EP3009767B1 - Heat pump device - Google Patents
Heat pump device Download PDFInfo
- Publication number
- EP3009767B1 EP3009767B1 EP13886812.0A EP13886812A EP3009767B1 EP 3009767 B1 EP3009767 B1 EP 3009767B1 EP 13886812 A EP13886812 A EP 13886812A EP 3009767 B1 EP3009767 B1 EP 3009767B1
- Authority
- EP
- European Patent Office
- Prior art keywords
- refrigerant
- heat transfer
- pipe
- gas cooler
- twist
- 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.)
- Active
Links
- 239000003507 refrigerant Substances 0.000 claims description 408
- XLYOFNOQVPJJNP-UHFFFAOYSA-N water Substances O XLYOFNOQVPJJNP-UHFFFAOYSA-N 0.000 claims description 145
- 239000007788 liquid Substances 0.000 claims description 32
- 238000010438 heat treatment Methods 0.000 claims description 12
- 238000007599 discharging Methods 0.000 claims 2
- 230000000694 effects Effects 0.000 description 9
- CURLTUGMZLYLDI-UHFFFAOYSA-N Carbon dioxide Chemical compound O=C=O CURLTUGMZLYLDI-UHFFFAOYSA-N 0.000 description 8
- 230000008859 change Effects 0.000 description 8
- 238000010586 diagram Methods 0.000 description 6
- 230000009467 reduction Effects 0.000 description 6
- 229910002092 carbon dioxide Inorganic materials 0.000 description 5
- 230000014509 gene expression Effects 0.000 description 5
- 239000000463 material Substances 0.000 description 5
- 239000008236 heating water Substances 0.000 description 4
- 239000001569 carbon dioxide Substances 0.000 description 3
- ATUOYWHBWRKTHZ-UHFFFAOYSA-N Propane Chemical compound CCC ATUOYWHBWRKTHZ-UHFFFAOYSA-N 0.000 description 2
- 230000008901 benefit Effects 0.000 description 2
- 239000012530 fluid Substances 0.000 description 2
- 230000017525 heat dissipation Effects 0.000 description 2
- 230000014759 maintenance of location Effects 0.000 description 2
- 238000005057 refrigeration Methods 0.000 description 2
- -1 R410A Chemical compound 0.000 description 1
- 230000002528 anti-freeze Effects 0.000 description 1
- 238000007664 blowing Methods 0.000 description 1
- 239000012267 brine Substances 0.000 description 1
- 230000006835 compression Effects 0.000 description 1
- 238000007906 compression Methods 0.000 description 1
- 230000003247 decreasing effect Effects 0.000 description 1
- 230000005484 gravity Effects 0.000 description 1
- 230000012447 hatching Effects 0.000 description 1
- 239000007791 liquid phase Substances 0.000 description 1
- 238000004519 manufacturing process Methods 0.000 description 1
- 239000000203 mixture Substances 0.000 description 1
- 239000012071 phase Substances 0.000 description 1
- 239000001294 propane Substances 0.000 description 1
- QQONPFPTGQHPMA-UHFFFAOYSA-N propylene Natural products CC=C QQONPFPTGQHPMA-UHFFFAOYSA-N 0.000 description 1
- 125000004805 propylene group Chemical group [H]C([H])([H])C([H])([*:1])C([H])([H])[*:2] 0.000 description 1
- HPALAKNZSZLMCH-UHFFFAOYSA-M sodium;chloride;hydrate Chemical compound O.[Na+].[Cl-] HPALAKNZSZLMCH-UHFFFAOYSA-M 0.000 description 1
- 229910000679 solder Inorganic materials 0.000 description 1
- 239000000243 solution Substances 0.000 description 1
- 238000013517 stratification Methods 0.000 description 1
- 230000007704 transition Effects 0.000 description 1
- 230000005514 two-phase flow Effects 0.000 description 1
- 238000011144 upstream manufacturing Methods 0.000 description 1
Images
Classifications
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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
- F25B30/00—Heat pumps
- F25B30/02—Heat pumps of the compression type
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
- F04B39/00—Component parts, details, or accessories, of pumps or pumping systems specially adapted for elastic fluids, not otherwise provided for in, or of interest apart from, groups F04B25/00 - F04B37/00
- F04B39/06—Cooling; Heating; Prevention of freezing
-
- 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
- F04C23/00—Combinations of two or more pumps, each being of rotary-piston or oscillating-piston type, specially adapted for elastic fluids; Pumping installations specially adapted for elastic fluids; Multi-stage pumps specially adapted for elastic fluids
- F04C23/008—Hermetic pumps
-
- 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
- F25B40/00—Subcoolers, desuperheaters or superheaters
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28D—HEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA DO NOT COME INTO DIRECT CONTACT
- F28D7/00—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall
- F28D7/0008—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall the conduits for one medium being in heat conductive contact with the conduits for the other medium
- F28D7/0016—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall the conduits for one medium being in heat conductive contact with the conduits for the other medium the conduits for one medium or the conduits for both media being bent
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28D—HEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA DO NOT COME INTO DIRECT CONTACT
- F28D7/00—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall
- F28D7/0066—Multi-circuit heat-exchangers, e.g. integrating different heat exchange sections in the same unit or heat-exchangers for more than two fluids
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28D—HEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA DO NOT COME INTO DIRECT CONTACT
- F28D7/00—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall
- F28D7/02—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall the conduits being helically coiled
- F28D7/024—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall the conduits being helically coiled the conduits of only one medium being helically coiled tubes, the coils having a cylindrical configuration
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/02—Tubular elements of cross-section which is non-circular
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/12—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element
- F28F1/34—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element and extending obliquely
- F28F1/36—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element and extending obliquely the means being helically wound fins or wire spirals
-
- 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
- F25B2339/00—Details of evaporators; Details of condensers
- F25B2339/04—Details of condensers
- F25B2339/047—Water-cooled condensers
-
- 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
- F25B2400/07—Details of compressors or related parts
- F25B2400/072—Intercoolers therefor
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F2275/00—Fastening; Joining
- F28F2275/06—Fastening; Joining by welding
Definitions
- the refrigerant may be refrigerant making it possible to supply high temperature hot-water such as, for example, carbon dioxide, R410A, propane, or propylene, but not limited to them.
- the hot-water supply pipe 16 When hot water is supplied from the hot-water supply pipe 16 to the hot-water supply terminal, the high temperature hot water in the hot water storage tank 2a is supplied through the hot water delivery pipe 14 to the hot-water supplying mixing valve 2c, and low temperature water is supplied through the water supply branch pipe 15 to the hot-water supplying mixing valve 2c.
- the high temperature hot water and the low temperature water are mixed by the hot-water supplying mixing valve 2c, and then supplied through the hot-water supply pipe 16 to the hot-water supply terminal.
- the hot-water supplying mixing valve 2c has a function of adjusting a mixture ratio between the high temperature hot water and the low temperature water so as to reach a hot-water supply temperature set by a user.
- the first refrigerant heat transfer pipes 42a, 42b, 42c are, respectively, fitted in the grooves 411a, 411b, 411c and wound helically along shapes of the grooves 411a, 411b, 411c. Such a configuration can increase a contact heat transfer area between the first twist pipe 41 and the first refrigerant heat transfer pipe 42.
- An inner diameter SRi of the first twist pipe 41 is herein defined as a length of a portion in Figure 4 .
- the inner diameter SRi of the first twist pipe 41 refers to an inner diameter of a portion with a smallest inner diameter in the first twist pipe 41.
- the first gas cooler 4 has a higher oil circulation rate than the second gas cooler 5, mainly flowing medium is the refrigerant.
- the heat exchanger constituting the first gas cooler 4 preferably has a configuration of a general heat exchanger for a refrigerant rather than of an oil cooler type heat exchanger.
- the first gas cooler 4 preferably uses a twist pipe like the second gas cooler 5.
- the ratio of the refrigerant pressure loss of the first gas cooler 4 with respect to the refrigerant pressure loss of the overall gas coolers is about 10%.
- setting the value of the inner diameter ratio di1/di2 to 1.4 sufficiently reduces the refrigerant pressure loss of the first gas cooler 4 in the relationship with the ratio of the channel length.
- too large a value of the inner diameter ratio di1/di2 that is, too large an inner diameter di1 of the first refrigerant heat transfer pipe 42 may cause the negative effects as described above such as an excessive length of the first twist pipe 41 or an increase in the retention of refrigerator oil in the first gas cooler 4.
- the value of the inner diameter ratio di1/di2 of 1.4 or less the inner diameter di1 of the first refrigerant heat transfer pipe 42 is not too large, thereby reliably preventing the negative effects.
- the refrigerant pressure loss of the first gas cooler 4 can be reliably prevented to reduce input power for the compressor 3 and improve a COP.
- the refrigerant pressure loss of the second gas cooler 5 can be sufficiently reduced.
- the inner diameter di2 of the second refrigerant heat transfer pipe 52 or the sectional area of each second refrigerant heat transfer channel in the second gas cooler 5 being relatively small increases the refrigerant flow speed in the second refrigerant heat transfer pipe 52, that is, in each second refrigerant heat transfer channel, thereby increasing a heat transfer coefficient of the refrigerant.
- the ratio of the total sectional area of the first refrigerant heat transfer channels in the first heat exchanger to the total sectional area of the second refrigerant heat transfer channels in the second heat exchanger is (1.2) 2 ⁇ 1.4. If the inner diameter ratio dil/di2 is 1.4, the ratio of the total sectional area of the first refrigerant heat transfer channels in the first heat exchanger to the total sectional area of the second refrigerant heat transfer channels in the second heat exchanger is (1.4) 2 ⁇ 2.
Landscapes
- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Physics & Mathematics (AREA)
- Thermal Sciences (AREA)
- Geometry (AREA)
- Heat-Pump Type And Storage Water Heaters (AREA)
- Heat-Exchange Devices With Radiators And Conduit Assemblies (AREA)
Description
- The present invention relates to a heat pump device.
