refrigeration manual system design p-4 ae104

2002 Copeland Corporation.Printed in the U.S.A.

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AE-104 January, 1986Part 4SYSTEM DESIGN

Section 17. BASIC APPLICATIONRECOMMENDATIONS

Fundamental Design Principles ...............17-1Compressor Selection ..............................17-1System Balance ....................................... 17-2Refrigerant ............................................... 17-2Compressor Cooling ................................17-2Compressor Lubrication ........................... 17-3Oil Pressure Safety Controls.................... 17-4Oil Separators .......................................... 17 4Suction Line Accumulators ...................... 17-5Pumpdown System Control ..................... 17-6Crankcase Heaters ..................................17-6Crankcase Pressure Regulating Valves... 17-7Low Ambient Head Pressure Control ....... 17-7Liquid Line Filter-Driers ............................ 17 8Sight Glass and Moisture Indicator ..........17-8Liquid Line Solenoid Valves ..................... 17-8Heat Exchangers...................................... 17-9Thermostatic Expansion Valves ...............17-9Evaporators .............................................. 17-10Suction Line Filers ....................................17-11High and Low Pressure Controls .............17-11Interconnected Systems .......................... 17-12Electrical Group Fusing ............................ 17-12

Section 18. REFRIGERATION PIPING

Basic Principles of RefrigerationPiping Design ........................................... 18-1Copper Tubing for Refrigerant Piping ...... 18-2Fittings for Copper Tubing ....................... 18-2Equivalent Length of Pipe ........................ 18-5Pressure Drop Tables ..............................18-5Sizing Hot Gas Discharge Lines ..............18-9Sizing Liquid Lines ...................................18-14Sizing Suction Lines .................................18-17Double Risers........................................... 18-22Suction Piping for Multiplex Systems ....... 18-23Piping Design for Horizontal andVertical Gas Lines ....................................18-24Suction Line Piping Design at theEvaporator................................................ 18-25Receiver Location ....................................18-25Vibration and Noise ..................................18-27Recommended Line Sizing Tables ..........18-28

Section 19. LOW TEMPERATURESYSTEMS

Single Stage Low Temperature Systems.19-1Two Stage Low Temperature Systems ....19-2Volumetric Efficiency................................19-2Two Stage Compression andCompressor Efficiency ............................. 19-4Compressor Overheating at ExcessiveCompression Ratios.................................19-6Basic Two Stage System .........................19-6Two Stage System Components .............19-7Piping on Two Stage Systems .................19-9Cascade Refrigeration Systems ..............19-13

Section 20. TRANSPORT REFRIGERATION

Compressor Cooling ................................20-1Compressor Speed ..................................20-1Compressor Operating Position...............20-2Compressor Drive ....................................20-2Refrigerant Charge...................................20-3Refrigerant Migration ...............................20-3Oil Charge................................................ 20-3Oil Pressure Safety Control .....................20-4Oil Separators ..........................................20-4Crankcase Pressure Regulating Valve ....20-4Condenser................................................ 20-4Receiver ...................................................20-5Purging of Air From System .....................20-5Liquid Line Filter-Drier..............................20-6Heat Exchanger .......................................20-6Liquid Line solenoid Valve .......................20-6Suction Line Accumulator ........................20-6Crankcase Heaters ..................................20-7Pumpdown Cycle .....................................20-7Forced Air Evaporator Coils .....................20-7Thermostatic Expansion Valves...............20-7Defrost Systems.......................................20-8Thermostat............................................... 20-8High-Low Pressure Control ......................20-8Eutectic Plate Applications .......................20-9Refrigerant Piping ....................................20-12Vibration ...................................................20-12Electrical Precautions...............................20-13Installation................................................ 20-13Field Troubleshooting on Transport Units20-14

Section 21. CAPACITY CONTROL


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2Revised 4/02

Copeland is a registered trademark of Copeland Corporation. Emerson isa registered trademark and service mark of Emerson Electric Co.

Compressors with Unloaders ................. 21-1Hot Gas Bypass ..................................... 21-3Bypass Into Evaporator Inlet .................. 21-3Bypass Into Suction Line ....................... 21-5Solenoid Valves for Positive Shut-offand Pumpdown Cycle ............................ 21-5Desuperheating Expansion Valve .......... 21-5Typical Multiple EvaporatorControl System ...................................... 21-6Power Consumption withHot Gas Bypass ..................................... 21-7

Section 22. LIQUID REFRIGERANTCONTROL IN REFRIGERATION AND AIRCONDITIONING SYSTEMS

Refrigerant-Oil Relationship ................... 22-1Refrigerant Migration ............................. 22-2Liquid Refrigerant Flooding .................... 22-2Liquid Refrigerant Slugging .................... 22-2

Tripping of Oil Pressure Safety Control . 22-2Recommended Corrective Action .......... 22-3

Section 23. ELECTRICAL CONTROLCIRCUITS

Typical Lockout Control Circuit .............. 23-1Control Circuit for CompressorProtection Against LiquidRefrigerant Flooding............................... 23-1Control Circuits to Prevent Short Cycling22-4Control Circuits for Compressors withCapacity Control Valves ......................... 23- 4


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INDEX OF TABLES

Table 21 Recommended Minimum Low Pressure Control Setting .......................................... 17-11Table 22 Dimensions and Properties of Copper Tube ............................................................. 18-3Table 23 Weight of Refrigerant in Copper Lines ...................................................................... 18-4Table 24 Equivalent Length in Feet of Straight Pipe For Valves and Fittings .......................... 18-5Table 25 Pressure Drop Equivalent for 2 F. Change in Saturation Temperature at

Various Evaporating Temperatures .......................................................................... 18-7Table 26 Maximum Recommended Spacing Between Pipe Supports for Copper Tubing ...... 18-28Table 27 Recommended Liquid Line Sizes.............................................................................. 18-29Table 28 Recommended Discharge Lines Sizes ..................................................................... 18-30Table 29 Recommended Suction Line Sizes, R-12, 40 F....................................................... 18-31Table 30 Recommended Suction Line Sizes, R-12, 25 F....................................................... 18-32Table 31 Recommended Suction Line Sizes, R-12, 15 F....................................................... 18-33Table 32 Recommended Suction Line Sizes, R-12, -20 F. .................................................... 18-33Table 33 Recommended Suction Line Sizes, R-12, -40 F. .................................................... 18-34Table 34 Recommended Suction Line Sizes, R-22, 40 F....................................................... 18-34Table 35 Recommended Suction Line Sizes, R-22, 25F........................................................ 18-35Table 36 Recommended Suction Line Sizes, R-22, 15 F....................................................... 18-36Table 37 Recommended Suction Line Sizes, R22, -20 F....................................................... 18-37Table 38 Recommended Suction Line Sizes, R-502, 25 F..................................................... 18-37Table 39 Recommended Suction Line Sizes, R-502, 15 F..................................................... 18-38Table 40 Recommended Suction Line Sizes, R-502, -20 F. .................................................. 18-39Table 41 Recommended Suction Line Sizes, R-502, -40 F. .................................................. 18-40Table 42 Efficiency Comparison of Single Stage vs. Two Stage Compression Typical

Air Cooled Application with Refrigerant R-502 .......................................................... 19-6Table 43 Recommended Discharge Line Sizes for Two Stage Compressors ......................... 19-10Table 44 Recommended Liquid Line Sizes for Two Stage Compressors................................ 19-10Table 45 Recommended Suction Line Sizes for Two Stage Compressors, -60 F. ................ 19-11Table 46 Recommended Suction Line Sizes for Two Stage Compressors, -60 F. ................ 19-11Table 47 Recommended Suction Line Sizes for Two Stage Compressors, -80 F. ................ 19-12Table 48 Recommended Suction Line Sizes for Two Stage Compressors, -80 F. ................ 19-12


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SECTION 17BASIC APPLICATION RECOMMENDATIONS

FUNDAMENTAL DESIGN PRINCIPLES

There are certain fundamental refrigeration de-sign principles which are vital to the proper func-tioning of any system.

1. The system must be clean, dry, and free fromall contaminants.

2. The compressor must be operated within safetemperature, pressure, and electrical limits.

3. The system must be designed and operated sothat proper lubrication is maintained in thecompressor at all times.

4. The system must be designed and operated sothat excessive liquid refrigerant does not enterthe compressor. Refrigeration compressorsare designed to pump refrigerant vapor, andwill tolerate only a limited quantity of liquidrefrigerant.

5. Proper refrigerant feed to the evaporator mustbe maintained, and excessive pressure drop inthe refrigerant piping must be avoided.

If these give steps are accomplished, then opera-tion of the system is reasonably certain to betrouble free. If any one is neglected, then eventualoperating problems are almost certain to occur.These basic fundamentals are closely inter-re-lated, and must always be kept in mind with regardto the application of any component, or wheneverany change in system operation is contemplated.

COMPRESSOR SELECTION

The compressor must be selected for the capacityrequired at the desired operating conditions inaccordance with the manufacturers recommen-dations for the refrigerant to be used. StandardCopeland single stage compressors are approvedfor operation with a given refrigerant in one of thefollowing operating ranges.

