EP0184181B1 - Wärmepumpe - Google Patents

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Publication number
EP0184181B1
EP0184181B1 EP85115297A EP85115297A EP0184181B1 EP 0184181 B1 EP0184181 B1 EP 0184181B1 EP 85115297 A EP85115297 A EP 85115297A EP 85115297 A EP85115297 A EP 85115297A EP 0184181 B1 EP0184181 B1 EP 0184181B1
Authority
EP
European Patent Office
Prior art keywords
pressure
heat
stages
working medium
condenser
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.)
Expired - Lifetime
Application number
EP85115297A
Other languages
German (de)
English (en)
French (fr)
Other versions
EP0184181A2 (de
EP0184181A3 (en
Inventor
Arpád Dr. Bakay
György Bergmann
Géza Hivessy
Istvan Dr. Szentgyörgyi
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.)
Energiagazdalkodasi Intezet
Original Assignee
Energiagazdalkodasi Intezet
Priority date (The priority date 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 date listed.)
Filing date
Publication date
Application filed by Energiagazdalkodasi Intezet filed Critical Energiagazdalkodasi Intezet
Priority to AT85115297T priority Critical patent/ATE57763T1/de
Publication of EP0184181A2 publication Critical patent/EP0184181A2/de
Publication of EP0184181A3 publication Critical patent/EP0184181A3/de
Application granted granted Critical
Publication of EP0184181B1 publication Critical patent/EP0184181B1/de
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
    • F25B11/00—Compression machines, plants or systems, using turbines, e.g. gas turbines
    • F25B11/02—Compression machines, plants or systems, using turbines, e.g. gas turbines as expanders
    • 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
    • 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
    • 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
    • F25B9/00—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • F25B9/006—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant the refrigerant containing more than one component

