Document yp2anD699gjYe14GYKJ5x1aod

644 CHAPTER 38 1965 Guide And Data Book Table 2 .... Comparison of Compressor Performance Compressor Speed (rpm) 1735 3450; ' Operating conditions Condensing t^p F. Liquid temp F Evaporator temp F Suet, pressure poa :Suct. gas temp F 135 111 ' 10 29.33 HO . 130 115 . 45 91.81 95- ' Capacity (Cooling effect at the evap oratormeasured on the calorimeter). 980 . 21,800 ' ' Coefficient of performance.. 4.'45 9.6 ` Power (Watts) - 220+ ; 2274 ` Power Distribution Work of compression Motor losses .- Compressor losses ....... > / - 116 ft 1580 67 r`*" 450 37.4 . 245 Fig. 7 .... Typical Performance Curves for a Rotary Compressor wall meets the front head rface and rear head face. The hydrodynamic sealing is dependent on clearance, surface speed, oil visoosity and surface finish of-the parts. Smoother finish and closer clearance is used when low viscosity oil is used, most often in the gmaH machines. Larger machines havegreater clearance and therefore are'operated with a higher viscosity oiL If the finish, clearance, alignments and oil viscosity are correctly chosen, internal leakage is reduced to a trace. There is, however, an unavoidable leakage that is due to the presence of refrigerant diluted in die oil which flows from the high pressure side to the suction side of the machine. Re-expansion losses are small since it is inherently a ample matter to make. the clearance space negligibly small. It is de sirable however to purposely introduce a controlled amount of re-expansion sufficient to retain in the cylinder some of the oil that Is held along the leading edge of the blade. This is done by means of transfer slots in the throat of the machine. The slots perform, the two important functions of supplying more oil to the cylinder wall and of providing a larger exit to the discharge ports. If these features are combined with suitable values for the radius ratio, overcompression losses are also minimised. For a given value of radius ratio, overoompresoon is increased as,the shaft speed is increased, therefore smaller radius ratios are required in the two-pole machines than may be used in the four-pole machines. Fortunately it is not neces sary to accept a severe penalty of high overcompression in the rotary machine if the compressor dimensions are selected with this requirement in mind. Noise ' The ability to achieve an acceptable noise level is an impor tant consideration in the design of any small compressor par ticularly where it is intended for use' on products which go into the home. Fig. 8 illustrates a convenient method for analyzing and evaluating a noise condition. This chart is typical for rotary compressors in a number of respects) Since the machine is inherently well-balanced little or no difficulty, arises in the first, second arid thinT frequency bands. In the middle frequency band the noise* is in large discrete frequen cies associated with the discharge gas. The upper frequencies,' particularly the seventh and eighth hand.*, are .wide band fre quencies mechanically induced by the moving parts. In prac-, tice various means for further reducing noise' level are em ployed. The machine may be internally spring mounted in the hermetic shell Such designs seek to contain the mounting problem completely in the compressor assembly. .Hie more usual form however, is to employ an external spring mounting:' In this form two additional requirements must be met. First, all possible .means of transmitting vibration from the'dis1' charge gas column into' the cabinet sheet metal or. in some' cases the condenser and tubing as well, must be'scnipuously avoided. Second) when high frequency hands'exhibit a high noise level the sound is absorbed either by enclosing the ma-' chine compartment or metalling acoustical'sound absorbing materials; a combination is.usually employed. The flow of suction gas is relatively smooth and produces so little noise that a'suction side muffler is not'required.)A muffler on the discharge side can.be used effectively particularly, on larger sizes. One of the most important benefits of a discharge Table 3 . ... Breakdown of Compressor Lasses* ' Compressor Size (Btuh) (Rated)' Speed (rp>") . Displacement (cu in. per rev) .: 975 (R12): 12,500 (R22); . 1750 - -3500 0.545 ... l.W* . Operating Conditions ' 10 F Evap` 45 F Eva'p 135 F Cond 130 F Cond > Total.Compressor Losses 37.4 Watts 245 Watts' -. -100% .-100%-; 1. Discharge Side'Losses (%) . A. Overcompressioo , B. Re-expansion ... ' ' 12!3. ' 2. Suction Side Losses.(%)" A'.'Pressure drop B. Heat transfer 6.42. 3. Internal Leakage (%)i..: A. Gas carried by oil circulation . B. Leakage past clearance 6.42 4. Oil pumping (%) 2.67" 5. Power to Overcome Friction (%) A. Bearing friction losses 19.25 B. Blade friction losses. - 10.16 C. Sealing film tones 42.8 51 ,. 12.25". 3.47 ` ' -2.04 ' ! 10.2 14.7'.,.. 16-32 * Tbe croup of mkchme* adeeted far this ttady ins mods up entirely of rrhiTrit--d r--H pirf-T-rf----t1 1 *M si Compressors 645 OCTAVE PASS BANOS IN CYCLES PER. SCCONO: ->7 1-75 - 1*0-30-80-1 eoo-24oo-4aoo-+soo 'dl-------1 E------- 3E--3E-------3t FOR rrpic AL 17 SO RP U CO MPRES SOR A = FF EC- s ACC; COUP RESSOR ONI-Y . * l ICROF HONE AT 1 a inc HES) B ST AKOAJW) REFRICERATOF TES r k-. \ r (AME ( OMPR ESSOR TESTCO IN A) :\ : k; i -^ " :h*Y ^ :k ; : ks k L- A: Ld N r l ej .