Document oDoYdEKLMLnLaQr3kXkK6rvVr

634 CHAPTER **8 1965 Guide And Date Bock Table 1 .... Design Featu^w.pf Reciprocating Compressors Amawua Cobj.-: -tt 1J-'S' ' . rt- - ,-.v . .. a `O- o ''' 1. Number of cvlinH#**--On* tn .16 12 2. HK Kangs H to 3. Cylinder Arrangement -150 - 150 a. VerticaLV or W, Radial Xf X n - b. Radial, Horizontal Opposed -r c. Horizontal, Vertical, V or.W 4,.Drive a. Hermetic Compressors, In duction Electric Motor ~'x' - ' b. Open Compressors--Direct Drive, V Belt, Chain, Gear, by Electric Motor or En gine X 5. Lubrication--Splash- or Force Feed, Flood XX , 6. Suction and Discharge Valves, --Ring Plate or Ring or Reed Flexing X 7. Ruction and Discharge Valve Arrangement . -! a. Suction and Discharge Valves in Head XX b. Uniflow--Suction Valves in Top of Piston, Suction ,. Gas Entering Thra .CyliSh; der Walls. Discharge Valves in Head. " '8. Cyiroder Cooling a.' Suction Gas Cooled b. Water Jacket Cylinder- Wall, Head, or Cylinder Wall and Head - - - c. Air Cooled d. Refrigerant Cooled Heads 9. Cylinder Head --- a. Spring Loaded Safetv ' X * x b. Bolted Head 10. Bearings--Sleeve, Anti-Frio-. tion X 1 6 12 Hto Up to 10 150-~x. X 'x ' X XX XX r- - . X V.: ' ` X- ' X X.. _ X __ Iftm Hcdocorboo Compressor Am- Comprettor /.. V -TT . U-*"-8> $ fI aS.' O 11. Capacity Control, if Provided --Manual or Automatic a. Suction Valve Lifting XX b. By-Pass--Cylinder Heads to Suction X e. Closing Inlet > ; . - x -X d. Adjustable Clearance - .X e. Variable Stroke /. Variable Speed x" 12. Materials Motor Insulations and Rub-: ber Materials must be Com patible with Refrigerant 22 and Oil Mixtures.' Otherwise no Restrictions. *x; X X No Copper or Brass - 13. Oil Return a. Crankcase Separated from Suction Manifolds, Oil Re- turn Check Valves, Equal. . izere, Spinnere, Foam .' . ' Breakers . X .X b. Crankcase' common with < Suction Manifold 14. Speeds 250, 1750. 1750 15. Pistons to. . to to 3600 3600 3600 :i u -v a. Aluminum or Cast Iron -X . : b. Ringless ...... .X c. Compression and Qil Cq'p- X .X trol Rings 18; Connecting Rod - > -- * Split Rod - with Removable Cap or Solid , . Eccentric Strap 17. Mounting XX Internal Spring Mount- X External Spring Mount * . Rigidly Mounted on Base-, - X __ X X1 X X 250. to 1800 X b. Suction strainer. - c. Motor (hermetic compressor).- d. Suction end discharge manifold, i e. Suction and diaoharge Valves. .* . /. Internal muffler. . ,, 2. Heat gain of refrigerant in compressor from 'the following sources: - a. Hermetic motor. b. Friction. c. Host of compression--heat exchange within the compressor. 3. Mechanical action of valves--such aa inertia ana valve spring rate. . J.-. < ' . 4. Gas leakage past the piston. 6. Oil circulation. Since oil is used in the'crankcase of recipro* eating compressors for lubrication of the bearing surfaces, it is virtually impossible to prevent small quantities from passing out with the discharge gas. Actually, some oil circulation u desirable because it reduces noise and lubricates parts not otherwise ac cessible. However, excessive oQ circulation effects' the`'efficiency of heat transfer surfaces, the compressor..overall.efficiency, and the refrigerant oil mixture properues.,The.effect'on evaporator capacity is shown in Fig; "1: The'slmoimt of"oil'circulation'is normally held below 3'percent by weight.'-The'refrigeration sys tem must be designed to effectively return all oil put into circula tion by the compressor. . . , ..\p . 6. Clearance. All of the gas cannot be pushed.,oyt of the cylinder during the compression stroke. Some will' remain in' pockets in the cylinder, in the discharge ports,' and in the clear ance space between the piston at top dead center and the valve plate. This gas is called the clearance gas and it will expand on the suction stroke until it reaches the'suetion ga pressure. Only at this point can additional gas be drawn into the'cylinder.-For-this reaeon, the entire compressor displacement is not available for pumping, , ir.................... . .. '7. Deviation from isentropic compression- When considering the ideal compressor, an isentropic compression cycle is assumed. In the. actual compressor,. the compression cycle deviates from isentropic compression due mainly to fluid and m<wh*nif*l fric tion within the cylinder..The actual compression cycle and work of compression must bo determined from an indicator card. It would be a difficult matter to consider all of them* losses individually. They can, however, be grouped;together *nH considered by category.'Their effect upon ideal compressor performance is measured by the following efficiencies: Compression Efficiency considers only, what occurs within the cylinder. It is ,a measure of the deviation of the actual compression from the isentropic .compression. Compression efficiency is defined as the ratio of the work required for isen tropic compression of the gas to the work done within the cylinder (as.obtained by an indicator.diagram).. Mechanical Efficiency