EvaporatingTemperature

High Temperature 45 F. to 0 F.or 55 F. to 0 F.

Medium Temperature 25 F. to -5 F.

Low Temperature 0 F. to -40 F.

Extra Low Temperature -20 F. to -40 F.

Operation at evaporating temperaturesabove the approved operating range may overloadthe compressor motor. Operation at evaporatingtemperatures below the approved operating rangeis normally not a problem if the compressor motorcan be adequately cooled, and discharge tempera-tures can be kept within allowable limits. Evaporat-ing temperatures below -40 F. are normally be-yond the practical lower limit of single stage opera-tion because of compressor inefficiencies and ex-cessive discharge gas temperatures. Because ofproblems of motor cooling or overloading, somemotor-compressors may have approval for opera-tion at limited condensing or evaporating tempera-tures within a given range, and if so, these limita-tions will be shown by limited performance curveson the specification sheet.

A given compressor may be approved in twodifferent operating ranges with different refriger-ants, for example, high temperature R-12 and lowtemperature R-502. Since the power require-ments for a given displacement with both R-22 andR-502 are somewhat similar, in some cases acompressor may be approved in the same operat-ing range for either of these refrigerants.

Two stage compressors may be approved forevaporating temperatures as low as -80 F., butindividual compressor specifications should beconsulted for the approved operating range.Operation at temperatures below -80 F. is nor-mally beyond the practical efficiency range ofCopeland two stage compressors, and for lowerevaporating temperatures, cascade systems shouldbe employed.


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Compressors with unloaders have individually es-tablished minimum operating evaporating tem-peratures since motor cooling is more critical withthese compressors. As the compressor is un-loaded, less refrigerant is circulated through thesystem, and consequently less return gas is avail-able for motor cooling purposes.

Copeland motor-compressors should never beoperated beyond published operating limits with-out prior approval of the Copeland ApplicationEngineering Department.

SYSTEM BALANCE

If the compressor or condensing unit selected for agiven application is to satisfactorily handle therefrigeration load, it must have sufficient capacity.However, over capacity can be equally as unsatis-factory as under capacity, and care must be takento see than the compressor and evaporator bal-ance at the desired operating conditions. Check-ing the proposed system operation by means of acompressor-evaporator-condenser balance chartas described in Section 16 is recommended.

If fluctuations in the refrigeration load are to beexpected, which could result in compressor opera-tion at excessively low suction pressures, thencome means of capacity control must be providedto maintain acceptable evaporating temperatures.If compressors with unloaders are not available orsuitable, and if the load cannot be adequatelyhandled by cycling the compressor, a hot gasbypass circuit may be required.

REFRIGERANT

Copeland compressors are primarily designed foroperation with Refrigerants 12, 22, and 502. Op-eration with other refrigerants in cascade systemsmay be satisfactory if the proper motor and dis-placement combination is selected, adequate lu-brication can be maintained, and if adequate com-pressor protection is provided. All applications withrefrigerants other than R-12, R-22, and R-502must be approved by the Copeland ApplicationEngineering Department.

R-502 is highly recommended for all single stagelow temperature applications, and particularly where

evaporating temperatures of -20 F. and belowmay be encountered. Because of the undesirablehigh discharge temperatures of R-22 when oper-ated at high compression ratios, R-22 should notbe used in single stage low temperature compres-sors 5 HP and larger.

Different expansion valves are required for eachrefrigerant, so the refrigerants are not interchange-able in a given system, and should never be mixed.If for some reason it is desirable to change fromone refrigerant to another in an existing system, itis usually possible to convert the system by chang-ing expansion valves and control settings provid-ing the existing piping sizes and component work-ing pressures are compatible. In some cases theexisting motor-compressor may be satisfactoryfor example, in converting from R-22 to R-502. Ifthe conversion will result in higher power require-ments as is the case in changing from R-12 toR502, then it may also be necessary to change themotor-compressor.

The refrigerant charge should be held to the mini-mum required for satisfactory operation, since anabnormally high charge will create potential prob-lems of liquid refrigerant control.

COMPRESSOR COOLING

Refrigerant cooled motor-compressors are depen-dent on return suction gas for motor cooling, and toa considerable extent, on both air and refrigerantcooled motor-compressors, the discharge gas tem-perature is directly related to the temperature of thereturn suction gas. Discharge temperatures above325 F. to 350 F. contribute to oil breakdown andvalve plate damage, and to avoid compressordamage, operating temperatures must be keptbelow this level. Peak temperatures occur at thedischarge valves, and normally the temperature ofthe discharge line will be from 50 F. to 100F.below the temperature at the valve plate. There-fore the maximum allowable discharge line tem-peratures from 225F. to 250F.

Suction gas entering the compressor should be nohigher than 65F. under low temperature loadconditions, or 90F. under high temperature loadconditions, and must never exceed 100F. Onsome abnormally critical low temperature applica-


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tions it may be desirable to insulate the suctionlines and return the suction gas to the compressorat lower than normal temperatures to prevent thedischarge temperatures from exceeding safe lim-its, but this is not normally necessary on commer-cial application where the saturated evaporatingtemperature is -40F. or above. The low dischargetemperature characteristics of R-502 have madepossible much more trouble free operation in singlestage low temperature applications.

Air cooled motor-compressors must have a suffi-cient quantity of air impinging directly on the com-pressor body for motor cooling. Refrigerant cooledmotor-compressors are cooled adequately by therefrigerant vapor at evaporating temperaturesabove 0F., but at evaporating temperatures below0F., additional motor cooling by means of air flowis necessary.

On air cooled condensing units, adequate coolingcan normally be accomplished by locating thecompressor in the discharge air blast from thecondenser fan. For proper cooling, the fan mustdischarge air directly against the compressor, sincethe compressor usually cannot be adequatelycooled by air pulled through a compartment inwhich the compressor is located. If the compres-sor is not located in the condenser discharge airstream, cooling must be provided by means of anauxiliary fan discharging air directly again the com-pressor body. On compressors with multiple headssuch as the Copeland 4R and 6R models, auxiliaryhorizontal airflow may not provide satisfactory cool-ing, and vertical cooling fans are required.

Water cooled compressors are provided with awater jacket are wrapped with a copper water coil,and water must be circulated through the compres-sor cooling circuit before entering the condenser.

Two-stage compressors are equipped with adesuperheating expansion valve for interstagecooling, and no auxiliary cooling is required.

If compressors or condensing units are located ina machine room, adequate ventilation air must beprovided to avoid an excessive temperature rise inthe room. To allow for peak summer temperaturesa 10F. temperature rise is recommended, al-though a 15F. rise in cooler ambients might beacceptable.

The most accurate calculation is to determine thetotal heat to be rejected by adding the compressorrefrigerating capacity at the design operating con-dition to the heat equivalent of the motor input. TheCFM can then be calculated by the formula

BTU/HRCMF = TD

For example, determine the machine room ventila-tion for an air cooled condensing unit operating at-25F. evaporator, 120F. condensing with a netrefrigeration capacity of 23,000 BTU/HR, 6,400watts input to the compressor motor, and a 1 H.P.condenser fan motor.

Compressor capacity 23,000BTU/HRHeat equivalent 6400 watts x 3.413 21,843BTU/HRHeat equivalent 1 H.P. fan motor 3,700 BTU/HRTotal Heat to be Rejected 48,543BTU/HR

With remote condensers, approximately 10% ofthe heat rejected is given off by the compressorcasting and the discharge tubing, and the ventila-tion can be calculated accordingly.

For convenience, table 20A gives a quick estimateof the ventilation air requirement if only the com-pressor capacity is known.

COMPRESSOR LUBRICATION

An adequate supply of oil must be maintained inthe crankcase at all times to insure continuouslubrication. The normal oil level should be main-tained at or slightly above the center of the sightglass while operating. An excessive amount of oilmust not be allowed in the system as it may resultin slugging and possible damage to the compres-sor valves.

Compressors leaving the Copeland factory arecharged with naphthenic 150 viscosity refrigera-tion oil, and the use of any other oil must bespecifically cleared with the Copeland ApplicationEngineering Department. The naphthenic base oilhas definite advantages over paraffinic base oilsbecause separation of refrigerant from paraffinicoils occurs at substantially higher temperatureswith the same oil-refrigerant concentration. When


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this separation or two phase condition exists the oilfloats on top of the refrigerant and the oil pump inletat the bottom of the sump is fed almost purerefrigerant at start up. The resulting improper lubri-cation can result in bearing failure. Because of thelower separating temperature of naphthenic oil, thepossibility of two-phasing is greatly reduced.

Copelametic compressors are shipped with a gen-erous supply of oil in the crankcase. However thesystem may require more or less oil depending onthe refrigerant charge and the system design. Onfield installed systems, after the system stabilizesat its normal operating conditions, it may be neces-sary to add or remove oil to maintain the desiredlevel.

