Heat Transfer Enhancement in Externally Finned Tubes and Internally Finned Tubes and Annuli pp 7-30 | Cite as
Round Tubes Having Plain-Plate Fins
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Abstract
The performance characteristics of round tubes having plain-plate fins are discussed in this chapter. The concepts of wavy fins, louvred fins, multi louvred fins, fin numbering, etc. are explained in detail. Staggered and in-line arrangement of tube bundles and fin spacing are the other topics covered here. The correlations for configurations are also presented.
Keywords
Plain-plate fins Wavy fin Louvred fin Multi-louvred fin In-line arrangement Staggered arrangementHaghighi et al. (2018) conducted an experimental investigation in natural heat convection on thermal performance and convective heat transfer coefficient of plate fins and plate cubic pin-fin heat sink. He conducted the experimental investigation for Rayleigh number range of 8 × 10^{6} to 9.5 × 10^{6}, heat input range of 10–120 W. Fin spacing and fin numbers were varied between 5 and12 mm and 5 and 9, respectively. They investigated the effect of fin spacing and number of fins of plate fins and plate cubic pin fin on thermal resistance and heat transfer. They found that plate cubic pin fin having 8.5 mm fin spacing and seven fins was better than plate-fin heat sink. They developed empirical correlations for average Nusselt number as a function of number of fin plates, fin spacing to height ratio as well as Rayleigh number.
Dimensions of test fins (Haghighi et al. 2018)
Fin type | Fin shape | Fin number | Fin spacing (mm) | S/H |
---|---|---|---|---|
Type A | Plate pin fin | 5 | 12 | 24/90 |
Type B | Plate pin fin | 7 | 8.5 | 17/90 |
Type C | Plate pin fin | 9 | 5 | 10/90 |
Type D | Plate cubic pin fin | 5 | 12 | 24/90 |
Type E | Plate cubic pin fin | 7 | 8.5 | 17/90 |
Type F | Plate cubic pin fin | 9 | 5 | 10/90 |
Heat transfer coefficient correlations for different fin patterns and pitch ratios for a given fin height of 10 mm (Didarul Islam et al. 2008)
Pattern | \( {\overline{Nu}}_{\mathrm{overall}}=c{\mathit{\operatorname{Re}}}^{0.7} \) | ||
---|---|---|---|
c | |||
PR = 2 | PR = 3 | PR = 3.5 | |
Co-angular | 0.163 | 0.153 | 0.149 |
Zigzag | 0.175 | 0.176 | 0.168 |
Co-rotating | 0.261 | 0.223 | 0.212 |
Co-counter rotating | 0.191 | 0.169 | 0.191 |
This is because of strong flow interactions accompanied with vortex attack on the end wall and fin surface. The fins with co-angular pattern have showed minimum pressure drop. They have observed the smoke flow pattern around the fins and oil titanium oxide flow pattern on the end wall. They reported that they were both in good agreement. They have also observed that the flow over co-angular fin plate was governed by horseshoe vortices while wavy flow behaviour was dominant in the case of zigzag fin pattern. This is because the diverging fin pairs generate longitudinal vortices which attack the end wall and fin surfaces together. In case of co-counter rotating fin pattern, the flow was only slightly disturbed due to converging fin pairs. They concluded that the fin with co-rotating pattern with pitch ratio 2 and fin height 10 mm has shown the best thermal performance among all the fins considered. Also, the heat transfer using co-rotating fin pattern was noted to be threefold that of the duct without fins.
Torii and Yanagihara (1997) worked with vortex generators; Sparrow et al. (1982, 1983) studied rectangular fin arrays; and Kadle and Sparrow (1986), Turk and Junkhan (1986), Oyakawa et al. (1993), Molki et al. (1995), El-Saed et al. (2002) and Bilen and Yapici (2002) have carried out similar investigations for heat transfer enhancement and pressure drop characteristics. Fabbri (1998, 1999), Zeitoun and Hegazy (2004), Olson (1992), Alam and Ghoshdastidar (2002), Saad et al. (1997), Kumar (1997), Yu et al. (1999), Liu and Jensen (1999), Sarkhi and Nada (2005), Wang et al. (2008a, b, c), Eckert and Irvine (1960), Yu and Tao (2004), Shih et al. (1995) and Park and Ligrani (2005) have carried out similar works.