is defined.as the ratio, of.the work delivered to.the gas (as obtained by an indicator,diagram) to the work delivered to the compressor shaft, Volumetric Efficiency is defined as the volume of fresh vapor Compressors 635 entering the cylinder per stroke divided by the piston dis placement Volumetric efficiency based on clearance volume alone csd be calculated as: ., - 1 + c(l - ) ,P> where j * volumetric efficiency. TT C" clearance, percent f Pt - absolute.pressure at end of compression.' Pi -- absolute, pressure at beginning of compression. " Experience indicates that the overall volumetric efficiency will run 15 to 25 percent lower, depending on the compression ratio, than the theoretical volumetric efficiency based on clearance alone. ...... Brake Horsepower is a function of the power input to the ideal compressor and the compression,' mechanical,- and volu metric efficiencies .of the compressor.,-.l .. Converting the ideal power ;input.per pound of gas.to horsepower: \ ' \`.n *s tcAcre 1 - thp 2544.65' (thp) x'frj >Ce| bhp " V, X St x. e. ;(3) (4) bhp brake.horsepower: ' *, ^ thp -- theoretical adiabatic horsepower per pound of gas per hour. V4 m displacement of compressor; cubic feet per-hour.- I .1 Vi -- specific volume of gas, cubic feet per pound! '' '. ; ei = volumetric efficiency. ""- - i = compression efficiency. 'j Cl * merhamrjQ *ffipi*n<y ... .< < . . Liquid Subcooling is not accomplished by the compressor, and the increase in capacity resulting from liquid subcooiing should not be credited to the compressor. .... ,'j; Suction Gas Superheat^ - The'amount'ofsuctiohgas superheat has several'effects oh compressor performance.-Lett'tKan'&^lb^-'d^^superheit-is generally undesirable because it results in continuous liquid bleed back to the compressor with resulting oil dilution and foaming and possibly liquid slugging. Unless all liquid-is evaporated in or before it enters the cylinder, it will evaporate during re-expansion and thereby reduce the capacity. It'also matron expansion valve control difficult. Where liquid carry over is necessary in a flooded evaporator to insure oil return to the'compressor, heat exchangers should be used to insure that dry gas enters'the compressor. Beyond 10 F deg an in crease in'superheat has very little effect on the compressor capacity and efficiency. Increase in superheat will increase tiie discharge temperature and, in hermetic compressors, the motor temperature. Less refrigerant is dissolved in the oil, and compression becomes noisier. Tests on several compressors showed an increase in capacity of 0.3 to 1.0 percent per 10 F deg increase in superheat in the range from 10--50 F deg super heat. . Power input was .reduced approximately 0.5 percent per 10 F deg superheat. The superheat was measured at the suction shut-off valve before the gas passed over the motor. Heat Re|ection The heat rejected by the condenser is the sum of the re frigeration effect and the heat equivalent of the power input to the compressor. Id a hermetic compressor, this heat rejec tion includes the inefficiency of the motor. Compressor heat rejection factors are necessary for sizing condensers. Motor Efficiency, Performance, and Starting Torque .The motor efficiency usually represents a compromise between cost and size. A motor can generally be made more efficient the larger it is for a given rating; Accepted efficiencies range from approximately 80 percent for a 3-hp motor to 92 percent for a 100-hp motor. Uneven loading has a marked effect.on motor efficiency; It is important that cylinders be spaced evenly. Also, the more cylinders there ore the smaller the impulses become. Greater moments of inertia of moving parts and higher speeds will reduce the impulse effect. Small and evenly spaced impulses are also beneficial from a noise and vibration standpoint. Since many compressors start against load, it is desirable to estimate starting torque requirements. The following equation is for. a single cylinder compressor and neglects the friction, the additional torque required to force discharge gas out of the.cylinder, and the fact that the tangential force at .the crankpin is not always equal to the normal force at the piston. This equation also assumes considerable gas leakage at the discharge valves, little or .no leakage, past the piston rings, and will yield a conservative estimate of starting torque. (P,-P,)A> 24dj/d* (P. - P,)A, " * 24Nt/Ni -.(5) (6> where,, . ; t Tt --'starting torque, pound feet. Pt -- discharge pressure, pounds per square inch, absolute.. Pi -- suction pressure, pounds per. square inch, absolute. > ' - A.--area of cylinder, square inches.- -- stroke of compressor, inches.... . <- - Nt - motor speed, revolutions per minute. *..... . ffx -- compressor speed, revolutions per minute.'.! . r Equation 5 shows 'that if pressures in the system are bal anced or almost equal, (Pt TM Pi)'the torque requirements are considerably reduoed. Thus, by using a pressure balancing 1 j