US3885402A - Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use - Google Patents

Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use Download PDF

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US3885402A
US3885402A US433418A US43341874A US3885402A US 3885402 A US3885402 A US 3885402A US 433418 A US433418 A US 433418A US 43341874 A US43341874 A US 43341874A US 3885402 A US3885402 A US 3885402A
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compressor
injection
liquid refrigerant
working fluid
port
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US433418A
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Jr Harold W Moody
Donald D Schaefer
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MARSHALL INDUSTRIES Inc
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Dunham Bush Inc
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Priority to US433418A priority Critical patent/US3885402A/en
Priority to CA201,635A priority patent/CA1001432A/en
Priority to SE7407434A priority patent/SE404952B/xx
Priority to GB2523174A priority patent/GB1464996A/en
Priority to ZA00743672A priority patent/ZA743672B/xx
Priority to AU69960/74A priority patent/AU479659B2/en
Priority to FR7422738A priority patent/FR2257880B3/fr
Priority to IT69180/74A priority patent/IT1014495B/it
Priority to JP49079773A priority patent/JPS50102909A/ja
Priority to DE2434873A priority patent/DE2434873A1/de
Priority to BR7762/74A priority patent/BR7407762D0/pt
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Publication of US3885402A publication Critical patent/US3885402A/en
Assigned to BT COMMERCIAL CORPORATION reassignment BT COMMERCIAL CORPORATION SECURITY INTEREST (SEE DOCUMENT FOR DETAILS). (ASSIGNS THE ENTIRE INTEREST). Assignors: DUNHAM-BUSH, INC.
Assigned to CONNECTICUT BANK AND TRUST COMPANY, N.A., THE, A CORP. OF DE reassignment CONNECTICUT BANK AND TRUST COMPANY, N.A., THE, A CORP. OF DE SECURITY INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: DUNHAM BUSH INC.
Assigned to MARSHALL INDUSTRIES, INC. reassignment MARSHALL INDUSTRIES, INC. CHANGE OF NAME (SEE DOCUMENT FOR DETAILS). Assignors: DUNHAM-BUSH, INC.
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C28/00—Control of, monitoring of, or safety arrangements for, pumps or pumping installations specially adapted for elastic fluids
    • F04C28/10—Control of, monitoring of, or safety arrangements for, pumps or pumping installations specially adapted for elastic fluids characterised by changing the positions of the inlet or outlet openings with respect to the working chamber
    • F04C28/12—Control of, monitoring of, or safety arrangements for, pumps or pumping installations specially adapted for elastic fluids characterised by changing the positions of the inlet or outlet openings with respect to the working chamber using sliding valves
    • F04C28/125—Control of, monitoring of, or safety arrangements for, pumps or pumping installations specially adapted for elastic fluids characterised by changing the positions of the inlet or outlet openings with respect to the working chamber using sliding valves with sliding valves controlled by the use of fluid other than the working fluid
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C29/00—Component parts, details or accessories of pumps or pumping installations, not provided for in groups F04C18/00 - F04C28/00
    • F04C29/04—Heating; Cooling; Heat insulation
    • F04C29/042—Heating; Cooling; Heat insulation by injecting a fluid
    • 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
    • F25B1/00—Compression machines, plants or systems with non-reversible cycle
    • F25B1/04—Compression machines, plants or systems with non-reversible cycle with compressor of rotary type
    • F25B1/047—Compression machines, plants or systems with non-reversible cycle with compressor of rotary type of screw type

