REMOVING OIL FROM A COMPRESSOR

  Occasionally problems in line sizing or system operation may caused oil to trap in the evaporator or suction line, and large amounts of oil may be added to the system in an effort to maintain a satisfactory oil level in the compressor. When the basic oil logging problem is corrected, the excess oil will return to the compressor crankcase, and unless removed from the system, can cause oil slugging, excessive oil pumping, and possible compressor damage. Also in cases where the system has been contaminated, for example by a broken water tube in a water cooled condenser, or in cleaning  a system after a bad motor burn, it may be necessary to completely remove the oil from the compressor crankcase.

  To some extent choice of a method for removing oil depends on the degree of system contamination. For removing excess oil or on system with only slight contamination, almost any method is acceptable. However if the system is badly contaminated, it may be advisable to remove the compressor bottom plate and thoroughly the interior of the crankcase.

1. Removing by Oil Drain Plug
  Some compressor are equipped with oil drain plugs. If so, this provides an easy method for removing oil.


  Close the suction service valve, and operate the compressor until the crankcase pressure is reduced to approximately 1 to 2 psig. Stop the compressor and isolate the crankcase by closing the discharge service valve. Carefully loosen the oil drain plug, allowing any pressure to bleed off before the threads are completely disengaged. Drain oil to the desired level by seepage around the threads without removing the plug.

  When draining is complete, tighten the drain plug, open the compressor valves, and restore the compressor to operation. The oil seal at the drain hole and the residual refrigerant pressure in the crankcase will effectively block the entrance of any measurable quantities of air or moisture into the system.

2. Removing by Oil Fill Hole
  If a drain plug is not convenient or is not furnished on the compressor, oil may be removed by means of the oil fill hole.

  Close the compressor suction service valve, reduce the crankcase pressure to 1 to 2 psig, and isolate by closing the discharge service valve.

  Carefully loosen the oil fill plug, allowing any pressure to bleed off before the threads are completely disengaged. Remove the oil fill plug, and insert a 1/4¨O.D. copper tube, so that the end is at or near the bottom of the crankcase. If possible use a tube of sufficient length so that the external end can be bent down below the crankcase, thus forming a syphon arrangement. Wrap a waste rag tightly around the oil fill opening, and crack the suction service valve, pressurizing the crankcase to approximately 5 psig, and then reclose the valve.

  Oil will be forced out the drain line, and will continue to drain by syphon effect until the crankcase is emptied. If the syphon arrangement is not possible, repressurize the crankcase as necessary to remove the desired amount of oil.

  The residual refrigerant pressure in the crankcase will prevent the entrance of any serious amounts of moisture or air into the system. Purge the crankcase by cranking the suction service valve off plug, tighten, open the compressor valves, and restore the compressor to operation.

  In large system where a large amount of excess oil must be removed, or where oil must be removed at intervals over a prolonged period, considerable time can be saved by brazing a dip tube in valve so that oil can be removed as desired as long as the crankcase pressure is above 0 psig. To speed up separation, the oil should be removed to a 1/4 sight glass level. After oil removal is complete, the oil level may then be raised to the normal operating level.


3. Removal by Means of Baseplate
  On accessible compressor, it may be necessary to remove the base plate if complete crankcase cleaning is necessary.

  Pump the system down to isolate the compressor, remove the base plate, clean as necessary, and reinstall with new gasket. Since both air and moisture can enter the crankcase during this operation, the crankcase should be evacuated with a vacuum pump before restoring to operation. In an emergency, the crankcase may be purged by cracking the suction service valve and venting through the oil fill hole and the discharge service port. Replace the plug in the oil fill hole and jog the compressor a few times by starting and stopping, discharging through the discharge service port. Cap the discharge service port, the discharge valve, and the compressor can be restored to operation.

4. Removing Oil From Welded Compressors
  If the oil must be removed from a welded compressor, for example to recharge with a measured amount of oil, the compressor must be removed from the system, and the oil drained out the suction line stub tilting the compressor

  After compressor is reinstalled the system must then be evacuated by means of an access valve or the process tube before recharging with refrigerant and restoring to operation.

