WO2016104147A1 - 冷凍又は空調装置及びその制御方法 - Google Patents

冷凍又は空調装置及びその制御方法 Download PDF

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WO2016104147A1
WO2016104147A1 PCT/JP2015/084522 JP2015084522W WO2016104147A1 WO 2016104147 A1 WO2016104147 A1 WO 2016104147A1 JP 2015084522 W JP2015084522 W JP 2015084522W WO 2016104147 A1 WO2016104147 A1 WO 2016104147A1
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refrigerant
gas
stages
heat exchanger
temperature
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PCT/JP2015/084522
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English (en)
French (fr)
Japanese (ja)
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展海 猪野
宜三 関屋
敏和 寒風澤
明登 町田
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株式会社前川製作所
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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
    • F25B1/00—Compression machines, plants or systems with non-reversible cycle
    • 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

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  • the present invention relates to a refrigeration or air conditioner using a reverse Ericsson cycle as a refrigeration cycle and a control method thereof.
  • the theoretical reverse Ericsson cycle consists of two isothermal processes and two isobaric processes, which are heat exchange processes.
  • refrigeration cycle components such as a compressor, a condenser, an expansion valve and an evaporator provided in series in a refrigerant circuit, gas refrigerant and condensation from the evaporator outlet toward the compressor inlet
  • a reverse Ericsson cycle similar to the theoretical reverse Ericsson cycle is possible.
  • Patent Document 1 discloses a refrigeration or air conditioning system using the above-mentioned approximate reverse Ericsson cycle.
  • this approximate reverse Ericsson cycle the partial stroke of the isothermal compression performed in the superheated steam region in the isothermal heat release stroke is replaced by a multistage adiabatic compression stroke and a multistage isostatic heat release stroke, and the regenerative heat exchanger inlet from the evaporator outlet
  • the thermal efficiency can be improved by controlling the dryness of the gas refrigerant flowing into the
  • Patent No. 4726258 gazette
  • the specific heat of the gas refrigerant is smaller than that of the liquid refrigerant, so there is a tendency for the liquid refrigerant side to be insufficiently supercooled, and improvement in thermal efficiency can not be expected much
  • the thermal efficiency is to be improved by controlling the dryness (wetness) of the gas refrigerant flowing from the evaporator into the regenerative heat exchanger. It is difficult to control the refrigerant state at the outlet of the evaporator accurately to the appropriate wet state. Since the optimum control point is only one point, the coefficient of performance of the cycle decreases even if the refrigerant state at the outlet of the evaporator is too wet or too wet.
  • At least one embodiment of the present invention aims to make it possible to further improve the thermal efficiency in a refrigeration or air conditioning system using the aforementioned approximate reverse Ericsson cycle.
  • a refrigeration or air conditioner is (1) a plurality of stages of compressors, a condenser, a first expansion means, and an evaporator provided in series in the refrigerant circulation path;
  • a gas cooler for cooling the discharge gas refrigerant of each compressor, provided in the refrigerant circulation path connecting the outlet of each compressor and the inlet of the compressor of the next stage with the plurality of stages of compressors;
  • a regenerative heat exchanger for exchanging heat between the low pressure side gas refrigerant traveling from the evaporator to the compressor and the high pressure side liquid refrigerant traveling from the condenser to the first expansion means.
  • a reverse Ericsson cycle which comprises an isothermal heat release process mainly by the condenser, an isothermal heat absorption process by the evaporator, an isobaric heat release process by the high pressure side liquid refrigerant in the regenerative heat exchanger, and an equal pressure heat absorption process by the low pressure side gas refrigerant.
  • a partial stroke of isothermal compression performed in the superheated steam region of the isothermal heat release stroke is replaced with a plurality of adiabatic compression strokes by the plurality of compressors and a plurality of equal pressure heat release strokes by the gas cooler.
  • An approximate isothermal process is formed by The low pressure side gas refrigerant in a wet state was introduced to the regenerative heat exchanger at the start point of the isothermal heat absorption stroke, and the low pressure side gas refrigerant in a saturated state was introduced to the regeneration heat exchanger at the start point of the equal pressure heat absorption stroke.
  • the controller is further provided with a control device for controlling the end point of the equal pressure heat-releasing stroke to be located within the subcooling region up to the wet limit supercooling point where the heat release stroke ends.