Definitions

  • the port is located at the optimum point of injection by determining; the minimum built-in volume ratio of the compressor, the rotor bore within which injection is to take place, the minimum and maximum system or operating compression ratios and by correlating these parameters with the pressure versus wrap angle plot for the selected compressor rotor bore.
  • the acceptable injection point for the injection port is thus determined in terms of the wrap angle from the suction side of the compressor for the screw rotor of the selected bore where continuous pressure injection of liquid refrigerant occurs within the working space regardless of undercompression or overcompression of the working fluid without significantly compromising compressor efficiency.
  • This invention relates to helical screw rotary compressors for operation within a refrigeration system, and more particularly. to a helical screw rotary compressor in which liquid refrigerant is injected within the compressor working space to cool the gaseous refrigerant working fluid during compression thereof.
  • Helical screw rotary compressors for compressing an elastic working medium generally comprise a housing or casing containing a working space formed by two parallel and intersecting bores with a low pressure or suction port at one end and a high pressure or discharge port at the other.
  • Two intermeshing helical screw rotors are rotatably mounted in respective bores and, if desired, a slidahle unload valve member may form part of the wall of the working space, whose position axially determines the capacity of the machine.
  • Refrigerant liquid injection cooling of compressor eliminates the need for the usual oil cooler.
  • Oil cooler heat injection load is usually a sizable percentage ofthc total heat rejection due to the amount of oil circulated. This eliminates the need for an oil cooler, piping and oil cooler control means.
  • FIG. 1 is a sectional, elevational and partially schematic view of a helical screw rotary screw compressor incorporating the liquid refrigerant injection scheme of the present invention.
  • FIG. 2 is a schematic, perspective view of the intermeshed male and female helical screw rotors of the compressor of FIG. 1.
  • FIG. 3 is a plot of the pressure curve for the screw compressor of FIG. I and a two dimensional representation of the variation of the change of the volume of the compressed working fluid within the compressor from suction to discharge.
  • FIG. 4a is a PV plot for the screw compressor of FIG. 1 under ideal conditions.
  • FIG. 4b is a PV plot for the compressor of FIG. I under overeompression conditions.
  • FIG. 4c is a PV plot of the compressor of FIG. I illustrating undercompression conditions.
  • FIG. 5 is an inlet end plane representation of the intermeshed screw rotors showing the creation of a closed pocket with the suction port superimposed thereon.
  • FIG. 6 is an outlet end plane representation of the compressor of FIG. I illustrating the angular relationship between the intermeshed screw rotors and one of the chambers sealed off from the suction and discharge sides of the compressor with the axial discharge port superimposed thereon.
  • FIG. 7 is a pressure versus male rotor wrap angle plot in the outlet end plane for the compressor of FIG. 1.
  • FIG. 8 is a pressure versus female rotor wrap angle plot in the outlet end plane for the compressor of FIG. 1.
  • the helical screw rotary compressor constitutes a positive displacement compressor in the same fashion of the typical reciprocating type compressor wherein a piston reciprocates within a cylinder toward and away from the head. with the gas pocket in this case being of maximum size at the end of the suction stroke, wherein the piston is at bottom dead center and with the chamber being compresssed to its smallest volume as the piston reaches top dead center under the compression stroke.
  • the rotor is cccentrically mounted with respect to the cylindrical chamber within which it rotates and the sliding vanes move radially in the rotor with a chamber determined between the outer cylindrical casing and the cylindrical rotor along with any two vanes and being quite large on the suction side of the machine. but being reduced con siderably in volume after approximately 180 rotation, wherein that chamber opens up to the discharge port.
  • the pictorial presentation of the change in volume of the chamber that is, the reduction in size from suction to discharge may be readily seen in two dimensions by a simple sectional view at right angles to the axis of rotation of the rotor in the case of the sliding vane type rotary compressor and a sectional view parallel to the reciprocating axis of a typical reciprocating type compressor.
  • the screw compressor while having fewer moving parts and constituting a rather simple operational concept, is not capable of providing a simple two dimensional presentation in terms of the change in volume of the individual gas chambers determined by the helical screws and the housing, but requires considerations of depth or third dimension.
  • FIG. 1 illustrates in two dimensions and in partial schematic form, a sectional view of a typical helical screw compressor.
  • the typical screw compressor indicated generally at 10 comprises a three part housing or casing as at 12, 14 and 16, the housing 12 being the suction side of the machine and being provided with an intake passage 18, housing 12 leading to the center housing 14 which houses the two intermeshed helical screw rotors, only one of which is shown in block form at 20 in the sectional elevational view of FIG. 1.