OIL PRESSURE SAFETY CONTROL

A major percentage of all compressor failures arecaused by lack of proper lubrication. Improperlubrication or the loss of lubrication can be due toa shortage of oil in the system, logging of oil in theevaporator or suction line due to insufficient refrig-erant velocities, shortage of refrigerant, refrigerantmigration or floodback to the compressor crank-case, failure of the oil pump, or improper operationof the refrigerant control devices.

Regardless of the initial source of the difficulty, thegreat majority of compressor failures due to loss oflubrication could have been prevented. Althoughproper system design, good preventive mainte-nance, and operation within the systems designlimitations are the only cure for most of theseproblems, actual compressor damage usually canbe averted by the use of an oil pressure safetycontrol.

An oil pressure safety control with a time delay of120 seconds is a mandatory requirement of theCopeland warranty on all Copelametic compres-sors having an oil pump. The control operates onthe differential between oil pump pressure andcrankcase pressure, and the two minute delayserves to avoid shut down during short fluctuationsin oil pressure during start-up.

A trip of the oil pressure safety switch is a warningthat the system has been without proper lubricationfor a period of two minutes. Repeated trips of the

oil pressure safety control are a clear indicationthat something in the system design or operationrequires immediate remedial action. On a welldesigned system there should be no trips of the oilpressure safety control, and repeated trips shouldnever be accepted as a normal part of the systemoperation.

The oil pressure safety control will not protectagainst all lubrication problems. It cannot detectwhether the compressor is pumping oil or a combi-nation of refrigerant and oil. If bearing trouble isencountered on systems where the oil pressuresafety control has not tripped, even though inspec-tion proves it to be properly wired, with the properpressure setting, an in good operating condition,marginal lubrication is occurring which probably isdue to liquid refrigerant floodback.

OIL SEPARATORS

Proper refrigerant piping design and operation ofthe system within its design limits so that adequaterefrigerant velocities can be maintained are theonly cure for oil logging problems, but an oil sepa-rator may be a definite aid in maintaining lubrica-tion where oil return problems are particularlyacute.

For example, consider a compressor having an oilcharge of 150 ounces, with the normal oil circula-tion rate being 2 ounces per minute. This meansthat on a normal system with proper oil return atstabilized conditions, two ounces of oil leave thecompressor through the discharge line everyminute, and two ounces return through the suctionline. If a minimum of 30 ounces of oil in thecrankcase is necessary to properly lubricate thecompressor, and for some reason oil logged in thesystem and failed to return to the compressor, thecompressor would run out of oil in 60 minutes.Under the same conditions with an oil separatorhaving an efficiency of 80%, the compressor couldoperate 300 minute or 5 hours before running outof oil.

As a practical matter, there seldom are conditionsin a system when no oil will be returned to thecompressor, and even with low gas velocities,some fraction of the oil leaving the compressor willbe returned. If there are regular intervals of full load


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conditions or defrost periods when oil can bereturned normally, an oil separator can help tobridge long operating periods at light load condi-tions. Oil separators are mandatory on systemswith flooded evaporators controlled by a float valve,on all two stage and cascade ultra-low temperaturesystems, and on any system where oil return iscritical.

Oil separators should be considered as a systemaid but not a cure-all or a substitute for good systemdesign. They are never 100% efficient, and in factmay have efficiencies as low as 50% depending onsystem operating conditions. On systems wherepiping design encourages oil logging in the evapo-rator, an oil separator can compensate for systemoil return deficiencies only on a temporary basis,and may only serve to delay lubrication difficulties.

If a system is equipped with a suction accumulator,it is recommended that the oil return from theseparator be connected to the suction line justahead of the accumulator. This will provide maxi-mum protection against returning liquid refrigerantto the crankcase. If the system is not equipped witha suction accumulator, the oil return line on suctioncooled compressors may be connected to thesuction line if more convenient than the crankcase,but on air cooled compressors, oil return must bemade directly to the crankcase to avoid damage tothe compressor valves.

If the separator is exposed to outside ambienttemperatures, it must be insulated to prevent re-frigerant condensation during off periods, resultingin return of liquid to the compressor crankcase.Small low wattage strap-on heaters are availablefor oil separators, and if any problem from liquidcondensation in the separator is anticipated, acontinuously energized heater is highly recom-mended.

SUCTION LINE ACCUMULATORS

If liquid refrigerant is allowed to flood through arefrigeration or air conditioning system and returnto the compressor before being evaporated, it maycause damage to the compressor due to liquidslugging, loss of oil from the crankcase, or bearingwashout. To protect against this condition onsystems vulnerable to liquid damage a suctionaccumulator may be necessary.

The accumulators function is to intercept liquidrefrigerant before it can reach the compressorvalves or crankcase. It should be located in thesuction line near the compressor, and if a reversingvalve is used in the system, the accumulator mustbe located between the reversing valve and thecompressor. Provisions for positive oil return to thecrankcase must be provided, but a direct gravityflow which will allow liquid refrigerant to drain to thecrankcase during shut-down periods must beavoided. The liquid refrigerant must be meteredback to the compressor during operation at acontrolled rate to avoid damage to the compressor.

Some systems, because of their design, will peri-odically flood the compressor with liquid refriger-ant. Typically, this can occur on heat pumps at thetime the cycle is switched from cooling to heating,or from heating to cooling. The coil which has beenserving as the condenser is partially filled withliquid refrigerant, and when suddenly exposed tosuction pressure, the liquid is dumped into thesuction line. On heat pumps equipped with expan-sion valves, there may be further flooding due tothe inability of the expansion valve to effectivelycontrol refrigerant feed for a short period after thecycle change until the system operation is againstabilized.

A similar situation can occur during defrost cycles.With hot gas defrost, when the defrost cycle isinitiated, the sudden introduction of high pressuregas into the evaporator may force the liquid refrig-erant in the evaporator into the suction line. If thedefrost cycle is such that the evaporator can fill withcondensed liquid during defrost, or on systemsutilizing electric defrost without a pumpdown cycle,an equally dangerous situation may exist at thetermination of the defrost cycle.

On systems with a large refrigerant charge, or onany system where liquid floodback is likely tooccur, a suction line accumulator is strongly rec-ommended. On heat pumps, truck applications,and on any system where liquid slugging can occurduring operation, a suction line accumulator ismandatory for compressor protection unless oth-erwise approved by the Copeland Application En-gineering Department. The actual refrigerant hold-ing capacity needed for a given accumulator is


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governed by the requirements of the particularapplication, and the accumulator should be se-lected to hold the maximum liquid floodback antici-pated.

PUMPDOWN SYSTEM CONTROL

Refrigerant vapor will always migrate to the coldestpart of the system, and if the compressor crank-case can become colder than other parts of thesystem, refrigerant in the condenser, receiver, andevaporator will vaporize, travel through the sys-tem, and condense in the compressor crankcase.

Because of the difference in vapor pressures of oiland refrigerant, refrigerant vapor is attracted torefrigeration oil, and even though no pressure ortemperature difference exists to cause a flow,refrigerant vapor will migrate through the systemand condense in the oil until the oil is saturated.During off cycles extending several hours or more,it is possible for liquid refrigerant to almost com-pletely fill the compressor crankcase due to the oilattraction. For example, in a system using R-12refrigerant which is allowed to equalize at an ambi-ent temperature of 70 F., the oil-refrigerant mix-ture in the crankcase will end up about 70% refrig-erant before equilibrium is reached.

The most positive and dependable means of keep-ing refrigerant out of the compressor crankcase isthe use of a pumpdown cycle. By closing a liquidline solenoid valve, the refrigerant can be pumpedinto the condenser and receiver, and the compres-sor operation controlled by means of a low pres-sure control. The refrigerant can thus be isolatedduring periods when the compressor is not inoperation, and migration of refrigerant to the com-pressor crankcase is prevented.

Pumpdown control can be used on all thermostaticexpansion valve systems with the addition of aliquid line solenoid valve, provided adequate re-ceiver capacity is available. Slight refrigerantleakage may occur through the solenoid valve,causing the suction pressure to rise gradually, anda recycling type control is recommended to repeatthe pumpdown cycle as required. The occasionalshort cycle usually is not objectionable.

A pumpdown cycle is highly recommended when-ever it can be used. If a non-recycling pumpdowncircuit is required, then consideration should begiven to the use of a crankcase heater in additionto the pumpdown for more dependable compres-sor protection.

CRANKCASE HEATERS

On some systems operating requirements, noiseconsiderations, or customer preference may makethe use of a pumpdown system undesirable, andcrankcase heaters are frequently used to controlmigration.

By warming the oil, the absorption of refrigerant bythe oil is minimized, and under mild weather con-ditions, any liquid refrigerant in the crankcase canbe vaporized and forced out of the compressor.For effective protection, heaters must be ener-gized several hours before starting the compres-sor. It is recommended that they be energizedcontinuously, independent of compressor opera-tion. Improperly sized heaters can overheat the oil,and heaters used on Copeland compressors mustbe specifically approved by the Copeland Applica-tion Engineering Department.