CONDENSER

  Condenser construction must be rigid and rugged, and the fin surface should be treated for corrosion resistance unless the metal is corrosion resistant. The area in which the condensers mounted on the skirt of road splash, while those mounted high on the nose of a truck or trailer are in a somewhat cleaner atmosphere. If the condenser is mounted beneath a trailer facing in the direction of travel, a mud guard should be provided. The type of tube and fin construction affects the allowable fin spacing, but n general, fin spacing of no more than 8 fins to the inch is recommended, although as high as 10 and 12 per inch.

  Since the unit will operate for extended periods when the vehicle cannot be considered in designing  for adequate air flow, but the condenser fan should be located so that the ram air effect aids rather than opposes condenser air flow. It also should be born in mind that often many trucks will be operating side by side at a loading dock, and the air flow pattern should be such that one unit will not discharge hot air directly into the intake of the unit on the next vehicle.

  Since the space available for condenser face area is limited in transport refrigeration applications, the condenser tube circuiting should be designed for maximum efficiency.

  Low head pressure during cold weather can result in lubrication failure of compressors. With trucks operating or parked outside or in unheated garages in the winter months, this condition can frequently occur. A decreased pressure differential across the expansion valve will reduce the refrigerant flow, resulting in decreased refrigerant velocity and lower evaporator pressure, permitting oil to trap in the evaporator. Frequently the feed will be decreased to the point that short-cycling of the compressor results. The use of a reverse acting pressure control for cycling the condenser fan, or some other type of pressure stabilizing device to maintain reasonable head pressure is highly recommended.

REFRIGERANT CHARGE

  Refrigerant R-12 is used in most transport system at the present time, but R-502 is well suited for low temperature applications, and its use is increasing. Since R-502 creates a greater power requirement for a given compressor displacement than R-12, the motor-compressor must be properly selected for the refrigerant to be used. Different expansion valve are required for each refrigerant, so the refrigerant are not interchangeable in a given system and should never maximum working pressures than those used with R-12, so normally it is not feasible to attempt to convert an existing R-12 unit for the use of R-502.

 The refrigerant charge should be held to the minimum required for satisfactory operation. An abnormally high refrigerant charge will create potential problems of liquid refrigerant migration, oil slugging, and loss of compressor lubrication due to bearing washout or excessive refrigerant foaming in the crankcase.

  Systems should be charged with the minimum amount of refrigerant necessary to insure a liquid seal ahead of the expansion valve at normal operating temperatures. For an accurate indication of refrigerant charge, a sight glass is recommended at the expansion valve inlet, and a combination sight glass and moisture indicator is essential for easy field maintenance checking. It should be born in mind that bubbles in the refrigerant sight glass can be caused by pressure drop or restrictions in the liquid line, as well as inadequate nominal working charge data should be used only as a general guide, since each installation will vary in its charge requirements.

ELECTRICAL PRECAUTIONS

  Electrical failures are a common field maintenance problem due to the wet environment, shock and vibration, and the possibility of improper power from an engine generator set.

  For the safety of operating and maintenance personnel, the electrical system should be grounded to the frame, and the frame in turn grounded by means of a chain or metal link to the ground if a generator  set in is mounted on the vehicle. All components should be grounded from one to the other, such as the generator set to condensing section to evaporator section. Cables to remote sources of power should carry an extra wire for grounding purposes at the supply plug.

  At the time of manufacture, each system should be given a high potential test to insure against electrical flaws in the wiring. All relays and terminals should be protected against the weather, and all wiring should be covered with protective loom to guard against abrasion. All switches should be of the sealed type, recommended by the manufacturer for use in wet environments. Plus type line connectors should be of the waterproof type. Electrical cables connecting split units should have a watertight cable cover, or should be run in conduit. All wiring should be fastened securely to prevent chafing, and should be clearly identified by wire marking and/or following the color code specified by the National Electrical Code.

