Heat transfer in solar receiver heat exchanger with combined punched-V-ribs and chamfer-V-grooves

International Journal of Heat and Mass Transfer 143 (2019) 118486

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Heat transfer in solar receiver heat exchanger with combined punched-V-ribs and chamfer-V-grooves Pongjet Promvonge a, Sompol Skullong b,⇑ a

Department of Mechanical Engineering, Faculty of Engineering, King Mongkut’s Institute of Technology Ladkrabang, Bangkok 10520, Thailand Energy Systems Research Group, Department of Mechanical Engineering, Faculty of Engineering at Sriracha, Kasetsart University Sriracha Campus, 199 M.6, Sukhumvit Rd., Sriracha, Chonburi 20230, Thailand b

a r t i c l e

i n f o

Article history: Received 6 March 2019 Received in revised form 16 June 2019 Accepted 25 July 2019

Keywords: Punched-rib V-groove Vortex generators Solar air heater Thermal performance

a b s t r a c t An experimental work has been carried out to investigate the influence of combined turbulence promoters (or turbulators) on forced convection and fluid flow resistance behaviors in a solar air heater duct. Two turbulators included V-ribs with punched holes and chamfered V-grooves were introduced. The V-rib and the V-groove having the attack angle of 45° were mounted repeatedly on the absorber plate with their arrangements for V-tip pointing upstream and pointing downstream. Air as the test fluid flowed into the duct with Reynolds number (Re) ranging from 5300 to 23,000. The rib parameters were three relative rib-pitches (RP = 1.0, 1.5 and 2.0), three inclination angles (b = 45°, 0° and
45°) of ribpunched holes having a single relative rib height or blockage ratio, RB = 0.5. The groove parameters included three relative groove-pitch lengths (RP = 1.0, 1.5 and 2.0) similar to the V-rib case. Influences of this newly designed absorber plate on Nusselt number (Nu) and friction factor (f) have been examined and compared with similar results of the smooth duct alone. The experimental results demonstrated that the combined turbulators at b = 45° and RP = 1.0 provide the maximum Nu and f, especially for the V-up case due to stronger vortex flows and the impinging flows from the punched holes over the absorber plate. Further, a new thermal enhancement factor (TEF) at similar pumping power has been proposed and it indicates that the combined V-up rib-groove with b = 45° and RP = 1.5 has the highest TEF of about 2.47. Ó 2019 Elsevier Ltd. All rights reserved.

1. Introduction Heat transfer augmentation in thermal systems used in many industries, such as electronics cooling, gas turbine blade cooling, drying process, transportation, heat exchanger and solar air heater (SAH) is very important for saving the energy resources, the high heat dissipation at high performance and the demand for compact systems. SAHs are cheap because of their inherent simplicity, minimal use of materials and cost and are widely applicable for many applications such as space heating, drying of agricultural, seasoning of timber, textile and marine products, etc. A conventional SAH as shown in Fig. 1 consists of the (1) blower, (2) flat-plate absorber to absorb the solar energy, (3) airflow channel, (4) glass cover, and (5) insulation. The thermal performance of a SAH is generally considered to be somewhat poor because of low heat transfer rate during air flowing through the smooth or flat absorber plate. To make a SAH more ⇑ Corresponding author. E-mail addresses: [email protected], [email protected] (S. Skullong). https://doi.org/10.1016/j.ijheatmasstransfer.2019.118486 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.

efficient, its thermal performance needs to improve by enhancing the heat transfer rate. The conventional methods for improving the heat transfer rate as well as thermal performance are the passive method such as rough surfaces/extended surfaces/turbulators, and the active method which requires extra external power sources, for example, fluid suction and injection, fluid vibration, mechanical aids, and the use of electrostatic fields [1]. However, several researches have been focused on the passive heattransfer enhancement method due to low cost, easy maintenance and manufacture. Turbulence promoters (or called ‘‘turbulators”) such as; wire coil [2,3], conical ring/nozzle [4–6], snail [7], twisted tape [8–11], rib [12,13], groove/dimple [14–16], fin [17,18], baffle [19,20], wing [21–23], winglet [24–26], etc. form an important group of the passive augmentation technique. For three decades, the passive heat transfer enhancement techniques characterized by different-type turbulators have been developed rapidly. Turbulators not only disrupt the boundary layer development of the fluid flow but also generate the longitudinal vortex flows that assist to wash up the reverse flow trapped behind the turbulators into the core flow leading to higher the heat

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P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

Nomenclature A Ac AR b Cp,a D d e f fR H h I k L _ m Nu NuR Pl Pw DP Pr Re RB RP Q T U V

heat transfer surface area of duct [m2] cross-sectional area of smooth duct [m2] duct aspect ratio [=W/H] rib height [mm] specific heat capacity of air [J/kgK] hydraulic diameter of duct [mm] hole diameter on ribs [mm] groove depth, [mm] friction factor [-] friction factor ratio [=f/f0] duct height [mm] average heat transfer coefficient [W/m2K] electrical current [amp] thermal conductivity of air [W/mK] length of test duct [mm] mass rate of airflow [kg/s] Nusselt number [=hDh/k] Nusselt number ratio, [=Nu/Nu0] streamwise or longitudinal pitch of ribs [mm] wetted perimeter of smooth duct [m] pressure drop [Pa] Prandtl number [=Cpl/ka] Reynolds number [=UDh/m] rib blockage ratio [=b/H] relative rib/groove pitch ratio [=Pl/H] rate of heat transfer [W] temperature [K] mean velocity [m/s] electrical voltage, [volt]

transfer rate. Researches on the applications of ribs, wings/winglets and dimples/grooves in a flat-surface duct/channel are found to be rich for enhancing the heat transfer in thermal systems while other turbulators are suitable for a round tube/duct. Iacovides et al. [27] presented turbulent forced convection in straight cooling passages with inclined ribs on two opposite walls. They suggested that the strong influence of the rib-induced secondary motion weakens the influence of turbulence on the flow and thermal development. Lu and Jiang [28] investigated experimentally and numerically the turbulent convection heat transfer in a channel with inclined/