  • FIG. 1 is a schematic view of a reverse Ericsson cycle in a TS diagram
  • FIG. 2 is a schematic view of the reverse Ericsson cycle in a Mollier diagram. 1 and 2, the symbol x is a saturated liquid line of the refrigerant, y is a saturated vapor line of the refrigerant, t is an isotherm, Tv is the discharge temperature of the single-stage compressor, and the broken line p in FIG.
  • a broken line S in FIG. 2 indicates an isentropic line corresponding to the vertical line bb 'in FIG.
  • the theoretical reverse Ericsson cycle is indicated by abgcda (solid line).
  • the theoretical reverse Ericsson cycle consists of two isothermal strokes (d ⁇ a and b ⁇ c) and two equal pressure strokes (a ⁇ b and c ⁇ d).
  • Two equal pressure strokes ab and cd are heat exchange processes in the regenerative heat exchanger.
  • the partial isothermal stroke (b ⁇ g) is a theoretical isothermal compression stroke. In the practical cycle, it is necessary to replace the adiabatic compression stroke (b.fwdarw.b ') and the equal pressure radiation stroke (b'.fwdarw.g), so that an additional compression work .DELTA.
  • the equal pressure stroke b'g is carried out by means of a gas cooler.
  • the partial stroke (isothermal stroke b ⁇ g) performed in the superheated steam region of the above-described equal pressure heat radiation stroke is divided into adiabatic compression strokes of multiple stages by multiple stages of compressors, and gas It is possible to approximate an isothermal process (b ⁇ g) by replacing the cooler and the sensible heat portion of the final stage with a plurality of stages of equal pressure heat dissipating processes with a condenser for cooling.
  • the compression work ⁇ can be reduced, and the COP (coefficient of performance) can be improved.
  • An outlet point d at which the liquid refrigerant outlet temperature of the regenerative heat exchanger is equal to the evaporation temperature is defined as a "wet limit supercooling point".
  • the opening degree of the first expansion means for example, expansion valve
  • the opening degree of the first expansion means may be increased to increase the amount of liquid supplied to the evaporator.
  • FIG. 3 and 4 show the relationship between the regenerative heat exchanger and coefficient of performance of the reverse Ericsson cycle using a single-stage compression stroke and a single-stage expansion stroke.
  • the vertical axes of the figures show the coefficient of performance (COP), and the horizontal axes show the liquid refrigerant outlet temperature (FIG. 3) and the gas refrigerant outlet temperature (FIG. 4) of the regenerative heat exchanger.
  • the display symbols such as ⁇ , ⁇ , ⁇ , ⁇ on the graph line indicate the positions of theoretical calculation points and do not indicate experimental results. This characteristic forms the main part of the present invention, and will be described in some detail below taking the refrigerant R600a (isobutane) of symbol B as an example.
  • FIG. 3 shows the characteristics of the liquid side refrigerant of the regenerative heat exchanger separately for the refrigerant.
  • a point c (state point c in FIGS. 1 and 2) on the liquid refrigerant outlet temperature 40 ° C. indicates the degree of subcooling, ie, COP in a state where the regenerative heat exchanger is not used.
  • the performance at point d (state point d in FIGS. 1 and 2) on the liquid refrigerant outlet temperature ⁇ 40 ° C. indicates COP at the maximum subcooling temperature. From FIG. 3, when the outlet temperature of the liquid refrigerant in the regenerative heat exchanger is known, the COP at that time can be known.
  • the outlet temperature of the liquid refrigerant is uniquely determined by the inlet temperature and the degree of humidity of the gas refrigerant. For example, if the gas refrigerant inlet is in a saturated gas state, the liquid refrigerant outlet is at point h (state point h in FIGS. 1 and 2), and COP is at a maximum. At that time, the solution temperature is about -13 ° C. according to FIG. This state point h is the aforementioned "sensible heat limit supercooling point".
  • the subcooling temperature of the liquid refrigerant can be lowered to the state point d ( ⁇ 40 ° C.).
  • the state point d is the "wet limit supercooling point”.
  • the inlet point of the gas refrigerant at this time (state point a 'in FIGS. 1 and 2) is defined as "wetness limit point a'".
  • the degree of wetness of the gas refrigerant inlet is increased above the wet limit point a ′, the temperature of the liquid refrigerant outlet remains unchanged, and the COP is lowered.
