 were obtained for in all the cases. The
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maximum thermal performance factor was computed as ~2.9 and ~2.3 in the PR=2 and AR= 0.125 for between 3000 and 7500 Reynolds number, respectively. The thermal performance factor for PR=2 and AR = 0.125 and 0.375 were obtained as ~2.9 and 2.6 for 3000 Reynolds number. The thermal performance factor was raised as ~ 8% with decreasing AR. The thermal performance
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factor for PR=3.5 was computed as ~2.5 and ~1.9 with AR= 0.125 for between 3000 and 7500 Reynolds number, respectively. On the other hand; the thermal performance factor for AR=0.125 and PR=2 and 3.5 were computed as ~2.9 and 2.5 for 3000 Reynolds number, respectively. The
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thermal performance factor was increased as ~ 13% with decreasing PR. The average thermal performance factor obtained with the lowest angle ratio (AR= 0.125) was higher than AR= 0.25
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and 0.375 at around 1% and 7%, respectively, for PR=2. The average thermal performance factor with lowest pitch ratio (PR= 2) was higher than PR = 2.8 and 3.5 at around 13% and 14%, respectively, for AR= 0.125.
4.5. Correlations Nusselt number, friction number and thermal performance factor are strong function of a system and operation parameter. Functional relationship for Nusselt number, friction number and thermal performance factor can be written as below: 11
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(15)
f = f 2 (Re, PR, AR)
(16)
η = f 3 (Re, PR, AR)
(17)
Nusselt number, friction factor and thermal performance factor correlations was developed for
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range of system and operation parameter with regression analysis. The correlations obtained from experimental results in a range of Reynolds number between 3000 and 7500 was shown in Eqs. (18)–(20), respectively. The predicted results from Eqs (18)–(20) were compared with
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experimental results and the obtained comparatively results were illustrated in Figs. 7–9. The
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predicted values obtained empricial equations have indicated in good agreement with experimental results. The maximum deviation of the between predicted and experimental results were within ± 10% for Nusselt number, friction factor and thermal performance factor, respectively.
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Nu = 0.127 Re 0.764 ( PR ) −0.356 ( AR ) −0.160 f = 50 .089 Re −0.713 ( PR ) −0.297 ( AR ) −0.409
(19) (20)
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η = 10 .006 Re −0.155 ( PR ) −0.230 ( AR ) −0.05
(18)
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5. Comparison with previous work In this section compared the thermal performance characteristics of a SAHs fitted with circular type turbulators with previous SAHs studies. In previous SAHs studies, various designs with different rib roughness geometry (transverse, V-shape, V shaped inclined discrete W-shape and inclined rib with gap viz.) and dimensions of the air flow passage in flat plate type solar air collectors were developed in order to increase the thermal performance of SAHs. In this study, a new type of SAHs having various type turbulators was designed as differently from previous SAHs studies. Turbulators as swirl generators which are conical-ring, conical-nozzle, V-nozzle, twisted-tape and screw-tape were used to improve of heat transfer with constant heat flux electric 12
ACCEPTED MANUSCRIPT heater on heat exchangers without solar radiation in earlier investigations [14-16, 23]. In order to see the advantage of turbulators of SAHs having semi circular type’s absorber plate, the present work was performed and compared with previous SAHs works as shown in Table 3. As investigated in Table 3, the thermo-hydraulic performance parameters of the present study have
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indicated in a good performance compared to previous studies. Thus, the use of the turbulator as well as the different rib roughness geometry used in SAHs to heat transfer enhancement was
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proposed and investigated.
6. Conclusions
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In this study, the effects of circular type turbulators having various obstacle relief angles (AR) and pitch ratio (PR) on heat transfer, flow friction and thermal performance factor characteristics in SAHs were investigated. The main results of this study are summarized below:
The heat transfer enhancement in the SAHs having turbulators was significantly higher
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than obtained in conventional plain tube. The maximum heat transfer enhancement obtained with decreasing pitch and angle ratios in the PR=2 and AR= 0.125. The maximum increase in the heat transfer was in a range of 416 % and 258 % for about 3000
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and 7500 Reynolds numbers, respectively, according to plain tube.
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The friction factor decreased with the increasing Reynolds number, PR and AR. The maximum friction factor was obtained as 511 % and 295 % in a range of between 3000 and 7500 Reynolds number, in the PR=2 and AR= 0.125 according to conventional plain tube.
The thermal performance factor decreased with the increasing Reynolds number, PR and AR. The maximum thermal performance factors computed as ~2.9 and ~2.3 in the PR=2 and AR= 0.125 for between 3000 and 7500 Reynolds number, respectively.