  Adequately sized extension cords, plugs, and receptacles must be used to avoid excessive voltage drop. Voltage at the compressor terminals must be within 10% of the nameplate rating, even under starting conditions. Many single phase starting problems on small delivery trucks can be traced to the fact that power is supplied to the compressor from household type wiring circuits through long extension cords, neither of which are sized properly for the electrical load. Single phase and type motors which are used for belt driver a compressor during over-the-road operation must be equipped with a relay to break the capacitor circuit, rather than a centrifugal switch. The variable speed operation experienced during truck operation may cause a centrifugal switch to fail because of excessive wear at low operating speeds. All start capacitors must be equipped with bleed resistors to permit the capacitor charge to bleed off rapidly. preventing arcing and overheating of the relay contacts.

  When units are operated from several power sources be sure all plugs and receptacles are wired in the same sequence, so that the compressor rotation will not be reversed.

COMPRESSOR SPEED

  Open type compressor operating from a truck engine by means of a power take-off or by a belt driver are subject to extreme speed ranges. A typical truck engine may idle at 500 RPM to 700 RPM, run at 1,800 RPM at 30 MPH, and run at 3,600 RPM over the highway at high speed. Whatever the power take-off or belt ratio, this means the compressor must operate through a speed ratio range of 6 to 1 or greater unless it is disconnected from the power source by some means.

  The compressor speed must be kept within safe limit to avoid loss of lubrication and physical damage. Operation within the physical limitations of the compressor may be possible for example from 400 RPM to 2,400 RPM. It may be possible to use a cut-out switch ta disconnect the compressor from the power source at a given speed. The compressor manufacturer should be contacted for minimum and maximum speeds of specific compressors.

  If the compressor is of  the accessible-hermetic type, there is no problem concerning speed so long as the electrical source is operating at the voltage and frequency for which the motor was designed. If the speed of the generator is varied in order to obtain variable speed operation, the voltage and frequency on the normal alternating current generator will vary proportionally. Since the compressor speed and motor load will vary directly with the frequency, it is often possible to operate over a wide speed range with satisfactory results.

  However, it should be born in mind that increasing the frequency and voltage of the generator above the level for which the compressor motor was designed will increase the load on the compressor, may overload the motor, and can result in bearing or other compressor damage. Operation at speeds too low to provide adequate compressor lubrication must also be avoided, although normally lubrication can be maintained on Copelametic compressors down to 600 RPM and possibly lower speeds.

  Each new application involving operation of the compressor at a voltage and frequency differing from its nameplate rating should be submitted to the Copeland Application Engineering Department for approval.

  One other problem that may arise with operation from a variable speed generator is the operation of electrical contactors, relays, etc. on voltages below or above their nameplate rating. Field tests have shown that the winding design and physical construction of electrical components can cause wide variation in voltage tolerance. The drop-out voltage of various types of commercially available 220 volt contactors may vary from 145 volts to 180 volts depending on construction. If it is planned to operate at variable voltage and frequencies, the electrical  components which are to be used should be extensively tested at the electrical extremes in cooperation with the manufacture to insure proper operation

TWO STAGE COMPRESSION AND COMPRESSOR EFFICIENCY

  In order to increase operating efficiency at low evaporating temperature, the compression can be done in two steps or stages. for two stage operation, the total compression ratio is the product of the compression ratio of each stage. In other words, for a total compression ratio of 16 to 1, the compression ratio of each stage might be 4 to 1; or compression ratios of 4 to 1 and 5 to 1 in separate stage will result in total compression ratio of 20 to 1.

  Two stage compression may be accomplished with the use of two compressor with the discharge of one pumping into the suction of the second, but because of the difficulty of maintaining proper oil levels in the two crankcases, it is more satisfactory to use one compressor with multiple cylinder. On Copeland two stage compressors, the ratio of low stage ti high stage displacement is 2 to 1. The greater volume of the low stage cylinder is necessary because of the difference in specific volume of the refrigerant vapor at low and interstage pressure. While the compression ratios of the two stage are seldom exactly equal, they will be approximately the same. A typical 6 cylinder two the compressor with its external manifold and desuperheating expansion valve is shown in figure, and a typical 3 cylinder two stage compressor with external manifold is shown

  Shows a comparison of five different volumetric efficiency curves. The threes straight lines are typical single stage curves one for an air conditioning compressor, one for a typical multi-purpose compressor, and one for a low temperature compressor. There are some variations in compressor design involved, but the primary difference in characteristics is due to clearance volume.