-
Patent Literature 1 discloses a hot-water supply cycle device including: a gas cooler having high temperature side refrigerant piping, low temperature side refrigerant piping, and water piping; and a hot-water supply compressor having a sealed container, a compressing element, an electric actuating element, an intake pipe, a discharge pipe, a refrigerant reintroduction pipe, and a refrigerant redischarge pipe. In this device, the intake pipe guides low pressure refrigerant directly to the compressing element, the discharge pipe discharges high pressure refrigerant compressed by the compressing element directly to an outside of the sealed container without releasing the high pressure refrigerant into the sealed container, the refrigerant reintroduction pipe guides the refrigerant resulting from the high pressure refrigerant having passed through the high temperature side refrigerant piping and been subjected to heat exchange into the sealed container, and the refrigerant redischarge pipe redischarges the refrigerant having passed through the electric actuating element in the sealed container to the outside of the sealed container, and feeds the refrigerant to the low temperature side refrigerant piping. -
- Patent Literature 1: Japanese Patent Laid-Open No.
2006-132427 - Patent Literature 2: Japanese Patent Laid-Open No.
2004-108616 - Patent Literature 3: Japanese Patent Laid-Open No.
2008-309361 - Patent Literature 4: Japanese Patent Laid-Open No.
2009-168383 - In the conventional device described above, refrigerator oil is supplied into a compression chamber of the compressing element in order to lubricate and seal a slide portion and reduce friction and gap leakage. Thus, a large amount of refrigerator oil together with a compressed refrigerant gas is discharged from the discharge pipe of the compressor out of the compressor, and is circulated to the high temperature side refrigerant piping. On the other hand, the refrigerant discharged from the refrigerant redischarge pipe of the compressor contains a significantly smaller amount of refrigerator oil than that discharged from the discharge pipe.
- The refrigerator oil has a much higher viscosity than the refrigerant. Thus, in the conventional device described above, the large amount of refrigerator oil together with the refrigerant is circulated to the high temperature side refrigerant piping, thereby increasing pressure loss of the refrigerant. This increases discharge pressure of the compressor and increases input power for the compressor, thereby reducing a coefficient of performance (COP).
- The present invention is achieved to solve the above described problems, and has an object to improve a COP of a heat pump device including a compressor having a first discharge passage and a second discharge passage, a mass flow rate of refrigerator oil discharged together with a refrigerant from the first discharge passage being higher than a mass flow rate of refrigerator oil discharged together with a refrigerant from the second discharge passage.
- A heat pump device of the invention is defined by
claim 1. - The heat pump device according to the present invention can reliably reduce pressure loss of the refrigerant in the first heat exchanger to which the refrigerant and the refrigerator oil are circulated, the refrigerant and the refrigerator oil being discharged from the first discharge passage with a large discharge amount of refrigerator oil. This can reduce input power for the compressor, and improve a COP.
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- [
Figure 1] Figure 1 is a configuration diagram of a heat pump device according toEmbodiment 1 of the present invention. - [
Figure 2] Figure 2 is a configuration diagram of a storage type hot-water supply system including the heat pump device inFigure 1 . - [
Figure 3] Figure 3 is a perspective view of essential portions of a first gas cooler included in the heat pump device inEmbodiment 1 of the present invention. - [
Figure 4] Figure 4 is a sectional view of essential portions of the first gas cooler included in the heat pump device inEmbodiment 1 of the present invention. - [
Figure 5] Figure 5 is an enlarged sectional view of essential portions of the first gas cooler and a second gas cooler included in the heat pump device inEmbodiment 1 of the present invention. - [
Figure 6] Figure 6 shows temperature changes of refrigerant and water in the first gas cooler and the second gas cooler as a whole, and a split position between the first gas cooler and the second gas cooler. - [
Figure 7] Figure 7 shows density change of the refrigerant in the first gas cooler and the second gas cooler as a whole. - [
Figure 8] Figure 8 shows ratios of refrigerant pressure losses of the first gas cooler and the second gas cooler in a case where the first gas cooler and the second gas cooler have the same shape other than their channel lengths. - [
Figure 9] Figure 9 is a configuration diagram of a conventional heat pump device. - [
Figure 10] Figure 10 shows a relationship between a ratio of a twist pitch p to an inner diameter SRi of a first twist pipe and a heat transfer coefficient on water side. - [
Figure 11] Figure 11 shows a relationship between the ratio of the twist pitch p to the inner diameter SRi of the first twist pipe and a required length of the first twist pipe. - [
Figure 12] Figure 12 shows a relationship between the ratio of the twist pitch p to the inner diameter SRi of the first twist pipe and a required length of a first refrigerant heat transfer pipe. - [
Figure 13] Figure 13 shows a relationship among a refrigerant pressure loss of the first gas cooler, the ratio of the twist pitch p to the inner diameter SRi of the first twist pipe, and an inner diameter di1 of the first refrigerant heat transfer pipe. - [
Figure 14] Figure 14 shows a relationship between the ratio of the twist pitch p to the inner diameter SRi of the first twist pipe in the first gas cooler and a length of the first twist pipe in each of cases inFigure 13 . - [
Figure 15] Figure 15 shows change in the refrigerant pressure loss of the first gas cooler when an inner diameter ratio di1/di2 of the first refrigerant heat transfer pipe and a second refrigerant heat transfer pipe is changed at p/SRi of 1.8 of the first twist pipe. - [
Figure 16] Figure 16 shows change in heat transfer coefficient on water side in a case where the twist pitch p of the first twist pipe is equal to a twist pitch p2 of the second twist pipe and inner diameters SRi of the first twist pipe and the second twist pipe are equal. - Now, with reference to the drawings, embodiment of the present invention will be described. Throughout the drawings, common components are denoted by the same reference numerals and overlapping descriptions will be omitted. In the description below, a channel length is sometimes simply referred to as "length" for simplicity.
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Figure 1 is a configuration diagram of a heat pump device according toEmbodiment 1 of the present invention.Figure 2 is a configuration diagram of a storage type hot-water supply system including the heat pump device inFigure 1 . As shown inFigure 1 , theheat pump device 1 according toEmbodiment 1 includes a refrigerant circuit including acompressor 3, afirst gas cooler 4 as a first heat exchanger, asecond gas cooler 5 as a second heat exchanger, anexpansion valve 6 as expansion means, and an evaporator 7, connected by refrigerant piping. Thefirst gas cooler 4 includes a first refrigerant heat transfer channel and a first liquid heat transfer channel, and performs heat exchange between the first refrigerant heat transfer channel and the first liquid heat transfer channel. Thesecond gas cooler 5 includes a second refrigerant heat transfer channel and a second liquid heat transfer channel, and performs heat exchange between the second refrigerant heat transfer channel and the second liquid heat transfer channel. Theheat pump device 1 causes a liquid to be a heat medium or an object to be heated to flow through the first liquid heat transfer channel in thefirst gas cooler 4 and the second liquid heat transfer channel in thesecond gas cooler 5, and heats the liquid. In the heat pump device according toEmbodiment 1, the liquid to be heated is water. The evaporator 7 in Embodiment 1 is constituted by an air-refrigerant heat exchanger for performing heat exchange between air and refrigerant. Theheat pump device 1 according to Embodiment 1 further includes afan 8 for blowing air to the evaporator 7, and a high and lowpressure heat exchanger 9 for performing heat exchange between high pressure refrigerant and low pressure refrigerant. During heating operation for heating water, theheat pump device 1 actuates thecompressor 3 to operate a heat pump cycle (refrigeration cycle). - As shown in
Figure 2 , theheat pump device 1 according toEmbodiment 1 may be combined with thetank unit 2 and used as a storage type hot-water supply system. In thetank unit 2, a hotwater storage tank 2a for storing hot water and water, and awater pump 2b are provided. Theheat pump device 1 and thetank unit 2 are connected by apipe 11 and apipe 12 through which water flows, and electric wires (not shown). One end of thepipe 11 is connected to awater inlet 1a of theheat pump device 1. The other end of thepipe 11 is connected to a lower portion of the hotwater storage tank 2a in thetank unit 2. Thewater pump 2b is provided in a middle of thepipe 11 in thetank unit 2. One end of thepipe 12 is connected to awater outlet 1b of theheat pump device 1. The other end of thepipe 12 is connected to an upper portion of the hotwater storage tank 2a in thetank unit 2. Instead of the shown configuration, thewater pump 2b may be placed in theheat pump device 1. - As shown in
Figure 1 , thecompressor 3 in theheat pump device 1 includes a sealedcontainer 31, a compressingelement 32 and anelectric actuating element 33 provided in the sealedcontainer 31, afirst intake passage 34, afirst discharge passage 35, asecond intake passage 36, and asecond discharge passage 37. Low pressure refrigerant sucked from thefirst intake passage 34 directly flows into the compressingelement 32 without being released into aninternal space 38 of the sealedcontainer 31. The compressingelement 32 is driven by theelectric actuating element 33, and compresses the low pressure refrigerant into high pressure refrigerant. The high pressure refrigerant compressed by the compressingelement 32 is discharged through thefirst discharge passage 35 directly out of the sealedcontainer 31 without being released into theinternal space 38 of the sealedcontainer 31. The high pressure refrigerant discharged from thefirst discharge passage 35 flows through apipe 10 into thefirst gas cooler 4. The high pressure refrigerant having passed through thefirst gas cooler 4 flows through apipe 17 to thesecond intake passage 36 of thecompressor 3. The high pressure refrigerant sucked from thesecond intake passage 36 into thecompressor 3 is released into theinternal space 38 of the sealedcontainer 31. InEmbodiment 1, the compressingelement 32 is placed below theelectric actuating element 33. An outlet of thesecond intake passage 36 opens into theinternal space 38 of the sealedcontainer 31 at a height between theelectric actuating element 33 and the compressingelement 32. An inlet of thesecond discharge passage 37 opens into theinternal space 38 of the sealedcontainer 31 at a height above theelectric actuating element 33. The high pressure refrigerant released from the outlet of thesecond intake passage 36 into theinternal space 38 of the sealedcontainer 31 passes through a gap or the like between a rotor 331 and astator 332 of theelectric actuating element 33 to a top of theelectric actuating element 33, and is discharged through thesecond discharge passage 37 out of the sealedcontainer 31. The high pressure refrigerant discharged from thesecond discharge passage 37 flows through apipe 18 into thesecond gas cooler 5. The high pressure refrigerant having passed through thesecond gas cooler 5 passes through apipe 19 to theexpansion valve 6. The high pressure refrigerant passes through theexpansion valve 6 to turn into low pressure refrigerant. The low pressure refrigerant flows through apipe 20 into the evaporator 7. The low pressure refrigerant having passed through the evaporator 7 flows through apipe 21 to thefirst intake passage 34 of thecompressor 3, and is sucked into thecompressor 3. The high and lowpressure heat exchanger 9 performs heat exchange between the high pressure refrigerant passing through thepipe 19 and the low pressure refrigerant passing through thepipe 21. The high pressure refrigerant discharged from thefirst discharge passage 35 is reduced in pressure due to pressure loss while returning through thefirst gas cooler 4 to thesecond intake passage 36. Thus, pressure PH2 of the high pressure refrigerant in theinternal space 38 of the sealedcontainer 31 is lower than pressure PH1 of the high pressure refrigerant discharged from thefirst discharge passage 35. Specifically, the discharge pressure PH1 of thefirst discharge passage 35 is higher than the discharge pressure PH2 of thesecond discharge passage 37. - The
heat pump device 1 further includes awater channel 23 for guiding water having flowed in from thewater inlet 1a to a water inlet of thesecond gas cooler 5, and awater channel 26 for guiding water (hot water) having flowed out of a water outlet of thefirst gas cooler 4 to thewater outlet 1b. A water outlet of thesecond gas cooler 5 is connected to a water inlet of thefirst gas cooler 4. During heating operation, water having flowed in from thewater inlet 1a flows through thewater channel 23 into thesecond gas cooler 5, and is heated by heat from the refrigerant in thesecond gas cooler 5. Hot water generated by heating in thesecond gas cooler 5 flows into thefirst gas cooler 4, and is further heated by heat from the refrigerant in thefirst gas cooler 4. The hot water further increased in temperature by being further heated in thefirst gas cooler 4 passes through thewater channel 26 to thehot water outlet 1b, and is fed through thepipe 12 to thetank unit 2. - The refrigerant may be refrigerant making it possible to supply high temperature hot-water such as, for example, carbon dioxide, R410A, propane, or propylene, but not limited to them.