Definitions

  • the invention relates to a heat pump, the working medium of which consists of the mixture of readily soluble media with different boiling points and in which the condensation and evaporation take place at a changing temperature.
  • the heat source is the medium designated by reference number 2, which can be cooled from a temperature T 2 'to a temperature T 2 ".
  • the task of the heat pump is to change the medium labeled 1 from a temperature T 1 ' to a temperature Temperature T j "to warm up. These changes in state of the two media are shown by continuous lines.
  • the power factor of the heat pump i.e. the quotient of the useful heat and the mechanical work used, can be expressed as follows: The power factor can be increased if the necessary mechanical work, ie the area enclosed by the cycle, can be reduced.
  • the theoretically optimal heat pump cycle would actually be the cycle represented by the chain-dotted line, which completely conforms to the curve of the temperature profile of the heat-emitting medium.
  • AECF heat is taken up with variable temperature on the route AE, is entropic compression on the route EC, heat emission with variable temperature on the route CF and isentropic expansion on the route FA.
  • the working medium can only absorb a quantity of heat from the medium 2 if its temperature is lower than that of the latter, that is, the curve path AE runs under the curve of the medium 2. But if the heat capacity of the two media is the same and the heat exchange surface is infinitely large, then it goes to heat Transfer necessary temperature difference back to an infinitely low value, that is, the curve section AE nestles against the curve of the medium 2. Similarly, it can be seen that, under the theoretical conditions mentioned, the path CF of the cyclic process conforms to the curve of the medium 1 from above.
  • a heat transfer medium from a single component is always used in the units (evaporators, condensers) of the conventional heat pumps (with compressors or absorption heat pumps), so that the evaporation and condensation always run at a constant temperature, i.e. the actual circular processes correspond to a certain extent to the circular process designated by dashed lines in FIG. 1.
  • the power factor can also be improved in those heat pumps whose working medium contains only one component; however, this requires several stages.
  • Fig. 2 the theoretical operation of a three-stage heat pump is shown on a TS diagram.
  • the cooling of the medium 2 and the heating of the medium 1 are indicated here by continuous lines. From this figure, 'it can readily be seen that the working surface of the three stages shown by dotted lines (the joint surface of the cycle processes AX'Y'Z', W "X" Y “Z” and W "'X”' CZ "' ) is smaller than that of the one-step cycle ABCD and much better than the latter to get to the theoretically optimal cycle AECF.
  • a circular process with a heat transfer that has a variable temperature sequence is most advantageously implemented by the previously known technical solutions by the so-called hybrid heat pump described in EP-B 0 021 205.
  • the circuit of the hybrid heat pump shown in Fig. 3 is reminiscent of the conventional heat pumps with compressor, differs from them in that a working medium is circulated from two components that can be easily detached from one another in the entire cycle.
  • the media pair does not evaporate completely, but a mixture of a vapor which is rich in the medium with a low boiling point and a liquid which is poor in the medium with the low boiling point occurs. out, and this mixture enters the compressor 3.
  • the compressor transfers this two-phase working medium from two components in the form of a so-called "wet compression" to a higher pressure level. From here, the vapor and the liquid phase reach a condenser (absorber) 4, where the vapor rich in the medium with a lower boiling point condenses and gradually dissolves in the accompanying liquid phase.
  • the working medium returns to the evaporator (degasser) 6 via an expansion valve 5. With the help of an internal heat exchanger 7, the power factor of the cycle can be improved.
  • FIG. 4 The actual sequence of the above cycle is shown in a T-S diagram by FIG. 4.
  • the letters designating the individual states correspond to the designations of FIG. 3.
  • the internal heat exchanger is not shown and isentropic expansion or compression is assumed.
  • the temperature change of the working medium is AT 2 in the evaporator (section AB) and LlT1 in the condenser (section CD). These two values are almost the same. This results from the peculiarity of the working media consisting of two components (from a solution) that, in the TS diagram of a given concentration, the curves for constant pressures are approximately parallel.
  • the hybrid heat pump can only work with a really favorable power factor if the temperature change of the heat-emitting and the heat-absorbing medium is almost the same, and the temperature change of the working medium in the evaporator and in the condenser is adapted to these temperature changes.