  The two vertical curved lines represent the comparative efficiency of a two stage compressor. Actually each separate stage would have a straight line characteristic similar to the single stage curves, but to enable comparison with single stage compressor, the overall volumetric efficiency has been computed on the basis of the total displacement of the compressor, not just the low stage displacement.

SINGLE STAGE LOW TEMPERATURE SYSTEMS

  Low temperature single stage system become increasingly critical from a design and application standpoint as the desired evaporating temperature is decreased. The combination of high compression ratios, low operating temperatures, and rarified  return gas can cause lubrication and overheating problems, and make the compressor more vulnerable to damage from moisture and contaminants in the system.

  The compressor selection, suction temperature, and application must be such that the temperature of the discharge line measured within 1¨ to 6¨of the discharge line service valve does not exceed 230° F. for Refrigerants 12,22 and 502. Under these conditions, the estimated average temperature at the discharge port (measured at the valve retainer on the valve plate) will be approximately 310° F. for R-12 and R-502, and 320° F. for R-22.

  The compressor displacement, pressure limiting devices, and quantity of cooling air or water must be selected to prevent the motor temperature from exceeding the limits stated below:

 A. 210° F. when protected by inherent protectors affected by line current and motor temperature.

 B. 190° F. when protected by motor starters.

  The temperature of the motor should be determined by the resistance method and should be determined when the compressor is tested in the highest ambient in which it is expected to operate, at 90 per cent of rated voltage, with 90° F. return suction gas temperature of 170° F. to 190° F. are highly recommended.

  In order to prevent the discharge and motor temperatures from exceeding recommended limits, it is very desirable, and in some instances absolutely necessary, to insulated the suction lines and return the suction gas to the compressor at a lower than normal temperature. This is particularly important with suction cooled compressor when R-22 is used. (Approximately 30° F. superheat suggested.)

  Suction cooled compressors required auxiliary cooling by means of an air below 0° F. evaporator temperature.

  Either the evaporator must be properly designed, or a pressure limiting device such as a pressure limiting expansion valve or crankcase pressure regulating valve must be provided to prevent motor overloading during pulldown period, or after defrost.

  Copeland now recommends R-502 for all single stage low temperature applications where evaporating temperature of -20° F. and below may be encountered. Now that R-502 is readily available, R-22 should not be used in single stage low temperature compressor, 5 H.P. and larger. The lower discharge temperature of R-502 have resulted in much more trouble-free operation.

  An adequate supply of oil must be maintained in the crankcase at all times to insure continuous lubrication. If the refrigerant velocity in the system is so low that rapid return of the oil is not assured, on adequate oil separator must be used. The normal oil level should be maintained at or slightly above the center of the sight glass. An excessive amount of refrigerant or oil must not be allowed in the system as it may result in excessive liquid slugging and damage to the compressor valve, pistons, or cylinders.

VIBRATION AND NOISE

  No matter how well the compressor is isolated, some noise and vibration will be transmitted through the piping, but both can be minimized by proper design and support of the piping.

  On small units a coil of tubing at the compressor may provide adequate protection against vibration. On larger units, flexible metallic hose is supported by vibration absorbing mounts allowing compressor movement, refrigerant lines should not be anchored solidly at the unit, but at a point beyond the vibration absorber, so the vibration can be isolated and not transmitted into the piping system.

  The noise characteristics of a large refrigeration or air conditioning system, particularly when installed with long refrigerant lines and remote condensers, are not predictable. Variation in piping configuration, the pattern of gas flow, line sizes, operating, all can affect the noise generated by the system. Occasionnlly a particular combination of gas flow and piping will result in a resonant frequency to an undesirable level. Gas pulsation from the compressor may also be amplified in a similar manner.

  If gas pulsation or resonant frequencies are encountered on a particular application, a discharge line muffler may be helpful in correcting the problem. The purpose of a muffler is to damper the pulses of gas in the discharge line and to change the frequency to a level which is not objectionable. A muffler normally depends on multiple internal baffles and/or pressure drop to obtain an even flow of gas. In general, the application range of a muffler depends on the volume and density of the refrigerant gas discharged from the compressor are both factor in muffler performance.