V_ W w

volume rate of flow [m3/s] duct width [mm] groove width, [mm]

Abbreviations LVG longitudinal vortex generator pp pumping/blowing power SAH solar air heater TEF thermal enhancement factor VG vortex generator WVG winglet-typed vortex generator Greek letters a attack angle of rib/groove, [°] b inclination angle of hole, [°] / groove chamfer angle, [°] d thickness of rib, [m] l viscosity of air, [kg/ms] m kinematic viscosity of air, [m2/s] q density of air, [kg/m3] Subscripts a air b bulk 0 smooth duct conv convection i inlet o out s surface

angled ribs. Promvonge and Thianpong [29] studied the influence of thermal performance assessment in a ribbed channel (including triangular, wedge and rectangular rib-shapes) by heating up the top wall of the channel only to achieve a constant heat-fluxed wall. They reported that the wedge rib yields the highest increase in both Nusselt number and friction factor but the triangular rib gives the maximum thermal performance. Influences of combined ribs and winglet-typed vortex generators (WVGs) on turbulent convection heat transfer in a channel were examined by Promvonge et al. [30]. They found that the combination of ribs and WVGs leads to a

Fig. 1. Schematic of conventional SAH system.

P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

double increase in heat transfer over the use of rib or WVG alone. Song et al. [31] explored the three-dimensional flow and heat transfer in a channel with crossed discrete double inclined ribs. They showed that Nusselt number for the 45° attack angle is higher than that for the 30° and 60°. Prasad [32] investigated the thermal performance in a solar air heater duct fitted with small wires and suggested that the artificially roughened solar air heaters provided better thermal performance compared to the smooth duct alone. Alam et al. [33] examined the heat transfer and flow friction characteristics of a rectangular solar air heater duct fitted with circularity of perforation holes in V-shaped blockages. They revealed that the highest increase in Nusselt number was observed at a relative pitch of 8. The effect of rib size and arrangement on forced convective heat transfer and flow friction behaviors in a solar air heater channel for turbulent flow was reported by Skullong et al. [34]. An experimental study for heat transfer augmentation in a solar air heater using multiple arcs with gap was conducted by Pandey et al. [35]. They found that the maximum increment in the heat transfer rate and friction loss was, respectively, 5.85 and 4.96 times in comparison with the flat plate duct. An experimental investigation of turbulence forced convection in a horizontal triangularsectional duct with V-grooved was conducted by Leung et al. [36]. Jin et al. [37] carried out an experimental work to investigate the heat transfer in triangular grooved channel using particle image velocimetry to analyze the flow mechanism inside. Eiamsa-ard and Promvonge [38] investigated numerically the heat transfer of turbulent channel flow over periodic grooves and found that the grooved channel provides a considerable increase in heat transfer at about 158% over the smooth channel. Ma et al. [39] accounted for rectangular block as longitudinal vortex generators (LVG) in a narrow rectangular channel, investigated experimentally the single-phase heat transfer using water as the working fluid. They found that the increases in heat transfer and the flow resistance were about 100.9% and 11.4% for laminar regime and around 87.1% and 100.3% for turbulent regime, respectively. Tang et al. [40] studied the 3-dimensional turbulent convection in a channel with different groove configurations (P-, V- and W-typed grooves). They reported that the heat transfer of the P-groove improved at about 45.8–65.4% while the friction loss was increased by 95.1–114.8% compared with the plain duct. Liu et al. [41] computed the heat transfer in a cooling dimpled channel with secondary hemispherical protrusions and suggested that the dimpled channel with secondary protrusions gave higher heat transfer compared with the conventional dimpled channel. Several efforts in earlier studies on various turbulators in the form of ribs and dimples/grooves are listed in Table 1 and 2, respectively. Table 1 provides the result summary for different rib-turbulators used in ducts/channels [12,13,27–29,32–35] having the maximum NuR, fR and TEF around 4.7, 158 and 1.53, respectively. Also, Table 2 gives the result summary for various dimple/groove turbulators [15,16,36–38,40,41] with the highest NuR, fR and TEF around 1.8, 1.98 and 1.47, respectively. It is clearly seen that thermal performance of using a single turbulator is not much high, (TEF < 1.6). The literature review above indicates that the rib turbulator can assist to augment the rate of heat transfer and pressure drop in a heating/cooling duct rather than the groove turbulator. Nonetheless, the application of combined VG devices is hoped to improve thermal performance in such a duct. There have been several investigations on using ribs/grooves alone but the study on thermal characteristics of V-ribs in common with V-groove has rarely been reported, especially for the punched ribs. The punched rib is utilized to reduce the pressure drop in the flow system, leading to higher thermal performance. Hence, the key purpose in the current work is to examine the flow friction and thermal characteristics in a rectangular channel fitted with both the punched V-ribs and the chamfered-V-grooves. The V-ribs with various punched-

