A B C D E F G H I J K L M N O P Q R S T U V W X Y Z

SURFACE CONDENSERS

DOI: 10.1615/AtoZ.s.surface_condensers

In a surface condenser vapor is brought into contact with a solid surface which is cooled to a temperature below the saturation temperature of the vapor at its prevailing partial pressure. The surface is usually in the form of a "nest" or "bundle" of metal tubes, the coolant flowing inside the tubes and the vapor condensing on the outer or "shell-side."

The Overall Heat Transfer Coefficient, U (based on the condensing surface area), may be expressed in terms of local coefficients, a, via the sum of thermal resistances:

where D0,Di are the tube outer and inner diameters and λw is the thermal conductivity of the tube material.

Heat exchange in single phase flows is usually impeded by the presence of insulating boundary layers. On condensation, however, the large reduction in volume as the vapor turns to condensate results in an inflow of vapor towards the surface; the heat transfer is impeded only by a thin film of condensate on the surface. As a result, the condensation heat transfer coefficient is usually higher than that on the liquid coolant side, the latter becoming the controlling process. The thermal resistance, F, of fouling on the coolant side is therefore an important consideration. Condensing heat transfer rates are typically two orders of magnitude higher than the rates for a gas on the shell side, so condensers do not generally require extended surface tubing. In some applications spirally grooved tubing has been used to aid drainage of the condensate film and to increase internal heat transfer by turbulating the coolant flow. If nonwetting can be maintained (e.g., by surfactants), dropwise condensation occurs, reducing the areas of condensate film and resulting in a heat transfer coefficient some four times that of filmwise condensation.

Vapor and condensate heat transfer coefficients, αv and αf, were first combined by Nusselt in 1916 [see, e.g., McAdams (1954)] in an expression for the effective condensing coefficient for a single horizontal tube assuming that the condensate forms a laminar film on the tube surface. However, the Nusselt formulation is less successful in predicting tube nest mean heat transfer coefficient. Vapor shear and inundation by condensate from tubes higher in the nest affects the condensate film. More importantly frictional pressure losses as the vapor passes between the closely pitched tubes lowers the partial pressure—and hence saturation temperature—reducing the heat exchange driving temperature difference. (See also Condensation and Condensers.)

The essential objective in surface condenser design is to provide equal access of vapor to all the surface. Early attempts to provide overall heat transfer coefficients for the design of steam condensers took no account of the detailed layout of the tubenest, e.g., HEI (1978). Modern practice is to model the proposed nest on computer to calculate thermal performance and to ensure that any non-condensable gases present are extracted at the point of lowest pressure, e.g., Rhodes and Marsland (1993). Poorly designed nests may suffer excessive frictional pressure losses and contain regions where the tubes are blanketed by noncondensable gases.

Power station condensers are some of the largest heat exchangers in existence, typically containing some 20,000 tubes of 25 mm outer diameter and 20 m in length. Computer models usually represent the nest as an array of cells, as shown in Figure 1, solving simultaneously the equations of continuity, energy and momentum for each cell. For such models detailed correlations for local heat transfer and frictional pressure loss are required; developments in this complex field are reviewed by Davidson (1987). Surprisingly, there appears to be no optimum arrangement of tubes, the wide variety of nest configurations in service being shown by Lang (1987).

Figure 1. 

Where cooling water is not available, condensers are being designed for direct cooling by air. Arrays of large diameter fans blow the air across banks of finned tubing containing the condensing vapor, e.g., Knirsh (1990).

Surface condensers for process industries are usually smaller than power plant condensers and may contain a region for desuper-heating the incoming vapor. Unlike power plant condensers, they may also be designed to subcool the condensate. Again a wide variety of arrangements are in use incorporating baffled shell-and-tube, gasketed plate and spiral plate constructions and plate-fin exchangers. Useful discussions of developments are given by Bell (1983).

REFERENCES

Bell, K. J. (1983) Trends in design and application of condensers in the process industries, Condensers: Theory and Practice, The Institution of Chemical Engineers Symposium Series No. 75.

Davidson, B. J. (1987) Thermal design of condensers for large turbines, Aerothermodynamics of Low Pressure Steam Turbines and Condensers, Ch. 8, M. J. Moore and C. H. Sieverding, Eds., Hemisphere Publishing Corp.

H. E. I. (1978) Standards for Steam Surface Condensers, Heat Exchange Institute, New York.

Knirsh, H. (1990) Design and construction of large direct cooled units for thermal power plants, PWR-26, ASME Joint Power Conference.

Lang, H. V. (1987) Steam condenser developments, Aerothermodynamics of Low Pressure Steam Turbines and Condensers, Ch. 8, M. J. Moore and C. H. Sieverding, Eds., Hemisphere Publishing Corp.

McAdams, W. H. (1954) Condensing Vapors, Heat Transmission, Ch. 13, McGraw-Hill Book Co. Inc.

Rhodes, N. and Marsland, C. (1993) Improvement of Condenser Performance using CFD, European Conference on the Engineering Applications of Computational Fluid Dynamics, Inst. Mech. Engrs 7-8 September 1993, London.

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