US3763659A - Refrigeration process, apparatus and method - Google Patents

Refrigeration process, apparatus and method Download PDF

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US3763659A
US3763659A US00222733A US3763659DA US3763659A US 3763659 A US3763659 A US 3763659A US 00222733 A US00222733 A US 00222733A US 3763659D A US3763659D A US 3763659DA US 3763659 A US3763659 A US 3763659A
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compressor
evaporator
orifice means
refrigeration system
refrigerant
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P Hover
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Tecumseh Products Co
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Tecumseh Products Co
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B49/00—Arrangement or mounting of control or safety devices
    • F25B49/02—Arrangement or mounting of control or safety devices for compression type machines, plants or systems
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B31/00—Compressor arrangements
    • F25B31/02—Compressor arrangements of motor-compressor units

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  • ABSTRACT A method in which a fixed size orifice is selected and inserted between the evaporator outlet and the suction inlet of a compressor in a refrigeration system to restrict gas flow to the compressor and thereby decrease the peak torque required to operate the compressor.
  • the disclosed apparatus is a hermetic compressor unit with such a calibrated orifice installed in each of the compressor intake tubes, thereby allowing the use of an electric drive motor which otherwise would be incapable of developing sufficient peak load torque after starting to operate the compressor in a refrigeration system with an evaporator at ambient temperature in the absence of such orifices.
  • a vapor-compression refrigeration system with a hermetically sealed compressor and a refrigerant such as R-12 is used in most conventional home refrigerators and freezers.
  • the gaseous refrigerant is discharged from the evaporator and received at the suction inlet of the compressor at a pressure in the neighborhood of pounds per square inch absolute (p.s.i.a.).
  • the pressure of the refrigerant at the suction inlet to the compressor may increase approximately three-to-five fold.
  • the abnormally high mass flow rate of refrigerant through the compressor 7 during start-up of a pressure and/or temperature equalized refrigeration system is substantially decreased by coupling the discharge of the evaporator to the inlet of the compressor through an inexpensive orifice of a preselected fixed size.
  • the term orifice as referred to here and throughout the remainder of this specification is intended to mean one or more nozzle or sharp-edged orifices (including thin plate orifices), with plural orifices being arranged in parallel.
  • the orifice automatically throttles abnormally high mass fiow rates of the refrigerant so that the compressor is not unduly loaded during this temporary load condition and yet allows a sufficient mass of refrigerant to flow into the compressor during normal operation of the refrigeration system.
  • Apparatus constructed pursuant to the method of this invention includes a calibratedorifice of a given selected size .coupled to the inlet of a conventional compressor gas pump so as to enable it to be driven by an electric motor of a selected reduced capacity; i.e., a motor having a maximum torque rating below that otherwise required to initially start-up and drive the compressor in a refrigeration system with an abnormally large mass flow rate condition were the orifice not present, as described in greater detail hereafter.
  • Object of this invention are to provide a method of constructing a compressor for a mechanical vapor compression refrigeration system which substantially decreases the torque load imposed on a compressor in a pressure and/or temperature equalized refrigeration system, and a motor-driven compressor unit constructed pursuant to the method for use in such a refrigeration system which is compact and of economical and reliable construction.
  • FIG. 1 is an end view partially in vertical section of a hermetically sealed compressor unit with an electric drive motor for use in a vapor compression refrigeration system embodying this invention.
  • FIGS. 2 and 3 are side and end views, respectively, of a restricted orifice of the hermetically sealed compressor unit of FIG. 1.
  • FIG. 4 is an isometric view partially in section of the compressor unit of FIG. 1 with the hermetic shell and various other component parts removed therefrom.
  • FIG. 5 is a semi-schematic drawing of a vapor compression refrigeration system embodying this invention with an evaporator and condensor respectively connected to the inlet and outlet of the hermetically sealed compressor unit of FIG. 1.