____l SO 3--L * a_jl 00 >4S$-Jt) too -200 500 1000 2000- SOOO 10,000'--, FREQUENCY IN CYCLES PER SECOND Fig. 8 .... Typical Sound Level Analysis of Rotary Compressor muffler is that it prevents the pulsating discharge gas stream from vibrating the connecting tubing) and .the condenser.' A common cause ofnoiseis-the:vibration transmitted by these members to the cabinet sheet metal panels.Tn general it:will be observed-that the noise level-for a rotary'compressor is also related to horsepower and to speed. Machines.of suc cessively larger horsepower, rating exhibit successively higher noise levels than machines of smaller horsepower, in the same design.- Increasing-the speed, for em-mpte from 4-pole .speed to 2-pole speed, increases the noise level but not uniformly in all the frequency bands. In this case, increased noiseTevel is most prominent in the high frequency bands and there is very little change in most of toe low and middle frequency, bands. .. Shafts - -.1 - FEATURES-. The requirements for rigidity in the shaft to meet deflection allowances that'Bre compatible with minimum permissible film thickness in the bearing, results in a shaft which surpasses the strength requirements based on the other'load and'torque, considerations. An exception'to this criterion may.occur when an. outboard bearing is used, particularly in small machines when a~ unMtmiumrffr shaft wnw adequate1 so'far as the compressor itself is concerned: Where a smal+diahieter shaft is to be considered, particular attention! must be given to avoid flexing at the armature end! .The flexing'of a small diameter shaft permits a gyrating motion of the armature' to develop and permits toe attendent forces together frith; the eccentric pull in the'air gap to build up excessively unless enough rigidity is provided by the shaft. - The selection of the shaft materials and shaft dimensions is based bn toe stiffness required and a' deflection 'ahalysb as well as the harrinoga and finish requirements of the journals. The'joumals are ground as accurateiy as possible and honed or polished to a finish of 5 raicroinch RMS .or better; The hardness should be 56 R* or harder. ;'- In small machines the cantilever loaded shaft has been' used very successfully. In larger sizes and'for higher head pressure, an outboard bearing should`.secure support on both sides !of the load.' .'r.--i : Journals and Bearings Stated briefly the bearing must support the rotating mem ber under all conditions over toe intended range of use. The bearing must be completely reliable, immune to possible dam age during run in, trouble free during a very long service life, require-no attention or maintenance of any kind, and exhibit a minimum of wear. Fortunately, a well designed sleeve bear ing will meet these requirements. The chief disadvantage of the sleeve bearing is that there is that there'-is no fast simple way of designing one.' A pro cedure that relies on past experience and merits of the bearing in another situation, judged to be sufficiently similar, is not likely to provide toe best design for toe application.. Experience shows that a cast-iron bearing with a hardened steel journal possesses superior qualities. The reason for this is not completely understood but the experience is verified by actual test. One possible explanation may be that this combi nation of materials maintains' a tough film of metallic com pounds on the bearing and journal surfaces and thereby retard oil degradation and prevent metallic contact. The grey cast-iron bearing must be a high quality fully peariitic nonporous casting. It should be ASTM Type A graphite and contain no ferrite, with a desired hardness range 187 to 220 BHN. Blades Blades are designed for maximum reliability. Particular attention is given to the choice of materials and lubrication, The slots are hardened, ground and polished to the best finish obtainable; above 56 R. and about 3 micro-inch RMS. A cast iron carefully selected for resistance to wear seems to be one of- the best blade materials. It lubricates well and has a modulus of elasticity that permits adequate load distribution at the slot edge. One material that performs well is a grey cast iron, ASTM Type A graphite; fine grain full peariitic, ferrite free, and in the hardness range 245-320 BHN. Peak loads in excess of 200 lb per inch of cylinder height have been used satisfactorily. <- From the designers point of view a decision is first made as to toe number of blades to be used and the blade dimension required for toe blade loading at maximum'conditions.-The slots are made long enough to support the blade adequately at maximum blade extension. This usually requires that the slot be made as long as possible. - ' The usual number of blades is two although three or more blades-may be used where the analysis of the compression cycle shows that discharge from the compression chamber can begin before over-compression occurs. Blade thickness is made sufficent to limit the deflection under load to 0.00005 in. or less per inch of blade length. The shape of the blade tip must be made to conform to surface of generation along the cylinder wall through 360 deg of rotation of the shaft since neither too large a radius nor too small a radius will lubricate properly or seal well Springs or other forms of mechanical linkage to control blade movement are not used in the sliding vane machine! Better performance and greater reliability have been achieved through hydrodynamic control of blade dynamics. The general requirement is that the blade tip maintain continuous contact with the cylinder walL The under blade hydrodynamic pressure is the most important of the several forces acting on the blade against the force of compression tending -to raise tire tip from the cylinder wall The underblade pressure is cyclic due-to the motion of the blade in the slot1and:the presence of oil behind toe -blade. The average underblade