It would be a mistake to assume that crankcaseheaters are a dependable cure for all migrationproblems. As the ambient conditions contributingto migration worsen, the ability of the crankcaseheater to keep refrigerant out of the crankcasedecreases. If the suction line slopes toward thecompressor, and the temperature to which thesuction line is exposed is sufficiently lower than thetemperature of the oil, refrigerant may condense inthe suction line and flow back to the compressor bygravity at a rate sufficient to offset the heat intro-duced by the heater. Heaters will not protectagainst liquid slugs or excessive liquid flooding.However, where operating conditions are not toosevere, crankcase heaters can provide satisfac-tory protection against migration.

Where a pumpdown cycle is not used, crankcaseheaters are mandatory on heat pumps, and onother air conditioning applications if the refrigerantcharge exceeds the established Copeland limits,unless tests prove the compressor is adequatelyprotected by other means.


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To prevent possible damage in shipment, crank-case heaters are not installed on compressors atthe factory.

CRANKCASE PRESSURE REGULATINGVALVES

In order to limit the power requirement of thecompressor to the allowable operating limit, acrankcase pressure regulating valve may be nec-essary. This most frequently occurs on low tem-perature compressors where the power require-ment during pulldown periods or after defrost maybe greatly in excess of the compressor motorscapabilities. Copeland compressors should not beoperated at suction pressure in excess of thepublished limits on compressor specification sheetswithout approval of the Copeland Application Engi-neering Department.

Since any pressure drop in the compressor suctionline lowers the system capacity, the CPR valveshould be sized for a minimum pressure drop. Inorder to restrict pull down capacity as little aspossible, the valve setting should be as high as themotor power requirement will allow.

Thermal expansion valves of the pressure limitingtype are not recommended when a CPR valve isused, particularly if the pressure settings are fairlyclose, because of the possibility of the action of thetwo valves coming in conflict in their response tosystem pressures.

LOW AMBIENT HEAD PRESSURE CONTROL

Within the operating limitations of the system, it isdesirable to take advantage of lower condensingtemperatures whenever possible for increasedcapacity, lower discharge temperatures, and lowerpower requirements. However, too low a dis-charge pressure can produce serious malfunc-tions. Since the capacity of capillary tubes andexpansion valves is proportional to the differentialpressure across the capillary tube or valve, areduction in discharge pressure will reduce itscapacity and produce a drop in evaporating pres-sure.

Low discharge pressures can result in starving theevaporator coil with resulting oil logging, shortcycling on low pressure controls, reduction of sys-tem capacity, or erratic expansion valve operation.

Systems with water cooled condensers and cool-ing towers require water regulating valves, or someother means of controlling the temperature or thequantity of water passing through the condenser.

If air cooled air conditioning systems are requiredto operate in ambient temperatures below60 F., a suitable means of controlling head pres-sures must be provided. Refrigeration systems arealso vulnerable to damage from low head pressureconditions, and adequate head pressure controlsshould be provided for operation in ambient tem-peratures below 50F.

Several proprietary control systems are availablefor low ambient operation, most of which maintainhead pressure above a preset minimum by par-tially flooding the condenser and thus reducing theeffective surface area. Methods of this type cancontrol pressures effectively, but do require a con-siderable increase in refrigerant charge and ad-equate receiver capacity must be provided.

Air volume dampers on the condenser operatedfrom refrigerant discharge pressure provide asimple, economical, and effective means of controlwhich is widely used.

Adequate protection at lowest cost can often beprovided by a reverse acting high pressurecontrol which senses discharge pressure, and actsto disconnect the condenser fan circuit when thehead pressure falls below the controls minimumsetting. The proper adjustment of the off-on differ-ential is particularly important to avoid excessivefan motor cycling, and the resulting fluctuations indischarge pressure may contribute to uneven ex-pansion valve feeding. In cold ambient tempera-tures the condenser must be shielded from thewind.

LIQUID LINE FILTER-DRIER

A liquid line filter-drier must be used on all fieldinstalled systems, and on all systems opened in thefield for service. Filter-driers are highly recom-


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mended for all systems, but are not mandatory onfactory assembled and charged units where care-ful dehydration and evacuation is possible duringmanufacture. Precharged systems with quick con-nect fittings having a rupture disc are considered tobe the equivalent of factory charged systems.

Moisture can be a factor in many forms of systemdamage, and the reduction of moisture to an ac-ceptable level can greatly extend compressor lifeand slow down harmful reactions. The desiccantused must be capable of removing moisture to alow end point and further should be of a type, whichcan remove a reasonable quantity of acid. It ismost important that the filter-drier be equipped withan excellent filter to prevent circulation of carbonand foreign particles.

SIGHT GLASS AND MOISTURE INDICATOR

A combination sight glass and moisture indicator isessential for easy field maintenance on any sys-tem, and is required on any field installed systemunless some other means of checking the refriger-ant charge is provided.

A sight glass is a convenient means of determiningthe refrigerant charge, showing bubbles when thereis insufficient charge, and a solid clear glass whenthere is sufficient charge. However, the operatorshould bear in mind that under some circum-stances even when the receiver outlet has a liquidseal, bubbles or flash gas may show in the sightglass. This may be due to a restriction or excessivepressure drop in the receiver outlet valve, a par-tially plugged drier or strainer, or other restriction inthe liquid line ahead of the sight glass. If theexpansion valve feed is erratic or surging, theincreased flow when the expansion valve is wideopen can create sufficient pressure drop to causeflashing at the receiver outlet.

Another source of flashing in the sight glass may berapid fluctuations in compressor discharge pres-sure. For example, in a temperature controlledroom, the sudden opening of shutters or the cyclingof a fan can easily cause a reduction in the con-densing temperature of 10F. to 15F. Any liquid inthe receiver may then be at a temperature equiva-lent to the lower condensing pressure, and flashingwill continue until the system has stabilized at thenew condensing temperature.

While the sight glass can be a valuable aid inservicing a refrigeration or air conditioning system,a more positive liquid indicator is desirable, and thesystem performance must be carefully analyzedbefore placing full reliance on the sight glass as apositive indicator of the system charge.

LIQUID LINE SOLENOID VALVE

A liquid line solenoid valve is recommended on allfield installed systems with large refrigerant charges,particularly when the system has a charge inexcess of three pounds of refrigerant per motorHP. The solenoid valve will prevent continued feedto the evaporator through the expansion valve orcapillary tube when the compressor is not operat-ing, and will control migration of liquid refrigerantfrom the receiver and condenser to the evaporatorand compressor crankcase.

If a pumpdown cycle is not used, the liquid linesolenoid valve should be wired to the compressormotor terminals so that the valve will be de-ener-gized when the motor is not operating.

HEAT EXCHANGER

A liquid to suction heat exchanger is highly recom-mended on all refrigeration systems, and is re-quired on package water chillers and water towater heat pumps because of the low operatingsuperheat. On medium and low temperature appli-cations, a heat exchanger increases system ca-pacity, helps to eliminate flashing of liquid refriger-ant ahead of the expansion valve, and aids both inpreventing condensation on suction lines and inevaporating any liquid flooding through the evapo-rator.

On small systems, soldering the liquid and suctionlines together for several feet makes an effectiveheat exchanger.

THERMOSTATIC EXPANSION VALVES

Thermostatic expansion valves must be selectedand applied in accordance with the manufacturersinstructions. Either internally equalized or exter-nally equalized valves will feed properly if appliedcorrectly. If the thermal expansion valve is of the


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externally equalized type, the external equalizerline must be connected, preferably at a point be-yond the expansion valve thermal bulb. Do not capor plug the external equalizer connection as thevalve will not operate without this connection.

Valve superheat should be preset by the valvemanufacturer, and field adjustment should be dis-couraged. Valves in need of adjustment should beset to provide 5F. to 10F. superheat at thethermal bulb location. Too high a superheat settingwill result in starving the evaporator, and can causepoor oil return. Too low a superheat setting willpermit liquid floodback to the compressor.

A minimum of 15F. superheat at the compressormust be maintained at all times to insure the returnof dry gas to the compressor suction chamber, anda minimum of 20F. superheat is recommended.Note that this is not superheat at the expansionvalve, but should be calculated from pressuremeasured at the suction service valve and thetemperature measured 18 from the compressoron the bottom of the horizontal run of suction linetubing. Lower superheat can result in liquid refrig-erant flooding back to the compressor during varia-tions in the evaporator feed with possible compres-sor damage as a result. Excessively wet refriger-ant vapor continually returning to the compressorcan reduce the lubricating qualities of the oil andgreatly increase compressor wear, as well as re-sulting in a loss of capacity.

It is important that users realize that flash gas in theliquid line can seriously affect expansion valvecontrol. So long as a head of pure liquid refrigerantis maintained at the expansion valve, its perfor-mance is relatively stable. But if flash gas is mixedwith liquid refrigerant fed to the valve, a largerorifice opening is required to feed to the sameweight of liquid refrigerant. The only way the orificeopening can be increased is by an increase insuperheat, and as the percentage of flash gasincreases, the superheat increases, the valve openswide, and the evaporator is progressively morestarved.