- The high temperature and high pressure refrigerant gas discharged from the
first discharge passage 35 of thecompressor 3 releases heat and is reduced in temperature while passing through thefirst gas cooler 4. InEmbodiment 1, the refrigerant reduced in temperature while passing through thefirst gas cooler 4 is sucked from thesecond intake passage 36 into theinternal space 38 of the sealedcontainer 31 to cool theelectric actuating element 33. Thus, a temperature of theelectric actuating element 33 and a surface temperature of the sealedcontainer 31 can be reduced. This can increase motor efficiency of theelectric actuating element 33, and reduce heat dissipation loss from a surface of the sealedcontainer 31. The refrigerant gas sucked into theinternal space 38 of the sealedcontainer 31 draws heat from theelectric actuating element 33 and is increased in temperature. The refrigerant gas is then discharged from thesecond discharge passage 37 and flows into thesecond gas cooler 5, and releases heat and is reduced in temperature while passing through thesecond gas cooler 5. The high pressure refrigerant reduced in temperature heats the low pressure refrigerant while passing through the high and lowpressure heat exchanger 9, and then passes through theexpansion valve 6. The refrigerant passes through theexpansion valve 6, and is thus reduced in pressure into a low pressure gas-liquid two-phase state. The refrigerant having passed through theexpansion valve 6 absorbs heat from outside air while passing through the evaporator 7, and is evaporated and gasified. The low pressure refrigerant coming out of the evaporator 7 is heated by the high and lowpressure heat exchanger 9, and then sucked from thefirst intake passage 34 into thecompressor 3. - If the high pressure refrigerant pressure is not less than critical pressure, the refrigerant in the
first gas cooler 4 and thesecond gas cooler 5 is reduced in temperature and releases heat still in a supercritical state without gas-liquid phase transition. If the high pressure refrigerant pressure is not more than the critical pressure, the refrigerant is liquefied and releases heat. InEmbodiment 1, carbon dioxide or the like is preferably used as the refrigerant to bring the high pressure refrigerant pressure to the critical pressure or more. If the high pressure refrigerant pressure is not less than the critical pressure, the liquefied refrigerant can be reliably prevented from flowing from thesecond intake passage 36 into theinternal space 38 of the sealedcontainer 31. This can reliably prevent the liquefied refrigerant from adhering to theelectric actuating element 33, and reduce rotational resistance of theelectric actuating element 33. In addition, the liquefied refrigerant does not flow from thesecond intake passage 36 into theinternal space 38 of the sealedcontainer 31, thereby preventing the refrigerator oil from being diluted by the refrigerant. - As shown in
Figure 2 , awater supply pipe 13 is further connected to a lower portion of the hotwater storage tank 2a of thetank unit 2. Water supplied from an external water source such as a water supply flows through thewater supply pipe 13 into the hotwater storage tank 2a and is stored. The hotwater storage tank 2a is always filled with water flowing from thewater supply pipe 13. A hot-water supplying mixingvalve 2c is further provided in thetank unit 2. The hot-water supplying mixingvalve 2c is connected via a hotwater delivery pipe 14 to the upper portion of the hotwater storage tank 2a. A watersupply branch pipe 15 branching off from thewater supply pipe 13 is connected to the hot-water supplying mixingvalve 2c. One end of the hot-water supply pipe 16 is further connected to the hot-water supplying mixingvalve 2c. The other end of the hot-water supply pipe 16 is connected to a hot-water supply terminal such as a tap, a shower, or a bathtub (not shown). - During heating operation for heating water stored in the hot
water storage tank 2a, the water stored in the hotwater storage tank 2a is fed by thewater pump 2b through thepipe 11 to theheat pump device 1, and heated in theheat pump device 1 to be high temperature hot water. The high temperature hot water generated in theheat pump device 1 returns through thepipe 12 to thetank unit 2, and flows into the hotwater storage tank 2a from above. By such heating operation, in the hotwater storage tank 2a, hot water and water are stored so as to form temperature stratification with a hot upper side and a cold lower side. - When hot water is supplied from the hot-
water supply pipe 16 to the hot-water supply terminal, the high temperature hot water in the hotwater storage tank 2a is supplied through the hotwater delivery pipe 14 to the hot-water supplying mixingvalve 2c, and low temperature water is supplied through the watersupply branch pipe 15 to the hot-water supplying mixingvalve 2c. The high temperature hot water and the low temperature water are mixed by the hot-water supplying mixingvalve 2c, and then supplied through the hot-water supply pipe 16 to the hot-water supply terminal. The hot-water supplying mixingvalve 2c has a function of adjusting a mixture ratio between the high temperature hot water and the low temperature water so as to reach a hot-water supply temperature set by a user. - The
heat pump device 1 includes acontrol unit 50. Thecontrol unit 50 is electrically connected to actuators and sensors (not shown) included in theheat pump device 1 and thetank unit 2, and user interface devices (not shown), and functions as control means for controlling operation of the storage type hot-water supply system. InFigure 2 , thecontrol unit 50 is provided in theheat pump device 1, but thecontrol unit 50 may be provided other than in theheat pump device 1. Thecontrol unit 50 may be provided in thetank unit 2. Thecontrol unit 50 may be provided in theheat pump device 1 and thetank unit 2 in a divided manner so as to be able to mutually communicate. - During heating operation, the
control unit 50 performs control so that a temperature of the hot water supplied from theheat pump device 1 to the tank unit 2 (hereinafter referred to as "hot water delivery temperature") reaches a target hot water delivery temperature. The target hot water delivery temperature is set to, for example, 65°C to 90°C. InEmbodiment 1, thecontrol unit 50 adjusts a rotation speed of thewater pump 2b to control the hot water delivery temperature. Thecontrol unit 50 detects the hot water delivery temperature using a temperature sensor (not shown) provided in thewater channel 26. If the detected hot water delivery temperature is higher than the target hot water delivery temperature, the rotation speed of thewater pump 2b is corrected to be higher, and if the hot water delivery temperature is lower than the target hot water delivery temperature, the rotation speed of thewater pump 2b is corrected to be lower. As such, thecontrol unit 50 can perform control so that the hot water delivery temperature matches the target hot water delivery temperature. The hot water delivery temperature may be controlled by controlling a temperature of the refrigerant discharged from thefirst discharge passage 35 of thecompressor 3, a rotation speed of thecompressor 3, or the like. - An oil reservoir (not shown) that stores refrigerator oil is located in a lower portion of the
internal space 38 of the sealedcontainer 31 of thecompressor 3 inFigure 1 . In order to lubricate and seal a slide portion to reduce friction and gap leakage, the refrigerator oil is supplied from the oil reservoir into the compressingelement 32. The refrigerator oil supplied into the compressingelement 32 together with the compressed high temperature and high pressure refrigerant gas is discharged from thefirst discharge passage 35. Thus, a relatively large amount of refrigerator oil is discharged from thefirst discharge passage 35. The refrigerant gas and the refrigerator oil discharged from thefirst discharge passage 35 form a gas-liquid two-phase flow, which flows through thefirst gas cooler 4 to thesecond intake passage 36, and is released from thesecond intake passage 36 into theinternal space 38 of the sealedcontainer 31. - The refrigerator oil has a higher density than the refrigerant gas. Thus, the refrigerator oil having flowed from the
second intake passage 36 into theinternal space 38 of the sealedcontainer 31 falls by gravity, and is stored in the oil reservoir in the lower portion of theinternal space 38 of the sealedcontainer 31. As such, the refrigerant is separated from the refrigerator oil. However, a part of the refrigerator oil is atomized and mixed in the refrigerant gas. A part of the refrigerator oil as a liquid film may be also raised and spattered by a flow of the refrigerant gas when the refrigerant and the refrigerator oil are released from an outlet of thesecond intake passage 36 into theinternal space 38 of the sealedcontainer 31. Thus, a small amount of refrigerator oil is mixed in the refrigerant gas passing through the gap between the rotor 331 and thestator 332 of theelectric actuating element 33 to a top of theelectric actuating element 33. A part of the mixed refrigerator oil is separated from the refrigerant gas by a centrifugal force caused by rotation of the rotor 331. The remaining refrigerator oil together with the refrigerant gas is discharged through thesecond discharge passage 37 out of the sealedcontainer 31. From the above, a mass flow rate of the refrigerator oil discharged from thefirst discharge passage 35 is higher than a mass flow rate of the refrigerator oil discharged from thesecond discharge passage 37. On the other hand, a mass flow rate of the refrigerant discharged from thefirst discharge passage 35 is equal to a mass flow rate of the refrigerant discharged from thesecond discharge passage 37. - A large amount of refrigerator oil together with the refrigerant gas is circulated to the first refrigerant heat transfer channel in the
first gas cooler 4. On the other hand, a smaller amount of refrigerator oil is circulated to the second refrigerant heat transfer channel in thesecond gas cooler 5 as compared to thefirst gas cooler 4. The refrigerator oil has a much higher viscosity than the refrigerant. Thus, the large amount of refrigerator oil being circulated to thefirst gas cooler 4 easily increases refrigerant pressure loss. The increase in refrigerant pressure loss of thefirst gas cooler 4 increases discharge pressure of thecompressor 3, and increases input power for thecompressor 3, thereby reducing a COP (coefficient of performance). In order to solve this problem, inEmbodiment 1, a total sectional area of the first refrigerant heat transfer channel(s) in thefirst gas cooler 4 through which the refrigerant and the refrigerator oil discharged from thefirst discharge passage 35 pass is larger than a total sectional area of the second refrigerant heat transfer channel(s) in thesecond gas cooler 5 through which the refrigerant and the refrigerator oil discharged from thesecond discharge passage 37 pass. - A sectional area of the channel herein refers to an area of a range of a flowing fluid in a section perpendicular to a flow direction of the fluid. If there are a plurality of first refrigerant heat transfer channels in the