  • the temperature of the gas refrigerant inlet reaches an overheat temperature of ⁇ 40 ° C. or more, the liquid refrigerant outlet state has a higher temperature than the state point h, and therefore, the COP is lower than the result of FIG.
  • FIG. 4 shows the characteristics of the gas side refrigerant of the regenerative heat exchanger separately for the refrigerant.
  • COP on the temperature -40 ° C line is a saturated gas state a at the gas refrigerant inlet of the regenerative heat exchanger at -40 ° C (state point a in Figs. 1 and 2) and a saturated gas state at the gas refrigerant outlet -40 ° C This indicates the COP when in the state, in other words, the COP when the regenerative heat exchanger is not used.
  • COP at point b (state point b in FIG. 1 and FIG. 2) on the gas refrigerant outlet temperature 40 ° C.
  • FIG. 4 is COP when the liquid refrigerant outlet temperature becomes “sensible heat limit supercooling point h” in FIG. 1 and FIG. Indicates From FIG. 4, when the outlet temperature of the gas refrigerant of the regenerative heat exchanger is known, the COP at that time can be known immediately. As the outlet temperature of the gas refrigerant decreases from the condensation temperature of 40 ° C., the COP decreases.
  • the performance improvement effect by the regenerative heat exchanger is This means that the difference is maximum when the sensible heat difference between the inlet (saturated) and the outlet (superheated) is used to the maximum.
  • the outlet temperature of the liquid refrigerant is uniquely determined by the inlet temperature and the humidity of the gas refrigerant. The problem is that the influence of the inlet humidity of the gas refrigerant is not shown at all in FIG. The reason is that even if the humidity of the gas refrigerant inlet moves from the state a in FIGS.
  • the inlet temperature is constant (evaporation temperature), so the influence of the humidity on the coordinates in FIG. Because there is no way to show that. From this, maintaining the inlet of the gas refrigerant of the regenerative heat exchanger in a wet state is irrelevant to theoretical performance improvement, but has important meaning from the viewpoint of the control method for maximizing the COP. This will be described later.
  • the theoretical coefficient of performance of the conventional refrigeration cycle is COPtc (COP of each refrigerant at 40 ° C in Fig. 3 and COP of each refrigerant at -40 ° C in Fig. 4), the theoretical coefficient of performance of the reverse Ericsson cycle is COPte (the above (1)
  • COPteh the broken line COP in (2) above
  • the actual coefficient of performance of the conventional refrigeration cycle under the same operating conditions is COPpc
  • Ericsson cycle Let COPpe be the actual machine performance coefficient of.
  • the actual performance coefficient COPpe of the Ericsson cycle can be predicted from the following equation using the actual performance coefficient COPpc of the conventional cycle.
  • COP los in the formula (c) represents the amount of COP loss due to the regenerative heat exchanger loss.
  • the actual machine performance prediction value according to equation (a) has an error of 1%. There is no possibility of exceeding, and it can be considered to be within the accuracy range of the experimental results.
  • the solid line coefficient of performance (COP) of the refrigerant R 600 a indicates the performance when the heat transfer loss of the regenerative heat exchanger is zero.
  • the supercooling temperature reaches -40.degree. C.
  • Td -40.degree. C.
  • COP 2.30, respectively.
  • the performance when the temperature difference between the liquid refrigerant and the gas refrigerant at the cold end side and the warm end side inlet / outlet is set to 5 ° C. is shown by a horizontal broken line in consideration of the heat transfer loss of the regenerative heat exchanger.
  • the performance other than the broken line is indicated by a single solid line because the performance is the same regardless of the heat transfer loss.
  • the temperature range of the liquid side refrigerant is ⁇ 35 ° C. to 40 ° C.
  • the temperature range of the gas side refrigerant is ⁇ 40 ° C. to 35 ° C. Since the cold end side temperature difference of the regenerative heat exchanger is 5 ° C., the liquid refrigerant outlet temperature in FIG.
  • the “sensible heat limit supercooling point” moves to the high temperature side and the maximum COP decreases as the heat transfer loss of the regenerative heat exchanger increases.
  • Refrigerant A (R717: ammonia) has cycle performance significantly different from that of the other exemplified refrigerants in a reverse Ericsson cycle having a single-stage compression stroke and a single-stage expansion stroke. That is, the performance is rapidly reduced due to the subcooling of the liquid refrigerant by the regenerative heat exchanger.