Atayılmaz and Teke (2009, 2010), Ahmadi et al. (2014), Taghilou et al. (2014), Park et al. (2014), An et al. (2012), Al-Arabi and Khamis (1982), Popiel et al. (2007), Na and Chiou (1980), Chae and Chung (2011), Qiu et al. (2013), Chen and Hsu (2007), Haldar et al. (2007), Mokheimer (2002), Elenbaas (1942) and Beckwith et al. (1990) have all worked with fins for natural convection heat exchanger applications.
Murali and Katte (2008) presented the performance of radiating pin fin having threads, grooves and taper on the outer surface. They concluded that the heat transfer rate from the radiator using threaded, grooved and tapered fin was about 1.2–3.7 times more than that from a solid radiating pin fin. Wilkins (1960), Kumar and Venkateshan (1994), Krishnaprakas (1996), Ramesh and Venkateshan (1997), Krikkis and Razelos (2002, 2003), Chung and Nguyen (1987), Schnurr et al. (1976), Black and Schoenhals (1968), Black (1973), Gorchakov and Panevin (1975, 1976), Bhise et al. (2002), Srinivasan and Katte (2004) and Holman (2000) have also worked on radiating fins.
Both pressure drop contributions are evaluated at the same minimum area mass velocity. On many occasions, it may happen that Reynolds number based on hydraulic diameter will not correlate the effect of fin pitch.
McQuiston (1978), Gray and Webb (1986), Kim et al. (1999) and Wang et al. (2000) correlated j and f data versus Reynolds number for plain fins on staggered tube arrangements; the accuracy level of predictions, however, widely vary.
The friction factor with the tubes is obtained from a correlation for flow normal to a staggered bank of plain tubes. Zukauskas (1972) and Incropera and Dewitt (2001) give the tube bank correlation. McQuiston (1978) correlation based on the same data, however, does a poor job.
Dimensions of tube bundles (Mon and Gross 2004)
Staggered | In-line | |||||||
---|---|---|---|---|---|---|---|---|
s1 | s2 | s3 | s4 | s5 | il | i2 | i3 | |
Tube outside diameter, d | 24 | 24 | 24 | 24 | 24 | 24 | 24 | 24 |
Fin diameter, d_{f} | 34 | 34 | 34 | 44 | 44 | 34 | 34 | 34 |
Fin height, h_{f} | 5 | 5 | 5 | 10 | 10 | 5 | 5 | 5 |
Fin thickness, t_{f} | 0.5 | 0.5 | 0.5 | 0.5 | 0.5 | 0.5 | 0.5 | 0.5 |
Fin spacing, s | 1.6 | 2 | 4 | 0.7 | 2 | 1.6 | 2 | 4 |
Fin pitch, S_{f} = s + t_{f} | 2.1 | 2.5 | 4.5 | 1.2 | 1.2 | 2.1 | 2.5 | 4.5 |
Transverse tube pitch, S_{t} | 40.8 | 40.8 | 40.8 | 52.8 | 52.8 | 40.8 | 40.8 | 40.8 |
Longitudinal tube pitch, S_{1} | 35.33 | 35.33 | 35.33 | 45.73 | 45.73 | 40.8 | 40.8 | 40.8 |
Number of rows, n | 4 | 4 | 4 | 4 | 4 | 4 | 4 | 4 |
For three or more rows, the entrance length-based correlations do not do justice to the data over the entire Reynolds number range (200 < Re_{Dh} < 800).
Use of smaller diameter finned tube heat exchanger is the recent trend; Kim et al. (1999) improved Gray and Webb (1986) correlation by including the data of Wang and Chi (2000) and Youn (1997) for heat exchangers having smaller diameter tubes. The improvement in prediction was noteworthy. A more general correlation is that of Wang et al. (2000).
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