  • FIG. 1 illustrates a hermetically sealed compressor unit 10 embodying apparatus for performingthe method of this invention in a vapor compression refrigeration system.
  • Hermetic compressor unit 10 has an outer shell 12 encasing a compressor 14 resiliently suspended therein which is driven by an electric motor 16.
  • Compressor I4 is of the positive displacement reciprocating piston type with a discharge leaf valve 18 mounted in a discharge chamber 19 of a cylinder head 20 secured to a cylinder block 22 by bolts 24.
  • compressor 14 is a low-side casing type and hence has a discharge outlet coupling 26 fixed to a side wall of shell 12 which is coupled with the interior of discharge chamber 19 through tubular conduits 28 and 30.
  • a refrigerant such as R-l2 (dichlorodifluoromethane) is received in the casing space defined by hermetic shell 12 through an inlet coupling 32, and the bottom portion of the shell provides a reservoir for lubricating oil.
  • the refrigerant gas to be compressed is supplied from the interior of shell 12 to inlet valve ports 36 of compressor 14 through a pair of upright intake tubes 38 which are mounted at their lower ends in cylinder block 22 and each connected via a muffler chamber 40, passageway 41, passageway 42 and suction chamber 44 with ports 36 in cylinder head 20.
  • Intake mufflers 40 and passages 41 and 42 may be cored or machined in cast iron cylinder block 22.
  • Each intake tube 38 is press fit and/or silver soldered or adhesively secured in a counter bore 46 in the upper end of passageway 41.
  • the compressor construction described above may be conventional and hence these and other details thereof will be well understood by those skilled in the art.
  • a calibrated orifice is selected and installed in the inlet passageway leading to ports 36.
  • this orifice is preferably provided in the form of hollow restrictor plugs 50 which are received and fixed one in each of the upper ends of intake tubes 38, as by press fitting the plugs therein.
  • each plug 50 is generally cylindrical and has a shoulder 52 adjacent one end which limits the extent to which the plug can be inserted into the associated tube 38.
  • Each plug 50 has a fixed flow restriction in the form of a coaxial restricted orifice 54 therethrough which has a minimum cross-sectional area of less than one third and preferably in the range of one fourth to one tenth of the cross-sectional area of the inside diameter of its associated intake tube 38.
  • the orifice 54 is fixed in size, it nevertheless provides a non-linear resistance to flow of gas therethrough; i.e., the pressure drop across orifice 54 varies directly with the mass flow rate of refrigerant gas therethrough in accordance with the known fluid dynamics of restrictive orifices.
  • the sun of the volumes of both intake tubes 38, both passages 41, both mufflers 40, both passages 42 and the common suction chamber 44 is preferably in the range of two-to-five times the volumetric displacement of reciprocating piston compressor 14.
  • this chamber consists of the usual suction muffler chamber 40 and associated passages 41, 42 and the suction chamber 44 formed in the head of the compressor plus the interior of the suction tube 38 itself downstream from the restriction plug 50.
  • the optimum point at which plug 50 should be located from the standpoint of the best starting characteristics is at the entrance to the suction tube. However, in some installations from the standpoint of noise reduction, it may be better to install plug 50 at the downstream end of suction tube 38 where it enters the muffler chamber proper.
  • Orifice 54 may also be located at the point where the return line from the refrigeration system enters the compressor casing; i.e., in coupling 32, rather than in the suction tube. However, this latter variation would not give the noise reduction properties of the illustrated design.
  • compressor unit with restricted orifices 54 installed in the intake tubes thereof is used in a vapor compressor refrigeration system with a condenser 56, capillary tube 58 and an evaporator 60 adapted for removing heat from a refrigerator cabinet 62.
  • Discharge outlet coupling 26 is connected to the inlet of condenser 56 and the outlet of the condenser is connected to the inlet of capillary tube 58.
