Document OzxXXdZ52d2kd5Qwj84ERpEgK

654 a nrrn s o uwiCk JO 963 Guide And Dota Book For those who prefer to'deal with isentropic analysis,'as efficiency can be defined as the ratio of the isentropic energy rise to the actual energy rise between suction conditions and the discharge (or condensing) pressure. If the suction and dis charge temperatures and pressures -.are known, it is then possible to determine the isentropic work and efficiency as follows: HS-.hj-'h. q. - ff.'/ff. '-/ (17) where = isentropic energy rise) Btu per pound.,' V a isentropic. enthalpy-at- discharge pressure and suction entropy, Btu per pound' ' - fig. 18 shows a typical-single-stage characteristic with polytropic head, H, plotted- against volume flow.Q, with lines of constant efficiency, all at various compressor speeds. i , A direct plot of head vs inlet volume flow can only be drawn for a given refrigerant at a fixed suction condition. A more universal plot, usable for the common halocarbon refrigerants under any conditions,', can)Tie developed'Ky"relating'flow, head, and impeller tip speed to the inlet, sonic velocity a. of the gas being-handled:-The shape.of the performance plot related to inlet sonic velocity is similar to that shown in Fig. 18, with the parameters being: Q/oc, H/af, and t/j/o* For optimum performance, the selection should be made in the high efficiency region of the compressor. Maximum sta bility for part load operation will be obtained when'the'seleo- tion can be made as far as possible from the surge line. For this reason, a selection'as shown at point A on'FIg. 18 is preferable to that at point B, although'either flow will give the same efficiency. ' - - ' . Compressor selections. for - two-,' three-,"or' four-stage' ma chines are performed'in a manner-similar'-to that-outlined above. The actual-inlet conditions'to each stage'are calculatorf in these cases,'-based on'performance of the previous stagedFor economizer cycles, such as shown in Fig. 10, it is necessary to calculate weight flows and inlet conditions separately for each stage by mixing the appropriate flows. - - The input gas horsepower'for each stage is then calculated by: WfH, 33,000 " where > P,' input gas horsepower. : - Wf = weight flow, pounds per minute. (18) . The total heat rejection to-the condenser is therefore: ' ' . ` . <j. **!`g + 0.212 pj ^ , ;(19) q,' -- "condenser load; tons. q evaporator (cooler), load,-tons.' '- . ' . I r: ..It may be necessary'toperform-the.calculations for more than one refrigerant, in order to select one with the most suit able characteristics. After a refrigerant has been selected by the approximate methods outlined, it is possible to derighor adapt the heat exchangers, reing the information given in Chapters'42 and 43. In standard water chillers uring Moling towers, both evaporators and condensers are usually HwagnaH for approaches from 6 to -10 F deg, with a nominal fouling factor of 0.0005. Actual selection of the design approach is governed by the,requirements.of the particular.application and will depend on the relative importance of first cost and operating cost, and also upon the equipment available. As noted earlier in this chapter, one of the more effective and commonly used forms of capacity-control to meet the normal system requirement is the variable inlet guide vane. To obtain full effectiveness from such guide vanes, they should be designed to operate efficiently over a wide range of angles. If they are mounted in an axial flow duct, they are very similar to the blade row of an flow compressor and may bede- rigned in accordance with axial flow compressor design methods. In order to minimize starting torque, prewhirls should close reasonably tight, but should maintain sufficient clearance to admit enough gas to prevent)overheating of the compressor. : - In the case of standard water chilling machines, manu facturers' selection tables have been developed, which make possible the immediate selection of a combination of heat ex changers and compressors to meet the requirements. In such tables, the net rated capacities are usually tabulated for sev eral leaving condenser water temperatures to cover the most common range of design conditions. Interpolation may th^u be used for intermediate values. For each listed condenser leaving water temperature, the net rated capacities are tabulated for various combinations of pass arrangements in both evaporator and condenser and for the most common range of leaving chilled-water temperature. Interpolation may again be used for the leaving chilled-water temperature falling within the rang** given. Opposite each evaporator and condenser pass arrangement are usually listed the allowable limits of water flow, both minimum and maximum, for any given selected pass arrangement. The specified water flow rates must fail within the - tabulated ranges for the pass arrangements selected. Curves are also . usually provided to give the water pressure drops as functions of the number of passes and the water flow: .The tabulated net capacity ratings are usually based on rated motor-brake horsepower and kilowatt input, unless otherwise indicated. A nominal fouling factor of 0.0005 is usually applied throughout! A' number of selections using different combinations will usu ally be required in order to obtain the most economical selec tion for the specified requirement. BIBLIOGRAPHY G. F. Wislicenus: Fluid Mechanics of Tvrbomachinery (M6-_ Graw-Hill Book Co., New York, 1947). =- A. H. Church: Centrifugal Pumps and Blowers (John Wiley and SonSj-Inc., New York, 1950). A. J. otepanoff: Turboblowers (John WQey and Sons, Inc;,' New York, 1955). - D. G. Shepherd: Principles of Tvrbomachinery (The Macmillan Company, New York, 1956).' T. B. Ferguson: The Centrifugal Compressor Stage (Butterworth and Company, London, 1963). ` "........... ..... F. J. Wiesner and H. E. Caswell: How refrigerant properties affect impeller dimensions (ASHRAE. Joubnal, 'October' 1959, p. 31). F. J.