If the valve has been operating with a large per-centage of flash gas entering the expansion valve,and a head of pure liquid refrigerant is suddenlyrestored, the orifice opening will be larger than

required for the load, and liquid will flood throughthe system to the compressor until the valve againregains control. Conventional expansion valveswith the thermal bulb strapped to the suction linemay be somewhat sluggish in response, and it maybe several minutes before control can be restoredto normal.

Typically, changes in the quality of liquid refrigerantfeeding the expansion valve can occur quickly andfrequently because of the action of head pressurecontrol devices, sudden changes in the refrigera-tion load, hunting of the expansion valve, action ofan unloading valve, or rapid changes in condens-ing pressure.

On systems with short suction lines and low super-heat requirements, quick response thermal bulbsor wells in the suction line may be essential to avoidperiodic floodback to the compressor.

Temperatures and pressure alone may not give atrue picture of the actual liquid refrigerant control ina system. Excessive oil circulation has the effectof increasing the evaporating temperature of therefrigerant. The response of the expansion valveis based on the saturation pressure and tempera-ture of pure refrigerant. In an operating system, thechanged pressure-temperature characteristics ofthe oil rich refrigerant will give the expansion valvea false reading of the actual superheat, and canresult in a somewhat lower actual superheat thanapparently exists, causing excessive liquid refrig-erant floodback to the compressor. The only realcure for this condition is to reduce oil circulation toa minimum. Normally excessive oil in the evapo-rator can only result from an excessive system oilcharge or other factors which could cause exces-sive oil circulation, or from low velocities in theevaporator which result in oil logging. In lowtemperature applications where proper oil circula-tion cannot be maintained, an oil separator may berequired.

Vapor charged valves are satisfactory for air con-ditioning usage, and are desirable in many casesbecause of their inherent pressure limiting charac-teristics. For all refrigeration applications, liquidcharged valves should be used to prevent conden-sation of the charge in the head of the valve and theresulting loss of control in the event the headbecomes colder than the thermal bulb.


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A pressure limiting type valve may be helpful inlimiting the compressor load, and also preventsexcessive liquid refrigerant floodback on start-up.On systems using hot gas defrost, the defrost loadis normally greater than the refrigeration load, andsome other means of limiting the compressorpower input must be used if required.

The thermostatic expansion valve must be sizedproperly for the load. Although a given valvenormally has a wide operating capacity range,excessively undersized or oversized valves cancause system malfunctions. Undersized valvesmay starve the evaporator, and the resulting ex-cessive superheat may adversely affect the sys-tem performance. Oversized valves can causehunting, alternately starving and flooding the evapo-rator, resulting in extreme fluctuations in suctionpressure.

The thermal bulb should normally be located on ahorizontal section of the suction line, close to theevaporator outlet, on the evaporator side of anysuction line trap or heat exchanger. Do not underany circumstances locate the thermal bulb in alocation where the suction line is trapped since thiscan result in erratic feeding. Satisfactory perfor-mance can usually be obtained with the bulbstrapped to the suction line at the 3 oclock position.Mounting on the top of the suction line will de-crease sensitivity, and may allow possible liquidflooding. Mounting on the bottom of the suction linecan cause erratic feeding due to the rapid tempera-ture changes that can result from even smallamounts of liquid refrigerant reaching the thermalbulb location. Particular attention should be givento the location of the thermal bulb on multipleevaporator systems to insure that the refrigerantreturning from one evaporator does not affect thecontrol of another evaporator.

EVAPORATORS

Evaporators must be properly selected for therefrigeration load. Too large an evaporator mightresult in low velocities and possible oil logging. Toosmall an evaporator will have excessive tempera-ture differentials between the evaporating refriger-ant and the medium to be cooled. The allowableTD between the entering air and the evaporatingrefrigerant may also be dictated by the humiditycontrol required.

Internal volume of the evaporator tubing should beat a minimum to keep the system refrigerant chargeas low as possible, so the smallest diameter tubingthat will give acceptable performance should beused. Since pressure drop at low evaporatingtemperatures is critical so far as capacity is con-cerned, multiple refrigerant circuits with fairly shortruns are preferred. At the same time, it is essentialthat velocities of refrigerant in the evaporator behigh enough to avoid oil trapping.

Vertical headers should have a bottom outlet toallow gravity oil drainage.

SUCTION LINE FILTERS

A heavy duty suction line filter is recommended forevery field installation. The filter will effectivelyremove contaminants from the system at the timeof installation, and serves to keep the compressorfree of impurities during operation. In the event ofa motor burn, the filter will prevent contaminationfrom spreading into the system through the suctionline.

The suction line filter should be selected for areasonable pressure drop, and should be equippedwith a pressure fitting just ahead of the filter,preferably in the shell, to facilitate checking pres-sure drop across the filter during operation.

HIGH AND LOW PRESSURE CONTROLS

Both high and low pressure controls are recom-mended for good system design on all air cooledsystems 1 HP and larger, and are essential on allfield installed air cooled systems and on all watercooled systems.

When used for low temperature unit operationcontrol, the low pressure control must not be setbelow the minimum operating limits of the com-pressor or the system. One of the most frequentcauses of motor overheating and inadequate lubri-cation is operation of the compressor at exces-sively low suction pressures. Copeland specifica-tion sheets list the approved compressor operatingrange, and recommended minimum low pressurecontrol settings for various operating ranges areshown in Table 21.


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High pressure controls may be either manual orautomatic reset as desired by the customer. If ofthe manual reset type, provision must be made toprevent liquid refrigerant flooding through the sys-tem to the compressor in the event of a trip of thehigh pressure control.

Internal automatic reset pressure relief valves(Copelimit) are provided in most Copelaweld com-pressors 1 HP and larger. On factory assembledpackage systems, the internal Copelimit valve maysatisfy U.L. and code requirements without the useof an external high pressure control. A similar highside to low side automatic reset pressure reliefvalve is installed in all Copelametic compressorswith displacements of 3,000 CFH or greater.

On factory assembled and charged package sys-tems, such as room air conditioners, where loss ofcharge protection is not considered critical, orwhere the motor protection device can provide lossof charge protection, low pressure controls maynot be essential although recommended.

INTERCONNECTED SYSTEMS

When the crankcases of two or more compressorsare interconnected for parallel operation on a singlerefrigeration system, serious problems of oil returnand vibration may be encountered unless the sys-tem is properly designed. The tandem compressorconsisting of two individual compressors with aninterconnecting housing replacing the individualstator covers provides a simple, trouble free solu-tion to this problem.

Because of the potential operating problems, inter-connection of individual compressors is not ap-proved with the exception of factory designed,tested, and assembled units specifically approvedby the Copeland Application Engineering Depart-ment.

ELECTRICAL GROUP FUSING

Individual circuit breakers or fuses should be pro-vided for each compressor motor. Group fusing,where two or more compressors are installed onone fused disconnect, is not recommended sincean electrical failure in one compressor would nottrip the fuse, and extensive electrical damagecould result.


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SECTION 18REFRIGERATION PIPING

Probably the first skill that any refrigeration appren-tice mechanic learns is to make a soldered joint,and running piping is so common a task that oftenits critical importance in the proper performance ofa system is overlooked. It would seem elementaryin any piping system that what goes in one end ofa pipe must come out the other, but on a systemwith improper piping, it is not uncommon for aserviceman to add gallons of oil to a system, and itmay seemingly disappear without a trace. It is ofcourse lying on the bottom of the tubing in thesystem, usually in the evaporator or suction line.When the piping or operating condition is cor-rected, the oil will return and those same gallons ofoil must be removed.

Refrigeration piping involves extremely complexrelationships in the flow of refrigerant and oil. Fluidflow is the name given in mechanical engineeringto the study of the flow of any fluid, whether it mightbe a gas or a liquid, and the inter-relationship ofvelocity, pressure, friction, density, viscosity, andthe work required to cause the flow. These rela-tionships evolve into long mathematical equationswhich form the basis for the fan laws which governfan performance, and the pressure drop tables forflow through piping. But 99% of the theories in fluidflow textbooks deal with the flow of one homog-enous fluid, and there is seldom even a mention ofa combination flow of liquid, gas, and oil such asoccurs in any refrigeration system. Because of itschanging nature, such flow is just too complex to begoverned by a simple mathematical equation, andpractically the entire working knowledge of refrig-eration piping is based on practical experience andtest data. As a result, the general type of gas andliquid flow that must be maintained to avoid prob-lems is known, but seldom is there one exactanswer to any problem.

BASIC PRINCIPLES OF REFRIGERATION PIP-ING DESIGN

The design of refrigeration piping systems is acontinuous series of compromises. It isdesirable to have maximum capacity, minimumcost, proper oil return, minimum power consump-tion, minimum refrigerant charge, low noise level,

proper liquid refrigerant control, and perfect flex-ibility of system operations from 0 to 100% ofsystem capacity without lubrication problems.Obviously all of these goals cannot be satisfied,since some are in direct conflict. In order to makean intelligent decision as to just what type ofcompromise is desirable, it is essential that thepiping designer clearly understand the basic ef-fects on system performance of the piping designin the different parts of the system.