  • the description is omitted because it deviates from the gist of the present invention, it is possible to improve the performance to the same extent as that of the refrigerant R600a even with the refrigerant R717 by increasing the number of compression stages and the number of expansion stages.
  • the adiabatic compression strokes of the plurality of stages Let the compression ratio r between multiple stages of compressors be (Pc / Pe) 1 / n ,
  • the control device is configured to cool the inlet gas refrigerant (condensing temperature + ⁇ ) to a temperature close to the condensing temperature in the plurality of stages of equal-pressure heat release strokes by the gas cooler.
  • ⁇ is a temperature rising range by compression of the suction gas of each stage compressor, and ⁇ can be decreased by increasing the number of stages. For this reason, ⁇ should be selected in consideration of the mechanical use temperature range of the compressor and the performance improvement by multistage formation.
  • the compression ratio r between the plurality of stages of compressors is set to (Pc / Pe) 1 / n .
  • the discharge pressure and the discharge temperature of the gas refrigerant can be made substantially equal, and the protrusion of the discharge pressure and the discharge temperature can be suppressed in some compressors. By this, it is possible to suppress the deterioration of the lubricating oil and the burning of the packing material and the like.
  • a first liquid gas separator provided in the refrigerant circuit between the regenerative heat exchanger and the first expansion means;
  • a second expansion means provided in the refrigerant circulation path on the inlet side of the first liquid gas separator;
  • the enthalpy difference is increased by providing the intermediate cooling device, and the corresponding change in total compression work occurs, but as a result, the COP is improved. Therefore, the COP of the refrigeration or air conditioner can be further improved.
  • a second liquid gas separator provided in the refrigerant circuit between the regenerative heat exchanger and the first expansion means; Third expansion means provided in the refrigerant circuit at the inlet side of the second liquid gas separator;
  • An economizer device further comprising: an economizer gas line for supplying gas refrigerant of the second liquid gas separator to an intermediate pressure region of the lower stage compressor among the plurality of stages of compressors.
  • the number of stages of the plurality of stages of compressors is two or three.
  • the number of stages of the plurality of stages of compressors is four or more, two stages or three stages become a practical number of stages because they do not contribute to the improvement of COP in spite of the increase in equipment cost.
  • a control method of a refrigeration or air conditioner (6) A plurality of stages of compressors, a condenser, a first expansion means, and an evaporator provided in series in the refrigerant circulation path; A gas cooler for cooling the discharge gas refrigerant of each compressor, provided in the refrigerant circulation path connecting the outlet of each compressor and the inlet of the next compressor with the plurality of stages of compressors; A control method of a refrigeration or air conditioner comprising: a regenerative heat exchanger for exchanging heat between a gas refrigerant traveling from the evaporator toward the compressor and a liquid refrigerant traveling from the condenser toward the first expansion means, A reverse Ericsson including an isothermal heat release process mainly by the condenser, an isothermal heat absorption process by the evaporator, an isobaric heat release process in a liquid area performed by heat exchange in the regenerative heat exchanger, and an isobaric heat absorption process in
  • the partial stroke performed in the superheated steam region of the isothermal heat release stroke is replaced with a plurality of adiabatic compression strokes by the plurality of stages of compressors and a plurality of equal pressure heat release strokes by the gas cooler and the condenser.
  • the first process The refrigerant in a wet state in the equal pressure heat absorption process is introduced into the regenerative heat exchanger, and the saturated gas refrigerant is introduced into the regenerative heat exchanger in the equal pressure heat absorption process, which is the end point of the isothermal heat release process Supercooling of the region from the sensible heat limit supercooling point to the wetness limit supercooling point where the end point of the adiabatic expansion stroke by the first expansion means is located on the saturated refrigerant liquid line and the end point of the isothermal heat release stroke And a second step of controlling the end point of the isothermal heat-releasing stroke at a point.
  • the practical cycle can be brought close to the theoretical reverse Ericsson cycle by the first process, COP can be improved and temperature increase of the gas refrigerant can be suppressed.
  • burning of the elastomer constituting the packing material and the like can be prevented.
  • the COP of the refrigeration or air conditioner can be maintained at the maximum, and the optimum control conditions are not point setting but range setting, and control is extremely easy.
