EP0374688A2 - Système réfrigérateur à deux évaporateurs pour réfrigérateurs ménagers - Google Patents

Système réfrigérateur à deux évaporateurs pour réfrigérateurs ménagers Download PDF

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Publication number
EP0374688A2
EP0374688A2 EP89122896A EP89122896A EP0374688A2 EP 0374688 A2 EP0374688 A2 EP 0374688A2 EP 89122896 A EP89122896 A EP 89122896A EP 89122896 A EP89122896 A EP 89122896A EP 0374688 A2 EP0374688 A2 EP 0374688A2
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EP
European Patent Office
Prior art keywords
compressor
refrigerant
gas
evaporator
expansion valve
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Granted
Application number
EP89122896A
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German (de)
English (en)
Other versions
EP0374688A3 (fr
EP0374688B1 (fr
Inventor
Heinz Jaster
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
General Electric Co
Original Assignee
General Electric Co
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Filing date
Publication date
Application filed by General Electric Co filed Critical General Electric Co
Publication of EP0374688A2 publication Critical patent/EP0374688A2/fr
Publication of EP0374688A3 publication Critical patent/EP0374688A3/fr
Application granted granted Critical
Publication of EP0374688B1 publication Critical patent/EP0374688B1/fr
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

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Classifications

    • 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
    • F25B1/10—Compression machines, plants or systems with non-reversible cycle with multi-stage compression
    • 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
    • F25B5/00—Compression machines, plants or systems, with several evaporator circuits, e.g. for varying refrigerating capacity
    • F25B5/04—Compression machines, plants or systems, with several evaporator circuits, e.g. for varying refrigerating capacity arranged in series
    • 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
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/13—Economisers
    • 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
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/23—Separators