In general, pressure drop in refrigerant lines tendsto decrease capacity and increase power require-ments, and excessive pressure drops should beavoided. The magnitude of the pressure dropallowable varies depending on the particular seg-ment of piping involved, and each part of thesystem must be considered separately. There areprobably more tables and charts available cover-ing line pressure drop and refrigerant line capaci-ties at a given pressure drop than on any othersingle subject in the field of refrigeration.

It is most important, however, that the piping de-signer realize that pressure drop is not the onlycriteria that must be considered in sizing refriger-ant lines, and that often refrigerant velocities ratherthan pressure drop must be the determining factorin system design. In addition to the critical natureof oil return, there is no better invitation to systemdifficulties than an excessive refrigerant charge. Areasonable pressure drop is far more preferablethan over-sized lines which can contain refrigerantfar in excess of the systems needs. An excessiverefrigerant charge can result in serious problems ofliquid refrigerant control, and the flywheel effect oflarge quantities of liquid refrigerant in the lowpressure side of the system can result in erraticoperation of the refrigerant control devices.

The size of the service valve supplied on a com-pressor, or the size of the connection on a con-denser, evaporator, accumulator, or other acces-sory does not determine the size of line to be used.Manufacturers select a valve size or connectionfitting on the basis of its application to an averagesystem, and such factors as the type of application,length of connecting lines, type of system control,


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variation in load, and other factors can be majorfactors in determining the proper line size. It isquite possible the required line size may be eithersmaller or larger than the fittings on various systemcomponents. In such cases, reducing fittings mustbe used.

Since oil must pass through the compressor cylin-ders to provide lubrication, a small amount of oil isalways circulating with the refrigerant. Refrigera-tion oils are soluble in liquid refrigerant, and atnormal room temperatures they will mix com-pletely. Oil and refrigerant vapor, however, do notmix readily, and the oil can be properly circulatedthrough the system only if the mass velocity of therefrigerant vapor is great enough to sweep the oilalong. To assure proper oil circulation, adequaterefrigerant velocities must be maintained not onlyin the suction and discharge lines, but in theevaporator circuits as well.

Several factors combine to make oil return mostcritical at low evaporating temperatures. As thesuction pressure decreases and the refrigerantvapor becomes less dense, the more difficult itbecomes to sweep the oil along. At the same timeas the suction pressure falls, the compression ratioincreases, and as a result compressor capacity isreduced, and the weight of refrigerant circulateddecreases. Refrigeration oil alone becomes theconsistency of molasses at temperatures below0F., but so long as it is mixed with sufficient liquidrefrigerant, it flows freely. As the percentage of oilin the mixture increases, the viscosity increases.

At low temperature conditions all of these factorsstart to converge, and can create a critical condi-tion. The density of the gas decreases, the massvelocity flow decreases, and as a result more oilstarts accumulating in the evaporator. As the oiland refrigerant mixture becomes more viscous, atsome point oil may start logging in the evaporatorrather than returning to the compressor, resultingin wide variations in the compressor crankcase oillevel in poorly designed systems.

Oil logging can be minimized with adequate veloci-ties and properly designed evaporators even atextremely low evaporating temperatures, but nor-mally oil separators are necessary for operation atevaporating temperatures below -50F. in order to

minimize the amount of the oil in circulation.

COPPER TUBING FOR REFRIGERANT PIPING

For installations using R-12, R-22, and R-502,copper tubing is almost universally used for refrig-erant piping. Commercial copper tubing dimen-sions have been standardized and classified asfollows:

Type K Heavy WallType L Medium WallType M Light Wall

Only types K or L should be used for refrigerantpiping, since type M does not have sufficient strengthfor high pressure applications. Type L tubing ismost commonly used, and all tables and data inthis manual are based on type L dimensions.

It is highly recommended that only refrigerationgrade copper tubing be used for refrigeration appli-cations, since it is available cleaned, dehydrated,and capped to avoid contamination prior to instal-lation. Copper tubing commonly used for plumbingusually has oils and grease or other contaminantson the interior wall, and these can cause seriousoperating problems if not removed prior to installa-tion.

Table 22 lists the dimensions and properties ofstandard commercial copper tubing in the sizescommonly used in refrigeration systems, and Table23 lists the weight of various refrigerants per 100feet of piping in liquid, suction and discharge lines.

FITTINGS FOR COPPER TUBING

For brazed or soldered joints, the required elbows,tees, couplings, reducers, or other miscellaneousfittings may be either forged brass or wroughtcopper. Cast fittings are not satisfactory since theymay be porous and often lack sufficient strength.

EQUIVALENT LENGTH OF PIPE

Each valve, fitting, and bend in a refrigerant linecontributes to the friction pressure drop because ofits interruption or restriction of smooth flow. Be-cause of the detail and complexity of computing thepressure drop of each individual fitting, normal


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practice is to establish an equivalent length ofstraight tubing for each fitting. This allows theconsideration of the entire length of line, includingfittings, as an equivalent length of straight pipe.Pressure drop and line sizing tables and charts arenormally set up on the basis of a pressure drop per100 feet of straight pipe, so the use of equivalentlengths allows the data to be used directly.

The equivalent length of copper tubing for com-monly used valves and fittings is shown in Table24.

For accurate calculations of pressure drop, theequivalent length for each fitting should be calcu-lated. As a practical matter, an experienced pipingdesigner may be capable of making an accurateoverall percentage allowance unless the piping isextremely complicated. For long runs of piping of100 feet or greater, an allowance of 20% to 30% ofthe actual lineal length may be adequate, while forshort runs of piping, an allowance as high as 50%to 75% or more of the lineal length may be neces-sary. Judgment and experience are necessary inmaking a good estimate, and estimates should bechecked frequently with actual calculations to in-sure reasonable accuracy.

For items such as solenoid valves and pressureregulating valves, where the pressure drop throughthe valve is relatively large, data is normally avail-able from the manufacturers catalog so that items

of this nature can be considered independently oflineal length calculations.

PRESSURE DROP TABLES

Figure 76, 77, and 78 are combined pressure dropcharts for refrigerants R-12,R-22, and R-502. Pressure drops in the dischargeline, suction line, and liquid line can be determinedfrom these charts for condensing temperaturesranging from 80F. to 120F.

To use the chart, start in the upper right handcorner with the design capacity. Drop verticallydownward on the line representing the desiredcapacity to the intersection with the diagonal linerepresenting the operating condition desired. Thenmove horizontally to the left. A vertical line droppedfrom the intersection point with each size of coppertubing to the design condensing temperature lineallows the pressure drop in psi per 100 feet oftubing to be read directly from the chart. Thediagonal pressure drop lines at the bottom of thechart represent the change in pressure drop due toa change in condensing temperature.

For example, in Figure 78 for R-502, the dotted linerepresents a pressure drop determination for asuction line in a system having a design capacity of5.5 tons or 66,000 BTU/hr operating with an evapo-rating temperature of -40F. The 2 5/8 O.D. suctionline illustrated has a pressure drop of 0.22 psi per100 feet at 85F. condensing temperature, but thesame line with the same capacity would have apressure drop of 0.26 psi per 100 feet at 100F.condensing, and 0.32 psi per 100 feet at120F.condensing.

In the same manner, the corresponding pressuredrop for any line size and any set of operatingconditions within the range of the chart can bedetermined.

SIZING HOT GAS DISCHARGE LINES

Pressure drop in discharge lines is probably lesscritical than in any other part of the system. Fre-quently the effect on capacity of discharge linepressure drop is over-estimated since it is as-sumed the compressor discharge pressure andthe condensing pressure are the same. In fact,


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there are two different pressures, the compressordischarge pressure being greater than the con-densing pressure by the amount of the dischargeline pressure drop. An increase in pressure drop inthe discharge line might increase the compressordischarge pressure materially, but have little effecton the condensing pressure. Although there is aslight increase in the heat of compression for anincrease in head pressure, the volume of gaspumped is decreased slightly due to a decrease involumetric efficiency of the compressor. Thereforethe total heat to be dissipated through the con-denser may be relatively unchanged, and the con-densing temperature and pressure may be quitestable, even though the discharge line pressuredrop and therefore the compressor discharge pres-sure might vary considerably.

The performance of a typical Copelametic com-pressor, operating at air conditioning conditionswith R-22 and an air cooled condenser indicatesthat for each 5 psi pressure drop in the dischargeline, the compressor capacity is reduced less than of 1%, while the power required is increasedabout 1%. On a typical low temperatureCopelametic compressor operating with R-502and an air cooled condenser, approximately 1% ofcompressor capacity will be lost for each 5 psipressure drop, but there will be little or no changein power consumption.

As a general guide, for discharge line pressuredrops up to 5 psi, the effect on system performancewould be so small as to be difficult to measure.Pressure drops up to 10 psi would not be greatlydetrimental to system performance provided thecondenser is sized to maintain reasonable con-densing pressures.