  • the second stroke is performed by controlling the number of revolutions of a motor driving at least the plurality of stages of compressors and controlling the first expansion means.
  • the COP can be improved by the simple control conventionally performed.
  • the second process is A first step of detecting a temperature of the liquid refrigerant at the outlet of the regenerative heat exchanger from the condenser toward the first expansion means; Controlling the temperature detection value detected in the first step to a temperature between the refrigerant temperature at the sensible heat limit supercooling point and the evaporation temperature of the refrigerant in the evaporator.
  • the sensible heat limit overcooling point or the wet limit overcooling point can be easily determined from the detection values of the temperature sensor provided in the evaporator or the regenerative heat exchanger. .
  • FIG. 5 is a TS diagram of an approximate reverse Ericsson cycle. It is a Mollier diagram of an approximate reverse Ericsson cycle. It is a diagram which shows the liquid refrigerant exit temperature of a regeneration heat exchanger, and the relationship of COP. It is a diagram which shows the relationship between the gas refrigerant exit temperature of a regeneration heat exchanger, and COP. It is a systematic diagram of a freezing device concerning one embodiment.
  • FIG. 6 is a TS diagram showing an approximate reverse Ericsson cycle of the refrigeration system shown in FIG. 5; It is a systematic diagram of a freezing device concerning one embodiment. It is a TS diagram which shows the approximate reverse Ericsson cycle of the freezing apparatus shown in FIG.
  • FIG. 11 is a TS diagram showing an approximate reverse Ericsson cycle of the refrigeration apparatus shown in FIG. 10; It is a systematic diagram of a freezing device concerning one embodiment. It is a systematic diagram of a freezing device concerning one embodiment. It is a TS diagram which shows the approximate reverse Ericsson cycle which the refrigerating apparatus shown in FIG. 13 comprises. It is a chart showing COP (calculation value) of a freezing apparatus concerning the embodiment.
  • FIG. 6 is a diagram showing COPs of an approximate reverse Ericsson cycle and a conventional refrigeration cycle.
  • expressions that indicate that things such as “identical”, “equal” and “homogeneous” are equal states not only represent strictly equal states, but also have tolerances or differences with which the same function can be obtained. It also represents the existing state.
  • expressions representing shapes such as quadrilateral shapes and cylindrical shapes not only represent shapes such as rectangular shapes and cylindrical shapes in a geometrically strict sense, but also uneven portions and chamfers within the range where the same effect can be obtained. The shape including a part etc. shall also be expressed.
  • the expressions “comprising”, “having”, “having”, “including” or “having” one component are not exclusive expressions excluding the presence of other components.
  • FIG. 5 to 12 show several embodiments of the refrigeration or air conditioning device according to the present invention.
  • FIG. 5 shows a refrigeration system 10A that performs three-stage compression and three-stage expansion
  • FIG. 6 is a TS diagram of an approximate reverse Ericsson cycle configured by the refrigeration system 10A
  • FIG. 7 shows a refrigeration system 10B that performs three-stage compression and two-stage expansion
  • FIG. 8 is a TS diagram of an approximate reverse Ericsson cycle configured by the refrigeration apparatus 10B
  • FIG. 9 is a Mollier wire of the approximate reverse Ericsson cycle.
  • FIG. FIG. 10 shows a refrigeration system 10C that performs three-stage compression single-stage expansion
  • FIG. 11 is a TS diagram of an approximate reverse Ericsson cycle configured by the refrigeration system 10C
  • FIG. 12 shows a refrigeration system 10D that performs three-stage compression and two-stage expansion.
  • Refrigerating apparatuses 10A to 10D shown in FIGS. 5 to 12 have a plurality of stages (three stages) of compressors 13a, 13b and 13c, a condenser 16, a first expansion valve 18, and evaporation which constitute a refrigeration cycle in a refrigerant circuit 12.
  • the vessels 20 are provided in series.
  • Each compressor is rotationally driven by the electric motors 14a, 14b and 14c, respectively.
  • a positive displacement compressor such as a screw compressor is used.
  • turbo compressors may be used in applications where fluctuations in operating conditions are small.
  • the gas refrigerant paths 12a and 12b connecting the outlet and the inlet of each compressor are provided with gas coolers 22a and 22b for cooling the gas refrigerant discharged from the compressor 13a or 13b with a cooling medium such as cooling water.