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hole inclination angles (b) were attached on the absorber plate that its surface was formed to be chamfered V-grooves at different relative pitches (RP = Pl/H). The vital aim for the inclination of holes is to avert the deterioration of vortex flows behind the ribs by the air jet issuing from the hole. On the basis of experimental results, the inclination angle of punched holes and the relative rib-groove pitches have been optimized. 2. Experimental system 2.1. Experimental setup and procedure The experimental system as depicted in Fig. 2 consisted of the test section, instrumentation to measure the temperatures and pressures, data acquisition system (Fluke 2680A), air blower, electrical power input, clamp meter (Fluke 376 FC), U-tube/inclined manometer, stainless-steel tank or settling tank, and orifice-type flowmeter (built according to an ASME standard [42]). The total length of the SAH duct was 3300 mm which included the 2200-mm entrance (for fully developed internal turbulent flow), 800-mm test and 300-mm exit sections as shown in Fig. 2. In the experimental set-up, a 2-kW blower was used as the source of supplying air at room temperature (maintained at about 25 °C) through a circular pipe into the settling tank at which an orifice flow-meter was on this pipeline. The settling tank was employed as an adaptor to change the circular duct to become the rectangular cross-sectional duct before connecting to the test section. The airflow rate was measured by the orifice flow-meter which was calibrated beforehand by using both hot-wire and vane-type anemometers (Testo 480). The test section was maintained at a constant wall heat-flux using the electrical plate-type heater for heating the top wall of the duct. The heat flux could be varied from 0 to 3500 W/m2 with the help of AC power supply (variac transformer). Proper insulation was provided to minimize convective heat loss to the surrounding air. Two PT-100, RTD-type temperature sensors (4 wire) were used to measure the inlet and outlet bulk temperatures while thirty T-Type thermocouples were employed for measuring the wall temperatures of the absorber. The thermocouple voltage outputs were fed into a data acquisition system (Fluke 2680A) and then recorded via a personal computer. The pressure drop of the test section was measured by a digital differential pressure manometer. To estimate the reliability in the experiment, the uncertainties of measured data were calculated based on Ref. [43]. The data uncertainties by the instruments of friction factor, Reynolds number and Nusselt number were presented in Table 3. The uncertainty equations of the experimental data were included in Appendix A. 2.2. Turbulators and range of parameters A graphic view of the SAH system equipped with combined punched-V-ribs and chamfered-V-grooves is displayed in Fig. 3. The rectangular duct was made of a 10-mm thick (t) aluminum plate with its cross-section dimensions of 200-mm (W) by 25mm (H). The ribs made of 5-mm thick (d) aluminum bar had 12.5-mm height (b), equivalent to the blockage ratio, RB = b/ H = 0.5 and four punched holes on a 45° V-rib were spaced evenly with three inclination angles of (b = 45°, 0° and
45°) at a single hole diameter (d = 5 mm), as seen in Fig. 4(a). A single value of the blockage ratio, RB and the hole size, d mentioned earlier was selected since they gave the maximum TEF for using the winglets as reported in Ref. [24]. Also, the 45° V-ribs was chosen rather than the 30° although the 30° yielded higher TEF than the 45° because of much higher NuR and fR which are the demand of this work. The measurement of surface temperatures at various axial locations

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Table 1 Turbulators in the form of ribs applied to SAH ducts. Authors and test section profiles

Turbulator Geometries

Conditions

Results NuR

fR

TEF

Thianpong et al. [12]

Isosceles,delta ribs

AR = 10, e/H = 0.13–0.26, RP = 1.33, Re = 5000–22,000

1.8–4.0

2–86

0.86–1.32

Skullong [13]

Inclined ribs

a = 45°–60°, RB = 0.1–0.3, Re = 5300–23,000

3.7–4.7

24.4–70.7

1.1–1.4

Iacovides et al. [27]

Inclined ribs

5.0 mm square ribs, spaced 50 mm apart and oriented at 45° with the duct axis

–

–

–

Lu and Jiang [28]

Angled ribs

a = 0°–90°, Re = 6000–14,000

–

–

0.7–1.55

Promvonge and Thianpong [29]

Delta, wedge and rectangular ribs

RB = 0.2, RP = 1.33, Re = 4000–16,000

2.6–4.4

32–158

0.79–1.12

Prasad [32]

Small wire

e/P = 10–40, e/D = 0.0092– 0.0279, Re = 2968– 12,238

–

–

–

Alam et al. [33]

Perforated holes in V-shaped blockages

w = 1–0.6, P = 4–12, e = 0.4–1.0, b = 5–25%, a = 30–75°,

–

–

–

Skullong et al. [34]

Square and thin ribs

RB = 0.1–0.4, RP = 0.5–1.33, Re = 5000–24,000

1.6–4.1

3–35

0.9–1.53

Pandey et al. [35]

Arcs with gaps

P/b = 4–16, b/D = 0.016–0.044, a = 30–75°, g/e = 0.5–2.0, d/x = 0.25–0.85, Re = 2100–21,000