Actually a reasonable pressure drop in the dis-charge line is often desirable to dampen compres-sor pulsation, and thereby reduce noise and vibra-tion. Some discharge line mufflers actually derivemuch of their efficiency from pressure drop throughthe muffler.

Discharge lines on factory built condensing unitsusually are not a field problem, but on systemsinstalled in the field with remote condensers, linesizes must be selected to provide proper systemperformance.

Because of the high temperatures existing in thedischarge line, oil flows freely, and oil circulationthrough both horizontal and vertical lines can bemaintained satisfactorily with reasonably low ve-locities. Since oil traveling up a riser usually creepsup the inner surface of the pipe, oil travel in verticalrisers is dependent on the velocity of the gas at thetubing wall. The larger the pipe diameter, thegreater will be the required velocity at the center ofthe pipe to maintain a given velocity at the wallsurface. Figures 79 and 80 list the maximumrecommended discharge line riser sizes for properoil return for varying capacities. The variation atdifferent condensing temperatures is not great, sothe line sizes shown are acceptable on both watercooled and air cooled applications.

If horizontal lines are run with a pitch in the directionof flow of at least in 10 feet, there is normally littleproblem with oil circulation at lower velocities inhorizontal lines. However, because of the rela-tively low velocities required in vertical dischargelines, it is recommended wherever possible thatboth horizontal and vertical discharge lines besized on the same basis.

To illustrate the use of the chart, assume asystem operating with R-22 at 40F. evaporat-ing temperature has a capacity of 100,000 BTU/hr. The intersection of the capacity and evapo-rating temperature lines at point X on Figure 80indicate the design condition. Since this isbelow the 2 1/8 O.D. line, the maximum size thatcan be used to insure oil return up a verticalriser is 1 5/8 O.D.

Oil circulation in discharge lines is normally aproblem only on systems where large variations insystem capacity are encountered. For example,an air conditioning system may have steps ofcapacity control allowing it to operate during peri-ods of light load at capacities possibly as low as25% or 33% of the design capacity. The samesituation may exist on commercial refrigerationsystems where compressors connected in parallelare cycled for capacity control. In such cases,vertical discharge lines must be sized to maintainvelocities above the minimum necessary to prop-erly circulate oil at the minimum load condition.


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For example, consider an air conditioning systemusing R-12 having a maximum design capacity of300,000 BTU/hr with steps of capacity reduction upto 66%. Although the 300,000 BTU/hr conditioncould return oil up a 3 1/8 O.D. riser, at light loadconditions the system would have only 100,000BTU/hr capacity, so a 2 1/8 O.D. riser must be used.In checking the pressure drop chart, Figure 76, atmaximum load conditions, a 2 1/8 O.D. pipe willhave a pressure drop of approximately 3 psi per100 feet at a condensing temperature of 120F.

One other limiting factor in discharge line sizing isexcessive velocity which can cause noise prob-lems. Velocities of 3,000 FPM or more may resultin high noise levels, and it is recommended thatmaximum velocities be kept well below this level.Figures 81 and 82 give equivalent discharge linegas velocities for varying capacities and line sizesover the normal refrigeration and air conditioningrange.

Because of the flexibility in line sizing that theallowable pressure drop makes possible, dischargelines can almost always be sized satisfactorilywithout the necessity of double risers. If modifica-tions are made to an existing system which resultin the existing discharge line being oversized atlight load conditions, the addition of an oil separatorto minimize oil circulation will normally solve theproblem.

To summarize, in sizing discharge lines, it is rec-ommended that a tentative selection of line size bemade on the basis of a total pressure drop ofapproximately 5 psi plus or minus 50%, the actualdesign pressure drop to a considerable degreebeing a matter of the designers judgment. CheckFigure 79 or 80 to be sure that velocities at mini-mum load conditions are adequate to carry oil upvertical risers, and adjust vertical riser size if nec-essary. Check Figure 81 or 82 to be sure velocitiesat maximum load are not excessive.

Recommended discharge line sizes for varyingcapacities and equivalent lengths of line are givenin Table 28, page .

SIZING LIQUID LINES

Since liquid refrigerant and oil mix completely,velocity is not essential for oil circulation in the

liquid line. The primary concern in liquid line sizingis to insure a solid liquid head of refrigerant at theexpansion valve. If the pressure of the liquidrefrigerant falls below its saturation temperature, aportion of the liquid will flash into vapor to cool theliquid refrigerant to the new saturation tempera-ture. This can occur in a liquid line if the pressuredrops sufficiently due to friction or vertical lift.

Flash gas in the liquid line has a detrimental effecton system performance in several ways. It in-creases the pressure drop due to friction, reducesthe capacity of the expansion device, may erodethe expansion valve pin and seat, can cause ex-cessive noise, and may cause erratic feeding of theliquid refrigerant to the evaporator.

For proper system performance, it is essential thatliquid refrigerant reaching the expansion device besubcooled slightly below its saturation tempera-ture. On most systems the liquid refrigerant issufficiently subcooled as it leaves the condenser toprovide for normal system pressure drops. Theamount of subcooling necessary, however, is de-pendent on the individual system design.

On air cooled and most water cooled applications,the temperature of the liquid refrigerant is normallyhigher than the surrounding ambient temperature,so no heat is transferred into the liquid, and the onlyconcern is the pressure drop in the liquid line.Besides the friction loss caused by flow through thepiping, a pressure drop equivalent to the liquidhead is involved in forcing liquid to flow up a verticalriser. A head of two feet of liquid refrigerant isapproximately equivalent to 1 psi. For example, ifa condenser or receiver in the basement of abuilding is to supply liquid refrigerant to an evapo-rator three floors above, or approximately 30 feet,then a pressure drop of approximately 15 psi mustbe provided for in system design for the liquid headalone.

On evaporative or water cooled condensers wherethe condensing temperature is below the ambientair temperature, or on any application where liquidlines must pass through hot areas such as boiler orfurnace rooms, an additional complication mayarise because of heat transfer into the liquid. Anysubcooling in the condenser may be lost in thereceiver or liquid line due to temperature rise alone


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unless the system is properly designed. On evapo-rative condensers where a receiver and subcoolingcoil are used, it is recommended that the refriger-ant flow be piped from the condenser to the re-ceiver and then to the subcooling coil. In criticalapplications it may be necessary to insulate boththe receiver and the liquid line.

On the typical air cooled condensing unit with aconventional receiver, it is probable that very littlesubcooling of liquid is possible unless the receiveris almost completely filled with liquid. Vapor in thereceiver in contact with the subcooled liquid willcondense, and this effect will tend toward a satu-rated condition.

At normal condensing temperatures, the followingrelation between each 1F. of subcooling and thecorresponding change in saturation pressure ap-plies.

Equivalent Changein Saturation

Refrigerant Subcooling Pressure

R-12 1 F. 1.75 psi

R-22 1 F. 2.75 psi

R-502 1 F. 2.85 psi

To illustrate, 5F. subcooling will allow a pressuredrop of 8.75 psi with R-12, 13.75 psi with R-22, and14.25 psi with R-502 without flashing in the liquidline. For the previous example of a condensing unitin a basement requiring a vertical lift of 30 feet orapproximately 15 psi, the necessary subcooling forthe liquid head alone would be 8.5F. with R-12,5.5F. with R-22, and 5.25F. with R-502.

The necessary subcooling may be provided by thecondenser used, but for systems with abnormallyhigh vertical risers, a suction to liquid heat ex-changer may be required. Where long refrigerantlines are involved, and the temperature of thesuction gas at the condensing unit is approachingroom temperatures, a heat exchanger located nearthe condenser may not have sufficient tempera-ture differential to adequately cool the liquid, andindividual heat exchangers at each evaporatormay be necessary.

In extreme cases, where a great deal of subcoolingis required, there are several alternatives. A spe-cial heat exchanger with a separate subcoolingexpansion valve can provide maximum coolingwith no penalty on system performance. It is alsopossible to reduce the capacity of the condenser sothat a higher operating condensing temperaturewill make greater subcooling possible. Liquidrefrigerant pumps may also be used to overcomelarge pressure drops.

Liquid line pressure drop causes no direct penaltyin power consumption, and the decrease in systemcapacity due to friction losses in the liquid line isnegligible. Because of this the only real restrictionon the amount of liquid line pressure drop is theamount of subcooling available. Most referenceson pipe sizing recommend a conservative ap-proach with friction pressure drops in the 3 to 5 psirange, but where adequate subcooling is available,many applications have successfully used muchhigher design pressure drops. The total frictionincludes line losses through such accessories assolenoid valves, filter-driers, and hand valves.

In order to minimize the refrigerant charge, liquidlines should be kept as small as practical, andexcessively low pressure drops should be avoided.On most systems, a reasonable design criteria is tosize liquid lines on the basis of a pressure dropequivalent to 2F. subcooling.