  • the regenerative heat exchanger 24 for heat-exchanging the gas refrigerant which goes to the compressor 13a from the evaporator 20 and the liquid refrigerant which came out of the condenser 16 is provided.
  • a first liquid gas separator 28 provided in the refrigerant circuit 12 between the regenerative heat exchanger 24 and the first expansion valve 18, and a first liquid gas separator
  • the second expansion valve 30 provided in the refrigerant circulation passage 12 on the inlet side of 28 and the gas refrigerant separated from the liquid refrigerant by the first liquid gas separator 28 are discharged from the first stage compressor 13a and the second stage compression It further comprises an intercooler 26 having an intermediate gas path 32 supplying the refrigerant path 12a connected to the inlet of the machine 13b.
  • the intermediate gas passage 32 is connected to the refrigerant passage 12a on the downstream side of the gas cooler 22a.
  • a second liquid gas separator 36 provided in the refrigerant circuit 12 between the regenerative heat exchanger 24 and the first expansion valve 18, and a second liquid gas separator
  • a third expansion valve 38 provided in the refrigerant circulation passage 12 on the inlet side of 36, and an economizer gas passage 40 for supplying the gas refrigerant of the second liquid gas separator 36 to the intermediate pressure region of the first stage compressor 13a
  • the system further comprises an economizer device 34.
  • the liquid refrigerant of the first liquid gas separator 28 is supplied to the second liquid gas separator 36 through the refrigerant circuit 12.
  • the regenerative heat exchanger 24 to the refrigerant circuit 12 are used.
  • the liquid refrigerant is supplied to the first liquid gas separator 28.
  • the symbol P indicates the arrangement of the pressure sensor
  • T indicates the arrangement of the temperature sensor
  • G indicates the arrangement of the level sensor of the refrigerant liquid level.
  • the control values of the temperature sensors, the pressure sensors, the level sensors, and the rotational speed sensors 44a, 44b and 44c provided in the drive motors 14a, 14b and 14c provided in each of the refrigeration apparatuses 10A to 10D are all control devices It is input to 42. Based on these detected values, the controller 42 opens the first expansion valve 18, the second expansion valve 30, and the third expansion valve 38, and the rotational speeds of the electric motors 14a, 14b and 14c, and the intermediate gas line 32.
  • the opening degree of the flow rate adjusting valve 46 provided in the control unit 50 and the opening degree of the flow rate adjusting valve 48 provided in the economizer gas passage 40 are controlled.
  • the refrigeration systems 10A to 10D constitute the aforementioned practical reverse Ericsson cycle that approximates the theoretical reverse Ericsson cycle.
  • the approximate reverse Ericsson cycle constructed by the refrigeration systems 10A to 10D is shown in the TS diagrams of FIGS. 6, 8 and 11, and the Mollier diagram of FIG.
  • the approximate reverse Ericsson cycle of the refrigeration systems 10A to 10D is mainly performed by the isothermal heat release process (b ⁇ c) by the condenser 16, the isothermal endothermic process (f ⁇ a) by the evaporator 20, and the liquid by the regenerative heat exchanger 24. It includes a region of equal pressure radiation stroke (c ⁇ h) and a superheated steam region of equal pressure absorption stroke (a ⁇ b).
  • the partial process performed in the superheated steam region is adiabatic compression process (b ⁇ b ′) in multiple stages by multiple stages of compressors 13a, 13b and 13c, and a gas cooler
  • a plurality of equal pressure radiation steps (b ′ ⁇ b) by 22 a and 22 b and the condenser 16 and a single equal pressure radiation stroke (b ′ ⁇ g) by the condenser 16 are replaced (first step).
  • the compression ratio between the compressors Let (Pc / Pe) 1 / n . Since the refrigeration apparatuses 10A to 10D include the three-stage compressors 13a, 13b and 13c, the compression ratio between the compressors is (Pc / Pe) 1/3 .
  • the discharge gas refrigerant temperature is (condensing temperature + ⁇ ), and a part of the compression is performed by cooling the ⁇ component in the equal pressure heat radiation process of each stage. Suppresses the discharge temperature of the machine to achieve uniform compression and heat dissipation.
  • the controller 42 controls the opening degree of the first expansion valve 18 directly connected to the inlet of the evaporator 20 so that the liquid side outlet temperature T lout of the regenerative heat exchanger 24 becomes the set value T set of the supercooled liquid outlet temperature. Control.