  • Housing 14 therefore contains the working space in the form of two intersecting bores, with bore 22 housing rotor 20 and wherein suction passage 18 terminates in a low pressure suction port at end plane 24.
  • the suction volume is illustrated schematically in quadrilateral form at 26 by lines 28, 29, 31 and 32.
  • a high pressure combined axial and radial discharge port at 34 at the opposite ends of the intermeshed screw rotors opens up into the compressor discharge passage 36 forming the principal portion of the housing 16.
  • Each intermeshed screw rotor, such as helical screw rotor 20 is supported for rotation by bearings 38 and 40, receiving the projecting ends of shafts 42 within respective housings I2 and 16.
  • the screw rotor 20 in this case is positively driven by drive shaft 44 which may form an integral part thereof.
  • the drive shaft 44 is driven by a prime mover (not shown).
  • Pertinent to the present invention. although conventional to rotary screw compressors. is an unloader valve indicated generally at 46 which is mounted for reciprocation within a bore 48 of central housing 14, the valve 46 defining part of the working space further defined by the bores such as bore 22 within casing I4 and the screw rotors themselves such as rotor 20.
  • the unloader valve 46 is provided at its discharge end with a scalloped or recessed leading edge 49 forming the radial discharge section of the discharge port 34 of the compressor.
  • the unloader valve 46 is of such axial length and cooperates with a stationary wall portion 50 of housing 14, which acts as a fixed stop for the unloader valve 46.
  • An unloader shaft 54 couples the U11 loader valve 46 to a piston 56 mounted for reciprocation within unloader cylinder 58 which is fixed to the housing section 12, the piston 56 being movable be tween the full line position shown and the dotted line position by the controlled application (not shown) of pressurized fluid such as lubricating oil bled from the lubricating system of the compressor to shift the unloader valve between load and unload positions as determined by refrigeration system load demand.
  • FIG. 1 illustrates in two dimensions the schematic change in volume of the gas chambers as formed between the helix of the rotors and the housing, this may be perhaps best appreciated by further reference to FIG. 2, which shows the essential change in volume of the gas volume open to suction at the left hand end and top of the intermeshed screws and the discharge volume at the lower right hand end of the compressor intermeshed screw rotors.
  • both the male helical screw rotor 20 and the female helical screw rotor 21 are shown as intersecting cylinders in dotted lines. with the intersecting cylinders indicative of the intermeshed lands and grooves,of respective rotors.
  • the unload valve 46 is shown as corresponding to full load position of FIG. 1.
  • the volume of the compressed gas open both to axial and radial porting leading to the discharge passage is indicated at 60 and takes an irregular form due to the configuration at the discharge end of the intermeshed screws.
  • the refrigerant in conventional fashion and in the manner of U.S. Pat. No. 3,795.l l7 travels in a closed loop from compressor discharge back to compressor suction.
  • the loop, FIG. 2 includes in order; condenser C where it is condensed (giving up heat), thermal expansion valve TXV where the liquid refrigerant expands. and evaporator E where it takes up heat by heat of vaporization.
  • the intermeshed screw rotors form individual gas chambers or pockets separated by vertical seal lines such as 62.
  • the chambers are identified in Roman numeral fashion at I. II. III. IV. V, VI and VII. chambers I, II. and III being open to suction and the vertical seal line 62' being the suction porting seal offline, that is, this line intersects lines 30 and 32 at point 64 corresponding to the same point 64 in FIG. 2 where none of the gas chambers to the right are open to suction. Compression occurs in a downstream direction, that is. to the right of cutoff point 64 and seal offline 62. In the schematic illustration, compression occurs within gas chambers IV. V and VI. with chamber VII being open to discharge. In this respect.
  • the size or volume of the gas chambers IV-VII are reduced thus causing compression of the gas contained therein as the chambers are moved in sequence from the suction zone to the discharge zone and in that respect obviously a new gas chamber replaces chamber I as it moves to position II as result of rotation of the screws in the manner illustrated in FIG. 2, due to the direct meshing of the helical lobes of respective rotors.
  • a gas chamber created by the intermeshed male and female rotors and the housing 14 may be more fully appreciated by reference to FIGS. 5 and 6.
  • the male rotor rotates at a speed one and one-half times that of the female rotor.
  • a horizontal center line passes through the axes of both rotors. FIG. 5, and intersects the contact point between lobe I of the male rotor and a point between lobes 1 and 2 of the six lobed female rotor 21 as viewed from the suction or inlet end.
  • the gas chambers formed thereby wrap 200 about female rotor 21 toward the discharge side of the machine, as indicated by the angle a, while correspondingly, that chamber portion of male rotor 20 wraps to angle (1 of 300 about the rotor axis as indicated on the outlet end plane representation of FIG. 6.