  • the outlet of capillary tube 58 is connected to the inlet of evaporator 60 and the outlet of evaporator 60 is connected to inlet coupling 32 so that the compressor unit can circulate a refrigerant through the system.
  • evaporator 60 In most automatic defrosting refrigerators, evaporator 60 is located directly in refrigeration cabinet 62, and a thermostat senses the temperature of the evaporator and cycles the compressor. In normal operation of such a refrigerator, after the compressor has been shut off by the thermostat, the thermostat prevents the compressor from being restarted until the temperature of the evaporator in the food compartment rises to approximately 4045F. This rise in temperature melts any ice formed on evaporator 60, thereby defrosting the evaporator. In addition, this entails sufficient down time to allow the refrigeration system pressures to equalize after every compressor pumping cycle.
  • frost-free refrigerator In the so-called frost-free refrigerator, evaporator 60 is located outside cabinet 62 and a fan circulates air over the evaporator and into the cabinet to remove heat therefrom and cool the contents thereof.
  • hermetic compressor unit 10 and the fan are shut off and an electric coil is used to heat the evaporator to approximately 65F thereby melting any ice formed thereon and also usually allowing sufficient time for the refrigeration system pressures to equalize.
  • frost-free refrigerators there is a defrosting cycle every time the compressor stops, which usually occurs two to four times per hour.
  • the entire refrigeration system for all refrigerators and freezers usually reaches room or ambient temperature whenever the motor of the compressor unit is disconnected from its source of power for any substantial period of time.
  • the evaporator of the refrigeration system becomes heated substantially above its normal operating temperature and the refrigeration system reaches temperature as well as pressure equilibrium for one reason or another.
  • the compressor unit suction inlet pressure at coupling 32 is substantially equal to the discharge pressure of evaporator 60 and is in the neighborhood of 20 p.s.i.a.
  • the discharge pressure of the compressor unit at coupling 26 is substantially equal to the inlet pressure to condenser 56 and is in the neighborhood of 200 p.s.i.a. if the refrigeration system is cycled so that the evaporator temperature is maintained in the range of-lOF. to 0F.
  • orifices 54 due to their non-linear resistance to gas flow present a disproportionately lower resistance to gas flow therethrough and hence present only a slight increase in the pressure drop thereacross compared to that existing through tubes 38 not equipped with restrictor plugs 50.
  • the invention is particularly well suited for use with low and medium back pressure refrigeration systems wherein the density of the refrigerant increases sufficiently due to the heating of the evaporator to provide an increase in the mass flow rate through the compressor of sufficient magnitude that the power input or torque required to start and drive .the compressor is sharply increased.
  • the back pressure i.e., the pressure of the refrigerant at the outlet of the evaporator and inlet to the compressor, is dependent on the particular refrigerant used in the system and the temperature of the evaporator during normal substantially steady state operation of the refrigeration system.
  • the pressure of the refrigerant at the evaporator outlet is generally in the range of 10 to 30 p.s.i.a.
  • the pressure of the refrigerant at the evaporator outlet is in the range of to 45 p.s.i.a. for R-l2 and to 70 p.s.i.a. for R-22.
  • Low back pressure systems using a refrigerant such as R'l2 with the evaporator normally operat ing in the range of 40F. to 10F. are commonly used in home refrigerators and freezers.
  • medium 7 back pressure systems using a refrigerant such as R-l2 or R-22 with the evaporator normally operating in the range of l0F. to 30F. are used in package or beverage dispensers and commercial refrigerators such as display cases.
  • the minimum surface area in radial cross section of orifice 54 through plug 50 is selected to provide an optimum balance between the conflicting parameters of maximum output capacity and operating efficiency of the hermetic motor compressor unit 10 during normal operation thereof versus limiting the maximum torque required to run the compressor under the abnormally large mass flow rate conditions.
  • the effect of the orifices as a resistance to quired to run the compressor under peakload conditions.
