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.