A limitation on liquid line velocity is possible dam-age to the piping from pressure surges or liquidhammer caused by the rapid closing of liquid linesolenoid valves, and velocities above 300 FPMshould be avoided when they are used. If liquid linesolenoids are not used, then higher velocities canbe employed. Figure 83 gives liquid line velocitiescorresponding to various pressure drops and linesizes.

To summarize, in sizing liquid lines, it is recom-mended that the selection of line size be made onthe basis of a total friction pressure drop equivalentto 2F. subcooling. If vertical lifts or valves withlarge pressure drops are involved, then the de-signer must make certain that sufficient subcoolingis available to allow the necessary pressure dropwithout approaching a saturation condition at whichgas flashing could occur. Check Figure 83 to be


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sure velocities do not exceed 300 FPM if a liquidline solenoid is used.

Recommended liquid line sizes for varying capaci-ties and equivalent lengths of line are given inTable 27, page .

SIZING SUCTION LINES

Suction line sizing is the most critical from a designand system standpoint. Any pressure drop occur-ring due to frictional resistance to flow results in adecrease in the pressure at the compressor suc-tion valve, compared with the pressure at theevaporator outlet. As the suction pressure isdecreased, each pound of refrigerant returning tothe compressor occupies a greater volume, andthe weight of the weight of the refrigerant pumpedby the compressor decreases. For example, atypical low temperature R-502 compressor at -40F. evaporating temperature will lose almost 6%of its rated capacity for each 1 psi suction linepressure drop.

Normally accepted design practice is to use as adesign criteria a suction line pressure drop equiva-lent to a 2F. change in saturation temperature.Equivalent pressure drops for various operatingconditions are shown in Table 25.

Evaporating Pressure Drop, PSITemperature R-12 R-022 R-502

45F 2.0 3.0 3.320F 1.35 2.2 2.40F 1.0 1.65 1.85

-20F .75 1.15 1.35-40F .5 .8 1.0

Of equal importance in sizing suction lines is thenecessity of maintaining adequate velocities toproperly return oil to the compressor. Studies haveshown that oil is most viscous in a system after thesuction vapor has warmed up a few degrees fromthe evaporating temperature, so that the oil is nolonger saturated with refrigerant, and this conditionoccurs in the suction line after the refrigerant vaporhas left the evaporator. Movement of oil throughsuction lines is dependent on both the mass andvelocity of the suction vapor. As the mass ordensity decreases, higher velocities are requiredto force the oil along.

Nominal minimum velocities of 700 FPM in hori-zontal suction lines and 1500 FPM in verticalsuction lines have been recommended and usedsuccessfully for many years as suction line sizingdesign standards. Use of the one nominal velocityprovided a simple and convenient means of check-ing velocities. However, tests have shown that invertical risers the oil tends to crawl up the innersurface of the tubing, and the larger the tubing, thegreater velocity required in the center of the tubingto maintain tube surface velocities which will carrythe oil. The exact velocity required in vertical linesis dependent on both the evaporating temperatureand the line size, and under varying conditions, thespecific velocity required might be either greater orless than 1500 FPM.

For better accuracy in line sizing, revised maxi-mum recommended vertical suction line sizes basedon the minimum gas velocities shown in the 1980ASHRAE Handbook have been calculated and areplotted in chart form for easy usage in Figures 84and 86. These revised recommendations super-seded previous Copeland vertical suction riserrecommendations. No change has been made inthe 700 FPM minimum velocity recommendationfor horizontal suction lines, and Figures 85 and 87cover maximum recommended horizontal line sizesfor proper oil return.

To illustrate, again assume a system operatingwith R-12 at 40F. evaporating temperature has acapacity of 100,000 BTU/hr. On Figure 84, theintersection of the evaporating temperature andcapacity lines indicate that a 2 1/8 O.D. line will berequired for oil return in the vertical suction risers.

Even though the system might have a much largerdesign capacity, the suction line sizing must bebased on the minimum capacity anticipated inoperation under light load conditions after allowingfor the maximum reduction in capacity from capac-ity control if provided.

Since the dual goals of low pressure drop and highvelocities are in direct conflict, obviously compro-mises must be made in both areas. As a generalapproach, in suction line design, velocities shouldbe kept as high as possible by sizing lines on thebasis of the maximum pressure drop that can be


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tolerated, but in no case should gas velocity beallowed to fall below the minimum levels necessaryto return oil. It is recommended that a tentativeselection of suction line sizes be made on the basisof a total pressure drop equivalent to a 2F. changein the saturated evaporating temperature. CheckFigures 84 or 86 to be sure that velocities in verticalrisers are satisfactory. Where refrigerant lines arelengthy, it may be desirable to use as large tubingas practical to minimize pressure drop, and Figure85 or 87 should be checked to determine themaximum permissible horizontal line size. Thefinal consideration must always be to maintainvelocities adequate to return oil to the compressor,even if this results in a higher pressure drop than isnormally desirable.

Recommended suction line sizes for varying ca-pacities and equivalent lengths of line are given inTables 29 to 41, starting on page .

DOUBLE RISERS

On systems equipped with capacity control com-pressors, or where tandem or multiple compres-sors are used with one or more compressorscycled off for capacity control, single suction linerisers may result in either unacceptably high or lowgas velocities. A line properly sized for light loadconditions may have too high a pressure drop atmaximum load, and if the line is sized on the basisof full load conditions, then velocities may not beadequate at light load conditions to move oil throughthe tubing. On air conditioning applications wheresomewhat higher pressure drops at maximum loadconditions can be tolerated without any majorpenalty in overall system performance, it is usuallypreferable to accept the additional pressure dropimposed by a single vertical riser. But on mediumor low temperature applications where pressuredrop is more critical and where separate risersfrom individual evaporators are not desirable orpossible, a double riser may be necessary to avoidan excessive loss of capacity.

A typical double riser configuration is shown inFigure 88. The two lines should be sized so that thetotal cross-sectional area is equivalent to the cross-section area of a single riser that would have bothsatisfactory gas velocity and acceptable pressuredrop at maximum load conditions. The two lines

normal are different in size, with the larger linetrapped as shown, and the smaller line must besized to provide adequate velocities and accept-able pressure drop when the entire minimum loadis carried in the smaller riser.

In operation, at maximum load conditions gas andentrained oil will be flowing through both risers. Atminimum load conditions, the gas velocity will notbe high enough to carry oil up both risers. Theentrained oil will drop out of the refrigerant gas flow,and accumulate in the P trap, forming a liquidseal. This will force all of the flow up the smallerriser, thereby raising the velocity and assuring oilcirculation through the system.

For example, assume a low temperature systemas follows:


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Maximum capacity 150,000 BTU/hr.Minimum capacity 50,000 BTU/hr.Refrigerant R-502Evaporating Temperature -40F.Equivalent length ofpiping, horizontal 125 ft.Vertical Riser 25 ft.Desired design pressure drop(equivalent to 2F.) 1 psi

A preliminary check of the R-502 pressure dropchart, Figure 78, indicates for a 150 foot run with150,000 BTU/hr capacity and a total pressure dropof approximately 1 psi, a 3 1/8 O.D. line is indi-cated. At the minimum capacity of 50,000 BTU/hr,Figure 87 shows a 3 5/8 O.D. horizontal suction lineis acceptable, but Figure 84 indicates that themaximum vertical riser size is 2 1/8 O.D. Referringagain to the pressure drop chart, Figure 78, thepressure drop for 150,000 BTU/hr through 2 1/8O.D. tubing is 4 psi per 100 feet, or 1.0 psi for the25 foot suction riser. Obviously, either a compro-mise must be made in accepting a greater pres-sure drop at maximum load conditions, or a doubleriser must be used.

If the pressure drop must be held to a minimum,then the size of the double riser must be deter-mined. At maximum load conditions, a 3 1/8 O.D.riser would maintain adequate velocities, so acombination of the sizes approximating the 3 1/8O.D. line can be selected for the double riser. Thecross sectional area of the line sizes to be consid-ered are:

3 1/8 O.D. 6.64 sq. in.

2 5/8 O.D. 4.77 sq. in.

2 1/8 O.D. 3.10 sq. in.

1 5/8 O.D. 1.78 sq. in.

At the minimum load condition of 50,000 BTU/hr.,the 1 5/8 O.D. line will have a pressure drop ofapproximately .5 psi, and will have acceptablevelocities, so a combination of 2 5/8 O.D. and 1

5/8 O.D. tubing should be used for the double riser.

In a similar fashion, double risers can be calculatedfor any set of maximum and minimum capacitieswhere single risers may not be satisfactory.

SUCTION PIPING FOR MULTIPLEX SYSTEMS

It is common practice in supermarket applicationsto operate several fixtures, each with liquid linesolenoid valve and expansion valve control, from asingle compressor. Temperature control of indi-vidual fixtures is normally achieved by means of athermostat opening and closing the liquid linesolenoid valve as necessary. This type of system,commonly called multiplexing, requires careful at-tention to design to avoid oil return problems andcompressor overheating.

Since the fixtures fed by each liquid line solenoidvalve may be cont

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