–

–

–

Re = 2000–20,000

of the test section was illustrated in Fig. 4(b). The hole on the rib was expected to inject the air jet on the absorber surface for b = 45°, apart from reducing the pressure drag from the V-ribs for different hole inclination angles. The grooved plate was made of 10-mm thick (t) aluminum plate with its groove dimensions of 5 mm (depth, e) by 3 mm (width, w), for three longitudinal pitch spacing (Pl = 25 mm, 37.5 mm and 50 mm), equivalent to relative pitch ratio, RP = 1.0, 1.5 and 2.0, at a single groove chamfer angle (/ = 45°). The punched V-rib was attached on the grooved absorber plate (located at the center between the grooves) by using superglue. Parameters of both the rib and the groove such as the attack angle and the longitudinal pitch were similarly given whereas the position (g) of grooves was set to a half rib pitch (g/Pl = 0.5). The 45° V-ribs-and-V-grooves with its V-tip pointing both upstream and downstream (so-called V-up and V-down) were mounted repeatedly on the absorber plate. The effect of three different longitudinal rib-groove pitch ratios (RP = Pl/H = 1.0, 1.5 and 2.0) and three punched-hole inclination angles (b = 45°, 0° and
45°) on thermal performance in the SAH duct have been examined. The

inclination angle (b) is defined as the angle between the normal of the rib surface and the line paralleled to the hole centerline (see Fig. 3). Turbulator arrangements and the range of parameters are listed in Table 4. 3. Theoretical analysis Heat transfer behaviors in a SAH duct equipped with combined turbulators at different relative pitch ratios, (RP = 1.0, 1.5 and 2.0) and punched-hole inclination angles (b = 45°, 0° and
45°), are presented in terms of Nusselt number (Nu), friction factor (f) and thermal enhancement factor (TEF). At steady state, the convection heat transfer rate (Qconv) can be expressed as

Q air ¼ Q conv

ð1Þ

where

_ p;a ðT o
T i Þ ¼ VI
heatloss Q air ¼ mC

ð2Þ

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P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486 Table 2 Turbulators in the form of grooves/dimples applied to SAH ducts. Authors and test section profiles

Turbulator Geometries

Conditions

NuR

fR

TEF

Zhang et al. [15]

Convex-dimples with grooves

H/Dh = 0.625 W/Dh = 2.5, e/Dh = 0.0625, Re = 5000–20,000

1.3–1.43

1.15–1.95

–

Sato et al. [16]

Dimples

d = 0.4d, D = 2d, Px = 3.24d, Pz = 3.24d, R = 1.45d, Re = 1000–10,600

0.9–7.0

–

0.8–1.2

Leung et al. [36]

V-grooves

0  h < 150° , Re = 2800–9500

–

–

–

Jin et al. [37]

Triangular groove

L/a = 2, L/H = 1.2, H = 15, W = 187.5

–

–

–

Eiamsa-ard and Promvonge [38]

Transverse grooves

B/H = 0.5–1.75 Re = 6000–18,000

–

–

0.98–1.35

Tang et al. [40]

Discrete grooves

15–75°

Liu et al. [41]

Dimples with secondary protrusions

h = 30–90°, Re = 5000–25,000

The heat loss can be examined from Eq. (2), where Qair was calculated from measured data and VI (=Qsupply) was obtained from the electrical power consumption using a watt-meter or powermeter. At a thermal equilibrium state, the heat supplied by the power (VI) through the test section is about 3–7% above the convection heat of the airflow owing to the heat loss to surrounding air. Hence by neglecting the heat loss, only the heat absorbed by the airflow is considered for computing the heat transfer coefficient by

  Q conv ¼ hAs Te s
T b

ð3Þ

in which

T b ¼ ðT o þ T i Þ=2

ð4Þ

The mean surface temperature on the absorber is obtained from averaging the local surface temperatures by

1 Te s ¼ 10  X T 1 þ T 11 þ T 21  T 2 þ T 12 þ T 22   þ 3 3   T 10 þ T 20 þ T 30 þ::: þ 3 Referring to Fig. 4(b), the mean local surface temperature at different locations is given as

Results

1.0–3

1.5–1.8

1.62–1.98

1.28–1.47

    T 1 þ T 11 þ T 21 T 2 þ T 12 þ T 22 ; T s2 ¼ ; ::: ; 3 3   T 10 þ T 20 þ T 30 ¼ 3

T s1 ¼ T s10

1 X ðT s1 þ T s2 þ ::: þ T s10 Þ Te s ¼ 10

ð5Þ

Thus, the average heat transfer coefficient (h) is calculated as

h¼

_ p;air ðT o
T i Þ mC   As Te s
T b

ð6Þ

Then, the average Nusselt number (Nu) is estimated via

Nu ¼

hD k

ð7Þ

The Reynolds number (Re) of airflow through the test duct is calculated by

Re ¼

UD

m

ð8Þ

where D is the hydraulic diameter of the smooth duct and defined as

D¼

4Ac 4W H ¼ 2  ðW þ H Þ Pw

ð9Þ

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P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

Fig. 2. Schematic sketch of the experimental setup.

Multiplying both sides of the above equation by (1/f0,Re), it becomes

Table 3 Uncertainties for measurements and dimensionless group parameters.

