QAVlQ G. T H O M A S P. H. HAYES
HighmPerformance Heat Coefficient increased by combining r e c t a n g u l a r fins o n condensing side and t w i s t e d t a p e i n s e r t o n flowing w a t e r side
overall heat transfer coefficient between conThedensing steam and water flowing in a vertical aluminum tube was increased to values 2.5 to 4 times the value for smooth tubes by combining rectangular fins on the condensing side and twisted tape insert on the flowing water side. The heat transfer coefficient increased with increasing water velocity and decreasing heat flux. The maximum heat transfer coefficient was 5400 Btu/hr ft2 O F (compared with 1300 Btu/hr ft2 O F for the smooth tube) when the water velocity was 20 ft/sec and the heat flux was 2 X lo4 Btu/hr ft2. The overall heat transfer coefficient between condensing steam and a thin film of evaporating water was increased to values 3 to 10 times the value for smooth tubes by combining rectangular fins on both the evaporating and condensing side. The performance of the double finned tube decreased with increasing heat flux. The maximum value of the heat transfer coefficient, 5000 Btu/hr ft2 O F , was observed at a heat flux of lo4 Btu/ hr ft2. The present results are compared with data from the literature for other high-performance surfaces to illustrate the range of options in surface shape and the magnitude of potential improvements available to the distillation plant designer. 4
INDUSTRIAL A N D ENGINEERING CHEMISTRY
Heat transfer surface area contributes some 40y0 to overall capital costs of long tube vertical evaporator (LTV) (5, 8, 74) and multistage flash (MSF) (8, 27) distillation plant costs. One of the most promising areas for water cost reduction is the development of highperformance heat transfer surfaces (74, 27). Among the techniques that have been developed for increasing the forced convection heat transfer coefficient when there is no change of phase are roughening the surface, deforming the surface, or placing an insert in the tube (7). Since there is no general theory for turbulence promoters, determination of optimum configuration depends on experiment. Among the most promising techniques are twisted tape inserts (70, 75, 7s) and deformation of the tube to produce a “roped” configuration (78). When there is a change of phase, as in the condensation of steam, the only technique available for increasing the heat transfer coefficient was promotion of dropwise condensation (37) prior to 1954 when Gregorig discovered (72) the pronounced increase caused by fluted surfaces. With the discovery (6) that fluted surfaces can also markedly increase thin film evaporation coefficients, significant increases in LTV plant performance became feasible. Since then a variety of enhanced per-
1)
Transfer Surfaces
Figure 7 . Finned tubes (partially disassembled to show insert) A , jnned outside with detached promoter inside; B, finned outside with twisted tape inside; C, both sidesfinned
formance heat transfer surfaces have been discovered. Fabrication experience, materials requirements, and economic evaluation are beyond the scope of the present paper. Information on some surfaces of current interest was presented at a recent symposium (25). The object of this paper is to present new results on heat transfer coefficients with improved surfaces and to compare these results with data from the literature to illustrate the range of options in surface shape and the magnitude of potential improvements available to the distillation plant designer.
Experimental Surfaces a n d Procedure Overall heat transfer coefficients were obtained for four different combinations of surface. I n all of the tests, the outside (condensing) surface was finned with fins of dimensions previously determined to be optimum (30). I n three of the tests, the heat was transferred by forced convection to flowing water: (1) a tube with no internal turbulence promoters; (2) a tube with internal detached turbulence promoters (37); and ( 3 ) a tube with an internal twisted tape insert (70, 19). I n the fourth series of tests the heat was transferred from condensing steam to evaporating water, and the tube was finned on both sides. Figure 1 is a photograph of the high-performance tubes used in this study.