  • the heat source is waste heat with a low temperature level, e.g. a waste water of 30 ° C or a heated cooling water that can be cooled to a maximum of + 5 ° C without risk of freezing, i.e. the temperature change is 25 ° C.
  • the task is to produce domestic hot water with a temperature of 85 ° C from the tap water available at 15 ° C for the purposes of the food industry.
  • the temperature change is 70 ° C, i.e. several times the other value.
  • Fig. 6 the temperature course of the media 1 and 2 is indicated by continuous lines.
  • the figure shows ideal circular processes (isentropic compression and expansion, infinitely large heat exchange surfaces).
  • the Carnot process with a dashed line and the theoretical cycle of the hybrid heat pump with a dash-dotted line are shown, the latter being adapted to the medium 2. It can be clearly seen from the figure that the area enclosed by the cyclic process of variable temperature and thus the necessary mechanical work are considerably less than in the Carnot process, but much larger than the theoretically necessary minimum work. This deficiency cannot be remedied either by adapting the Kries process to the medium 1 or by using an intermediate variant.
  • the object of the invention is such a further development of the hybrid heat pump, which makes it possible, independently of one another, to adapt the temperature flow of the evaporator and the condenser between very wide limits to the temperature flow of the heat-emitting or heat-absorbing medium, so that the theoretically largest possible power factor is approximated to the maximum can be.
  • the task is at a heat pump with a compressor, an evaporator, a condenser and a pressure reducer and pipelines connecting these units, the working medium of the heat pump being in heat exchange in the evaporator and the condenser with external heat transfer media while achieving condensation or evaporation at variable temperature consists of a mixture of easily soluble media with different boiling points, the compressor being designed as a unit having multiple suction and / or pressure ports, the port of which has several pressure stages for simultaneous suction at more than one pressure level and / or for delivery tion to more than one pressure level, and the evaporator and / or the condenser is multi-stage, the number of pressure stages of the evaporator being equal to the number of suction-side pressure levels, and the number of pressure stages of the condenser being the number of pressure-side pressure levels, according to the invention solved in that under Broadening the temperature range of the condensation and / or evaporation of the working medium with regard to the external heat transfer media, the evapor
  • pressure-reducing elements e.g. Expansion valves are installed in such a way that a pressure-reducing element is arranged between each two adjacent pressure stages when the pressure stages of the compressor are arranged successively according to the level of the pressure levels.
  • the multiple inlet and / or outlet ports are designed so that the turbine in accordance with the number of pressure stages of the compressor for receiving or discharging the working medium at several pressure levels simultaneously is capable.
  • an internal heat exchanger is installed for heat exchange between the media emerging from the condenser and the evaporator.
  • the heat pump works with a two-component working medium that evaporates and condenses at variable temperatures, at least the condenser and / or the evaporator working at more than one pressure level p 3 , p 4 , p 5 , which causes the temperature change of the working medium can be influenced as required.
  • An example of this is shown in FIG. 8.
  • the working medium exits the compressor 3 at three different pressure levels and a separate condenser is assigned to each outlet pressure level, so that the heat-absorbing medium 1 is heated in the three condensers 4a, 4b, 4c, at three different pressures.
  • the working medium from the condensers enters an expansion turbine 8 at three correspondingly different pressure levels and is fed therefrom at two different pressure levels to the two evaporators 6a and 6b, which are heated by the heat-emitting medium 2 and from which the working medium at two accordingly different inlet pressure levels is passed back into the compressor 3.
  • FIGS. 8 and 9 shows this cycle in a T-S diagram in the case of isentropic compression and expansion.
  • the temperature changes of media 1 and 2 are shown separately for infinitely large heat exchange surfaces on the right side of the figure.
  • the condenser and the evaporator in FIGS. 8 and 9 only have, for example, three or two pressure stages, since the number of pressure stages can be determined as required.
  • the actual switching of the heat pump according to the invention is more complicated, namely it preferably also contains internal heat exchangers 7 e.g. 10.
  • the expansion turbine 8 is only economical in very large systems, so that pressure-reducing elements (e.g. throttle valves) are generally used instead of these turbines.
  • pressure-reducing elements e.g. throttle valves
  • FIG. 10 Such an embodiment is shown in FIG. 10.