  A given muffler may work satisfactorily on a fairly wide range of compressor sizes, but it is also quite possible that a given system may require a muffler with a particular pressure drop to effectively dampen pulsations. On problem applications, trial and error may be the only final guide. While large muffler are often more efficient in reducing the overall level of compressor discharge noise, in order to satisfactorily dampen pulsations, smaller muffles with  a greater pressure drop are usually more effective. Adjustable mufflers are often helpful since they allow tuning of the muffler pressure characteristics to the exact system requirement.

  Occasionally, a combination of operating conditions, mounting and piping arrangement may result in a resonant condition, which tends to magnify compressor pulsation and cause a sharp vibration, although noise may not be a problem. For large Copelametic compressors, discharge muffler plates haven been developed for use when necessary to dampen excessive pulsation. The muffler plate fits between the discharge valve and the compressor body and has a number of muffling holes break up the pattern of gas flow  and create sufficient restriction to reduce the gas pulsation to a minimum.

  When piping passes through walls or floors, precaution should be taken to see that the piping does not touch any structural members and is properly supported by hangers in order o prevent the transmission of vibration into the building. Failure to do so may result in the building structure becoming a sounding board.

PIPING DESIGN FOR HORIZONTAL AND VERTICAL LINES

  Horizontal suction and discharge lines should be pitched downward in the direction of flow to aid in oil drainage, with a downward pith of at least  ​1⁄2 inch in 10 feet. Refrigerant lines should always be as short and should run as directly as possible.

  Piping should be located so that access to system components is not hindered, and so that any components, which could possibly  require future maintenance are easily accessible. If piping must be run through boiler rooms or other areas where they will be exposed to abnormally high temperatures, it may be necessary to insulate both the suction and liquid line to prevent excessive  heat transfer into the lines.

  Every vertical suction riser greater than 3 to 4 feet in height should have a ¨P¨ trap at the base to facilitate oil return up the riser as shown in figure. To avoid the accumulation of large quantities of oil, the trap should be of minimum depth and the horizontal section should be as short as possible. Prefabricated wrought copper traps are available, or a trap can be made by using two street ells and one regular ell. Traps at the foot of hot gas riser are normally not required because of the easier movement of oil at higher temperatures. However it is recommend that discharge line from the compressor be looped to the floor prior to being run vertically upward to prevent the drainage of oil back to the compressor head during shut down periods.

  For long vertical risers in both suction and discharge lines, additional traps are recommended for each full length of pipe (approximately 20 feet) to insure proper oil movement.

  In general, trapped section of the suction line should be avoided except where necessary for oil return. Oil or liquid refrigerant accumulating in the suction line during the off cycle can return to the compressor at high velocity as liquid slugs on start up, and can break compressor valves or cause other damage.

SUCTION PIPING FOR MULTIPLEX SYSTEMS

  It is common practice in supermarket application to operate several fixtures, each with liquid line solenoid valve and expansion valve control, from a single compressor. Temperature control of individual fixtures is normally achieved by meas of a thermostat opening and closing the liquid line solenoid valve as necessary. this type of system, commonly called multiplexing, requires careful attention to design to avoid oil return problems and compressor overheating.

  Since the fixtures fed by each liquid line solenoid valve may be each liquid line solenoid valve may be controlled individually, and since the load on each fixture is relatively constant during operation, individual suction lines and risers are normally run from each fixture or group of fixtures controlled by a liquid lines solenoid valve for minimum pressure drop and maximum efficiency in oil return. This provides excellent control so long as the compressor is operating at design suction pressure, but there may be periods of light load when most or all of the liquid line solenoid are closed. Unless some means of controlling compressor capacity is provided, this can result in compressor short cycling or operation at excessively low suction pressure, which can result only in overheating the compressor,

  Because of the fluctuations in refrigeration load caused by closing of the individual liquid line solenoid valves, some means of compressor capacity control must be provided. In addition, the means of capacity control must be such that it will not allow extreme variations in the compressor suction pressure.