I n all of the above cases, the condenser was a 4-ftlong, 1/2-in.-o.d. aluminum tube with 0.028-in. wall. The condenser tube was inserted into a n insulated glass pipe of 1 in. i.d. ; removal of insulation permitted verification that the tube was always operating with film condensation. Additional experimental details may be found elsewhere (30, 37). Finned condenser surface. The finned condenser surface was manufactured by clamping twelve 0.013 X 1/8 X 41-1/2-in.-long fins on the outside of the tube. The spacing between fins of 0.13 in. is near the optimum reported in a previous study (30). The width of 0.013 in. represents a balance between minimizing the amount of surface covered and structural requirements. T h e fin height is a function of flow rate of condensate a t the bottom of the tube and may be estimated (30) from
h
>
1.6
[-]v r D
SLPN
Clamping of fins to the surface was dictated by the nature of the optimization studies; the preferred method of manufacture would form the fins as a n integral part of the surface. Detached turbulence promoters. Detached turbulence promoters for the inside of the tube were fabricated by winding a wire in a spiral and welding four longitudinal wires on the outside of the spiral to hold the spiral away from the surface. Wire dimensions were based on the optimum values determined in a VOL. 6 2
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TABLE I.
CHARACTERISTICS
O F TUBES USED I N HIGH-PERFORMANCE HEAT TRANSFER STUDIES
Tube Characteristics Wall k/x, I.D., Length, Thick- Btulhr UP/Uo Material in. zn. ness ft2 OF. Min Max Enhanced Condensing Steam, No Enhancement of Flowing Water
High-Performance Surface
Rectangular finned condensing side Fluted condensing side Dropwise (organic additive) Dropwise (gold plate)
A1 cu cu 90-10 Cu-Ni
0.444 1/2 0.555 1/2
48 38 56-1,'s 5
0.028 1/16 0.035
51,000 34,000 75,000
0
1.83 1.55 1.70 1.34
2.86 1.56 2.73 1.62
Reference Present study Lustenader and Staub (1964) Bromley, et al. (1968) Erb, et al. (1965)
Both Condensing Steam and Flowing Water Enhanced
+ +
Finned tube detached turbulence promoter twisted tape Finned tube spiral Dropwise (gold plate) grooves Fluted tube twisted tape
+
+
A1 A1
0.444 0.444
48 48
0.028 0.028
51,000 51,000
2.17 2.50
3.31 Present study 4 . 2 3 Present study
Brass
0.652 1/2
a
38
0.049 1/16
15,000 34,000
2.08 2.57
2.20 2.67
Lawson, et al. (1966) Lustenader and Staub (1964)
11.3 5.00 6.82 4.31
Present study Thomas and Alexander (1969) Distillation Digest (1967) Thornas and Alexander (1969)
cu
Both Condensing and Evaporating Side Enhanced Double-finned tube Fluted-finned tube Textured-finned tube Double-fluted tube Q
A1 410-SS
0.444 2
48 36
a
e
a
0,028 0.004 0,035
cu
2
96
0,062
3.17 4.84 6.25 2.80
Not available.
-
, ..
. 4
5
20
30
previous study (37). For the 1/2-in. tube, this required the spiral to be wound with a pitch of 0.31 in. and a wire diameter of 0.035 in. The longitudinal runners were wires with a diameter of 0.035 in. Twisted tape promoters. The twisted tape insert was fabricated from 0.013-in.-thick metal and had a twist ratio (defined as length per 180" twist divided by inside tube diameter) of 3.8. This twist ratio approaches the smallest value that can be achieved without special fabrication techniques (75); prior studies have shown that the smaller the twist ratio, the greater the effect on the forced convection heat transfer coefficient. Evaporative side fins. There were 10 fins, 0.013 in. thick, intersecting at the center of the tube on the evaporative side. This gave substantially the same spacing as the fins on the condensing side. As in the condensing studies, this method of manufacture was dictated by the nature of the optimization studies. The preferred method of manufacture would result in fins as an integral part of the surface.
.