  • the condenser unit has three pressure stages, similar to the previous example, while the evaporator unit has two pressure stages. If necessary, a different number of pressure levels can also be selected.
  • the working medium passes from the compressor 3 with three different pressure levels p 3 , p 4 , p 5 into the three condensers 4a, 4b, 4c, where the heat-absorbing medium 1 is heated by the working medium.
  • Internal heat exchangers 7a, 7b, 7c are connected downstream of the condensers, where the working medium cools further under high pressure and transfers heat to the working medium with low pressure.
  • the outputs of the inner heat exchangers 7a, 7b, 7c are brought together with the interposition of one of two expansion valves 5c, 5d, which are followed by two further expansion valves 5a, 5b.
  • the pressure of the working medium is gradually reduced to the required level in the four expansion valves 5a, 5b, 5c, 5d, after which the working medium enters two of two evaporators 6a, 6b with two different pressure levels.
  • the evaporators 6a, 6b are heated by the heat-emitting medium 2.
  • the here heated and partially evaporated working medium continues to heat up in the inner heat exchangers 7a, 7b, 7c, two of which are connected in series to the one evaporator 6a and one to the other evaporator 6b, then it reappears at corresponding pressure levels p 1 and pp the compressor 3 a.
  • FIG. 11a If the construction of the compressor 3 is not suitable for having suction or pressure ports at different pressure levels, several compressors can also be provided according to FIG. 11a.
  • five compressors 3a, 3b, 3c, 3d, 3e are expediently installed in a row on a common axis, the common axis not being an essential condition.
  • the working medium enters the two first compressors 3a and 3b with two different pressures and with three different pressures from the three last compressors 3c, 3d and bw. 3e off.
  • the suction pressure p 2 is somewhat greater than the pressure p 3 on the pressure side.
  • the circuit of the internal heat exchangers 7a, 7b, 7c, in Fig. 10 is such that the working medium emerging from the evaporator with a pressure P2 from a liquid with a pressure p 5 , and the medium with a pressure p 1 from the liquids with the pressures p 3 and p 4 is heated.
  • the circuit shown in the figure is optimal for certain values of the media flows and the pressures. However, there may also be cases in which a circuit deviating from the figure is associated with a greater thermodynamic advantage, for example if the mass flows and the pressure levels are distributed differently among the individual condensers and evaporators, as a result of which the temperature sequences are also different.
  • FIG. 11c such a case is presented in FIG. 11c, in which the medium emerging from the evaporator 6a with the pressure p 1 in the inner heat exchanger 7a from a liquid with the pressure p 3 , and the medium with the pressure P2 in the inner heat exchangers is heated from media with the pressures P4 and p 5 7b and 7c.
  • FIG. 11 d it can also happen that it is worthwhile to divide the heat given off by the condensate with the pressure P4 under the media with the pressures p 1 and P2 , as can be seen from FIG. 11 d.
  • the medium with the pressure P4 is branched onto the internal heat exchangers 7b and 7c emitting its heat, so that they are connected in parallel, but such a case is also possible, for which it is more favorable to connect the inner heat exchangers 7b and 7c in series along the flow path of the medium with the pressure p 3 .
  • FIG. 12 A special case of realizing the inventive concept is shown in Fig. 12, wherein only the condenser operates at three pressure stages 4a, 4b, 4c and only one evaporator 6 is provided, i.e. the compressor only draws in at a single pressure level and supplies working media with three different pressure levels. This is necessary if the temperature change of the heat-absorbing medium is significantly greater than that of the heat-emitting medium.
  • FIG. 13 An opposite case can be seen from FIG. 13, according to which only one condenser stage 4 and three evaporator stages 6a, 6b, 6c are provided.
  • 10 shows the general solution of the task according to the invention, according to which the number of stages of the condensers and evaporators differs from one another. In a special case, this number of steps can also be the same, e.g. two pressure stages on the compressor 3 (i.e. two evaporator stages) and two pressure stages on the pressure side (i.e. two condenser stages).
  • the solution according to the invention can be traced back to the series connection of two independent cycle processes of the hybrid heat pump will.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Central Heating Systems (AREA)
  • Sorption Type Refrigeration Machines (AREA)
  • Compression-Type Refrigeration Machines With Reversible Cycles (AREA)
  • Control Of The Air-Fuel Ratio Of Carburetors (AREA)
  • Engine Equipment That Uses Special Cycles (AREA)
  • Heat-Pump Type And Storage Water Heaters (AREA)
EP85115297A 1984-12-03 1985-12-03 Wärmepumpe Expired - Lifetime EP0184181B1 (de)