Definitions

  • the present invention relates to household refrigerators operating with a vapor compression cycle and more particularly, to refrigerators with a two stage compressor.
  • the cycle includes a compressor , condensor , expansion valve , evaporator , and a two phase refrigerant.
  • a capillary tube acts as an expansion valve.
  • the capillary tube is placed in close proximity with the suction line of the compressor to cool the capillary tube.
  • the sub­cooling which occurs to the refrigerant in the capillary tube increases the cooling capacity per unit mass flow rate in the system thereby increasing system efficiency which more than compensates for the disadvantage of increasing the temperature of the gas supplied to the compressor.
  • the evaporator in Fig. 1 operates at approximately -10°F.
  • Refrigerator air is blown across the evaporator and the air flow is controlled so that part of the air flow goes to the freezer compartment and the remainder of the flow goes to the fresh food compartment.
  • the refrigerator cycle therefore, produces its refrigeration effect at a temperature which is appropriate for the freezer, but lower than it needs to be for the fresh food compartment. Since the mechanical energy required to produce cooling at low temperatures is greater than it is at higher temperatures, the simple vapor compres­sion cycle uses more mechanical energy than one which pro­duces cooling at two temperature levels.
  • a well known procedure to reduce mechanical energy use is to operate two independent refrigeration cycles, one to serve the freezer at low temperatures and one to serve the fresh food compartment at an intermediate temperature. Such a system, however, is very costly.
  • a refrigerator system suitable for use in a household refrigerator having a freezer compartment a fresh food compartment.
  • the refrigerator system includes a first expansion valve, a first evaporator for providing cooling to the freezer compartment, a first compressor, a second compressor, a condensor, a second expansion valve, and a second evaporator providing cooling to the fresh food compartment. All the above elements are connected together in series in that order, in a refrigerant flow relationship.
  • a phase separator connects the second evaporator to the first expansion valve in a refrigerator flow relationship and the phase separator provides intercooling between the first and second compressors.
  • a phase separator 27 shown in cross section in Figure 3 comprises a closed receptacle 31 having at the upper portion an inlet 33 for admitting liquid and gaseous phase refrigerant and having two outlets 35 and 37.
  • a screen 44 is located in the upper portion of the receptacle to remove any solid material carried along with the refrigerant when entering the inlet 33.
  • the first outlet 35 is located at the bottom of the receptacle 31 and provides liquid refrigerant 39.
  • the second outlet 37 is provided by a conduit which extends from the interior of the upper portion of the receptacle to the exterior. The conduit is in flow communication with the upper portion and is arranged so that liquid refrigerant entering the upper portion of the receptacle through inlet 33 cannot enter the open end of the conduit.
  • Two phase refrigerant from the outlet of the second evaporator 25 is connected to the inlet 33 of the phase separator 27.
  • the phase separator provides liquid refrigerant to the first expansion valve 11.
  • the phase separator also provides saturated refrigerant vapor which combines with vapor output by the first compressor 15 and together are connected to the inlet of the second compressor 17.
  • the first evaporator 13 contains refrigerant at a temperature of approximately -10°F for cooling the freezer compartment.
  • the second evaporator 25 contains the refrigerant at a temperature of approximately 25°F for cooling the fresh food compartment.
  • the first expansion valve 11 is adjusted to obtain just barely dry gas flow, which can be accomplished, for example, by observing a sight glass located in the conduit 26 between the first evaporator 13 and the first compressor 15.
  • the gas enters the first compressor 15 stage and is compressed.
  • the gas discharged from the first compressor is mixed with gas at the saturation temperature from the phase separator 27 and the two gases are further compressed by the second compressor 17.
  • the high temperature, high pressure discharge gas from the second compressor is condensed in condensor 21 with the expansion valve 23 adjusted to obtain some subcooling of the liquid exiting the condensor. This can be accomplished by observing a sight glass situated between the condensor 21 and the second expansion valve 23.
  • the liquid refrigerant condensed in the condensor 21 passes through the second expansion valve where it expands from the high pressure of the condensor 21 to a lower intermediate pressure in the second evaporator 25.
  • the expansion of the liquid causes part of the liquid to evaporate and cool the remainder to the second evaporator temperature.
  • the liquid and gas phase refrigerant enters the phase separator 27. Liquid refrigerant accumulates in the lower portion of the receptacle and gas accumulates in the upper portion.
  • the phase separator supplies the gas portion to be combined with the gas exiting the first stage compressor 15.
  • the gas from the phase separator is at approximately 25°F and cools the gas exiting from the first stage compressor, thereby lowering the gas temperature entering the second compressor 17 from what it would have otherwise have been without the intercooling.
  • the liquid of the two phase mixture from the second evaporator 25 flows from the phase separator 27 through the first expansion valve 11 causing the refrigerant to a still lower pressure.
  • the remaining liquid evaporates in the first evaporator 13 cooling the evaporator to ap­proximately -10°F.
  • a sufficient refrigerant charge is supplied to the system so that the desired liquid level can be maintained in the phase separator.
  • the pressure ratio of the two compressors is deter­mined by the refrigerant used and the temperatures at which the evaporators are to operate.
  • the pressure at the input to the first compressor 15 is determined by the pressure at which the refrigerant exists in two phase equilibrium at -10°F.
  • the pressure at the output of the first compressor is determined by the saturation pressure of the refrigerant at 25°F.
  • the temperature of the condensor 21 has to be greater than that of the ambient temperature in order to function as a heat exchanger under a wide range of operating conditions. If the condensor is to operate at 105°F, for example, then the pressure of the refrigerant at saturation can be determined.
  • the volume displacement capability of the compressors are determined by the amount of cooling capacity the system requires at each of the two temperature levels, which determines the mass flow rate of the refrigerant through the compressors.
  • the dual evaporator two-stage cycle requires less mechanical energy compared to a single evaporator single com­pressor cycle with the same cooling capacity.
  • the efficiency advantages come about due to the fact that the gas leaving the higher temperature evaporator is compressed from an intermediate pressure, rather than from the lower pressure of the gas leaving the lower temperature evaporator.
  • Also contributing to improved efficiency is the cooling of the gas exiting the first compressor by the addition of gas cooled to saturation temperature from the phase separator. The cooling of the gas entering the second compressor reduces the mechanical energy requirement of the second compressor.
  • FIG. 4 Another embodiment of the present invention is shown in Figure 4.
  • the system comprises a first expansion valve 51, a first evaporator 53, and a first compressor 55, all of which are connected in series in that order in a refrigerant flow relationship by conduit 57.
  • a second compressor 61, a condensor 63, a second expansion valve 65, and a second evaporator 67, are connected in series in that order, in a refrigerant flow relationship by conduit 69.