. 2
3
23
i E L O t I T Y Ifl/sec
Figure 2. Effect of high-Ferformance condensing side surface on overall heat transfer coeficient between condensing steam and flowing water
6
51,000 43,000 9,000 42,000
INDUSTRIAL A N D ENGINEERING CHEMISTRY
Results
The characteristics of the tubes used in the present study, as well as data on tubes used in other high-performance heat transfer tests (3, 7, 9, 78, 27, 32), are given in Table I. Values of the heat transfer coefficient for the high-performance tubes and for comparable smooth tubes (no enhancement) are given in Figures 2-4. The results in the table and figures are divided into three categories : 1. Only enhanced condensing side, no enhancement of flowing water side (potential decrease in pumping power because of shorter tube length)
c
5 XIO~
10
2
4
5 Xl03
IO"
2
4
VELOCITY ift/secl
7 XIO*~XIO~ 5 HEAT FLUX l 0 I u l h r f I 2 l
7 X t 0 4 3YI03 5
lo4
2
IO4
2
5
4
8Y.04
8 x104
HEAT FLUX iBlu/hr f f Z )
Figure 3. Combined effect of turbulence promoters and high-performance condensing side svrface on overall heat transfer coejicient between condensing steam andflowing water
Figure 4 . Effect of high-performance surfaces on the overall heat transfer coejicient between condensing steam and evaporating liquids
2. Both condensing side and flowing water side enhanced (maximum value of overall heat transfer coefficient between condensing steam and flowing water) 3. Both condensing and evaporating side enhanced T h e results for heat transfer from condensing steam to water flowing in smooth vertical tubes are shown in Figure 2 as overall heat transfer coefficient us. velocity of flowing water. I n all cases, the overall coefficient increased markedly as the water velocity increased, indicating that the water-side coefficient was controlling. At a velocity of 10 ft/sec, the maximum heat transfer coefficient was 2600 Btu/hr ft2 O F [finned tube of present study and fluted tube (27)]. The maximum ratio of promoted to unpromoted heat transfer coefficient was 2.7 to 2.8 [finned tube of present study and dropwise condensation (3)1. T h e results for heat transfer from condensing steam to water flowing in tubes with internal turbulence promoters are shown in Figure 3 as overall heat transfer coefficient us. velocity of flowing water. I n all cases, the overall coefficient increased appreciably as the water
velocity increased. Thus, even with turbulence promoters, the water-side coefficient was controlling. At a velocity of 10 ft/sec, the maximum heat transfer coefficient was 3500 Btu/hr ft2 O F at a heat flux of 2 X lo4 Btu/hr ft2 (finned tube with twisted tape insertpresent study). Enhancement of condensation by finned tubes is a function of heat flux (30) consequently increasing the heat flux from 2 X lo4 to l o 5 Btu/hr ft2 reduces the overall coefficient of the finned tube with twisted tape insert from 3500 to 2500 Btu/hr ft2 O F a t the 10 ft/sec velocity. Since the smooth tube results also decreased with increasing heat flux, the increase in relative promotion, Up/Uo,changed somewhat less, from 3.12 to 2.78.
AUTHORS D a v i d
G. T h o m a s and P. H. Hayes are a t the
O a k Ridge National Laboratory, O a k Ridge, T e n n . Research wasjointly sponsored by the O@ce of Saline W a t e r , U. S.Department of the Interior, and the U. Atomic Energy Commission, under contract with the Union Carbide Cor$.
s.
TABLE II. F R I C T I O N FACTOR INCREASES FOR WATER-SI D E ENHANCED PERFORMANCE TUBES Turbulence
Promoter
Friction Factor Ratio. f,/f Velocit_v,f t/sec Reference
2
5
10
20
Detached
3 2
3.3
3.35
3.4
Presentstudy
Twisted tape Spiral groove
3.3
3.2
3.0
2
2
3.1 2
a
Present study Lawson, et al. (1966)
Twisted tape
2.4
2.3
2.2
a
Lustenader and Staub (1964)
a
Not available.