  Where multiple compressor are used, cycling of individual compressor provides satisfactory control. Where multiplexing is done with a single compressor, a hot gas bypass system has proven to be the most satisfactory means of capacity reduction, since this allows the compressor to operate continuously at a reasonably constant suction pressure while compressor cooling can be safely controlled by meas of a desuperheating expansion valve

  In all cases, the operation of the system under all possible combinations of heavy load, light load, defrost, and compressor capacity must be studied carefully to be certain that operating condition will be satisfactory

DOUBLE RISERS

  On system equipped with capacity control compressor, or where tandem or multiple compressors are used with one or more compressors cycle off for capacity control, single suction line risers may result in either unacceptably high or low gas velocities. A line properly sized for light load conditions may have too high a pressure drop at maximum load, and if the lines is sized on the basis of full load conditions, then velocities may be adequate at light load conditions to move oil through the tubing. On air conditioning application where somewhat higher pressure drops at maximum load conditions can be tolerated without any major penalty in overall system performance, it is usually preferable to the additional pressure drop imposed by a single vertical riser. But on medium or low temperature application where pressure drop is more critical and where separate risers from individual evaporators are not desirable or possible, a double riser may be necessary to avoid on excessive loss of capacity.

  A typical double riser configuration is shown in figure. The two lines should be sized so that the total cross-sectional area is equivalent to the cross-sectional area of a single riser that would have both satisfactory gas velocity and acceptable pressure drop at maximum load conditions. The two lines normally are different in size, and the smaller lines must be sized to provide adequate velocities and acceptable pressure drop when the entire minimum load is carried in the smaller riser

  In operation, at maximum load conditions gas and entrained oil will be flowing through both riser. At minimum load conditions, the gas velocity will not be high enough to carry oil up both risers. The entrained oil will drop out of the refrigerant gas flow, and accumulate in the¨P¨ trap forming a liquid seal. This will force all of the flow up the smaller riser, thereby raising the velocity and assuring oil circulation through the system.

SIZING SUCTION LINES

  Suction line sizing is the most critical from a design and system standpoint. Any pressure drop occurring due to frictional resistance to flow results in a decrease in the pressure at the compressor suction valve, compared with the pressure at the evaporator outlet. As the suction pressure is decreased, each pound of refrigerant returning to the compressor occupies a greater volume, and the weight of the refrigerant pumped by the compressor decreases. For example, a typical low temperature R-502 compressor at -40° F. Evaporation temperature will lose almost 6% of its rated capacity for each 1 psi suction line pressure drop.

  Normally accepted  design practice is to use as a design criteria a suction line pressure line drop equivalent to a 2° F. change in saturation temperature. Equivalent pressure drop for various operating conditions are shown in table

  Pressure drop equivalent for 2° F. change in saturation temperature at various evaporating temperature

               Evaporating      pressure drop PSI
               Temperature        R-12          R-22       R-502

                 45° F.                  2.0              3.0         3.3
             
                 20° F.                  1.35            2.2         2.4

                   0° F.                  1.0              1.65       1.85

                -22° F.                   .75             1.15       1.35

                -40° F.                   .5                 .8         1.0

  Of equal importance in sizing suction lines is the necessity of maintaining adequate velocities to properly return oil  to the compressor. Studies have shown that oil is most viscous in a system after suction vapor has warmed up a few degrees from the evaporating temperature, so that the oil is no longer saturated with refrigerant, and this condition occurs in the suction line after the refrigerant vapor has left the evaporator. Movement of oil through suction lines is dependent on both the mass and velocity of the suction vapor. As the mass or density decreases, higher velocities are required to force the oil along.

  Nominal minimum velocities of 700 FPM in horizontal suction lines and 1500 FPM in vertical suction lines have been recommended and used successfully for many year as suction lines sizing design standards. Use of the one nominal velocity provided a simple and convenient means of checking velocities. However tests have shown that in vertical risers the oil tends to crawl up the inner surface of the tubing, and the larger the tubing, the greater velocity required in the center of the tubing to maintain tube surface velocities which will carry the oil. The exact velocity required in vertical lines is dependent on both the evaporating temperature and the lines size, and under varying condition, the specific velocity required might be either greater or less than 1500 FPM.