  • the screw compressor such as that illustrated in the figures has a fixed built-in volume ratio V,- V /V where V equals the volume of the suction and V equals the volume at discharge.
  • the screw compressor pressure ratio P,- in this case. is equal to V,-"' where k is the ratio of specific heat for the fluid being compressed.
  • K or n is ratio of specific heat.
  • the working fluid of the compressor may comprise refrigerants such as R22, where k equals 1.15, or ammonia (NH where k equals 1.29.
  • the system compression ratio C/R matches the compressor pressure ratio under ideal conditions, that is. P,- lJ/PS C /R. as evidenced in FIG. 4a.
  • a system compression ratio C/R is determined by relationships imposed by the system evaporator and condenser. This involves the heat exchange capability of each component. These relationships can change as follows:
  • FIG. 4b this is the situation where overcompression occurs and wherein the compressor pressure ratio P is greater than the system compression ratio C/R which is equal to P /P Overcompression occurs and lost work due to overcompression is graphically seen by the shaded area within the PV diagram and wherein opening up of a given chamber within the compressor to discharge drops the pressure within that chamber to discharge pressure which is significantly lower.
  • FIG. 4c the effects of undercompression are seen where P,- is less than P /R C/R, and wherein the shaded area within that plot is indicative of the lost work due to undercompression with the discharge having to further compress the working fluid within the compressor at the point where the chamber opens up to discharge pressure P,,.
  • FIG. 3 constitutes a composite of a graph of pressure across the compressor from suction to discharge and a schematic representation of the change in volume from the suction side to the discharge side of the compressor.
  • FIG. 3 further illustrates overcompression and undercompression at the dash line F and dash line P respectively.
  • the liquid refrigerant injection zone or injection window is identified as the zone existing between the two vertical arrows I and I and are superimposed again on a two dimensional schematic representation of the compression of the working fluid in the nature of the upper part of composite FIG. 3, while the pressure curves for the working fluid appear at the bottom of that figure.
  • the liquid refrigerant injection zone as partially defined by arrow I, lies just to the right of suction seal line 62, for gas chamber IV associated with the lead tip, approximately one seal line away from the suction seal off 62.
  • the injection Zone constitutes a zone within the compression region C of the screw compressor, FIG. 3, and while chamber IV just before lag tip clears a position where it closes off chamber IV to suctionfliquid refrigerant injection could occur just before cutoff, since due to the high velocity rotation of the screw rotors, as liquid is injected into the chamber, the kinetics of the machine require that the chamber be immediately thereafter sealed from the suction side and the gas within the same starts to be compressed as the chamber moves toward the point wherein it opens up to the discharge side of the machine.
  • the vertical arrow I indicates the other possible limit in the angular position of the liquid refrigerant injection port for the machine, that is, injection would have to cease just before the chamber opens up to the discharge side.
  • the theoretical maximum zone as defined by vertical arrows I and I may apply.
  • the zone is narrowed when over'compression occurs, since the liquid refrigerant pressure as defined by the compressor discharge pressure. is about equal to P which is the discharge manifold pressure (generally equal to refrigeration system condenser pressure from which the liquid refrigerant for injection is supplied) and must be higher than P the point P,,- on the pressure graph of FIG.
  • item 34 is the radial discharge port and in FIG. 2 item 60 is the axial discharge port.
  • the radial discharge port and axial discharge port are set up so that gas starts to discharge from both at the same instant, the radial and axial compressor ratio are the same. If gas begins discharging from the radial port first, the radial compression ratio is lower than the axial compression ratio. Gas will begin to discharge from the axial port after the compressor rotors have turned slightly further.
  • the true compressor compression ratio is defined. as the compression ratio determined by the port that has the lowest theoretical built-in," inother words, the port that begins discharging gas first.
  • the radial and axial compression ratio are usually the same.
  • the radial port usually leads the axial (radial begins discharging before axial) in order to have an improved part load operating power characteristics. For example, if it has been determined that V, radial is 3.0 for instance and V axial is 3.3, the minimum V, is 3.0 and this is the figure to be used in further calculations. It is this parameter which must be later applied to the plot of either FIG. 7 or FIG. 8.
  • Step number two is the establishment of which rotor bore liquid is to be injected into. This step determines the selection of the plot of FIG. 7 or FIG. 8 as the case may be. 7
  • Injection may be radial in a selected rotor bore or axial in an end plane.
  • the next step is the establishment of maximum and minimum system or operating compression ratios C/R as defined by the liquid refrigerant or other working liquid, and otherparameters related to the pressure and temperature of the working fluid at both the intake or suction side of the machine and the discharge side.