  • This reduction in maximum torque imposed by abnormal load conditions in turn permits a smaller size motor to be used, or for the same size motor it provides a higher rating for the compressor because of its ability to run under more severe pressure conditions.
  • the cost benefits achieved from the reduction in the peak torque required to run a compressor in low and medium back pressure refrigeration systems achieved by using a calibrated orifice pursuant to the present invention more than offsets the slight loss in overall efficiency, which may be in the range of onehalf to 2 percent.
  • the loss in capacity result ing from the installation of a calibrated orifice or orifices 54 can be readily compensated for in existing compressor designs by increasing the bore diameter of the cylinder or cylinders in compressor 14 to thereby increase the displacement of the compressor an amount sufficient to offset the reduction in pumping capacity caused by the orifices. This change can be accomplished at very little cost.
  • a restriction in the form of the orifice 54 provides better: characteristics for optimizing of peak torque versus running capacity as compared to a restriction in the form of a long narrow tube.
  • the calibrated orifice could be a plug 50 as shown or it could be a thin plate or a mere necking down of the suction tube 38 to provide the appropriate ratio of area change.
  • the compressor manufacturer may provide a standardized compressor design capable of handling a range of system capacities.
  • a calibrated orifice 54 may be installed in the compressor according to parameters disclosed herein to prevent overloading thereof to thereby match the compressor to said selected system.
  • Calibrated orifices 54 having a minimum crosssectional area providing a pressure drop across the orifices in the range of 2 to 6 percent, and preferably approximately four percent, of the absolute pressure of the refrigerant in shell 12 during normal substantially steady state operation of the refrigeration system (i.e., after the effects of the heated evaporator and pressure equalization have been dissipated) are believed to function satisfactorily.
  • a pressure drop in the range of 0.4 to 1.2 p.s.i., and preferably 0.8 p.s.i. is believed to be satisfactory.
  • the diameter in inches of the minimum surface area in cross section should preferably be substantially equal to
  • A is an empirical constant equal to 0.24
  • B is an empirical constant equal to 0.12
  • K is an empirical constant equal to 267
  • V is the specific volume in cubic feet per pound mass of the gaseous refrigerant in shell 12 of the hermetic compressor during normal steady state operation of the refrigeration system
  • P is the desired pressure drop across the orifice in pounds per square inch (it being understood that the value (P) is raised to the power expressed in the large brackets of the foregoing expression)
  • Q is the mass flow rate in pounds per hour through the orifice during normal substantially steady state operation of the refrigeration system.
  • compressor 14 can be and preferably is driven by an electric motor 16 incapable of producing sufficient maximum torque to initially start and run the compressor after the evaporator has been heated and the refrigeration system pressure equalized in the absence of orifices 54.
  • an electric motor 16 capable of producing sufficient maximum torque to initially start and run the compressor after the evaporator has been heated and the refrigeration system pressure equalized in the absence of orifices 54.
  • a hermeticcompressor unit constructed in accordance with this invention having a single cylinder positive displacement reciprocating piston-type compressor with a volumetric displacement of 1.067 cubic inches can be successfully started and run in a pressure equalized refrigeration system with a heated evaporator having an R-l 2 refrigerant by an electric motor capable of developing a maximum torque of approximately 34 ounce feet at l volts when calibrated orifices 54 are utilized.
  • This particular hermetic compressor unit had orifices providing a pressure drop of about 4 percent of the normal substantialy steady state pressure of the refrigerant within shell 12 which resulted in a reduction of more than 25 percent in the maximum torque required to start and run the compressor in a pressure equalized system with a heated evaporator.
  • each orifice 54 each having a minimum cross-sectional diameter of0. 125 inch in accordance with the above formula and. a length of three-eighths inch.
  • the pressure drop across each orifice was 0.8 pound per square inch when the upstream pressure was 19 p.s.i.a. and the mass flow rate through the compressor during substantially steady state operation was 24 pounds per hour.