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 sizing is to insure a solid liquid head of refrigerant at the expansion valve. If the pressure of the liquid refrigerant falls below its saturation temperature, a portion of the liquid will flash into vapor to cool the liquid refrigerant to the new saturation temperature. This can occur in a liquid if the pressure drops sufficiently due to friction or vertical lift.

  Flash gas in the has a detrimental effect on system performance in several ways. It increases the pressure drop due to friction, reduces the expansion device, may erode the expansion valve pin and seat, can cause excessive noise, and may cause erratic feeding of the liquid refrigerant to the evaporator.

  For proper system performance, it is essential that liquid refrigerant reaching the expansion device be subcooled slightly below its saturation temperature. On most system the liquid refrigerant is sufficiently subcooled as it leaves the condenser to provide for normal system pressure drops. The amount of subcooling necessary, however, is dependent on the individual system design.

  On air cooled and most water cooled applications, the temperature of the liquid refrigerant is normally higher that the surrounding ambient temperature, so no heat is transferred into the liquid, and the only concern is the pressure drop in the liquid line. Besides the friction loss caused by flow through the piping, a pressure drop equivalent to the liquid head is involved in forcing liquid to flow up a vertical riser. A head of two feet of liquid refrigerant is approximately equivalent to 1 psi. if a condenser or receiver in the basement of a building is to supply liquid refrigerant to an evaporator thee floors above, or approximately 30 feet, then a pressure drop of approximately 15 psi must be provided for in system design for the liquid head alone.

  On evaporative or water cooled condensers where the condensing temperature is below the ambient air temperature, or on any application where liquid lines must pass through hot areas such as boiler or furnace rooms, an additional complication may arise because of heat transfer into the liquid. Any subcooling in the condenser may be lost in the receiver or liquid line due to temperature rise alone unless the system is properly designed. On evaporative condensers where a receiver and subcooling coil are used, it is recommended that the refrigerant flow be piped from the condenser to the receiver and then to the subcooling coil. In critical applications it may be necessary to insulate both the receiver and the liquid line.

  On the typical air cooled condensing unit with a conventional receiver, it is probable that very little subcooling of liquid is possible unless the receiver is almost completely filled with liquid. Vapor in the receiver in contact with the subcooled liquid will condense, and this effect will tend toward a saturate condition.

  At normal condensing temperatures, the following relation between each 1° F, of subcooling and the corresponding in saturation pressure applies.


                                                                                            Equivalent Change
                                                                                                 in 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

SIZING HOT GAS DISCHARGE LINES

  Pressure drop in discharge lines is probably less critical than in any other part of the system.
Frequently the effect on capacity of discharge line pressure drop is over-estimated since it is assumed the compressor discharge pressure and the condensing pressure are the same. In fact, there are two different pressure, the compressor discharge  pressure being greater then the condensing pressure by the amount of the discharge line pressure drop. An increase in pressure drop in the discharge line increase the compressor discharge pressure materially, but have little effect on the condensing pressure. Although there is a slight increase in the heat of compression for an increase in head pressure, the volume of gas pumped is decreased slightly due to a decrease in volumetric efficiency of the compressor. Therefore the total heat to be dissipated through the condenser may be relatively unchanged, and the condensing temperature and pressure may be quite stable, even though the discharge line pressure drop and therefore the compressor discharge pressure might vary considerably.

  The performance of a typical Copelametic compressor, operating at air conditioning condition with R-22 and an air cooled condenser indicates that for each 5 psi pressure drop in the discharge line,the compressor capacity is reduced is increased about 1%. On a typical low temperature Copelametic compressor operating with R-502 and on air cooled condenser, approximately 1% of compressor capacity will be lost each 5 psi pressure in power consumption.

    As a general guide, for discharge line pressure drops up to 5 psi, the effect on system performance would be so small as to be difficult to measure. Pressure drops up to 10 psi would not be greatly detrimental to performance provided the condenser is sized to maintain reasonable condensing pressure.