  • C/R maximum and minimum system or operating compression ratios
  • the next step is picking the angle at which injection is to take place. Injection must not take place at an angle that exposes the injection port to the minimum built-in pressure ratio of the compressor, that is, for example, assuming that female bore injection has been selected by step two and a minimum built-in of 3.0 as determined by step one, injection should never take place at an angle in excess of 286.
  • Vertical lines X and Y which emanate from the compressor built in line V, 3.0 at P and P plots where the lag tip and lead tip of the lobes defining a given closed compressor thread indicate two unacceptable areas for liquid refrigerant injection.
  • Vertical line Z which intersects the lag tip plot line at the point where the horizontal line indicated minimum operating compression ratio for the compressor and system defines the minimum injection point P in terms of the angular location on the outlet of the selected female rotor bore. The rationale for this selection follows.
  • injection occurs between 226 line X and 286 line Y, injection will be intermittent, and reverse injection flow can occur if the compressor operates near its lowest designed compression ratio 2.7 as defined by step three, which can damage the injection modulation valve as well as limiting liquid injection flow.
  • line Z injection is assured, since the pressure of the liquid refrigerant within the line insofar as the liquid refrigerant is concerned, is well above that within the particular chamber in which injection is occurring.
  • the low end of the overcompression range (210 to 226) forms the injection window or zone for female rotor bore liquid refrigerant injection.
  • the discharge pressure minus the injection port pressure assuming there are no pressure losses in the discharge line, the liquid injection line and the TX valve.
  • the pressure of the trapped gas that is being rotated and compressed by the rotor is defined by the rotor turning angle and the theoretical pressure that exists at the lag rotor tip or lead rotor tip.
  • the average pressure for a given rotor lobe is a pressure between lead tip and lag tip and about one-third distance above lead tip curve.
  • AP is a term used for determining pressure difference available in order to have liquid flow into the compression chamber as described below, keeping, in mind that unless the pressure of the liquid refrigerant to be injected is above that of the closed thread receiving the same, no liquid refrigerant will be injected as. AP will either be zero or negative in terms of desired direction of flow of liquid refrigerant from the bleed line into the closed thread of the compressor.
  • the AP at the injection port under minimum compression ratio conditions for R22 at 40/l05 would be 59 psid and at 20/l45, referred to earlier, the differential would be 258 psid, determined as follows:
  • the liquid refrigerant injection pressure must be an acceptable amount above injection port pressure to insure continued injection.
  • the measured port pressure P follows the relationship (P min (P max P min)/3) the AP (liquid to port) is then (P P,). For example,
  • the match of the built-in pressure ratio to the actual operating pressure ratio is significant.
  • the actual compression ratio C/R is narrower in variation, there is more latitude in the location of the point of injection because on one end there is the concern with leak back to suction, and on the other end, there is the concern of overcompression prior to exposure of the discharge porting which locates the injection point further back towards the suction.
  • the best brake horsepower per ton obviously occurs if there is absolute match between the discharge port location to the actual operating parameters of the compressor, and that the point of liquid injection is not earlier than need be.
  • step (5) 6. ascertaining from the plot of step (5) subsequent to step (4) an acceptable injection point for said injection port in terms of wrap angle from suction of the screw rotor for the bore as determined from step (2) which results in continuous injection of liquid refrigerant within said working space regardless of undercompression or overcompression of the working fluid and the position of the slide valve within the limits of the parameters determined by steps (1), (2) and (3), without significantly compromising compressor efficiency.
  • a helical screw rotary compressor forming a component of a refrigeration system, said system including in order from the compressor, a condenser, a thermal expansion valve and an evaporator, and wherein said compressor includes: a housing forming a working space comprising two intersecting bores, a suction port at one end of said housing opening to said space for admitting gaseous refrigerant as the compressor working fluid to said space and a discharge port within said housing opening to said space for discharging said working fluid after compression, two intermeshing helical screw rotors rotatably mounted in said bores and defining with said bores working chambers wherein said compressor discharges compressed refrigerant gas into said condenser of the refrigeration system and a shiftable slide valve for returning a variable portion of the working fluid back to the suction port prior to compression, and said system includes means for bleeding off a portion of said refrigerant in liquid form from said condenser and an injection port opening up into said working space for injecting bled liquid refriger