Re30

Uncertainties in measurements Instrument Accuracy T-type thermocouple ± 0.1 °C RTD PT-100 ± 0.1 °C Anemometer ± 0.1 m/s Inclined-manometer ± 1 mm Digital-manometer ± 0.5 mm Uncertainties in dimensionless group parameters Parameters Re (Reynolds number) Nu (Nusselt number) f (friction factor)

% Error ± 0.5 ± 0.5 ±5 ±6 ±5

f ¼

DP qU = ðL=DÞ 2

Uncertainty (%) ±5% ±5% ±7%

ð10Þ

The performance evaluation criterion of a heat exchanger known as thermal enhancement factor, (TEF) under a similar blowing/pumping power (pp) constraint suggested by Webb and Kim [44] is derived. As mentioned earlier, the fluid power in the smooth duct is equal to that in the enhanced duct and hence, it is displayed as 0

ð11Þ

in which the subscript ‘‘0” denotes the smooth duct. Then, Eq. (11) can be rewritten in terms of friction factor and Reynolds number as below.



f Re3



0

 ¼ f Re3

¼ Re3

f f 0;Re

ð12Þ

Substituting the power-law equation like Blasius f-equation for

!

    V_ DP ¼ V_ DP

f 0;Re

smooth ducts,f 0 ¼ ReCm1 1 into Eq. (12),

in which Ac is the cross-sectional area of the smooth duct and Pw being its wetted perimeter. The friction factor (f) can be written by 2

!

f0

  3  Re0 C1 =Re0 m1 f

¼  f 0 Re Re C1 =Rem1 Rearranging;

 3
m1   Re0 f ¼ f 0 Re Re

By the definition of TEF; TEF ¼

    h Nu ¼ h0 pp Nu0 pp

ð13Þ

ð14Þ

Inserting power-law equation like Dittus-Boelter Nu-Eq. for smooth ducts,Nu0 ¼ C 2 Re0 m2 Pr n into Eq. (14),

TEF ¼ 

   m2  Re0 Nu Nu = ¼ n 2 m2 =Rem2 Þ Nu0 Re Rem2 C 2 Rem 0 Pr ðRe

ð15Þ

By inserting Eq. (13),

 TEF ¼

Nu Nu0



m2  3
m 1 f = f 0 Re Re

ð16Þ

where C1, C2, m1, m2 and n are constants. The constants, m1 and m2 appearing in Eq. (16) can be obtained from the correlations of the present smooth duct. 4. Validation of experimental results To validate the accuracy of the SAH system, f and Nu of the present smooth duct are verified with the published standard

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Fig. 3. Geometry of test section with combined punched-V-rib and chamfered-V-groove.

equations. This means that those measured results are, respectively, compared with data of Blasius and Petukhov correlations for f; and Dittus–Boelter and Gnielinski for Nu, found in Ref. [45] for turbulent duct flows as depicted in Fig. 5(a) and (b), respectively. The mentioned correlations are as follows: Correlations of Blasius,

f ¼ 0:316Re
0:25


2

ð18Þ

f-correlation of the current smooth duct,

0:351

ð19Þ

Re0:262

Correlations of Dittus–Boelter,

Nu ¼ 0:023Re0:8 Pr 0:4

ð20Þ

and Gnielinski,

ðf =8Þ ðRe
1000Þ Pr 1 þ 12:7ðf =8Þ

Nu Nu0



 0:755 f 2:738 = f 0 Re Re

ð16aÞ

5. Results and discussion

f ¼ ð0:79 ln Re
1:64Þ

Nu ¼

 TEF ¼

ð17Þ

and Petukhov,

f0 ¼

Thus, the calculation of TEF can be achieved by substituting m1 = 0.262 and m2 = 0.755 from Eqs. (19) and (22), into Eq. (16) and then it becomes

1=2

ðPr 2=3
1Þ

ð21Þ

Nu-correlation of the present smooth duct,

Nu0 ¼ 0:0327Re0 0:755 Pr0:4

ð22Þ

The deviations between the experimental and correlation’s data are within ±6% and ±8% for Dittus–Boelter and Gnielinski correlations and by ±8% and ±9% for Blasius and Petukhov equations, respectively. The current results agree quite well with the correlation results with deviation by ±9% each. Hence, the agreement of the two sets of values ensures the accuracy of the data collected in this experiment.

5.1. Surface temperature distribution of the absorber The distribution of surface temperatures (Ts) of the uniform wall heat-flux absorber mounted with the V-down rib-groove at RP = 1.0, 1.5 and 2.0, b = 45°, Re = 11785 is presented in Fig. 6. The axial variations of local wall temperatures come from 30 points of surface temperatures on the absorber, where x represents the first location of the upstream thermocouples placed on the absorber as seen in Fig. 4(b). From the figure, the profiles of local temperatures (Ts) have an increasing tendency with rising x/H for all RP values while show a gradual decline for the two endlocations (at x/H  35 and 39) owing to the outlet effect. The smallest value of Ts is found at RP = 1.0 and then, Ts at RP = 1.5 is found to be lower than that at RP = 2.0 because of high heat removal of the absorber. 5.2. Effect of V-down rib-and-groove 5.2.1. Heat transfer The heat transfer is presented in terms of the nondimensional convection heat transfer coefficient, Nusselt number (Nu) and Nusselt number ratio (NuR = Nu/Nu0). The variations of Nu and NuR with Re for the roughened SAH with combined the rib-groove

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Fig. 4. (a) Photograph of the absorber with combined punched-V-rib and chamfered-V-groove and (b) Schematic distribution of thermocouples on the absorber plate.