  • the crosssectional area of the opening of each orifice was approximately one-ninth of the cross-sectional area of its associated intake tubes and the combined total volume of the intake tubes 38, intake passageways 41 and 42, intake mufflers 40 and suction chamber 44 was approximately seven cubic inches.
  • the refrigeration system was cycled to maintain the temperature of the evaporator in the range of l0F. to OF., except during defrosting when the evaporator was heated to raise the suction pressure to p.s.i.a., and the compressor started and ran with l 10 volts to the motor.
  • hermetic compressor unit constructed and used in a vapor compression refrigeration system in accordance with this invention is a hermetic compressor unit having a positive displacement reciprocating piston-type compressor with a volumetric displacement of 1.067 cubic inches, a motor developing a miximum torque of approximately 34 ounce feet at 1 15 volts and two orifices 54 each having a diameter of .109 inch and a length of three-eighths inch.
  • the pres sure drop across each of the orifices 54 was approximately 1.0 p.s.i. when the upstream pressure was 19 p.s.i.a.
  • each restricted orifice was approximately one-twelfth of its associated intake tube and the combined total volume of the intake tubes 38, intake passages 41 and 42, suction mufflers 40 and suction chamber 44 was 7 cubic inches.
  • This hermetic compressor unit was operated in a refrigeration system with an R-l2 refrigerant having a normal substantially steady state mass flow rate of 22.8 pounds per hour through the compressor with a normal substantially steady state pressure of 19 p.s.i.a. in shell 12 with the hermetic compressor unit being interniittently operated to maintain the temperature of the evaporator between 10F. to 0F. except during defrosting when the evaporator was heated to raise the suction pressure to 104 p.s.i.a., and the compressor started and ran with volts to the motor.
  • a further example of a hermetic compressor unit constructed and used in a vapor compression refrigeration system in accordance with this invention is a hermetic compressor unit having a positive displacement reciprocating piston-type compressor with a volumetric displacement of 1.067 cubic inches, a motor developing a maximum torque of approximately 34 ounce feet at l 15 volts and two restricted orifices 54 each having a diameter of 0.140 inch and a length of three-eighths inch.
  • the pressure drop across each orifice was approximately 0.06 p.s.i. when the upstream pressure was 19 p.s.i.a.
  • each orifice 54 was approximately one-seventh of its associated intake tube- 38 and the combined total volume of the intake tubes 38, intake passages 41 and 42, suction mufflers 40 and suction chamber 44 was 7 cubic inches.
  • This hermetic compressor unit was operated in a refrigeration system with an R-l2 refrigerant having a normal substantially steady state mass flow rate of 24.3 pounds per hour through the compressor with a normal substantially steady state refrigerant pressure of 19 p.s.i.a. in shell 12 with the hermetic compressor unit being intermittently operated to maintain the temperature of the evaporator between --lF. and 0F. except during defrosting when the evaporator was heated to raise the suction pressure to 87 p.s.i.a., and the compressor started and ran with 1 l0 volts to the motor.
  • a vapor compression refrigeration apparatus and process is provided which substantially decreases the peak power or torque requirements of the compressor in a pressure and/or temperature equalized refrigeration system, particularly with a heated evaporator.
  • the resulting hermetic compressor unit utilizes an electric drive motor having a substantially smaller maximum running torque capacity.
  • the invention thus provides a hermetic compressor unit of more economical construction and assembly compared to prior art hermetic compressor units.
  • the present invention is particularly useful in applications involving low to medium back pressure refrigeration systems, although it is not necessarily limited to such systems.
  • the initial starting torque demand is less than the peak torque encountered by motor 16 after the compressor 14 has been brought up to running speed, which may be for example 3,400 rpm. It only requires a fraction of a second, say one-tenth to one-half second, for the motor to accelerate the compressor from zero velocity to its normal running speed.