  Actually a reasonable pressure drop in the discharge line is often desirable to dampen compressor pulsation, and thereby reduce noise and vibration. Some discharge line mufflers actually derive much of their efficiency from pressure drop

OIL SEPARATORS

  Proper refrigerant piping design and operation of the system within its design limits so that adequate refrigerant velocities can be maintained are the only cure for oil logging problems, but an oil separator may be a definite aid in maintaining lubrication where oil return problems are particularly acute.

  For example, consider a compressor having an oil charge of 150 ounces, with the normal oil circulation rate being 2 ounces per minute. This means that on a normal system with proper oil return at stabilized conditions, two ounces of oil leave the compressor through the discharge line every minute, and two ounces return through the suction line. If a minimum of 30 ounces of oil in the crankcase is necessary to properly lubricate the compressor, and for some reason oil logged in the system and failed to return the compressor, the compressor would run out of oil in 60 minutes. Under the same condition with an oil separator having an efficiency of 80%, the compressor could operate 300 minutes or 5 hour before running out of oil.

  As a practical matter, there seldom are condition in a system when no oil will be returned to the compressor, and even with low gas velocities, some fraction of oil leaving the compressor will be returned. If there are regular intervals of full load conditions or defrost periods when oil can help to bridge long operating periods at light load condition. Oil separators are mandatory on systems with flooded evaporators controlled by a float valve, on all two stage and cascade ultra-low temperature system, and on any system where oil return is critical.

  Oil separators should be considered as a system aid but not a cure-ail or a substitute for good system design. They are never 100% efficient, and in fact may have efficiencies as low as 50% depending on system operating conditions. On systems where piping design encourages oil logging in the evaporator, an oil separator can compensate for system oil return deficiencies only on 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 the separator be connected to the suction line just ahead of the accumulator. This will provide maximum protection against returning liquid refrigerant to the crankcase. If the system is not equipped with a suction accumulator, the oil return line on suction cooled compressors may be connected to the suction line if more convenient than the crankcase,but on air cooled compressor, oil return must be made directly to the crankcase to avoid damage to the compressor valves.

  If the separator is exposed to outside ambient temperatures, it must be insulated to prevent refrigerant condensation during off periods, resulting in return of liquid to the compressor crankcase. Small low wattage strap-on heaters are available for oil separators, and if any problem from liquid condensation inn the separator is anticipated, a continuously energized heater is highly recommended.

OIL PRESSURE SAFETY CONTROL

  A major percentage of all compressor failures are caused by lack of proper lubrication. improper lubrication or the loss of lubrication can be due to a shortage of oil in the system, logging of oil in the evaporator or suction line due to insufficient  refrigerant velocities, shortage of refrigerant, refrigerant migration or floodback to the compressor crankcase, failure of the oil pump, or improper operation of the refrigerant control devices.

  Regardless of the initial source of the difficulty, the great majority of compressor failures due to loss of lubrication could have been prevented. Although proper system design, good preventive maintenance, and operation within the system design  limitations are the only cure for most of these problems, actual compressor damage usually can be averted by the use of on oil pressure safety control.

  An oil pressure safety control with a timer delay of 120 seconds is a mandatory requirement of the Copeland warranty on all Copelametic compressor having an oil pump pressure and crankcase pressure, and the two minute delay serves to avoid shut down during short fluctuations in oil pressure during start-up.

  A  trip of the oil pressure safety switch is a warning that the system has been without proper lubrication for a period of two minutes. Repeated trips of the oil pressure safety control are a clear indication that something in the system design or operation requires immediate remedial action. On a well designed system, there should be no trips of the oil pressure safety control, and repeated trips should never be accepted as a normal pert of the system operation.

  The oil pressure safety control will not protect against all lubrication problems. It cannot detect whether the compressor is pumping oil or a combination of refrigerant and oil. If bearing trouble is encountered on system where the oil pressure safety control has not tripped, even though inspection proves it to be properly wired, wit the proper pressure setting, and in good operating condition, marginal lubrication is occurring which probably is due to liquid refrigerant floodback.