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ACCEPTED MANUSCRIPT Nusselt number, friction factor and thermal performance factor were correlated and developed the empirical equations. The predicted values obtained empirical equations have indicated in good agreement with experimental results within ±10 %.
As a result, the maximum heat transfer enhancement obtained using turbulators in SAHs
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was found at higher Reynolds number, lower pitch and angle ratios. Also, the effect of the
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heat enhancement factor of AR was higher than PR in the general evalutions.
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ACCEPTED MANUSCRIPT Nomenclature
Ap
Absorber plate area of collector, m2
AR
Angle ratio
cp
Specific heat of air, j/kg°C
Greek letters
D
Diameter, mm
α
Obstacle clearance angle
f
Friction factor
ν
Kinematic viscosity, m2/s
h
Heat transfer coefficient, W/m2°C
ρ
Fluid density, kg/m3
I
Solar radiation W/m2
k
Thermal conductivity, W/m2°C
L
Lenght, mm
m&
Mass flow rate, kg/s
Nu
Nusselt number
P
Fluid pressure, Pa
Pr
Prandtl number
PR
Pitch ratio
Re
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Air velocity, m/s
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U
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Subscripts
Air
b
Bulk
e
Environment
in
Inlet
out
Outlet
p
Plain tube
Reynolds number
s
Surface
Qu
Useful heat gain, W
t
Turbulator
T
Temperature, °C
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EP
TE D
a
15
ACCEPTED MANUSCRIPT REFERENCES 1. V.S. Hans, R.P. Saini, J.S. Saini, Heat transfer and friction factor correlations for a solar
air heater duct roughened artificially with multiple v-ribs, Solar Energy 84 (2010) 898–
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911.
2. J.L. Bhagoria, J.S. Saini, S.C. Solanki, Heat transfer coefficient and friction factor
correlations for rectangular solar air heater duct having transverse wedge shaped rib
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roughness on the absorber plate, Renewable Energy 25 (2002) 341–369.
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3. T. S. Kumar, V. Mittal, N. S. Thakur and A. Kumar, Heat Transfer and Friction Factor
Correlations for Rectangular Solar Air Heater Duct Having 60° Inclined Continuous Discrete Rib Arrangement, British Journal of Applied Science & Technology 1(3),( 2011)
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67-93.
4. A. P. Singh, Varun, Siddhartha, Heat transfer and friction factor correlations for multiple
arc shape roughness elements on the absorber plate used in solar air heaters, Experimental
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Thermal and Fluid Science, 54, (2014), 117–126
5. A.R. Jaurker, J.S. Saini, B.K. Gandhi, Heat transfer and friction characteristics of
rectangular solar air heater duct using rib-grooved artificial roughness, Solar Energy, 80 (2006), 895–907.
6. M. Sethi, Varun, N.S. Thakur, Correlations for solar air heater duct with dimpled shape
roughness elements on absorber plate, Sol. Energy, 86, (2012), 2852–2861.
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ACCEPTED MANUSCRIPT 7. R. Karwa, G. Chitoshiya, Performance study of solar air heater having v-down discrete
ribs on absorber plate, Energy, 55, (2013), 939–955. 8. A. Hobbi, K. Siddiqui, Experimental study on the effect of heat transfer enhancement
devices in flat-plate solar collectors, International Journal of Heat and Mass Transfer, 52,
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(2009), 4650–4658.
9. S. Skullong , S. Kwankaomeng, C. Thianpong, P. Promvonge, Thermal performance of
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turbulent flow in a solar air heater channel with rib-groove turbulators, International
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Communications in Heat and Mass Transfer, 50, (2014), 34–43.
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and Mass Transfer, 52, (2009), 2834–2848.
11. H. Benli, Experimentally derived efficiency and exergy analysis of a new solar air heater
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having different surface shapes, Renewable Energy 50 (2013) 58-67
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Transfer. Sixth edition, John-Wiley & Sons, 2006.
13. C. Muthusamy, M. Vivar, I. Skryabin, K. Srithar, Effect of conical cut-out turbulators with
internal fins in a circular tube on heat transfer and friction factor, International Communications in Heat and Mass Transfer, 44, (2013), 64–68.
14. C. Thianpong, K. Yongsiri, K. Nanan, S. Eimsa-ard, Thermal performance evalation of
heat exchangers fitted with twisted-ring turbulators, International Communications in Heat and Mass Transfer, 39, (2012), 861–868. 17
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15. P. Promvonge, S. Eimsa-ard, Heat transfer in a circular tube fitted with free spacing snail
entry and conical nozzle turbulators, International Communications in Heat and Mass
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Transfer, 34, (2007), 838–848.