  • the wetness at the evaporator outlet at this time is controlled to be the wetness at the state point (a 'to a).
  • the temperature of the regenerative heat exchanger liquid side outlet temperature Tout is the temperature range of the sensible heat limit supercooling point h and the wet limit supercooling point d It becomes equal to the set temperature T set set inside.
  • the rotational speed control of the electric motors 14a, 14b and 14c and the opening degree control of the first expansion valve 18 may be used in combination.
  • a temperature sensor 50 is provided in the liquid refrigerant passage 12 c at the outlet of the regenerative heat exchanger 24, and a pressure sensor 52 and a temperature sensor 54 are provided in the refrigerant passage 12 at the outlet of the evaporator 20.
  • the exemplary control method by the controller 42 first detects the temperature at the outlet of the regenerative heat exchanger 24 of the liquid refrigerant from the condenser 16 toward the first expansion valve 18 (first step). Next, the detected temperature is controlled to a temperature between the refrigerant temperature Th at the sensible heat limit supercooling point h and the evaporation temperature of the refrigerant in the evaporator 20 detected by the temperature sensor 54 (second step).
  • the discharge gas refrigerant (condensing temperature + ⁇ ) of the first stage compressor 13a and the second stage compressor 13b is cooled to near the condensation temperature by the gas coolers 22a and 22b.
  • the discharge gas refrigerant of the three-stage compressor 13 c is sent to the condenser 16.
  • the refrigerant temperature of the liquid gas separator 28 decreases from the expansion valve inlet temperature Th to the expansion valve outlet temperature Tk ′, passes through the first expansion valve 18 to become the evaporation temperature Te, and is supplied to the evaporator 20.
  • the multistage compression / expansion method causes a change in total compression work corresponding to the increase in enthalpy of ⁇ Hkf, but as a result, the COP is improved.
  • the gas refrigerant of the economizer unit 34 is sucked into the economizer port of the one-stage compressor 13a to lower the third expansion valve inlet temperature Tk 'to the expansion valve outlet temperature Tl'.
  • the enthalpy of ⁇ Hlk increases, and as a result, the COP can be improved.
  • the intermediate gas passage 32 and the economizer gas path 40 provided at the outlet of the expansion valve are gas at the sensible heat limit supercooling point h. It is necessary to connect to a lower stage compressor lower than the refrigerant pressure. That is, it is an area
  • the partial stroke (isothermal stroke b ⁇ g) performed in the superheated steam region is divided into a plurality of adiabatic compression strokes by the plurality of compressors 13a, 13b and 13c, a gas cooler 22a,
  • the compression work ⁇ can be reduced and COP can be improved by replacing it with a plurality of equal pressure radiation steps by 22 b and the condenser 16.
  • the control device 42 introduces a wet refrigerant from the evaporator 20 to the gas side inlet of the regenerative heat exchanger 24, and the end point of the isothermal heat release process at the liquid side outlet of the regenerative heat exchanger exceeds the sensible heat limit.
  • the discharge gas refrigerant at (condensing temperature + ⁇ ) is cooled to near the condensing temperature to suppress the protrusion of the discharge temperature in some of the compressors, thereby achieving homogeneous compression and the like. Achieve heat dissipation.
  • the temperature of the gas refrigerant can be lowered by multiple stages of equal-pressure radiation processes (b ' ⁇ b) and (b' ⁇ g) in the superheated steam region, deterioration of lubricating oil contained in the refrigerant, packing material, etc. Burnout of the elastomer can be prevented.
  • R717 (NH 3 ) refrigerant having a high specific heat ratio under the conditions of a condensing temperature of 40 ° C. and an evaporation temperature of -40 ° C. and a compressor suction gas refrigerant temperature of 40 ° C. about 380 ° C. in single-stage compression
  • the two-stage compression is about 180 ° C
  • the three-stage compression is about 120 ° C
  • the temperature rise width ⁇ is 340 ° C for a single stage
  • 140 ° C for a two-stage compressor and 80 ° C for a three-stage compressor, and the number of stages increases. Therefore, a drop in refrigerant gas temperature can be expected.
  • the compression ratio between the compressors is set to (condensing pressure Pc / evaporation pressure Pe) 1/3 , so that the discharge temperature of the gas refrigerant of each compressor is Since the control can be performed almost equally, it is possible to suppress the protrusion of the discharge temperature in some compressors.