  • a phase separator 71 shown in cross section in Figure 5, comprises a closed receptacle 73 having at the upper portion a first inlet 75 for admitting liquid and gaseous phase refrigerant, a second inlet 77 for introducing gas refrigerant below a liquid level 81 in the lower portion of the receptacle and two outlets 83 and 85.
  • a screen 87 is located in the upper portion of the receptacle to remove any solid material carried along with the refrigerant when entering the first inlet.
  • the first outlet 83 is located at the bottom of the receptacle and provides liquid refrigerant.
  • the second outlet 85 is provided by a conduit located in the upper portion of the receptacle and is arranged so that liquid refrigerant entering the first inlet cannot enter the open end of the conduit 85.
  • Two phase refrigerant from the outlet of the second evaporator 67 enters the first inlet 75 of the phase separator.
  • the phase separator provides liquid refrigerant to the first expansion valve 52 from outlet 83 of the phase separator.
  • the discharge gas refrigerant from the first compressor 55 is introduced into the receptacle 75 through the second inlet 77 where it mixes with the liquid refrigerant.
  • the second outlet 85 delivers gas at the saturation temperature of the liquid to the second compressor 61.
  • the first evaporator 53 contains refrigerant at a temperature of approximately -10°F for cooling the freezer compartment.
  • the second evaporator 67 contains refrigerant at a temperature of approximately 25°F for cooling the fresh food compartment.
  • the first expansion valve 51 is adjusted to obtain just barely dry gas flow such as by observing a sight glass installed in the conduit 57 between the evaporator 53 and the compressor 55.
  • the gas enters the first compressor stage 55 and is compressed.
  • the gas discharged from the first compressor is mixed with and is in direct contact with liquid refrigerant in the phase separator 71, reducing the gas temperature to the saturation temperature. Some of the liquid refrigerant is evaporated by the gas entering the second inlet.
  • the liquid refrigerant that evaporates cools the incoming gas from the first compressor 55 to the saturation gas temperature. Saturated gas from the upper portion of the phase separator flows into the inlet of the second compressor 61.
  • the high temperature, high pressure gas discharged by the second compressor 61 is condensed in a condensor 63 with the throttling adjusted by the second expansion valve 65 to obtain some subcooling. This can be accomplished, for example, by observing the sight glass situated between the condensor 63 and the second evaporator.
  • the liquid refrigerant condensed in condensor passes through the second expansion valve 65 where it expands from the high pressure in the condensor to a lower intermediate pressure in the second evaporator 67.
  • the expansion of the liquid causes part of the liquid to evaporate and cools the remainder to the second evaporator temperature.
  • the liquid and gas phase refrigerant enters the phase separator 71.
  • the liquid accumulates in the lower portion of the receptacle and the gas in the upper portion.
  • Liquid refrigerant from the phase separator flows through the first expansion valve 51 causing the refrigerant to expand to a still lower pressure.
  • the remaining liquid evaporates in the evaporator cooling the evaporator to approximately -10°F.
  • a sufficient refrigerant charge is supplied so that the desired liquid level can be maintained in the phase separator 71.
  • the dual evaporator two stage cycle requires 29% less mechanical energy compared to a single evaporator single compressor cycle with the same cooling capacity.
  • the effi­ciency advantages come about due to the fact that the gas leaving the higher temperature evaporator 67 (second evaporator) is compressed from an intermediate pressure rather than from the lower pressure of the gas leaving the lower temperature evaporator 53. Also contributing to the improved efficiency is the cooling of the gas leaving the first compressor 55 to the saturation temperature, before compression to the system's high pressure in the second compressor 61.
  • the cycle shown in Figure 2 is calculated to be more efficient than the cycle in Figure 4 by approximately 2%.
  • Figure 2 While the arrangement of Figure 2 results in a higher gas inlet temperature to the second compressor 61 and thereby requires greater compression work, the cycle of Figure 2 makes available more liquid at intermediate pressure for expansion to the low temperature evaporator 53 (first evap­orator) thereby increasing the cycle's efficiency.
  • Figure 2 has a higher inlet temperature to the second compressor since not all the gas supplied to the second compressor is cooled to the saturation temperature as is done in the cycle of Figure 4.
  • the compressors can be of the reciprocating type with hermetically sealed motors or of the rotary type with hermetically sealed motors or of any positive displacement type with hermetically sealed motors.
  • the first compressor when refrigerant R-12 is used can be very small and operates against a pressure ratio of only 2, which could allow the use of, for example, an inexpensive diaphragm compressor. Improved efficiency can be achieved by operating both compressors from a single motor. Since a larger motor can be more efficient than two smaller motors providing the same total power.
  • the evaporator exit saturation temperature was as­sumed to be -10°F, and have a pressure drop of 1 psi and an exit superheat of 0°.
  • the compressor adiabatic efficiency was assumed to be 0.61, motor efficiency 0.8 and additional heating of suction gas due to heat transfer from the compressor shell 43°F.
  • the capillary tube heat transfer to the suction line of the compressor results in suction gas of heating of 98°F.
  • the condensor entrance saturation temperature is assumed to be 130°F, the pressure drop 10 psi, and exit subcooling 5°F.
  • the motor discharge tem­perature is calculated to be 429°F, refrigerant flow rate 18.6 1bm/hr, compressor power 270 Watts and the coefficient of performance 1.09.
  • the first evapo­rator was assumed to have an exit saturation temperature of -10°F, with a pressure drop of 1 psi and an exit superheat of 0°F.
  • the second evaporator is assumed to have an exit tem­perature of 25°F and 0 psi pressure drop.
  • the first and second compressor have an adiabatic efficiency of 0.7 and a motor efficiency of 0.8.
  • the first compressor produces an additional superheating of suction gas due to heat transfer from the compressor shell of 5°F.
  • the second compressor has an additional superheating of suction gas of 10°F.
  • the condensor has an entrance saturation temperature of 130°F, a pressure drop of 10 psi and an exit subcooling of 5°F.
  • the cooling capacity of 1000 Btu/hr is divided equally between two evaporators.
  • the computed results from the above parameters for the cycle in Figure 2 are a second compressor discharge gas temperature of 208°F and a first stage compressor discharge gas temperature of 66°F.
  • the compressor flow rates of the first and second compressors are 8.0 lbm/hr and 24.7 lbm/hr, respectively.
  • the first and second compressor power consump­tions are 22.2 and 164 watts, respectively.
  • the coefficient of performance is 1.58.
  • the computed results for the cycle of Figure 4 us­ing the above parameters are a first and second compressor discharge gas temperature of 66 and 208°F, respectively.
  • the coefficient of performance was calculated to be 1.64.
  • the system of Figure 4 can be modified by changing the operation of the phase separator of Figure 5. If the second inlet 77 is connected to the conduit of the second outlet 85, then the gas from the outlet of the first compressor would not be in direct contact with the liquid refrigerant but would still be cooled, although not to the saturation temperature.
  • the phase separator would provide intercooling between the two compressors, operating as heat exchanger, but not as much cooling as when the gas is in direct contact with the liquid refrigerant.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Devices That Are Associated With Refrigeration Equipment (AREA)
EP19890122896 1988-12-23 1989-12-12 Système réfrigérateur à deux évaporateurs pour réfrigérateurs ménagers Expired - Lifetime EP0374688B1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
US28884888A 1988-12-23 1988-12-23
US288848 1988-12-23