- Wiesner: Practical Stage Performance Correlations for Cen-4 trifvgal Compressors (ASME Paper 60 HYD-17, March 1960). J. M. Schuhs: The polytropic analysis of centrifugal compres^ sors (ASA/E Transactions, January 1962. VoL 84, p. 69). H. E. Sheets: Nondimensional compressor performance for. a range of Mach numbers and molecular weights (ASME Trans- actions, January 1962, p. 93). O. E. Balje: A contribution to the problem of designing radial turbomachines (ASME Transactions, May 1952, p. 451). D. M. Church: Centrifugal gas compressors (Chemical Engi neering, March 5, March 19, nd April 2, 1962). Standard for Application and Ratings of Centrifugal Liquid- Chilling Packages (ARI Standard 555-63, Air-Conditioning and Refrigeration Institute, 1963). CHAPTER 39 STEAM-JET REFRIGERATION EQUIPMENT r Basic Concepts, System Components, Steam Ejector Design Considerations, Two-Fluid Systems, Performance HE stcam-jet.refrigeration,cycIe has been used.for oyer joyed wide application in the field of comfort air conditioning. Tfifty years, yet it is one of.the least commonly employed This apparent paradox can be understood through knowledge refrigeration cycles in use today.,The'lack of popularity of,the of the baric characteristics of the cycle.' . steam-jet cycle stems from some inherent-performance limita-, In certain process cooling applications, the steam-jet re tions despite the baric simplicity of the cycle' itself. Improved frigeration unit has proven to have outstanding character equipment design.is gradually broadening the.application of istics. In those applications -where direct vaporization is used this cycle.and its future popularity will depend upon the im-. for'concentration or drying of'foods and chemicals, the1 use provements that manufacturers can incorporate into the of heat exchangers is eliminated.1This is of particular advan equipment., '- : - ; / BASIC CONCEPTS ;; _ tage in processing heat-sensitive foods or chemicals where the cooling produced by evaporation reduces the processing tem perature and prevents product deterioration. The concentra : The steam-jet refrigeration ,cycle is quite similar to more conventional refrigeration cycles with an-evaporator, a com pression device, a condenser, and a refrigerant'as the baric system- components. Instead of a mechanical compression device, the system characteristically employs a steam ejector or booster to compress the refrigerant-to the condenser pre&r sureJeveL ' , , ,t . . .Water is'used as the refrigerant-and.the cooling effect.is produced in the steam-jet refrigeration cycle by the continu ous vaporization of a part of the water in the evaporator.at-a low absolute pressure level. When chilling water from 55 F to 45 F, about 1 percent of the water flowing through the evapo^ rator must be vaporized. The pressure in an evaporator operating at 45 F will be 0.30 in. of mercury absolute and, due to the high specific volume of steam at this low pressure, about 400 cfm of water vapor per ton of refrigeration must be compressed to the condenser pressure level. This should be compared to-a Refrigerant 12'system in which the suction pressure will be about 40 psig and the corresponding flow rate about 4 cfm per ton of refrigeration. Steam-jet ejectors arc ideally suited to handle such.large volume flows, although they must typically operate over a pressure ratio (absolute condenser pressure to absolute evaporator pressure) of 6 to 8 which is rather high for efficient ejector performance. Since tion of fruit juices, the freeze-drying process of preserving food, the dehydration of pharmaceutical products,- 6uch' as antibiotics, the crystallization of-many chemicals and foods, and-the chilling of leafy vegetables are typical examples of steam-jet system coolihg applications in which mechanical refrigeration is often either not suitable or not competitive. In'common practice, steam-jet refrigeration systems pro duce chilled water at temperatures between 35 to 70 F, which would include the ranges used-for comfort-air conditioning (35 to 50 F is used for this purpose). However, in process work where the processed substance (which may not be-water) is directly evaporated, vaporization at higher temperatures is common. t:;,. Since water is the refrigerant in a1 steam-jet refrigeration unit, it cannot be operated at evaporator temperatures below 32 F. due to freezing, and in comfort cooling applications evaporator temperatures above 50 F are impractical due to the inability to adequately control humidity in the condi tioned space. Thus, with water as a refrigerant, the useful range . of evaporator, temperature is. thermodynamically restricted. . , Steam-jet refrigeration systems are available commercially in 30 to "500 ton capacities, although they have been built in both smaller or larger sizes. . ..... the motive steam energizing the ejector must be condensed along with the water vapor drawn from the evaporator, it is SYSTEM COMPONENTS necessary to reject 2 to 3 times the amount of heat in the A steam-jet refrigeration system consists of an evaporator condenser that a conventional mechanical refrigeration cycle (or flash tank), a steam-jet ejector (or booster ejector), a con would require. Thus, a larger condenser and a relatively denser, and a two-stage - ejector noucondensable pump. higher cooling water flow must be employed in steam-jet units. ... Auxiliary components include-a..chilled water circulating pump, a cooling water circulating .pump, a condenser con In order to efficiently maintain a low- absolute pressure in densate pump, and necessary valves and control elements. the condenser (for example, 2.24 in. of mercury absolute at The pumps and;valves are the only system components which 105 F condensing temperature) it is common to employ a contain moving parts. Figs. 1. and 2 show two of the mpst small two-stage ejector package.which- removes air and other common baric sytfemsand their component arrangements. noncondensable gases from the condenser and pumps them up /to atmospheric pressure where they are discharged. Evaporators *. `' Despite the advantages of simplicity' ruggedness of design, freedom from vibratioh~'higbTreliability, low maintenance, aQd low cost, the steam-jet refrigeration cycle has not en- The evaporator is usually a large volume vessel which must provide a large water surface area for efficient evaporative cooling action.* Water sprays (Fig:. 1) or cascading, water in Th* gnrml responsibility for this chapter ia --ip--* to TC 5.6. Absorption sheets (Fig. 2) are two common'means, of maximizing-water surface. . 655