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Values of the ratio of friction factor for the enhanced tube to friction factor with smooth tube, fP/f, are given in Table I1 for velocities of 2, 5, 10, and 20 ft/sec. The enhancing devices on the flowing-water side caused a n increase in pressure drop of a factor of 2 to 3.4. Since the maximum heat transfer increase did not occur a t the minimum pressure drop increase, the choice among the different combinations of surfaces depends on a detailed economic analysis. I t is noteworthy that the spiral grooved tube and dropwise condensation resulted in a doubling of the overall heat transfer coefficient with only a doubling of the pressure drop, and the finned tube with twisted tape insert tripled the overall heat transfer coefficient with only a tripling of the pressure drop. T h e results for heat transfer from condensing steam to evaporating water are shown in Figure 4 as overall heat transfer coefficient us. heat flux. For the four surface combinations shown here, the dependence of the heat transfer coefficient on heat flux was strikingly different for the different tubes. T h e overall coefficient for the double-finned (present study) and double-fluted tubes ( 7 7 , 32, 35) decreased with increasing heat flux; the finned tube showed the stronger dependence. The overall coefficient of the fluted-finned tube (32) and the textured-finned tube (7) increased slightly with increasing heat flux, both in upflow and downflow modes of operation. The largest overall heat transfer coefficient observed was 5000 Btu/hr ft2 "F a t a heat flux of lo4 Btu/hr ft2 using the double finned tube (present study). Instead of comparing the surfaces a t constant heat flux, it is probably more realistic to make comparison a t constant temperature difference, AT. If a AT of 7°F is assumed to be a reasonable value for interstage temperature difference, then the double-finned tube in downflow (present study) and the textured-finned tube in upflow (7) both produce greater than an eight-fold increase in overall heat transfer coefficient over the value for a smooth tube. All surfaces produced greater than a 4-1/2-fold increase. I n addition to the major effects of tube design, fluid velocity, and heat flux (or equivalently, overall temperature difference) illustrated in Figures 2-4, the overall heat transfer coefficient between condensing steam and evaporating water may be affected in a secondary way by such factors as composition of the liquid feed or absolute temperature. For instance, in five-effect evaporator pilot plant tests (73) of spirally grooved tubes, the overall coefficient with sea water (-3.5% NaC1) was about 40% greater than coefficients measured in the same system using fresh water (no NaCl). In other tests (73) in which spiral flow in the liquid film was induced by a spiral insert in the vapor-phase region of the tube, the overall heat transfer coefficient was increased from 800 Btu/hr ft2 "F 8
INDUSTRIAL AND E N G I N E E R I N G CHEMISTRY
a
IL d
2
5
t cc
+
C
I J
2
2
>
102
100
120
lQO
160
180
200
220
140
EVAP09AT NG SIDE TEMPERATLRE I'F)
Figure 5. Effect of evaporating side temperature on the overall heat transfer coeJcient between condensing steam and evaporating water
with no NaCl in the feed to over 1400 Btu/hr*ft2.OF with 5 to 6% by wt NaCl in the feed. Similar results have also been obtained in laboratory studies with a double fluted tube (24). Such a n increase is greater than can be accounted for by change in thermal conductivity, specific heat, or viscosity of water by the addition of NaCl; since increases of this magnitude are not observed during thin-film evaporation on smooth tubes, it appears that the effect is associated with the nonuniform distribution of liquid on the surfaces. The overall heat transfer coefficient for both smoothand high performance surfaces increases with absolute temperature. Limited data available (Figure 5) indicate a straight-line relationship between the log of the overall heat transfer coefficient and the evaporating side temperature for both enhanced performance and smooth tubes. Thus, a temperature coefficient, B, defined by