Priority Applications (1)

Application Number Priority Date Filing Date Title
AT85115297T ATE57763T1 (de) 1984-12-03 1985-12-03 Waermepumpe.

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
HU844461A HU198328B (en) 1984-12-03 1984-12-03 Method for multiple-stage operating hibrid (compression-absorption) heat pumps or coolers
HU446184 1984-12-03

Publications (3)

Publication Number Publication Date
EP0184181A2 EP0184181A2 (de) 1986-06-11
EP0184181A3 EP0184181A3 (en) 1988-01-13
EP0184181B1 true EP0184181B1 (de) 1990-10-24

Family

ID=10968033

Family Applications (1)

Application Number Title Priority Date Filing Date
EP85115297A Expired - Lifetime EP0184181B1 (de) 1984-12-03 1985-12-03 Wärmepumpe

Country Status (9)

Country Link
US (1) US4688397A (no)
EP (1) EP0184181B1 (no)
JP (1) JPS61180861A (no)
AT (1) ATE57763T1 (no)
CA (1) CA1262057A (no)
DE (1) DE3580249D1 (no)
DK (1) DK161482C (no)
HU (1) HU198328B (no)
NO (1) NO164738C (no)

Families Citing this family (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
HU198329B (en) * 1986-05-23 1989-09-28 Energiagazdalkodasi Intezet Method and apparatus for increasing the power factor of compression hybrid refrigerators or heat pumps operating by solution circuit
HU210994B (en) * 1990-02-27 1995-09-28 Energiagazdalkodasi Intezet Heat-exchanging device particularly for hybrid heat pump operated by working medium of non-azeotropic mixtures
DE102014213542A1 (de) * 2014-07-11 2016-01-14 Siemens Aktiengesellschaft Verfahren zum Betrieb einer Wärmepumpe mit wenigstens zwei Verdampfern
DE102014213543A1 (de) * 2014-07-11 2016-01-14 Siemens Aktiengesellschaft Verfahren zum Betrieb einer Wärmepumpe mit wenigstens zwei Verflüssigern
WO2019169187A1 (en) * 2018-02-28 2019-09-06 Treau, Inc. Roll diaphragm compressor and low-pressure vapor compression cycles

Family Cites Families (16)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CH97319A (de) * 1921-05-20 1923-01-02 Escher Wyss Maschf Ag Kälteanlage mit Kreiselverdichter und mindestens zwei Verdampfern, die mit verschiedenen Drücken arbeiten.
DE712629C (de) * 1937-09-14 1941-10-22 Karl Glaessel Mehrfach wirkender Kompressor fuer Kaelteanlagen
DE830801C (de) * 1950-07-25 1952-02-07 E H Edmund Altenkirch Dr Ing Kompressions-Kaelteanlage
DE867122C (de) * 1950-08-29 1953-02-16 Edmund Dr-Ing E H Altenkirch Verfahren und Vorrichtung zum Heben der einem Waermetraeger entzogenen Waermemenge niedrigerer Temperatur auf eine hoehere Temperatur
DE1035669B (de) * 1954-08-09 1958-08-07 Frantisek Wergner Verfahren zum Betrieb einer Kompressor-Kuehlanlage mit mindestens zweistufiger Kompression eines in der Anlage umlaufenden Kaeltemittels sowie Kompressor-Kuehlanlage zur Durchfuehrung des Verfahrens
US2952139A (en) * 1957-08-16 1960-09-13 Patrick B Kennedy Refrigeration system especially for very low temperature
GB879809A (en) * 1960-08-03 1961-10-11 Conch Int Methane Ltd Refrigeration system
DE1241468B (de) * 1962-12-01 1967-06-01 Andrija Fuderer Dr Ing Kompressionsverfahren zur Kaelterzeugung
FR1566236A (no) * 1968-01-10 1969-05-09
FR1568871A (no) * 1968-01-18 1969-05-30
US3577742A (en) * 1969-06-13 1971-05-04 Vilter Manufacturing Corp Refrigeration system having a screw compressor with an auxiliary high pressure suction inlet
FR2337855A1 (fr) * 1976-01-07 1977-08-05 Inst Francais Du Petrole Procede de production de chaleur utilisant une pompe de chaleur fonctionnant avec un melange de fluides
HU186726B (en) * 1979-06-08 1985-09-30 Energiagazdalkodasi Intezet Hybrid heat pump
FR2497931A1 (fr) * 1981-01-15 1982-07-16 Inst Francais Du Petrole Procede de chauffage et de conditionnement thermique au moyen d'une pompe a chaleur a compression fonctionnant avec un fluide mixte de travail et appareil pour la mise en oeuvre dudit procede
JPS6176855A (ja) * 1984-09-19 1986-04-19 株式会社東芝 カスケ−ド結合ヒ−トポンプ装置
DE3565718D1 (en) * 1984-09-19 1988-11-24 Toshiba Kk Heat pump system

Also Published As

Publication number Publication date
NO854845L (no) 1986-06-04
CA1262057A (en) 1989-10-03
NO164738B (no) 1990-07-30
DE3580249D1 (de) 1990-11-29
HUT41526A (en) 1987-04-28
DK161482B (da) 1991-07-08
DK553885A (da) 1986-06-04
NO164738C (no) 1990-11-14
JPS61180861A (ja) 1986-08-13
ATE57763T1 (de) 1990-11-15
DK553885D0 (da) 1985-11-29
DK161482C (da) 1991-12-16
HU198328B (en) 1989-09-28
US4688397A (en) 1987-08-25
EP0184181A2 (de) 1986-06-11
EP0184181A3 (en) 1988-01-13

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