first gas cooler 4, that is, if the refrigerant and the refrigerator oil having flowed into thefirst gas cooler 4 are split into the plurality of first refrigerant heat transfer channels and flow in parallel, a total sectional area of the first refrigerant heat transfer channels refers to a sum of sectional area of each of the first refrigerant heat transfer channels. Similarly, if there are a plurality of second refrigerant heat transfer channels in thesecond gas cooler 5, that is, if the refrigerant and the refrigerator oil having flowed into thesecond gas cooler 5 are split into the plurality of second refrigerant heat transfer channels and flow in parallel, a total sectional area of the second refrigerant heat transfer channels refer to a sum of sectional area of each of the first refrigerant heat transfer channels. - As described below, in
Embodiment 1, the total sectional area of the first refrigerant heat transfer channel(s) in thefirst gas cooler 4 is larger than the total sectional area of the second refrigerant heat transfer channel(s) in thesecond gas cooler 5, thereby reliably preventing an increase in refrigerant pressure loss of thefirst gas cooler 4. This reduces discharge pressure of thecompressor 3, reduces input power for thecompressor 3, and improves a COP. -
Figure 3 is a perspective view of essential portions of thefirst gas cooler 4 inEmbodiment 1.Figure 4 is a sectional view of essential portions of thefirst gas cooler 4 inEmbodiment 1. As shown inFigures 3 and4 , thefirst gas cooler 4 includes onefirst twist pipe 41 and three first refrigerantheat transfer pipes 42.Figure 4 shows a section in a longitudinal direction of thefirst twist pipe 41. InFigure 3 , the three first refrigerantheat transfer pipes 42 are denoted byreference numerals Figure 3 , for easy distinction between the first refrigerantheat transfer pipes heat transfer pipes Figure 3 does not refer to sections. - In the
first gas cooler 4 inEmbodiment 1, the refrigerant and the refrigerator oil flow in the first refrigerantheat transfer pipe 42. Specifically, the first refrigerantheat transfer pipe 42 forms the first refrigerant heat transfer channel. Thefirst gas cooler 4 inEmbodiment 1 includes the three first refrigerantheat transfer pipes first gas cooler 4 are split into the three first refrigerantheat transfer pipes first gas cooler 4, that is, the first heat exchanger is not limited to three, but may be one, two, four or more. - The
first twist pipe 41 has ahelical groove 411 in an outer periphery thereof. The number of the groove(s) 411 is equal to the number of the first refrigerant heat transfer pipe(s) 42. Specifically, inEmbodiment 1, thefirst twist pipe 41 has threegrooves 411 in parallel. InFigure 3 , the threegrooves 411 are denoted byreference numerals grooves heat transfer pipes grooves grooves first twist pipe 41 and the first refrigerantheat transfer pipe 42. - In the
first gas cooler 4 inEmbodiment 1, thefirst twist pipe 41 forms a first liquid heat transfer channel through which water passes. InEmbodiment 1, onefirst twist pipe 41, that is, one first liquid heat transfer channel is provided in thefirst gas cooler 4. However, in the present invention, a plurality of first liquid heat transfer channels may be provided in thefirst gas cooler 4, that is, the first heat exchanger so that a liquid such as water is split into the first liquid heat transfer channels and flows in parallel. - Water flows through the
first twist pipe 41 from right to left inFigures 3 and4 . The refrigerant and the refrigerator oil flow helically through the first refrigerantheat transfer pipe 42 from left to right inFigures 3 and4 . Specifically, a flow direction of water is opposite to a traveling direction of the refrigerant flowing helically to form counter flows. - An inner diameter SRi of the
first twist pipe 41 is herein defined as a length of a portion inFigure 4 . Specifically, the inner diameter SRi of thefirst twist pipe 41 refers to an inner diameter of a portion with a smallest inner diameter in thefirst twist pipe 41. -
Figure 5 is an enlarged sectional view of essential portions of thefirst gas cooler 4 and thesecond gas cooler 5 inEmbodiment 1. InFigure 5 , (1) shows thefirst gas cooler 4. InFigure 5 , (2) shows thesecond gas cooler 5. As shown inFigure 5 , thefirst twist pipe 41 and the first refrigerantheat transfer pipe 42 are joined with aheat transfer material 60 such as solder. Thesecond gas cooler 5 includes asecond twist pipe 51 and a second refrigerantheat transfer pipe 52. Thesecond twist pipe 51 has ahelical groove 511 in an outer periphery thereof. In thesecond gas cooler 5 inEmbodiment 1, the second refrigerantheat transfer pipe 52 forms a second refrigerant heat transfer channel, and thesecond twist pipe 51 forms a second liquid heat transfer channel. Since thesecond gas cooler 5 has similar structure as thefirst gas cooler 4, drawings corresponding toFigures 3 and4 are omitted. The description above on thefirst gas cooler 4 also applies to thesecond gas cooler 5.Figure 5 shows a section in a longitudinal direction of thefirst twist pipe 41 or thesecond twist pipe 51. - As shown in
Figure 5 , in a case where the first refrigerantheat transfer pipe 42 or the second refrigerantheat transfer pipe 52 originally having a circular tubular shape is wound helically around thefirst twist pipe 41 or thesecond twist pipe 51, a sectional shape of the first refrigerantheat transfer pipe 42 or the second refrigerantheat transfer pipe 52 after being wound is not a circle, but is a flat or elliptical shape with a long side in an axial direction of thefirst twist pipe 41 or thesecond twist pipe 51. An inner diameter di1 of the first refrigerantheat transfer pipe 42 or an inner diameter di2 of the second refrigerantheat transfer pipe 52 herein refer to an inner diameter of a circular state before the refrigerant heat transfer pipe is wound around thefirst twist pipe 41 or thesecond twist pipe 51. - Generally, in the
first gas cooler 4 or thesecond gas cooler 5, an end of the first refrigerantheat transfer pipe 42 or the second refrigerantheat transfer pipe 52 is not wound around thefirst twist pipe 41 or thesecond twist pipe 51. Thus, in such a portion, the inner diameter di1 of the first refrigerantheat transfer pipe 42 or the inner diameter di2 of the second refrigerantheat transfer pipe 52 before being wound around thefirst twist pipe 41 or thesecond twist pipe 51 may be measured. - Instead of the above definition, the first refrigerant
heat transfer pipe 42 or the second refrigerantheat transfer pipe 52 wound around thefirst twist pipe 41 or thesecond twist pipe 51 may be regarded to have an elliptical shape, and an average value of a long diameter and a short diameter of the ellipse may be used as the inner diameter di1 of the first refrigerantheat transfer pipe 42 or the inner diameter di2 of the second refrigerantheat transfer pipe 52. - As shown in
Figure 5 , inEmbodiment 1, the inner diameter di1 of the first refrigerantheat transfer pipe 42 in thefirst gas cooler 4 is desirably larger than the inner diameter di2 of the second refrigerantheat transfer pipe 52 in thesecond gas cooler 5. In addition, a twist pitch p of thefirst twist pipe 41 in thefirst gas cooler 4 is desirably larger than a twist pitch p2 of thesecond twist pipe 51 in thesecond gas cooler 5. The twist pitch p of thefirst twist pipe 41 in thefirst gas cooler 4 and the twist pitch p2 of thesecond twist pipe 51 in thesecond gas cooler 5 are herein defined as lengths of portions inFigure 5 . Specifically, the twist pitch p of thefirst twist pipe 41 is a distance between centers of two peaks with thegroove 411 therebetween in a section in a longitudinal direction of thefirst twist pipe 41. Similarly, the twist pitch p2 of thesecond twist pipe 51 is a distance between centers of two peaks with thegroove 511 therebetween in a section in a longitudinal direction of thesecond twist pipe 51. - In an example described below, a case of using carbon dioxide as the refrigerant will be described. In the example described below, the number of the first refrigerant heat transfer channel(s) in the
first gas cooler 4 is equal to the number of the second refrigerant heat transfer channel(s) in thesecond gas cooler 5.Figure 6 shows temperature changes of the refrigerant and water in thefirst gas cooler 4 and thesecond gas cooler 5 as a whole, and a split position between thefirst gas cooler 4 and thesecond gas cooler 5. The axis of abscissa inFigure 6 represents a ratio to a total length of thefirst twist pipe 41 and the second twist pipe 51 (that is, a sum of lengths of the first liquid heat transfer channel and the second liquid heat transfer channel). An origin (0) on the axis of abscissa inFigure 6 represents a water outlet and a refrigerant inlet of thefirst gas cooler 4, and a right end (1) on the axis of abscissa represents a water inlet and a refrigerant outlet of thesecond gas cooler 5. - As described above, a large amount of refrigerator oil together with the refrigerant gas is circulated in the first refrigerant