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Table 4 Parameters of combined turbulators. Parameters

Range

Aspect ratio (AR = W/H) Rib height (b) Rib blockage ratio (RB) Hole diameter of V-rib Rib attack angle (a) Hole inclination (b) Groove attack angle (a) Relative rib pitch ratio (RP) Relative groove pitch ratio (RP) Groove position to rib pitch ratio (g/H) Depth of groove (e) Chamfer angle of groove (/)

8 12 mm (fixed) 0.5 (fixed) 5 mm (fixed) 45° (fixed) 45°, 0° and
45° 45° (fixed) 1.0, 1.5 and 2.0 1.0, 1.5 and 2.0 0.5 (fixed) 5 mm (fixed) 45° (fixed)

devices arranged by V-tip pointing downstream on the absorber plate are depicted in Fig. 7(a) and (b), respectively. In Fig. 7(a), it is seen that the combined turbulators yield the considerable heat transfer rate with similar tendency with the smooth duct and Nu increases with decreasing RP. Nu for the combined turbulators with b = 45° seems higher than that with b =
45° and 0°. This is because of stronger vortex strength and impingement flow over the absorber plate by air jet passing through the hole, aside from more interruption of the boundary layer development. A close examination reveals that the use of combined turbulators leads to higher Nu than that of the V-down groove acting alone. Fig. 7(b) presents the distribution of NuR with Re for the compound turbulators placed on the absorber. As seen from the figure, NuR shows the slight decrease with rising Re. NuR of b = 45° is, as expected, much higher than that of b = -45° and 0° at similar operating conditions. At b = 45°,
45° and 0°, NuR values are about 5.74–5.96, 5.44–5.67 and 4.87–5.10; 5.35–5.56, 5.10–5.37 and 4.57–4.80; and 5.08–5.31, 4.78–5.00 and 4.22–4.48 times for RP = 1.0, 1.5 and 2.0, respectively. For the groove alone, NuR values are, respectively, around 1.95–2.23, 1.75–2.06 and 1.30–1.68 for RP = 1.0, 1.5 and 2.0.

Fig. 6. Axial profiles of local absorber-surface temperatures for the V-down ribgroove at RP = 1.0, 1.5 and 2.0, b = 45°, Re = 11785.

5.2.2. Flow resistance The experimental results of pressure drop in a SAH with combined V-down rib-and-groove placed on the absorber plate are presented in terms of friction factor (f) and friction factor ratio (fR = f/ f0). The influence of the combined V-down rib-and-groove on f and fR plotted versus Re is displayed in Fig. 8(a) and (b), respectively. As seen in Fig. 8(a), f declines as Re increases for all the cases. f values for the compound devices are extremely larger than those for the smooth duct alone due to the flow obstruction from both devices, the rise of surface area, and the swirling flow induced by the ribgroove turbulators. The results reveal that the mean f of b = 45° is around 10% and 16% higher than that of b = -45° and 0°, respectively, and f trends to increase with decreasing RP. The maximum f of the combined devices at b = 45°, RP = 1.0 is some 37.35 times above the smooth duct alone, similar to the Nu condition.

Fig. 5. Verification of the smooth duct, (a) f and (b) Nu.

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P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

Fig. 7. (a) Nu and (b) NuR against Re for the combined V-down rib-and-groove.

As displayed in Fig. 8(b), fR shows the uptrend with increasing Re and its values at b = 45°,
45° and 0° are, respectively, ranging from 27.62–37.35, 22.36–29.40 and 18.42–24.16; 25.18–33.78, 20.93–26.95 and 16.57–21.69; and 23.68–31.58, 19.08–25.38 and 15.26–20.02 for RP = 1.0, 1.5 and 2.0. At RP = 1.0, 1.5 and 2.0, fR values for the groove alone are respectively, in a range of 7.0–7.6, 4.8–5.1, and 2.4–3.2. It is worth noting that fR of the compound V-down devices is much higher than that of the groove alone around 240–373%, 276–517%, and 383–915% for RP = 1.0, 1.5 and 2.0, respectively.

5.2.3. Thermal performance The relationship between TEF and Re for the V-down rib-andgroove is portrayed in Fig. 9. As seen in the figure, TEF for both devices, exhibits the decreasing tendency with rising Re. In general, TEF of the combined turbulators found to be above unity is higher than that of employing the groove alone. TEF with RP = 1.5 and b = 45° seems to be the best amongst all the turbulators used and is around 2.25 at lower Re. In the current study, both devices at RP = 1.5 provide TEF higher than the ones at PR = 1.0 and 2.0 around 2.6–3.8% and 5–6%, respectively. TEF of

Fig. 8. (a) f and (b) fR versus Re for the combined V-down rib-and-groove.

P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

11

increases with decreasing RP. Nu of the groove alone is also included for comparison. It is noted that at similar RP, the b = 45° provides the highest Nu and is, respectively, about 6–8% and 11– 13% above the b =
45° and 0°. This can be explained from the interruption of the boundary layer development of the flow by the turbulators and the impingement of the air jet passing through the hole on the absorber surface leading to the drastic enhancement of heat transfer above the b =
45° and 0° cases having no impinging air jets on the absorber wall. In Fig. 10(b), the general tendency of NuR for the combined V-up turbulators seems to decrease slightly for rising Re. The combined turbulators all show the superior heat transfer performance. The average increases in NuR for b = 45°,
45° and 0° are in the range of 6.23–6.52, 5.98–6.29 and 5.36–5.62; 5.77–6.10, 5.56–5.90 and 4.98–5.23; and 5.45–5.70, 5.22–5.47 and 4.65–4.88 times while the groove alone yields the NuR around 2.25–2.53, 2.04–2.35 and 1.56–1.92 times at RP = 1.0, 1.5 and 2.0, respectively. In the Re range studied, the combined V-up turbulators give NuR in a range of 4.65–6.52, higher than the groove alone around 138–278%.