Publications (3)

Publication Number Publication Date
EP0374688A2 true EP0374688A2 (fr) 1990-06-27
EP0374688A3 EP0374688A3 (fr) 1991-05-08
EP0374688B1 EP0374688B1 (fr) 1996-05-22

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Family Applications (1)

Application Number Title Priority Date Filing Date
EP19890122896 Expired - Lifetime EP0374688B1 (fr) 1988-12-23 1989-12-12 Système réfrigérateur à deux évaporateurs pour réfrigérateurs ménagers

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EP (1) EP0374688B1 (fr)
JP (1) JP2766356B2 (fr)
DE (1) DE68926533T2 (fr)

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2446062A (en) * 2007-01-26 2008-07-30 Grasso Gmbh Refrigeration Tech Carbon dioxide refrigeration system with compressors in two-stage arrangement
WO2024022500A1 (fr) * 2022-07-29 2024-02-01 山前（珠海）医疗科技有限公司 Équipement de stockage et procédé de réfrigération associé
WO2024022501A1 (fr) * 2022-07-29 2024-02-01 山前（珠海）医疗科技有限公司 Équipement de réfrigération et procédé de réfrigération associé

Families Citing this family (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
KR100609169B1 (ko) * 2004-11-02 2006-08-02 엘지전자 주식회사 캐스캐이드 냉동사이클

Citations (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4435962A (en) * 1980-06-20 1984-03-13 Shin Meiwa Industry Co., Ltd. Refrigerating apparatus
US4745777A (en) * 1986-03-31 1988-05-24 Mitsubishi Denki Kabushiki Kaisha Refrigerating cycle apparatus

Patent Citations (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4435962A (en) * 1980-06-20 1984-03-13 Shin Meiwa Industry Co., Ltd. Refrigerating apparatus
US4745777A (en) * 1986-03-31 1988-05-24 Mitsubishi Denki Kabushiki Kaisha Refrigerating cycle apparatus

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
E.H.M. B[CKSTR\M et al.: "K{ltetechnik", 3rd edition, 1965, pages 112-115,293-294, Verlag G. Braun, Karlsruhe, DE *

Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2446062A (en) * 2007-01-26 2008-07-30 Grasso Gmbh Refrigeration Tech Carbon dioxide refrigeration system with compressors in two-stage arrangement
GB2446062B (en) * 2007-01-26 2011-10-12 Grasso Gmbh Refrigeration Technology CO2 refrigeration system with compressors in two-stage arrangement
WO2024022500A1 (fr) * 2022-07-29 2024-02-01 山前（珠海）医疗科技有限公司 Équipement de stockage et procédé de réfrigération associé
WO2024022501A1 (fr) * 2022-07-29 2024-02-01 山前（珠海）医疗科技有限公司 Équipement de réfrigération et procédé de réfrigération associé

Also Published As

Publication number Publication date
JPH02225953A (ja) 1990-09-07
JP2766356B2 (ja) 1998-06-18
DE68926533D1 (de) 1996-06-27
DE68926533T2 (de) 1997-01-23
EP0374688A3 (fr) 1991-05-08
EP0374688B1 (fr) 1996-05-22

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