  • peak torque demand will be encountered normally from about to 30 seconds to about 1 minute after start-up of the compressor. This peak torque will normally last for a few minutes until the suction or back pressure (normally measured in the compressor housing 12 prior to entry of the refrigerant gas into the suction tubes 38) has been reduced approximately to normal operating values in the system.
  • the present invention enables a compressor motor of reduced capacity to drive through this period of peak loading by limiting the maximum gas pumping load which can be presented by the system to the compressor.
  • the peak load conditions presented by various types of vapor compression refrigeration systems will, of course, vary due to such factors as the system having a defrost cycle or frost-free mode of operation, as well as the type of condition encountered at start-up of the compressor.
  • a soak-out condition i.e., wherein the system is shut off for an extended period of time, normally 18 to 24 hours, so that both temperature and pressure equalization will occur throughout the system
  • an abnormal gas pumping load may be imposed on the compressor.
  • pressure equalization can also occur without temperature equilibrium having been established, as in the case of a defrost cycle.
  • the condenser is already hot, the condenser gas temperature being say in the range of F.
  • the lubricating oil present in the oil sump at the bottom of shell or housing 12 will drop in temperature. This factor, coupled:with the quiescent state of the oil, will allow the oil to absorb some of the refrigerant vapor.
  • the system When the vapor goes into solution in the oil, the system will be depleted of some of its refrigerant in vapor form, thereby reducing the gas pumping load encountered in pul ing down the system after start-up.
  • the oil remains too hot to absorb much of thie refrigerant vapor and hence more refrigerant in vapogr form will be present in the system.
  • the gas pressure can reach higher levels because the motor will "still be hot from its previous operating cycle, and there is more gas in the system because less of the gas can be absorbed by the higher temperature oil.
  • the condenser will be hotter than the ambient temperature and hence the discharge pressure of the compressor will be higher after start-up and the rise in discharge pressure will be accelerated.
  • the gas temperature is high enough in the system, there still may not be enough refrigerant present to produce saturated gas conditions.
  • short shutdown cycles may pose more severe gas loading problems than encountered after a soak-out condition has occurred.
  • the density of the refrigerant gas at the inlet to the compressor may be at least doubled compared to the density of the refrigerant gas during normal operation of the refrigeration system.
  • the evaporator temperature reaches about 80F., usually no liquid refrigerant will be left in the system.
  • saturated vapor conditions will no longer prevail and, therefore, above this temperature the pressure does not increase as fast due to the presence of superheated vapor conditions in the system. Nevertheless, until such superheated vapor conditions are reached, the aforementioned two-to-five fold density change or increase in back pressure can and does occur.
  • the provision of a calibrated orifice means of the present invention will reduce the maximum pressure and temperature conditions produced in the system when the condenser fan fails or the air flow through the condenser is reduced or restricted due to other adverse conditions, such as dirt or dust accumulation in the condenser, or when the ambient temperature condition of the condenser is abnormally high.
  • a method of decreasing the peak torque demand of a hermetically sealed positive displacement compressor in a vapor compression refrigeration system containing a refrigerant and wherein a low or medium back ressure exists at the outlet of an evaporator of the system during normal operation thereof comprising the steps of providing a compressor having a volumetric displacement rating in excess of the requirements of said system, selecting flow restriction orifice means having a fixed minimum cross-sectional area calibrated to limit the mass flow rate of said refrigerant into said compressor after it is just started in the operation of said system under an abnormal condition with said evaporator at an elevated temperature relative to its normal operating temperature, and locating said orifice means in said refrigeration system such that all of the refrigerant entering the suction inlet of the compressor does so exclusively via said orifice means, whereby the peak torque required to run the compressor under said abnormal condition is decreased compared to the torque required to run said compressor under the same conditions without said orifice means.