Fig. 9. TEF versus Re for the combined V-down rib-and-groove.

the V-down turbulators is about 58–75% over that of the groove alone. 5.3. Effect of V-up rib-and-groove 5.3.1. Heat transfer The distributions of Nu and NuR with Re for the SAH with the combined V-up rib-groove devices are, respectively, depicted in Fig. 10(a) and (b). According to Fig. 10(a), the employ of combined V-up turbulators yields the considerable increase of heat transfer with the same trend as the smooth duct and as observed, Nu

5.3.2. Flow resistance The distributions of f and f/f0 with Re for the roughened absorber plate are, respectively, displayed in Fig. 11(a) and (b). It is apparent in Fig. 11(a) that the combined V-up turbulators provide the substantial increase in f above the smooth duct alone. The high friction loss is mainly due to increasing the surface area, larger obstruction of flow and strongly reversing flow by the turbulators. The V-up turbulators at b = 45° give much higher f than the ones at b = -45° and 0°, however. A close examination reveals that f increases with the declines of RP and Re. In Fig. 11(b), fR shows the rising trend with the increase of Re for all cases. The average fR values for b = 45°,
45° and 0° are, respectively, about 35.26, 28.20 and 24.28; 32.64, 26.65 and 21.70; and 30.75, 25.52 and 19.78 times at RP = 1.0, 1.5 and 2.0. As expected, fR of the small RP is much higher than that of the high one. In comparison, fR values of the combined V-up turbulators at RP = 1.0, 1.5 and 2.0 are, respectively, in a range of 200–318%, 252–402% and 288–624% above that of the groove alone.

Fig. 10. (a) Nu and (b) NuR versus Re for the combined V-up rib-and-groove.

12

P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

Fig. 11. (a) f and (b) fR against Re for the combined V-up rib-and-groove.

5.3.3. Thermal performance of V-up turbulators The TEF variation with Re for the V-up turbulators is depicted in Fig. 12. In the figure, TEF generally seems to be above unity for both the combined V-up turbulators and the V-up groove alone, indicating that the use of both the turbulators and the groove alone results in the advantage over the smooth duct, especially for the combined devices. Throughout the experimental results, it is apparent that the b = 45° gives the highest TEF and its maximum is seen at RP = 1.5. In general, TEF is found to have the rising tendency with decreasing Re for all cases investigated and the peak around 2.47 is at b = 45°, RP = 1.5 for the combined V-up turbulators. The reason why TEF decreases with the increase of Re comes

Fig. 12. TEF versus Re for the combined V-up ribs-and-grooves.

from the fact that at lower Re, there are large NuR and small fR values as seen in Fig. 10b and Fig. 11b, leading to high TEF as calculated by Eq. (16a). On the other hand, at higher Re, NuR and fR values have the reversing trends, resulting in lower TEF. At b = 45°,
45° and 0°, the values of TEF are approximately 2.06–2.38, 2.13–2.47 and 2.02–2.31; 1.97–2.28, 2.03–2.35 and 1.95–2.24; and 1.90–2.18, 1.93–2.21 and 1.87–2.15 for RP = 1.0, 1.5 and 2.0, respectively. The application of the V-up turbulators leads to higher TEF than that of the groove alone around 56–77%. 5.4. Effect of turbulator arrangements For comparison purpose, the results of the combined punchedV-rib and chamfered-V-groove for both V-up and V-down arrangements at b = 45° and RP = 1.5 are compared with the combined solid-V-rib and chamfered-V-groove at similar RP at which the highest thermal performance appears. Thus, the turbulators at RP = 1.5 only are investigated by plotting the NuR, fR and TEF against Re, as depicted in Fig. 13. As seen in the figure, NuR and fR of the combined solid-V-rib and chamfered-V-groove are considerably higher than those of the combined punched-V-rib and chamfered-V-groove, especially for the V-up case. It is interesting to note that TEF values for the combined devices are above unity and have the reducing trend with rising Re. The V-up turbulators give higher TEF than the V-down ones for all cases. This can be explained from the fact that the V-up can create the commonflow-down vortices [24] which can induce the impingement flow behind the V-tip area aside from the impinging air-jet issuing from the downward hole, leading to drastic increase in the heat transfer rate. The V-down produces the common-flow-up vortices [24] which induce the impinging flow on the wall opposite the absorber which is the insulation wall, leading to lower NuR as can be observed in Fig. 13. Also, fR values for both the V-down and the V-up are nearly similar and are considerably lower than those for both the V-down and the V-up solid-rib-and-groove owing to the punched hole effect. The highest TEF around 2.47 is achieved for the V-up punched-V-rib-and-chamfered-V-groove and it is higher than the V-down one, the V-up and the V-down solid-V-rib-andchamfered-V-groove, around 7–9%, 14–15% and 19–20%, respec-

P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

13

Fig. 13. Effect of the rib-and-groove arrangement on NuR, fR and TEF at RP = 1.5.