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Thermal Sciences (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
  • Rotary-Type Compressors (AREA)
US433418A 1974-01-14 1974-01-14 Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use Expired - Lifetime US3885402A (en)

Priority Applications (11)

Application Number Priority Date Filing Date Title
US433418A US3885402A (en) 1974-01-14 1974-01-14 Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use
CA201,635A CA1001432A (en) 1974-01-14 1974-06-04 Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use
SE7407434A SE404952B (sv) 1974-01-14 1974-06-06 Skruvrotorkompressor
GB2523174A GB1464996A (en) 1974-01-14 1974-06-06 Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use
AU69960/74A AU479659B2 (en) 1974-01-14 1974-06-10 Optimized point of injection of liquid refrigerant in a helical screw rotary compressot for refrigeration use
ZA00743672A ZA743672B (en) 1974-01-14 1974-06-10 Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use
FR7422738A FR2257880B3 (it) 1974-01-14 1974-06-28
IT69180/74A IT1014495B (it) 1974-01-14 1974-07-09 Compressore rotativo
JP49079773A JPS50102909A (it) 1974-01-14 1974-07-13
DE2434873A DE2434873A1 (de) 1974-01-14 1974-07-19 Rotationsschraubenverdichter mit optimal angeordneter stelle zur injektion eines fluessigen kuehlmittels
BR7762/74A BR7407762D0 (pt) 1974-01-14 1974-09-18 Aperfeicoamento em compressor rotativo parafuso helicoida

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US433418A US3885402A (en) 1974-01-14 1974-01-14 Optimized point of injection of liquid refrigerant in a helical screw rotary compressor for refrigeration use

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JP (1) JPS50102909A (it)
BR (1) BR7407762D0 (it)
CA (1) CA1001432A (it)
DE (1) DE2434873A1 (it)
FR (1) FR2257880B3 (it)
GB (1) GB1464996A (it)
IT (1) IT1014495B (it)
SE (1) SE404952B (it)
ZA (1) ZA743672B (it)

Cited By (20)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4020642A (en) * 1973-11-19 1977-05-03 Hall-Thermotank Products Limited Compression systems and compressors
US4062199A (en) * 1975-06-24 1977-12-13 Kabushiki Kaisha Maekawa Seisakusho Refrigerating apparatus
US4234296A (en) * 1978-06-14 1980-11-18 Hitachi, Ltd. Screw compressor
US4261180A (en) * 1978-01-06 1981-04-14 Hitachi, Ltd. Refrigerator
US4375156A (en) * 1980-10-03 1983-03-01 Dunham-Bush, Inc. Closed loop compressed gas system with oil mist lubricated screw compressor
US4515540A (en) * 1983-11-22 1985-05-07 Frick Company Variable liquid refrigerant injection port locator for screw compressor equipped with automatic variable volume ratio
US4553911A (en) * 1983-11-22 1985-11-19 Frick Company Method of coding the oil in screw compressors equipped with automatic variable volume ratio
US4890461A (en) * 1987-07-21 1990-01-02 Bernard Zimmern Hermetic or semi-hermetic refrigeration motor-compressor unit
US4974427A (en) * 1989-10-17 1990-12-04 Copeland Corporation Compressor system with demand cooling
WO2007106090A1 (en) * 2006-03-13 2007-09-20 Carrier Corporation Slide valve with hot gas bypass port
US20070251256A1 (en) * 2006-03-20 2007-11-01 Pham Hung M Flash tank design and control for heat pumps
US20080078204A1 (en) * 2006-10-02 2008-04-03 Kirill Ignatiev Refrigeration system
US20080152524A1 (en) * 2005-06-29 2008-06-26 Mayekawa Mfg. Co., Ltd. Oil supply method of two-stage screw compressor, two-stage screw compressor applying the method, and method of operating refrigerating machine having the compressor
US20080236179A1 (en) * 2006-10-02 2008-10-02 Kirill Ignatiev Injection system and method for refrigeration system compressor
US7647790B2 (en) 2006-10-02 2010-01-19 Emerson Climate Technologies, Inc. Injection system and method for refrigeration system compressor
US20100086402A1 (en) * 2008-10-07 2010-04-08 Eaton Corporation High efficiency supercharger outlet
US8539785B2 (en) 2009-02-18 2013-09-24 Emerson Climate Technologies, Inc. Condensing unit having fluid injection
US20180023566A1 (en) * 2014-12-23 2018-01-25 Edwards Limited Rotary screw vacuum pumps
US20210100159A1 (en) * 2017-08-01 2021-04-08 Capstan Ag Systems, Inc. Systems and methods for suppressing vaporization of volatile fluids in agricultural fluid application systems
US11293438B2 (en) * 2016-12-15 2022-04-05 Carrier Corporation Screw compressor with magnetic gear