S = C + B t U is B = 0.0060 for smooth tubes ( 77, ZQ), 0.0027 for fluted tubes (24),and 0.012 for textured-finned tubes (27) when the temperature is given in degrees Fahrenheit. Implications
DESIGK.A variety of high-performance heat transfer surfaces and configurations are available now which permit plant designs with overall heat transfer coefficients in excess of 3000 Btiq'hr ft2 "F (compared with values of 600 to 800 Btu/hr ft2 "F for smooth tubes). Realization of these gains depends on use of thin wall, high thermal conductivity materials which are compatible with the intended service (2, 36) as well as reliable fouling control (76, 20, 22, 34) for extended periods of time. Optimum plant design may require combinations of different surfaces to take full advantage of the depen-
dence on heat flux flow velocity, temperature dependence of the heat transfer coefficient and the effect of feed composition on the evaporative heat transfer coefficient. RESEARCH. With high-performance surfaces (fluted, finned, corrugated, etc.) in which a thin film of liquid covers a t least part of the surface, gravity flow is probably not controlling. Although the mechanism is not well understood, it is believed (4, 72, 30, 33) that surface tension effects are primarily responsible for the high heat transfer coefficients. This is because surface tension tends to draw the bulk of the liquid in the film to grooves or corners in the surface where it drains away rapidly as a rivulet. Such thin liquid films are also known (23, 26, 28) to be vigorously agitated by surface tension gradients produced by small local concentration or temperature gradients. Differences in tube material or shape plus the rivulet of water on the surface may cause different heat flow patterns and the local temperature gradients required for development of local surface tension gradients. Thus, specially shaped surfaces which produce the marked improvement in thin filmevaporating and condensing coefficients may, because of this special shape, have substantially different performances when fabricated from materials with different thermal conductivity or with different metal wall thicknesses and shape. Limited data indicate that with high-performance surfaces the overall heat transfer coefficient responds differently to changes in the absolute temperature and brine concentration on the evaporating side from what it does in smooth tubes. Little is known of the mechanism for these effects or whether the 40 to 8070 increases observed (13, 24, 28) in evaporating saline solutions with some higher performance surfaces are the maximum possible increases. Additional fundamental studies of the mechanisms of thin-film evaporation and condensation are required to provide a guide for interpretation of experimental results and a firm basis for plant design. Acknowledgment
T h e authors wish to acknowledge the support and suggestions of G. Young and K. A. Kraus and the ingenuity and skill of W. A. Wilburn in fabricating the finned surfaces. Nomenclature = tube area, ft2 = temperature coefficient, 1 / O F = constant in Eq. 3 D = tube diameter, ft f = smooth tube friction factor, dimensionless fp = friction factor with turbulence promoters, dimensionless gL = gravitational acceleration, ft/secz h = fin height, ft = film heat transfer coefficient inside tube, Btu/hr ft2 O F h,
A B C
= film heat transfer coefficient outside tube, Btu/hr ft2 O F = thermal conductivity, Btu/hr ft2 OF/ft = number of fins, dimensionless = heat flux, Btu/hr ft2 R, = metal wall resistance, O F hr ft2/Btu R, = scale resistance, O F hr ft2/Btu t = temperature, O F A T = temperature difference, O F AU = overall heat transfer coefficient difference, Btu/hr ft2 O F U = overall heat transfer coefficient, Btu/hr ft2/ O F U p = enhanced overall heat transfer coefficient, Btu/hr ft2 O F U, = smooth tube overall heat transfer coefficient, Btu/hr ft2 O x = tube wall thickness, ft