heat transfer pipe 42 in thefirst gas cooler 4. In thefirst gas cooler 4, hot refrigerator oil is also subjected to heat exchange with water. Specific heat of the refrigerator oil being lower than specific heat of the refrigerant gas may cause a reduction in heating capability and a resulting reduction in hot-water supply efficiency. In a relationship between the temperature and the specific heat of the refrigerant gas and the refrigerator oil, the specific heat of the refrigerant gas significantly increases at a temperature of 20°C to 60°C, while the specific heat of the refrigerator oil is substantially constant irrespective of the temperature. In order to prevent a reduction in heating capability due to the refrigerant gas containing a large amount of refrigerator oil, the refrigerant gas needs to contain little refrigerator oil in a temperature zone with a significant increase in specific heat of the refrigerant gas. As shown inFigure 6 , a temperature of a pinch point at which temperatures of the refrigerant gas and water are closest is about 50°C. Thus, an upper limit temperature in a range with a rapid increase in specific heat of the refrigerant gas is about the temperature at the pinch point plus 10°C. Thus, if an outlet temperature (≈ a temperature of the second intake passage 36) of the first refrigerantheat transfer pipe 42 in thefirst gas cooler 4 is 10°C or more higher than the temperature at the pinch point, a reduction in heating capability can be prevented. If the outlet temperature of the first refrigerantheat transfer pipe 42 in thefirst gas cooler 4 is at least higher than the temperature at the pinch point, a significant reduction in heating capability can be prevented. From the above, the split position between thefirst gas cooler 4 and thesecond gas cooler 5 is desirably on a high temperature side of the pinch point at which the temperatures of the refrigerant gas and water are closest. In particular, inEmbodiment 1, as shown inFigure 6 , the length of thefirst twist pipe 41 in thefirst gas cooler 4 is desirably about 10% on the high temperature side of the total length of thefirst twist pipe 41 and thesecond twist pipe 51. -
Figure 7 shows density change of the refrigerant in thefirst gas cooler 4 and thesecond gas cooler 5 as a whole. The axis of abscissa inFigure 7 refers to the same as the axis of abscissa inFigure 6 . As shown inFigure 7 , the refrigerant at higher temperature has a lower density. - Pressure loss ΔP of the refrigerant in the refrigerant heat transfer pipe is expressed by the following
expression 1. Here, the sectional shape of the refrigerant heat transfer pipe is a circle for simplicity of description. -
- Here, for simplicity of description, it is assumed that the shapes and the refrigerant flow rates of the first refrigerant
heat transfer pipe 42 and the second refrigerantheat transfer pipe 52 are constant, and the pipe friction coefficient λ does not change. From the above expression, the refrigerant pressure loss ΔP per unit channel length is proportional to 1/p. - In
Embodiment 1, the refrigerant gas containing the large amount of refrigerator oil is circulated to thefirst gas cooler 4, and the refrigerant gas containing only the small amount of refrigerator oil is circulated to thesecond gas cooler 5. If a viscosity of a CO2 gas refrigerant in thefirst gas cooler 4 is 1, an average viscosity ratio of the refrigerator oil is 311. As such, the refrigerator oil has a significantly higher viscosity than the CO2 gas refrigerant. This increases pressure loss of the refrigerant gas containing the large amount of refrigerator oil. -
- The oil circulation rate OC is a ratio of the mass flow rate of the refrigerator oil with respect to a sum of the mass flow rate of the refrigerant and the mass flow rate of the refrigerator oil. In a rated operation state of the
heat pump device 1, the oil circulation rate OC of thefirst gas cooler 4 is preferably not less than 2%, and more preferably not less than 5%. In addition, in the rated operation state of theheat pump device 1, the oil circulation rate OC of thefirst gas cooler 4 is preferably not more than 20%, and more preferably not more than 10%. Setting the oil circulation rate OC of thefirst gas cooler 4 to the above-described lower limit value or more allows heat from the hot refrigerator oil in thecompressor 3 to be effectively used for heating water in thefirst gas cooler 4, improving heating capability. Setting the oil circulation rate OC of thefirst gas cooler 4 to the above-described upper limit value or less can reliably reduce the refrigerant pressure loss of thefirst gas cooler 4, and also reliably prevent an excessive reduction in the amount of the refrigerator oil in thecompressor 3. - In the rated operation state of the
heat pump device 1, the oil circulation rate OC of thesecond gas cooler 5 is preferably not less than 0.01%, and more preferably not less than 0.1%. In the rated operation state of theheat pump device 1, the oil circulation rate OC of thesecond gas cooler 5 is preferably not more than 1%, and more preferably not more than 0.5%. Setting the oil circulation rate OC of thesecond gas cooler 5 to the above-described upper limit value or less can reliably reduce the refrigerant pressure loss of thesecond gas cooler 5. If the oil circulation rate OC of thesecond gas cooler 5 is low and close to the above-described lower limit value, the refrigerator oil has little influence, and there is no need to further reduce the oil circulation rate OC of thesecond gas cooler 5 to be lower than the above-described lower limit value. Depending on operation conditions of theheat pump device 1, the oil circulation rate OC of thesecond gas cooler 5 may be lower than the above-described lower limit value. - If the oil circulation rate OC is about 5% to 10%, the refrigerant pressure loss is about 1.6 to 2.0 times larger than that when the oil circulation rate OC is 0.5% or less under the same other conditions.
-
Figure 8 shows ratios of refrigerant pressure losses of thefirst gas cooler 4 and thesecond gas cooler 5 in a case where thefirst gas cooler 4 and thesecond gas cooler 5 have the same shape other than their channel lengths.Figure 9 is a configuration diagram of a conventional heat pump device. First, a conventionalheat pump device 70 inFigure 9 will be described. Components common with those of theheat pump device 1 according toEmbodiment 1 are denoted by the same reference numerals and overlapping descriptions will be omitted. Theheat pump device 70 inFigure 9 includes acompressor 71 having one intake passage and one discharge passage instead of thecompressor 3 in theheat pump device 1 according toEmbodiment 1. Theheat pump device 70 includes asingle gas cooler 72 instead of thefirst gas cooler 4 and thesecond gas cooler 5. In theheat pump device 70, low pressure refrigerant sucked from thepipe 21 into thecompressor 71 is compressed by thecompressor 71 into high pressure refrigerant. The high pressure refrigerant is discharged from thecompressor 71 and passes through thepipe 10 and thegas cooler 72 to thepipe 19. - The case of "0.5% OR LESS IN OVERALL GAS COOLER(S)" in
Figure 8 refers to a case where, as in the conventionalheat pump device 70 inFigure 9 , thegas cooler 72 is not split into thefirst gas cooler 4 and thesecond gas cooler 5, and the refrigerant from which the refrigerator oil is separated in the sealed container of thecompressor 71 is caused to flow into thegas cooler 72. Specifically, this refers to a case of a conventional refrigeration cycle where the refrigerant is not returned into the sealedcontainer 31 of thecompressor 3 between thefirst gas cooler 4 and thesecond gas cooler 5. In this case, if the refrigerant pressure loss of theoverall gas cooler 72 is 1, a ratio of the refrigerant pressure loss of a portion corresponding to a channel length of 10% on a refrigerant high temperature side of a total channel length of thegas cooler 72 is 0.17. The ratio of the refrigerant pressure loss of a remaining portion corresponding to a channel length of 90% on a refrigerant low temperature side is 0.83. As shown inFigure 7 , on the high temperature side of the refrigerant gas, the refrigerant density is low, and thus the ratio of the refrigerant pressure loss of the portion corresponding to the channel length of 10% of the total channel length is 17% of the total refrigerant pressure loss, and higher than the ratio of the channel length. - The case of "OIL CIRCULATION RATE IS HIGH IN FIRST GAS COOLER AND 0.5% OR LESS IN SECOND GAS COOLER" in
Figure 8 refers to a case where the refrigerant pressure loss is twice larger than that when the oil circulation rate is 0.5% or less because of the oil circulation rate of about 5% to 10% of thefirst gas cooler 4. Here, the channel length of 10% on the refrigerant high temperature side of the total channel length of thefirst gas cooler 4 and thesecond gas cooler 5 corresponds to thefirst gas cooler 4. In this case, if the refrigerant pressure loss of theoverall gas cooler 72 is 1, the ratio of the refrigerant pressure loss of thefirst gas cooler 4 is 0.17 × 2 = 0.34. Thus, the ratio of the refrigerant pressure loss of thefirst gas cooler 4 and thesecond gas cooler 5 as a whole is 0.34 + 0.83 = 1.17. As such, if the refrigerant pressure loss is twice larger on the refrigerant high temperature side with high refrigerant pressure loss per unit channel length, the refrigerant pressure loss of the overall gas coolers is significantly influenced. Thus, the refrigerant pressure loss of the overall gas coolers is 1.17 times as compared to a case with a low oil circulation rate as a whole. The ratio of the refrigerant pressure loss of thefirst gas cooler 4 with respect to the overall gas coolers is 29% and high. - Although the
first gas cooler 4 has a higher oil circulation rate than thesecond gas cooler 5, mainly flowing medium is the refrigerant. Thus, the heat exchanger constituting thefirst gas cooler 4 preferably has a configuration of a general heat exchanger for a refrigerant rather than of an oil cooler type heat exchanger. For example, thefirst gas cooler 4 preferably uses a twist pipe like thesecond gas cooler 5. - From the above, the high oil circulation rate of the
first gas cooler 4 easily increases the refrigerant pressure loss of thefirst gas cooler 4, and increases discharge pressure of thecompressor 3. This easily increases input power for thecompressor 3 and reduces the COP. Thus, inEmbodiment 1, the refrigerant pressure loss of thefirst gas cooler 4 is reduced as described below. - A relationship among the inner diameter di1 of the first refrigerant
heat transfer pipe 42, the channel length L of the first refrigerantheat transfer pipe 42, and the refrigerant pressure loss of thefirst gas cooler 4 will be described. The refrigerant pressure loss ΔP in the first refrigerantheat transfer pipe 42 has the following proportional relationship from theexpressions 1 to 3 above, with a pipe friction coefficient, a refrigerant density, and a refrigerant flow rate being constant. - Thus, in order to reduce the refrigerant pressure loss of the
first gas cooler 4, it is advantageous to reduce the channel length L of the first refrigerantheat transfer pipe 42 and increase the inner diameter di1 of the first refrigerantheat transfer pipe 42. - Next, an advantage of an increase in the twist pitch p of the