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102012013267A1 (de) 2012-07-04 2014-01-09 Fraunhofer-Gesellschaft zur Förderung der angewandten Forschung e.V. Substrateinrichtung, Konservierungsgerät und Verfahren zur Kryokonservierung einer biologischen Probe

Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3073514A (en) * 1956-11-14 1963-01-15 Svenska Rotor Maskiner Ab Rotary compressors
US3210958A (en) * 1964-09-10 1965-10-12 Gen Electric Heat pump comprising rotary compressor including injection cooling arrangement
US3432089A (en) * 1965-10-12 1969-03-11 Svenska Rotor Maskiner Ab Screw rotor machine for an elastic working medium
US3795117A (en) * 1972-09-01 1974-03-05 Dunham Bush Inc Injection cooling of screw compressors

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3073514A (en) * 1956-11-14 1963-01-15 Svenska Rotor Maskiner Ab Rotary compressors
US3210958A (en) * 1964-09-10 1965-10-12 Gen Electric Heat pump comprising rotary compressor including injection cooling arrangement
US3432089A (en) * 1965-10-12 1969-03-11 Svenska Rotor Maskiner Ab Screw rotor machine for an elastic working medium
US3795117A (en) * 1972-09-01 1974-03-05 Dunham Bush Inc Injection cooling of screw compressors