h, k N Q/A
F
Greek Letters Y
r p
= kinematic viscosity = mass flow rate, lb,/sec = fluid density, Ib,/fts
ft
REFERENCES (1) Bergles, A. E., and Morton, H. L., M I T Engineering Projects Lab Report 5382-34. Feb. 1965. ( 2 ) Bohlmann, E. G., and Posey, F. A., Proc. First Int. Symp. on Water Desalination, Washington, D. C., Oct. 3-9, 1965, 1 , p 306. (3) Brornley,L. A., Porter, J. W., and Read, S. M., A.Z.Ch.E. J., 14,245 (1968). (4) Brown, C. E., 1967 Saline Water Conversion Rept., U. S. Department of I n terior, pp 87-88. ( 5 ) Burley, M. J., “Analytical $mnparison of Multistage Flash and Long T u b e Vertical Distillation Processes, Paper 16 presented a t the Second European Symposium on Fresh Water from the Sea, Athens, May 1967. (6) Diedrich, G. E., and ?tz, C. W., “Distillation Apparatus Having Corrugated HeatTransferSurfaces, U,S. Patent 3,291,704,Dec. 13, 1966. (7) Distillation Digest, 1 (l), 63-65 (Nov. 1967). (8) Ibid., 3 (l), 37 (1969). (9) Erb, R. A., and Thelen, E., IND.END.CHEM.,57 (lo), 49 (1965). (10) Gambill, W. R., and Bundy, R . D., “ A n Evaluation of the Present Status of Swirl Flow Heat Transfer,” A S M E Paper 62-HT-42. (11) General Electric Co., Office of Saline Water, Res. Develop. Progr. Rept. No. 181, April 1966. (12) Gregorig, Romano, Z A M P , 5, 36 (1954). (13) Hill, J. W., Office of Saline Water, Res. Develop. Progr. Rept. No. 367, July 1 , 1967. (14) Kays, D. D., Proc. Western Water Power Symp., Los Angeles, Calif., April 8-9, 1968, p C59. (15) Kidd, G. J., A.I.Ch.E. J . , 15, 58 (1969). (16) Knudsen J. G. “Industry’s Need for Research in Process Heat Transfer ” gaper presdted a t ;he Ninth National AIChE-ASME Heat Transfer Conferen&, eattle, Wash., Aug. 6-9, 1967. (17) Kroll, J. E., and McCutchan, J. W., Trans. A S M E J . Heat Transfer, 90, 201 (1968). (18) Lawson, C. G . , Kedl, R . J., and McDonald, R . E., ANS Trans., 9 (2), 565 (1966), also Office of Saline Water Ann. Rept., 1965, p 178. (19) Lo ina, R. F., and Bergles, A. E., Trans. A S M E J . Heat Transfer, 9 1 , 434 (19697. (20) Lu, C. H., and Fabuss, B. M., IND.ENO.CHEM.,PROCESS DES. DEVELOP.,7, 206 (1968). (21) Lustenader, E. L., and Staub, F. W., ZNCO Power Conf., Session 11, May 5-8, 1964. (22) McCutchan, J. W., and Sieder, E. N., Chem. Eng. Progr. Symp. Ser., 64 (go), 255-69 (1968). (23) Norman, W. S., andBinns, D . T., Trans. Inst. Chem. Eng., 38,294 (1960). (24) ORNL Staff, DistillatzonDie., 2 (l), 41 (1968). (25) OSW Symposium on Enhanced Tubes for Desalination Plants, March 11-12, 1969, Proc., to be published. (26) Ponter, A. B., Davies, G. A., Ross, T. K., and Thornley, P. G., Znt. J . Heat Mass Transfer, 10, 349-59 (1967). (27) Porter, J. W., Mech. Eng., 90 (l), 18 (1968). (28) Scphton, H . H., Office of Saline Water, Res. Develop. Progr. Rept. No. 361. (29) Standiford, F. C., and Bjork, H. F., “Advances in Chemistry,” 27, 115-27 (1960). (30) Thomas, D. G., A.I.Ch.E. J.,14, 644 (1968). 6, 385-90 (1967). (31) Thomas, D. G., IND.END.CHEM.PRDCESS DES.DEVELOP., (32) Thomas, D. G., and Alexander, L. G., Desalination, in press, (33) Thomas, D. G., and Young, Gale, unpublished, IND. ENG. CHEW, PRDCESS DES. DEVELOP.,in press. (34) Reitzer, B. J., IND.ENO.CHEM.PROCESS DES.DEVELDP.,3,345 (1964). (35) U.S. Department of Interior, Saline Water Conversion Rept., p 158 (1964). (36) Van Berkel, F. C. A. A., Van Hasseli, J. W., and Van Der Torren, J. H Paper SWD/94 presented a t the First Int. Symp. Water Desahnation, Washing! ton, D. C., Oct. 3-7, 1965. (37) 223, Williams, CE 367, Nov. A. G.,1968. Nandapurkar, S. S., and Holland, F. A,, Chem. Engr., No.
.,
(38) Zuber, Novak, and Staub, F. W., Ind. J . Heat Mass Transfer, 9, 897-905 (1966).
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