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twisted tape, Applied Thermal Engineering, 30, (2010), 1673-1682.
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17. WM. Kays, ME. Crowford, Convective heat and mass transfer. 3rd edition, New
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York:McGraw-Hill: 1993
18. E. Baysal, Comparison of Experimental and Numerical Heat Transfer Results in
Concentric Pipe Heat Exchangers, Journal of Polytechnic, Vol: 11 No: 4 pp.345-352, 2008. 19. S.J. Kline, F.A. McClintock, Describing uncertainties in single sample experiments, Mech.
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Eng. 75 (1953) 3–8.
20. B.N. Prasad, J.S. Saini, Effect of artificial roughness on heat transfer and friction factor in
a solar air heater, Sol. Energy, 1988;6: 555–60.
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21. S. Singh, S. Chander, J.S. Saini, Thermo-hydraulic performance due to relative roughness
pitch in V-down rib with gap in solar air heater duct—Comparison with similar rib
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roughness geometries, Renewable and Sustainable Energy Reviews, 43, (2015): 1159– 1166.
22. A.M. Ebrahim Momin, J.S. Saini, S.C. Solanki, Heat transfer and friction in solar air
heater duct with V-shaped rib roughness on absorber plate. Int. J. Heat Mass Transfer 2002;.45:33, 83–96. 23. S. Eiamsa-ard, V. Kongkaitpaiboon and K. Nanan, Thermohydraulics of Turbulent Flow
Through Heat Exchanger Tubes Fitted with Circular-rings and Twisted Tapes, Chinese Journal of Chemical Engineering, 21(6) 585—593 (2013).
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45°
0,125
90°
0,25
135°
0,375
Length (L),mm 100 140 175 100 140 175 100 140 175
Diameter (D),mm 50
50
Pitch Ratio (PR=L/D) 2 2,8 3,5 2 2,8 3,5 2 2,8 3,5
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Angle Ratio (AR=α/360)
50
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Obstacle relief angle (α)
Table 2. The uncertainty during the measurements of the parameters Unit
Collector inlet temperature Collector outlet temperature Absorber surface temperature Ambient temperature
Time Solar radiation
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Pressure loss
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Air flow velocity
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Parameter
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Uncertainty (δ)
0
C
± 0.10
0
C
± 0.10
0
C
± 0.10
0
C
± 0.10
m/s
± 0.02
min
± 0.10
W/m2
± 1,26
Pa
± 0,10
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Table 3. Comparison of optimum values of performance parameter for various enhancement geometries in SAHs. Enhancement Element
Nu
f
η
Present study
Circular turbulators and semi-circular absorber plate
5,16
6,11
2,82
Bhagoria et al. [2]
Transverse wedge shaped ribs
2.4
5.3
1.38
Singh and Saini [21]
W down rib
2.7
2.86
1.94
Jaurker et al. [5]
Rib groove
2.75
3.61
1.76
Prasad and Saini [20]
Transverse wire
2.38
4.25
1.78
Bopche and Tandale [10]
Inverted U shape rib
2.82
3.72
1.82
Momin et al. [22]
V sahaped rib
2.30
2.83
1.78
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FIGURE CAPTIONS
Fig. 1. Schematic view of experimental set-up Fig. 2. Sectional view of absorber plate in experimental setup
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Fig. 3. Schematic diagram with parameters of circular type turbulators
Fig.4. Changes of (a) Nusselt number and (b) Nut/Nup with Reynolds number. Fig.5. Changes of (a) Friction factor and (b) ft/fp with Reynolds number.
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Fig.6. Changes of thermal performance factor with Reynolds number Fig. 7. Comparison of experimental and predicted Nusselt number
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Fig. 9. Comparison of experimental and predicted thermal performance factor
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Fig. 1. Schematic view of experimental set-up.
Fig. 2. Sectional view of absorber plate in experimental setup
Fig. 3. Schematic diagram with parameters of circular type turbulators
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Fig.4. Changes of (a) Nusselt number and (b) Nut/Nup with Reynolds number.
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Fig.5. Changes of (a) Friction factor and (b) ft/fp with Reynolds number.
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Fig.6. Changes of thermal performance factor with Reynolds number
Fig. 7. Comparison of experimental and predicted Nusselt number
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Fig. 8. Comparison of experimental and predicted friction factor
Fig. 9. Comparison of experimental and predicted thermal performance factor
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ACCEPTED MANUSCRIPT Highlights
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The heat transfer enhancement in a new Solar Air Heaters (SAHs) having various picth and angle ratios were investigated. The highest heat transfer enhancement was provided with decreasing picth and angle
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ratios.
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The effect of the heat enhancement of pitch ratio was higher than angle ratio.
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