  • the fixed minimum cross-sectional area of said flow restriction means is generally circular and has a diameter D in inches substantially equal to where K equals 167, A equals 0.24, 8 equals 0.12, V equals the specific volume in cubic feet per pound mass of the refrigerant immediately upstream of the orifice, P equals the pressure drop across the flow restriction means in pounds per square inch, and Q equals the mass flow rate in pounds per hour through the compressor when the refrigeration system is operating under the conditions of Section 6-2 of Standard 520 of the American Refrigeration Institute, with the evaporator temperature at about --l0F., the gas temperature entering the compressor at about F., the compressor ambient temperature at about 90F., the liquid temperature at the expansion valve at about 90F., and the condensing temperature at about F.
  • a hermetic compressor unit for a refrigeration system having a hermetic casing with inlet means for supplying a refrigerant gas to the space in said casing, a positive displacement compressor with an inlet valve, said compressor being mounted in said casing and adapted to be driven by an electric motor, a muffier chamber having an outlet communicating with said compressor inlet valve and a suction pipe extending generally upright in said casing and communicating at its lower end with an inlet of said muffler chamber and at its upper end with the space in said casing, the improvement comprising restricted orifice means having a fixed minimum cross-sectional area in said suction pipe for restricting the mass flow rate of refrigerant therethrough, said orifice means being located such that all of the refrigerant enters said compressor exclusively via said restricted orifice means and being dimensioned to limit the mass flow rate of refrigerant to a given value through said compressor when said compressor is started under a given condition of abnormally high density of gaseous refrigerant in said casing of at
  • V the specific volume in cubic feet per pound mass of the refrigerant immediately upstream of the orifice
  • P the pressure drop across the orifice in pounds per square inch
  • Q the mass flow rate through the compressor in pounds per hour where the refrigeration system is operating under the conditions of Section 6-2 of Standard 520 of the American Refrigeration Institute, with evaporator temperature at l0F., gas temperature entering the compressor at about F compressor ambient temperature at about 90F, liquid temperature at the expansion valve at about 90F., and condensing temperature at about F.
  • said orifice means has a fixed minimum cross-sectional area and is located between the outlet of the evaporator and the suction inlet of the compressor so as to continuously couple the same through said orifice means.
  • a vapor compression refrigeration system containing a refrigerant and having a low to minimum back pressure at the outlet of an evaporator of the system during normal operation of said system, said system further having the inlet of a hermetically sealed positive displacement compressor coupled to the outlet of the evaporator exclusively via orifice means having a fixed minimum cross-sectional area calibrated to limit the maximum mass flow rate into the compressor after it is first started and an abnormal back pressure occurs with the evaporator at an elevated temperature compared to its normal operating temperature compared to its normal operating temperature whereby the torque required to operate the compressor is decreased compared to the torque required to operate the compressor under the same conditions without said orifice means.
  • said hermetically sealed positive displacement compressor comprises an electric motor, a positive displacement gas pump driven by said motor, and a casing encapsulating and hermetically sealing said motor and said gas pump therein, said electric motor being incapable of producing sufficient torque to operate said gas pump with said evaporator at said elevated temperature in the absence of said orifice means and being capable of developing sufficient torque to operate said gas pump with said orifice means.