Fig. 14. Predicted versus experimental data for V-down turbulators, (a) Nu and (b) f.

tively. This implies that the combined V-up punched-V-rib and chamfered-V-groove is the best choice as a promising compound device for enhancing heat transfer rate in practical use. 6. Development of correlations for Nu and f The empirical correlations are expected to be fruitful for the Nu and f prediction in a SAH system or for determining the optimal

configuration conditions in engineering applications. Both the measured data were fitted using a least-square regression analysis. The detail on establishing the empirical correlations can be found in Refs. [46–48]. Nu and f of the SAH with the combined turbulators correlated as a function of Re, Prandtl number (Pr), punchedhole inclination angles (b = 45°, 0° and
45°) and relative pitch ratio (RP = 1.0, 1.5 and 2.0) are formulated as summarized in Table 5. Also, the correlations of the V-groove alone are incorpo-

14

P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

Fig. 15. Predicted versus experimental data for V-up turbulators, (a) Nu and (b) f.

Table 5 Empirical correlations for all turbulators used in the current study. Correlation of turbulator

Error



Combined punched-V-rib and chamfered-V-groove, V-down arrangement Nu ¼ 0:238Re0:724 Pr 0:4 ð90 þ bÞ0:042 ðRP Þ
0:23

±10%

f ¼ 1:147Re
0:052 ð90 þ bÞ0:068 ðRP Þ
0:623 V-Groove alone, V-down

±10%

Nu ¼ 0:274Re0:629 Pr 0:4 ðRP Þ
0:516

±10%

f ¼ 3:809Re
0:292 ðRP Þ
1:448 Combined punched-V-rib and chamfered-V-groove, V-up arrangement

±10%

Nu ¼ 0:264Re0:72 Pr 0:4 ð90 þ bÞ0:045 ðRP Þ
0:211

±6.7%

f ¼ 1:302Re
0:055 ð90 þ bÞ0:056 ðRP Þ
0:583 V-Groove alone, V-up

±8%

Nu ¼ 0:252Re0:652 Pr 0:4 ðRP Þ
0:467

±6.7%

f ¼ 3:014Re
0:248 ðRP Þ
1:232

±8%

rated. Comparisons of Nu and f data predicted by the correlations and measurements for the combined turbulators with V-down and V-up arrangements are depicted in Figs. 14 and 15, respectively. The comparison signifies that data predicted by the correlations are found to be in good agreement with measured data. The discrepancies between both the measured and the predicted data are also summarized in Table 5. 7. Conclusions An experimental investigation on behaviors of thermal and flow resistance in a SAH duct with the combined punched-V-rib and chamfered-V-groove, and the V-groove alone has been carried out. The influences of operating Re, rib-groove arrangement, longitudinal rib-groove pitch ratios (RP) and punched-hole inclination angle (b) on thermal performance of the SAH duct are examined. Key findings in the current work can be summarized as follows:  The novel combined turbulators can generate LVG of the main flow that helps to increase the flow intensity and to move the









central flow to the near-wall flow regions while the air jet passing through the downward hole can provide the impinging flow on some area of the absorber. All turbulators perform the superior heat transfer to the smooth SAH duct alone. The highest heat transfer enhancement appears for the combined punched V-rib and chamfered V-groove with V-up arrangement at b = 45° and RP = 1.0. The combined turbulators with V-up arrangement are recommended because they have an advantage on thermal performance over the ones with V-down. The maximum TEF around 2.47 is found for the combined V-up punched-V-rib and chamfered-V-groove at b = 45° and RP = 1.5, where Nu and f increase at about 6.52 and 38.67 times above the smooth duct alone. The hole inclination angle must be downward, b = 45° to achieve the impinging jet on the absorber surface. The b = 0° is not recommended since the air jet from the hole will deteriorate the strength of vortex/swirl flow appearing behind the ribs resulting in poor heat transfer. In comparison, the combined punched-V-rib and chamfered-Vgroove provide the augmentation of thermal performance at about 14–15% higher than the combined solid-V-rib and chamfered-V-groove and also at about 56–77% above the groove alone. The Nu and f empirical correlations for the combined turbulators are developed for design.

Declaration of Competing Interest The authors declare that there are no conflicts of interest.

Appendix A The data uncertainties of friction factor, Reynolds number and Nusselt number are obtained from the expressions below.

P. Promvonge, S. Skullong / International Journal of Heat and Mass Transfer 143 (2019) 118486

For friction factor: Df f

¼

1 f

n

@f DðDPÞ @ ðD P Þ

¼

n

o2

DðDPÞ DP

where DðDP Þ DP

¼ Dhh and

DRe Re

¼

o2 n o2 n o2 0:5 @f @f þ @L DL þ @D DD þ @ ð@fReÞ DRe n

o2

h

þ

DL 2 L

_ 2 Dm _ m

þ

þ

3DD 2 D

þ

2DRe 2

0:5

Re

DD 2 i0:5 D

For Nusselt number: DNu Nu

¼

1 Nu

h

@ 2  @ 2 i0:5 þ @D ðNuÞDD þ @k ðNuÞDk n 2  o0:5 2 Dh ¼ þ DDD h

2 @ ðNuÞDh @h

in which h ¼ q00 =ðT s
T b Þ, then Dh h

¼

1 h

n

¼

@h @q00

Dq00

n

Dq00 q00

where q00 ¼ p0:5 DL

o2

o2

n

@h @T s

þ þ

n

h 2  V R

DT s

DT s T s
T b

o2

o2

þ

þ n

n

@h @T b

DT b T s
T b

DT b

o2 0:5

o2 0:5

i _ p ðT bo
T bi Þ þ mC

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