  • the discharge pressure and discharge temperature of the gas refrigerant discharged from the compressor can be reduced, thereby reducing the power of the compressor. It is possible to further improve the COP of the refrigeration system.
  • the control by the control device 42 is simply performed by controlling the rotational speed of the electric motors 14a, 14b and 14c for driving the compressors 13a, 13b and 13c in multiple stages and controlling the opening degree of the first expansion valve 18. It can be easily done by control.
  • the detection of the sensible heat limit supercooling point h and the wet limit supercooling point d can be made from the detection values of the temperature sensors 50 and 54 and the pressure sensor 52 provided at the outlet of the evaporator 20 and the outlet of the regenerative heat exchanger It can be easily obtained.
  • FIG. 13 shows a refrigeration system 10E that performs two-stage compression and two-stage expansion as another embodiment
  • FIG. 14 shows an approximate reverse Ericsson cycle configured by the refrigeration system 10E.
  • FIG. 15 is a diagram showing COPs of the approximate reverse Ericsson cycle and the conventional refrigeration cycle.
  • FIG. 15 shows the case where the intermediate cooling device and the economizer device are not attached. It can be seen that the COP obtained in each embodiment of the present invention is improved over the conventional refrigeration cycle.
  • the refrigerant is R717 (NH 3 ) in two-stage compression
  • COP is lower than that of the conventional refrigeration cycle, but as the number of compression stages increases, the discharge temperature of the gas refrigerant from each compressor decreases. Compressor power can be reduced. Therefore, the compression loss can be improved, and the COP can be improved more than in the conventional refrigeration cycle. Further, it can be understood from FIG. 15 that an increase in COP can not be expected so much even when the number of stages of the compressor is four or more.
  • FIG. 16 is a chart showing COP (calculated value) of the refrigerating apparatus according to the several embodiments shown in FIG. 5 to FIG. Among them, FIG. 05 and the present invention No. 04 is the same value as the COP of the approximate reverse Ericsson cycle of the single-stage compression and 3-stage compression shown in FIG.

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  • Air Conditioning Control Device (AREA)
PCT/JP2015/084522 2014-12-26 2015-12-09 冷凍又は空調装置及びその制御方法 WO2016104147A1 (ja)

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JP2019074250A (ja) * 2017-10-16 2019-05-16 株式会社デンソー ヒートポンプサイクル

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FR3118484A1 (fr) * 2020-12-28 2022-07-01 Commissariat A L’Energie Atomique Et Aux Energies Alternatives Système de compression à plusieurs étages de compression montés en série
WO2024189697A1 (ja) * 2023-03-10 2024-09-19 ダイキン工業株式会社 冷凍サイクル装置

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JPS5556567A (en) * 1978-10-20 1980-04-25 Takatama Mitsuko Method of refrigeration by vapor compression
JP2004061061A (ja) * 2002-07-31 2004-02-26 Matsushita Electric Ind Co Ltd 冷凍サイクル装置およびその運転方法
JP2008185327A (ja) * 2007-01-26 2008-08-14 Grasso Gmbh Refrigeration Technology 二段配置形式のオイルオーバフロー式スクリューコンプレッサを備えたｃｏ２冷凍装置
JP2009133547A (ja) * 2007-11-30 2009-06-18 Mitsubishi Electric Corp 冷凍サイクル装置
JP2009529123A (ja) * 2006-03-27 2009-08-13 株式会社前川製作所 蒸気圧縮式冷凍サイクル、その制御方法およびそれを用いた冷凍装置

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JPS5556567A (en) * 1978-10-20 1980-04-25 Takatama Mitsuko Method of refrigeration by vapor compression
JP2004061061A (ja) * 2002-07-31 2004-02-26 Matsushita Electric Ind Co Ltd 冷凍サイクル装置およびその運転方法
JP2009529123A (ja) * 2006-03-27 2009-08-13 株式会社前川製作所 蒸気圧縮式冷凍サイクル、その制御方法およびそれを用いた冷凍装置
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JP2019074250A (ja) * 2017-10-16 2019-05-16 株式会社デンソー ヒートポンプサイクル
CN111247378A (zh) * 2017-10-16 2020-06-05 株式会社电装 热泵循环
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