first twist pipe 41 will be described.Figure 10 shows a relationship between a ratio of the twist pitch p to the inner diameter SRi of thefirst twist pipe 41 and a heat transfer coefficient on water side.Figure 10 shows change in heat transfer coefficient on water side with a constant inner diameter SRi and an increased twist pitch p of thefirst twist pipe 41. InFigure 10 , the heat transfer coefficient on water side is represented by a ratio to the heat transfer coefficient on water side when p/SRi is 1. As shown inFigure 10 , with increasing p/SRi, that is, with increasing twist pitch p of thefirst twist pipe 41, the heat transfer coefficient on water side increases. -
Figure 11 shows a relationship between the ratio of the twist pitch p to the inner diameter SRi of thefirst twist pipe 41 and a required length of thefirst twist pipe 41. InFigure 11 , a length of thefirst twist pipe 41 required, when the twist pitch p is increased with a constant inner diameter SRi of thefirst twist pipe 41, for obtaining an equal amount of heat exchange is represented by a ratio to a reference length. Thefirst gas cooler 4 as a twist pipe type heat exchanger is configured so that the first refrigerantheat transfer pipe 42 is wound along thehelical groove 411 in thefirst twist pipe 41. Thus, increasing the twist pitch p of thefirst twist pipe 41 reduces the length of the first refrigerantheat transfer pipe 42 wound around thefirst twist pipe 41 per unit length, thereby reducing a contact area between the first refrigerantheat transfer pipe 42 and thefirst twist pipe 41. Thus, with increasing twist pitch p of thefirst twist pipe 41, the length of thefirst twist pipe 41 required for equalizing the amounts of heat exchange of the refrigerant and water is increased. However, as shown inFigure 10 , with increasing twist pitch p, the heat transfer coefficient on water side increases, thereby increasing heat exchange efficiency per unit length of thefirst twist pipe 41. These relationships determines the relationship inFigure 11 . -
Figure 12 shows a relationship between the ratio of the twist pitch p to the inner diameter SRi of thefirst twist pipe 41 and a required length of the first refrigerantheat transfer pipe 42. InFigure 12 , the length of the first refrigerantheat transfer pipe 42 required, when the twist pitch p is increased with a constant inner diameter SRi of thefirst twist pipe 41, for obtaining an equal amount of heat exchange is represented by a ratio to the length of the first refrigerantheat transfer pipe 42 required at p/SRi of 1. As described with reference toFigure 11 , with increasing twist pitch p of thefirst twist pipe 41, a required length of thefirst twist pipe 41 is increased. However, with increasing twist pitch p of thefirst twist pipe 41, the length of the first refrigerantheat transfer pipe 42 wound around thefirst twist pipe 41 per unit length is reduced. Thus, as shown inFigure 12 , with increasing twist pitch p of thefirst twist pipe 41, the required length of the first refrigerantheat transfer pipe 42 is reduced. However, in a region at p/SRi of about 1.8 or more, the required length of the first refrigerantheat transfer pipe 42 is less likely to be reduced. - Summarizing the above characteristics, if the twist pitch p of the
first twist pipe 41 is increased, the length of thefirst twist pipe 41 required for obtaining an equal amount of heat exchange is increased, while the ratio of the heat transfer coefficient on water side is increased, thereby relatively gently increasing the required length of thefirst twist pipe 41. As shown inFigure 12 , with increasing twist pitch p of thefirst twist pipe 41, the length of the first refrigerantheat transfer pipe 42 can be effectively reduced, which is advantageous for reducing the refrigerant pressure loss of thefirst gas cooler 4. -
Figure 13 shows a relationship among the refrigerant pressure loss of thefirst gas cooler 4, the ratio of the twist pitch p to the inner diameter SRi of thefirst twist pipe 41, and the inner diameter di1 of the first refrigerantheat transfer pipe 42. InFigure 13 and thereafter, a ratio di1/di2 of the inner diameter di1 of the first refrigerantheat transfer pipe 42 in thefirst gas cooler 4 to the inner diameter di2 of the second refrigerantheat transfer pipe 52 in the second gas cooler is referred to as "an inner diameter ratio".Figure 13 shows changes in refrigerant pressure loss of thefirst gas cooler 4 when the twist pitch p of thefirst twist pipe 41 is changed for each of cases where the inner diameter ratio di1/di2 is set to a plurality of values inFigure 13 with a constant amount of heat exchange in thefirst gas cooler 4. InFigure 13 , the refrigerant pressure loss of thefirst gas cooler 4 is represented by a ratio to the refrigerant pressure loss of thefirst gas cooler 4 when values of the inner diameter ratio di1/di2 and p/SRi are both 1. -
Figure 14 shows a relationship between the ratio of the twist pitch p to the inner diameter SRi of thefirst twist pipe 41 in thefirst gas cooler 4 and the length of thefirst twist pipe 41 in each of the cases inFigure 13 . InFigure 14 , the length of thefirst twist pipe 41 is represented by a ratio to the length of thefirst twist pipe 41 when the values of the inner diameter ratio dil/di2 and p/SRi are both 1. InFigures 13 and 14 , the ratio of the twist pitch p2 to the inner diameter SRi of thesecond twist pipe 51 in thesecond gas cooler 5 is about 1. The inner diameter SRi of thefirst twist pipe 41 in thefirst gas cooler 4 is equal to the inner diameter SRi of thesecond twist pipe 51 in thesecond gas cooler 5. - As shown in
Figure 13 , in a case where the inner diameter ratio di1/di2 is equal, with increasing p/SRi, that is, with increasing twist pitch p of thefirst twist pipe 41, the refrigerant pressure loss of thefirst gas cooler 4 is reduced. In a case where the twist pitch p of thefirst twist pipe 41 is equal, with increasing inner diameter ratio dil/di2, that is, with increasing inner diameter di1 of the first refrigerantheat transfer pipe 42, the refrigerant pressure loss of thefirst gas cooler 4 is reduced. - As described above, with increasing inner diameter ratio di1/di2, that is, with increasing inner diameter di1 of the first refrigerant
heat transfer pipe 42, the refrigerant pressure loss of thefirst gas cooler 4 is more effectively reduced. However, with increasing inner diameter di1 of the first refrigerantheat transfer pipe 42, the flow speed of the refrigerant in the first refrigerantheat transfer pipe 42 is reduced, thereby reducing a heat transfer coefficient in the first refrigerantheat transfer pipe 42. Thus, as shown inFigure 14 , with increasing inner diameter ratio di1/di2, that is, with increasing inner diameter di1 of the first refrigerantheat transfer pipe 42, the length of thefirst twist pipe 41 required for obtaining an equal amount of heat exchange is increased. In addition, as shown inFigure 14 , with increasing p/SRi, that is, with increasing twist pitch p of thefirst twist pipe 41, the required length of thefirst twist pipe 41 is increased. - If the length of the
first twist pipe 41 in thefirst gas cooler 4 is increased, a size of the overall gas coolers including thefirst gas cooler 4 and thesecond gas cooler 5 may be increased to increase a size of a casing of theheat pump device 1. In addition, if the length of thefirst twist pipe 41 in thefirst gas cooler 4 is increased, an amount of material required for thefirst twist pipe 41 is increased to increase weight and cost. In addition, if the length of thefirst twist pipe 41 forming the water channel is excessively increased, an amount of heat dissipation from thefirst gas cooler 4 out of theheat pump device 1 may be increased or the pressure loss on water side may be increased. - As described above, with increasing twist pitch p of the
first twist pipe 41 in thefirst gas cooler 4, the refrigerant pressure loss of thefirst gas cooler 4 is reduced, while the length of thefirst twist pipe 41 is increased. Thus, excessively increasing the twist pitch p of thefirst twist pipe 41 may excessively increase the length of thefirst twist pipe 41, thereby causing the negative effects as described above. In this view, p/SRi as the ratio of the twist pitch p to the inner diameter SRi of thefirst twist pipe 41, according to the invention, is not more than 1.8. As described above, in the region at p/SRi of more than 1.8, the required length of the first refrigerantheat transfer pipe 42 is less likely to be reduced by increasing the twist pitch p of thefirst twist pipe 41. Thus, in the region at p/SRi of more than 1.8, further increasing the twist pitch p of thefirst twist pipe 41 is less effective for further reducing the refrigerant pressure loss, and also easily causes the negative effects of the increased length of thefirst twist pipe 41. On the other hand, p/SRi of 1.8 or less can reliably prevent the negative effects of the increased length of thefirst twist pipe 41. - In addition, p/SRi of the
first twist pipe 41 in thefirst gas cooler 4, according to the invention, is not less than 1.1, more preferably not less than 1.2, and further preferably not less than 1.4. Setting p/SRi to preferably 1.1 or more, more preferably 1.2 or more, and further preferably 1.4 or more can effectively reduce the length of the first refrigerant heat transfer pipe 42 (seeFigure 12 ). This can more reliably reduce the refrigerant pressure loss of thefirst gas cooler 4. In short, p/SRi of thefirst twist pipe 41 in thefirst gas cooler 4 is, according to the invention, not less than 1.1 and not more than 1.8, more preferably not less than 1.2 and not more than 1.8, and further preferably not less than 1.4 and not more than 1.8. By setting p/SRi to such a range, markedly advantageously, increasing the twist pitch p of thefirst twist pipe 41 can sufficiently increase the effect of reducing the refrigerant pressure loss of thefirst gas cooler 4 and can reliably prevent the negative effects due to the increased length of thefirst twist pipe 41. - Next, a preferable maximum value of the inner diameter ratio di1/di2 of the first refrigerant