Cited By (39)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4020642A (en) * 1973-11-19 1977-05-03 Hall-Thermotank Products Limited Compression systems and compressors
US4062199A (en) * 1975-06-24 1977-12-13 Kabushiki Kaisha Maekawa Seisakusho Refrigerating apparatus
US4261180A (en) * 1978-01-06 1981-04-14 Hitachi, Ltd. Refrigerator
US4234296A (en) * 1978-06-14 1980-11-18 Hitachi, Ltd. Screw compressor
US4375156A (en) * 1980-10-03 1983-03-01 Dunham-Bush, Inc. Closed loop compressed gas system with oil mist lubricated screw compressor
US4515540A (en) * 1983-11-22 1985-05-07 Frick Company Variable liquid refrigerant injection port locator for screw compressor equipped with automatic variable volume ratio
US4553911A (en) * 1983-11-22 1985-11-19 Frick Company Method of coding the oil in screw compressors equipped with automatic variable volume ratio
US4890461A (en) * 1987-07-21 1990-01-02 Bernard Zimmern Hermetic or semi-hermetic refrigeration motor-compressor unit
US4974427A (en) * 1989-10-17 1990-12-04 Copeland Corporation Compressor system with demand cooling
US20080152524A1 (en) * 2005-06-29 2008-06-26 Mayekawa Mfg. Co., Ltd. Oil supply method of two-stage screw compressor, two-stage screw compressor applying the method, and method of operating refrigerating machine having the compressor
US8277207B2 (en) 2005-06-29 2012-10-02 Mayekawa Mfg. Co., Ltd. Oil supply method of two-stage screw compressor, two-stage screw compressor applying the method, and method of operating refrigerating machine having the compressor
US7722346B2 (en) * 2005-06-29 2010-05-25 Mayekawa Mfg. Co., Ltd. Oil supply method of two-stage screw compressor, two-stage screw compressor applying the method, and method of operating refrigerating machine having the compressor
US20100089078A1 (en) * 2005-06-29 2010-04-15 Mayekawa Mfg. Co., Ltd. Oil supply method of two-stage screw compressor, two-stage screw compressor applying the method, and method of operating refrigerating machine having the compressor
WO2007106090A1 (en) * 2006-03-13 2007-09-20 Carrier Corporation Slide valve with hot gas bypass port
CN101400889B (zh) * 2006-03-13 2012-10-03 开利公司 具有热气体旁路口的滑阀
US8221104B2 (en) 2006-03-13 2012-07-17 Carrier Corporation Screw compressor having a slide valve with hot gas bypass port
US20100272580A1 (en) * 2006-03-13 2010-10-28 Wilson Francis P Slide valve with hot gas bypass port
US8020402B2 (en) 2006-03-20 2011-09-20 Emerson Climate Technologies, Inc. Flash tank design and control for heat pumps
US8505331B2 (en) 2006-03-20 2013-08-13 Emerson Climate Technologies, Inc. Flash tank design and control for heat pumps
US20070251256A1 (en) * 2006-03-20 2007-11-01 Pham Hung M Flash tank design and control for heat pumps
US20080047292A1 (en) * 2006-03-20 2008-02-28 Emerson Climate Technologies, Inc. Flash tank design and control for heat pumps
US20080047284A1 (en) * 2006-03-20 2008-02-28 Emerson Climate Technologies, Inc. Flash tank design and control for heat pumps
US7827809B2 (en) 2006-03-20 2010-11-09 Emerson Climate Technologies, Inc. Flash tank design and control for heat pumps
US20110139794A1 (en) * 2006-03-20 2011-06-16 Emerson Climate Technologies, Inc. Flash tank design and control for heat pumps
US8181478B2 (en) 2006-10-02 2012-05-22 Emerson Climate Technologies, Inc. Refrigeration system
US8769982B2 (en) 2006-10-02 2014-07-08 Emerson Climate Technologies, Inc. Injection system and method for refrigeration system compressor
US20080078204A1 (en) * 2006-10-02 2008-04-03 Kirill Ignatiev Refrigeration system
US20100095704A1 (en) * 2006-10-02 2010-04-22 Kirill Ignatiev Injection System and Method for Refrigeration System Compressor
US7647790B2 (en) 2006-10-02 2010-01-19 Emerson Climate Technologies, Inc. Injection system and method for refrigeration system compressor
US20080236179A1 (en) * 2006-10-02 2008-10-02 Kirill Ignatiev Injection system and method for refrigeration system compressor
US20100086402A1 (en) * 2008-10-07 2010-04-08 Eaton Corporation High efficiency supercharger outlet
US8096288B2 (en) 2008-10-07 2012-01-17 Eaton Corporation High efficiency supercharger outlet
US8539785B2 (en) 2009-02-18 2013-09-24 Emerson Climate Technologies, Inc. Condensing unit having fluid injection
US9494356B2 (en) 2009-02-18 2016-11-15 Emerson Climate Technologies, Inc. Condensing unit having fluid injection
US20180023566A1 (en) * 2014-12-23 2018-01-25 Edwards Limited Rotary screw vacuum pumps
US10533552B2 (en) * 2014-12-23 2020-01-14 Edwards Limited Rotary screw vacuum pumps
US11293438B2 (en) * 2016-12-15 2022-04-05 Carrier Corporation Screw compressor with magnetic gear
US20210100159A1 (en) * 2017-08-01 2021-04-08 Capstan Ag Systems, Inc. Systems and methods for suppressing vaporization of volatile fluids in agricultural fluid application systems
US11672197B2 (en) * 2017-08-01 2023-06-13 Capstan Ag Systems, Inc. Systems and methods for suppressing vaporization of volatile fluids in agricultural fluid application systems

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IT1014495B (it) 1977-04-20
BR7407762D0 (pt) 1975-09-16
GB1464996A (en) 1977-02-16
CA1001432A (en) 1976-12-14
ZA743672B (en) 1975-06-25
SE7407434L (it) 1975-07-15
DE2434873A1 (de) 1975-07-17
JPS50102909A (it) 1975-08-14
SE404952B (sv) 1978-11-06
AU6996074A (en) 1975-12-11
FR2257880A1 (it) 1975-08-08
FR2257880B3 (it) 1977-05-06

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