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US4506519A (en) * 1983-08-24 1985-03-26 Tecumseh Products Company Hermetic compressor discharge line thermal block
US4549857A (en) * 1984-08-03 1985-10-29 Carrier Corporation Hermetic motor compressor having a suction inlet and seal
JPH0321544A (ja) * 1989-06-15 1991-01-30 Daihatsu Motor Co Ltd 配線装置
US5341654A (en) * 1993-04-16 1994-08-30 Copeland Corporation Suction gas conduit
US20060048529A1 (en) * 2003-12-22 2006-03-09 Shin Jong M Refrigerating system for refrigerator
US20070033965A1 (en) * 2005-08-09 2007-02-15 Carrier Corporation Refrigerant system with suction line restrictor for capacity correction
US20100095691A1 (en) * 2007-03-12 2010-04-22 Naoshi Kondou Cooling storage and method of operating the same
US20160281701A1 (en) * 2009-01-09 2016-09-29 Aurelio Mayorca Method and equipment for improving the efficiency of compressors and refrigerators
US11892211B2 (en) 2021-05-23 2024-02-06 Copeland Lp Compressor flow restrictor
US12422173B2 (en) 2022-08-19 2025-09-23 Copeland Lp Multiple-compressor system with oil balance control

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DE102016219535A1 (de) 2016-10-07 2018-04-12 Mahle International Gmbh Verdampfer

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US3264842A (en) * 1963-10-10 1966-08-09 Ranco Inc Refrigerating system and suction pressure responsive throttling valve therefor
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US2198258A (en) * 1937-01-21 1940-04-23 Crosley Corp Refrigeration system
US2133875A (en) * 1937-02-17 1938-10-18 Gen Electric Refrigerating machine
US2445527A (en) * 1945-10-01 1948-07-20 Brunner Mfg Company Compressor unloader
US2497668A (en) * 1947-10-03 1950-02-14 Gen Electric Overload limiting device
US2737030A (en) * 1951-12-21 1956-03-06 Nash Kelvinator Corp Refrigerating system having defrosting arrangement
US3264842A (en) * 1963-10-10 1966-08-09 Ranco Inc Refrigerating system and suction pressure responsive throttling valve therefor
US3401873A (en) * 1967-01-13 1968-09-17 Carrier Corp Compressor cylinder block

Cited By (15)

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US4506519A (en) * 1983-08-24 1985-03-26 Tecumseh Products Company Hermetic compressor discharge line thermal block
US4549857A (en) * 1984-08-03 1985-10-29 Carrier Corporation Hermetic motor compressor having a suction inlet and seal
JPH0321544A (ja) * 1989-06-15 1991-01-30 Daihatsu Motor Co Ltd 配線装置
US5341654A (en) * 1993-04-16 1994-08-30 Copeland Corporation Suction gas conduit
US20060048529A1 (en) * 2003-12-22 2006-03-09 Shin Jong M Refrigerating system for refrigerator
US7263849B2 (en) * 2003-12-22 2007-09-04 Lg Electronics Inc. Refrigerating system for refrigerator
WO2007021373A3 (en) * 2005-08-09 2007-05-03 Carrier Corp Refrigerant system with suction line restrictor for capacity correction
US7251947B2 (en) 2005-08-09 2007-08-07 Carrier Corporation Refrigerant system with suction line restrictor for capacity correction
US20070033965A1 (en) * 2005-08-09 2007-02-15 Carrier Corporation Refrigerant system with suction line restrictor for capacity correction
CN101243257B (zh) * 2005-08-09 2012-08-15 开利公司 具有用于能力修正的抽吸管线限制器的制冷剂系统
US20100095691A1 (en) * 2007-03-12 2010-04-22 Naoshi Kondou Cooling storage and method of operating the same
US20160281701A1 (en) * 2009-01-09 2016-09-29 Aurelio Mayorca Method and equipment for improving the efficiency of compressors and refrigerators
US10961995B2 (en) * 2009-01-09 2021-03-30 Aurelio Mayorca Method and equipment for improving the efficiency of compressors and refrigerators
US11892211B2 (en) 2021-05-23 2024-02-06 Copeland Lp Compressor flow restrictor
US12422173B2 (en) 2022-08-19 2025-09-23 Copeland Lp Multiple-compressor system with oil balance control

Also Published As

Publication number Publication date
FR2170030A1 (OSRAM) 1973-09-14
BE794727A (fr) 1973-07-30
FR2170030B3 (OSRAM) 1976-01-30
DE2303964A1 (de) 1973-08-09
JPS4888509A (OSRAM) 1973-11-20
BR7300764D0 (pt) 1973-09-25
NL7300013A (OSRAM) 1973-08-06
IT977032B (it) 1974-09-10
ES411219A1 (es) 1976-01-01

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