heat transfer pipe 42 and the second refrigerantheat transfer pipe 52 will be described.Figure 15 shows change in the refrigerant pressure loss of thefirst gas cooler 4 when the inner diameter ratio di1/di2 of the first refrigerantheat transfer pipe 42 and the second refrigerantheat transfer pipe 52 is changed at p/SRi of 1.8 of thefirst twist pipe 41. InFigure 15 , the refrigerant pressure loss of thefirst gas cooler 4 is represented by a ratio to a sum of the refrigerant pressure loss of thefirst gas cooler 4 and the refrigerant pressure loss of the second gas cooler 5 (that is, the refrigerant pressure loss of the overall gas coolers). As shown inFigure 15 , with increasing inner diameter ratio dil/di2, that is, with increasing inner diameter di1 of the first refrigerantheat transfer pipe 42, the refrigerant pressure loss of thefirst gas cooler 4 is reduced, and the ratio of the refrigerant pressure loss of thefirst gas cooler 4 to the refrigerant pressure loss of the overall gas coolers is reduced. However, as shown inFigure 14 , with increasing inner diameter ratio di1/di2, that is, with increasing inner diameter di1 of the first refrigerantheat transfer pipe 42, the length of thefirst twist pipe 41 is increased. Also, in thefirst gas cooler 4 to which the large amount of refrigerator oil is circulated, too large an inner diameter di1 of the first refrigerantheat transfer pipe 42 may reduce the refrigerant flow speed, thereby reducing flowage of the refrigerator oil. This may significantly increase retention of refrigerator oil in thefirst gas cooler 4. For these reasons, it is desirable to set the inner diameter ratio di1 of the first refrigerantheat transfer pipe 42 in the first gas cooler to a value that is not too large. - As shown in
Figure 6 , the channel length of thefirst gas cooler 4 is about 10% of the channel length of the overall gas coolers. Thus, if the ratio of the refrigerant pressure loss of thefirst gas cooler 4 with respect to the refrigerant pressure loss of the overall gas coolers can be reduced to about 10%, it can be said that the refrigerant pressure loss of thefirst gas cooler 4 is sufficiently reduced. It can be also said that further reducing the refrigerant pressure loss of thefirst gas cooler 4, that is, reducing the refrigerant pressure loss per unit channel length in thefirst gas cooler 4 to be smaller than the refrigerant pressure loss per unit channel length in thesecond gas cooler 5 is an excess. As shown inFigure 15 , if the inner diameter ratio di1/di2 is about 1.4, the ratio of the refrigerant pressure loss of thefirst gas cooler 4 with respect to the refrigerant pressure loss of the overall gas coolers is about 10%. Thus, it can be said that setting the value of the inner diameter ratio di1/di2 to 1.4 sufficiently reduces the refrigerant pressure loss of thefirst gas cooler 4 in the relationship with the ratio of the channel length. However, too large a value of the inner diameter ratio di1/di2, that is, too large an inner diameter di1 of the first refrigerantheat transfer pipe 42 may cause the negative effects as described above such as an excessive length of thefirst twist pipe 41 or an increase in the retention of refrigerator oil in thefirst gas cooler 4. On the other hand, with the value of the inner diameter ratio di1/di2 of 1.4 or less, the inner diameter di1 of the first refrigerantheat transfer pipe 42 is not too large, thereby reliably preventing the negative effects. - In addition, the value of the inner diameter ratio di1/di2 of the first refrigerant
heat transfer pipe 42 and the second refrigerantheat transfer pipe 52 is preferably not less than 1.1, and more preferably not less than 1.2. Setting the value of the inner diameter ratio dil/di2 to preferably 1.1 or more, and more preferably 1.2 or more can more reliably reduce the refrigerant pressure loss of the first gas cooler 4 (seeFigure 13 ). In short, the value of the inner diameter ratio di1/di2 is preferably not less than 1.1 and not more than 1.4, and more preferably not less than 1.2 and not more than 1.4. By setting the value of the inner diameter ratio di1/di2 to such a range, markedly advantageously, the negative effects described above due to the excessive increase in the inner diameter di1 of the first refrigerantheat transfer pipe 42 can be reliably prevented, and the refrigerant pressure loss of thefirst gas cooler 4 can be sufficiently reduced. - As described above, according to
Embodiment 1, the refrigerant pressure loss of thefirst gas cooler 4 can be reliably prevented to reduce input power for thecompressor 3 and improve a COP. - As shown in
Figure 7 , the refrigerant density in thesecond gas cooler 5 is higher than the refrigerant density in thefirst gas cooler 4. As described above, with increasing refrigerant density, the refrigerant pressure loss per unit channel length is reduced. Thus, assuming other conditions are equal, the refrigerant pressure loss per unit length of the second refrigerantheat transfer pipe 52 in thesecond gas cooler 5 is smaller than the refrigerant pressure loss per unit length of the first refrigerantheat transfer pipe 42 in thefirst gas cooler 4. Thus, even if the inner diameter di2 of the second refrigerantheat transfer pipe 52 or the sectional area of each second refrigerant heat transfer channel in thesecond gas cooler 5 is smaller than the inner diameter di1 of the first refrigerantheat transfer pipe 42 or the sectional area of each first refrigerant heat transfer channel in thefirst gas cooler 4, the refrigerant pressure loss of thesecond gas cooler 5 can be sufficiently reduced. In addition, the inner diameter di2 of the second refrigerantheat transfer pipe 52 or the sectional area of each second refrigerant heat transfer channel in thesecond gas cooler 5 being relatively small increases the refrigerant flow speed in the second refrigerantheat transfer pipe 52, that is, in each second refrigerant heat transfer channel, thereby increasing a heat transfer coefficient of the refrigerant. This can reduce the length of thesecond twist pipe 51, that is, the second liquid heat transfer channel in the second gas cooler. From the above, the inner diameter di1 of the first refrigerantheat transfer pipe 42 or the sectional area of each first refrigerant heat transfer channel in thefirst gas cooler 4 is preferably larger than the inner diameter di2 of the second refrigerantheat transfer pipe 52 or the sectional area of each second refrigerant heat transfer channel in thesecond gas cooler 5. -
Figure 16 shows change in heat transfer coefficient on water side in a case where the twist pitch p of thefirst twist pipe 41 is equal to the twist pitch p2 of thesecond twist pipe 51 and the inner diameters SRi of thefirst twist pipe 41 and thesecond twist pipe 51 are equal. The axis of abscissa inFigure 16 refers to the same as the axis of abscissa inFigure 6 . InFigure 16 , the heat transfer coefficient on water side is represented by a ratio to the heat transfer coefficient on water side at the water outlet of thefirst gas cooler 4. As shown inFigure 16 , with increasing distance from the refrigerant inlet and the water outlet of thefirst gas cooler 4, that is, with decreasing temperature of water, the heat transfer coefficient on water side is reduced. Thus, if the twist pitch p of thefirst twist pipe 41 is equal to the twist pitch p2 of thesecond twist pipe 51, and the inner diameters SRi of thefirst twist pipe 41 and thesecond twist pipe 51 are equal, the heat transfer coefficient on water side in thesecond gas cooler 5 is lower than the heat transfer coefficient on water side in thefirst gas cooler 4. In this view, in thesecond gas cooler 5, it is desirable to set a relatively small twist pitch p2 of thesecond twist pipe 51 to increase a contact area between the second refrigerantheat transfer pipe 52 and thesecond twist pipe 51. This can reduce the length of thesecond twist pipe 51 in thesecond gas cooler 5. In contrast, as described above, the twist pitch p of thefirst twist pipe 41 in thefirst gas cooler 4 is desirably relatively large. From the above, the twist pitch p of thefirst twist pipe 41 in thefirst gas cooler 4 is preferably larger than the twist pitch p2 of thesecond twist pipe 51 in thesecond gas cooler 5. - In
Embodiment 1, the inner diameter SRi of thefirst twist pipe 41 in thefirst gas cooler 4 is preferably equal to the inner diameter SRi of thesecond twist pipe 51 in thesecond gas cooler 5. If thesecond gas cooler 5 is placed near thefirst gas cooler 4, an upstream end of thefirst twist pipe 41 is connected to a downstream end of thesecond twist pipe 51. In this case, the inner diameter SRi of thefirst twist pipe 41 being equal to the inner diameter SRi of thesecond twist pipe 51 allows easy connection between thefirst twist pipe 41 and thesecond twist pipe 51. In addition, the inner diameter SRi of thefirst twist pipe 41 being equal to the inner diameter SRi of thesecond twist pipe 51 allows material and a manufacturing method used for thefirst twist pipe 41 and thesecond twist pipe 51 to be shared, thereby reducing cost. - In
Embodiment 1, the number of the first refrigerant heat transfer pipe(s) 42, that is, the number of the first refrigerant heat transfer channel(s) in thefirst gas cooler 4 is preferably equal to the number of the second refrigerant heat transfer pipe(s) 52, that is, the number of the second refrigerant heat transfer channel(s) in thesecond gas cooler 5. The number of the first refrigerant heat transfer pipe(s) 42 being equal to the number of the second refrigerant heat transfer pipe(s) 52 allows thefirst twist pipe 41 and thesecond twist pipe 51 to be similarly designed, thereby reducing cost. - In
Embodiment 1, the case where the first heat exchanger (first gas cooler 4) and the second heat exchanger (second gas cooler 5) are the twist pipe type heat exchangers has been described as an example. However, in the present invention, the first heat exchanger and the second heat exchanger are not limited to the twist pipe type heat exchanger, but various types of heat exchangers may be used. - As described above, the value of the inner diameter ratio di1/di2 of the first refrigerant
heat transfer pipe 42 and the second refrigerantheat transfer pipe 52 is preferably not less than 1.1 and not more than 1.4, and more preferably not less than 1.2 and not more than 1.4. If the inner diameter ratio di1/di2 is 1.1, the ratio of the total sectional area of the first refrigerant heat transfer channels in the first heat exchanger to the total sectional area of the second refrigerant heat transfer channels in the second heat exchanger is (1.1)2 ≈ 1.2. If the inner diameter ratio di1/di2 is 1.2, the ratio of the total sectional area of the first refrigerant heat transfer channels in the first heat exchanger to the total sectional area of the second refrigerant heat transfer channels in the second heat exchanger is (1.2)2 ≈ 1.4. If the inner diameter ratio dil/di2 is 1.4, the ratio of the total sectional area of the first refrigerant heat transfer channels in the first heat exchanger to the total sectional area of the second refrigerant heat transfer channels in the second heat exchanger is (1.4)2 ≈ 2. Thus, if a numerical range of the inner diameter ratio di1/di2 is replaced by a numerical range of the ratio of the channel sectional area, it can be said that the ratio of the total sectional area of the first refrigerant heat transfer channels to the total sectional area of the second refrigerant heat transfer channels is preferably not less than 1.2 and not more than 2, and more preferably not less than 1.4 and not more than 2. The ratio of the channel sectional area within such a range provides advantages similar to those described above. - In
Embodiment 1 described above, the case where the number of the first refrigerant heat transfer channels in the first heat exchanger (first gas cooler 4) is equal to the number of the second refrigerant heat transfer channels in the second heat exchanger (second gas cooler 5) has been mainly described, however, in the present invention, the number of the first refrigerant heat transfer channels may be larger than the number of the second refrigerant heat transfer channel(s). If the number of the first refrigerant heat transfer channels is larger than the number of the second refrigerant heat transfer channel(s), the total sectional area of the first refrigerant heat transfer channels can be larger than the total sectional area of the second refrigerant heat transfer channels with a simple configuration. If the number of the first refrigerant heat transfer channels is larger than the number of the second refrigerant heat transfer channel(s), for example, the sectional area of the first refrigerant heat transfer channel may be equal to the sectional area of the second refrigerant heat transfer channel. This allows the first refrigerantheat transfer pipe 42 in thefirst gas cooler 4 and the second refrigerantheat transfer pipe 52 in thesecond gas cooler 5 to be made of a common material, thereby reducing cost. - In
Embodiment 1, the heat pump device for heating water using the first heat exchanger and the second heat exchanger has been described as an example, but in the present invention, the liquid heated by the first heat exchanger and the second heat exchanger is not limited to water, but for example, may be brine, antifreeze liquid, or the like. -
- 1
- heat pump device
- 1a
- water inlet
- 1b
- water outlet
- 2
- tank unit
- 2a
- hot water storage tank
- 2b
- water pump
- 2c
- hot-water supplying mixing valve
- 3
- compressor
- 4
- first gas cooler
- 5
- second gas cooler
- 6
- expansion valve
- 7
- evaporator
- 8
- fan
- 9
- high and low pressure heat exchanger
- 10, 11, 12, 17, 18, 19, 20, 21
- pipe
- 13
- water supply pipe
- 14
- hot water delivery pipe
- 15
- water supply branch pipe
- 16
- hot-water supply pipe
- 23, 26
- water channel
- 31
- sealed container
- 32
- compressing element
- 33
- electric actuating element
- 34
- first intake passage
- 35
- first discharge passage
- 36
- second intake passage
- 37
- second discharge passage
- 38
- internal space
- 41
- first twist pipe
- 42, 42a, 42b, 42c
- first refrigerant heat transfer pipe
- 50
- control unit
- 51
- second twist pipe
- 52
- second refrigerant heat transfer pipe
- 60
- heat transfer material
- 70
- heat pump device
- 71
- compressor
- 72
- gas cooler
- 331
- rotor
- 332
- stator
- 411, 411a, 411b, 411c, 511
- groove
Claims (9)
- A heat pump device(1) comprising:a compressor(3) including a first discharge passage(35) for discharging refrigerant and first refrigerator oil, and a second discharge passage(37) for discharging the refrigerant and second refrigerator oil having a mass flow rate lower than a mass flow rate of the first refrigerator oil;a first heat exchanger(4) including one or a plurality of first refrigerant heat transfer channels through which the refrigerant and the first refrigerator oil discharged from the first discharge passage(35) pass, and one or a plurality of first liquid heat transfer channels through which a liquid passes, heat exchange being performed between the first refrigerant heat transfer channel and the first liquid heat transfer channel; anda second heat exchanger(5) including one or a plurality of second refrigerant heat transfer channels through which the refrigerant and the second refrigerator oil discharged from the second discharge passage(37) pass, and one or a plurality of second liquid heat transfer channels through which the liquid passes, heat exchange being performed between the second refrigerant heat transfer channel and the second liquid heat transfer channel, characterized in thata total sectional area of the one or a plurality of first refrigerant heat transfer channels is larger than a total sectional area of the one or a plurality of second refrigerant heat transfer channels, whereby the total sectional area of the one or a plurality of first refrigerant heat transfer channels is a sum of sectional area of each of the first refrigerant heat transfer channels, and the total sectional area of the one or a plurality of second refrigerant heat transfer channels is a sum of sectional area of each of the second refrigerant heat transfer channels, whereby the sectional area is in a section perpendicular to a flow direction,the first refrigerant heat transfer channel is formed of a first refrigerant heat transfer pipe(42, 42a, 42b, 42c),the second refrigerant heat transfer channel is formed of a second refrigerant heat transfer pipe(52),the first liquid heat transfer channel is formed of a first twist pipe(41) having a helical groove(411, 411a, 411b, 411c) in an outer periphery,the second liquid heat transfer channel is formed of a second twist pipe(51) having a helical groove(511) in an outer periphery,the first refrigerant heat transfer pipe(42, 42a, 42b, 42c) is located along the groove(411, 411a, 411b, 411c) in the first twist pipe(41),the second refrigerant heat transfer pipe(52) is located along the groove(511) in the second twist pipe(51),a twist pitch of the first twist pipe(41) is larger than a twist pitch of the second twist pipe(51),wherein when p is a twist pitch of the first twist pipe(41) and SRi is an inner diameter of the first twist pipe(41), p/SRi is not less than 1.1 and not more than 1.8.
- The heat pump device(1) according to claim 1, wherein the number of the first refrigerant heat transfer channels is equal to the number of the second refrigerant heat transfer channels, and
a sectional area of each of the one or a plurality of first refrigerant heat transfer channels is larger than a sectional area of each of the one or a plurality of second refrigerant heat transfer channels. - The heat pump device(1) according to claim 1 or 2, wherein a ratio of an inner diameter of the first refrigerant heat transfer pipe(42, 42a, 42b, 42c) to an inner diameter of the second refrigerant heat transfer pipe(52) is not less than 1.1 and not more than 1.4.
- The heat pump device(1) according to any one of claims 1 to 3, wherein an inner diameter of the first twist pipe(41) is equal to an inner diameter of the second twist pipe(51).
- The heat pump device(1) according to claim 1, wherein the number of the one or a plurality of first refrigerant heat transfer channels is larger than the number of the second refrigerant heat transfer channels.
- The heat pump device(1) according to any one of claims 1 to 5, wherein a ratio of the total sectional area of the one or a plurality of first refrigerant heat transfer channels to the total sectional area of the one or a plurality of second refrigerant heat transfer channels is not less than 1.2 and not more than 2.
- The heat pump device(1) according to any one of claims 1 to 6, wherein a ratio of the mass flow rate of the first refrigerator oil with respect to a sum of the mass flow rate of the refrigerant and the mass flow rate of the first refrigerator oil in the first heat exchanger(4) is not less than 2% and not more than 20%.
- The heat pump device(1) according to any one of claims 1 to 7, wherein a ratio of the mass flow rate of the second refrigerator oil with respect to a sum of the mass flow rate of the refrigerant and the mass flow rate of the second refrigerator oil in the second heat exchanger(5) is not less than 0.01% and not more than 1%.
- The heat pump device(1) according to any one of claims 1 to 8, wherein the liquid is water, and the heat pump device(1) has a function of supplying hot water obtained by heating the water.
Applications Claiming Priority (1)
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PCT/JP2013/066313 WO2014199479A1 (en) | 2013-06-13 | 2013-06-13 | Heat pump device |
Publications (3)
Publication Number | Publication Date |
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EP3009767A1 EP3009767A1 (en) | 2016-04-20 |
EP3009767A4 EP3009767A4 (en) | 2017-01-25 |
EP3009767B1 true EP3009767B1 (en) | 2020-12-09 |
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EP13886812.0A Active EP3009767B1 (en) | 2013-06-13 | 2013-06-13 | Heat pump device |
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EP (1) | EP3009767B1 (en) |
JP (1) | JP6075451B2 (en) |
WO (1) | WO2014199479A1 (en) |
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CN107429948B (en) * | 2015-03-19 | 2019-12-13 | 三菱电机株式会社 | heat pump system |
EP3370026B1 (en) * | 2017-01-18 | 2019-06-05 | Mitsubishi Electric Corporation | Twisted pipe heat exchanger |
JP7199842B2 (en) * | 2018-06-15 | 2023-01-06 | 三菱重工サーマルシステムズ株式会社 | water heat exchanger, gas cooler |
JP7113210B2 (en) * | 2018-12-17 | 2022-08-05 | パナソニックIpマネジメント株式会社 | heat pump system |
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JPS512325Y1 (en) * | 1970-11-07 | 1976-01-23 | ||
JPS5224805Y2 (en) * | 1971-10-05 | 1977-06-06 | ||
JPS512325B2 (en) * | 1971-11-19 | 1976-01-24 | ||
JP3477531B1 (en) * | 2002-07-04 | 2003-12-10 | 太平洋精工株式会社 | Heat exchanger and method for producing the same, and bath water heating system and floor heating system using such heat exchanger |
JP3877207B2 (en) | 2002-09-13 | 2007-02-07 | 株式会社前川製作所 | Hot water supply system for CO2 refrigeration cycle |
JP2004205077A (en) * | 2002-12-24 | 2004-07-22 | Sapporo Holdings Ltd | Liquid cooling device using flat pipe and liquid cooling supply device using the same |
JP2010091266A (en) * | 2004-08-26 | 2010-04-22 | Mitsubishi Electric Corp | Twisted tube type heat exchanger |
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JP4434924B2 (en) * | 2004-11-05 | 2010-03-17 | 三菱電機株式会社 | Compressor and hot water supply cycle device |
JP4200329B2 (en) * | 2005-06-06 | 2008-12-24 | パナソニック株式会社 | Heat exchange device and heat pump water heater using the same |
JP2008190787A (en) * | 2007-02-05 | 2008-08-21 | Furukawa Electric Co Ltd:The | Spiral tube and heat exchanger using the same |
JP2008241217A (en) * | 2007-03-29 | 2008-10-09 | Mitsubishi Electric Corp | Method of manufacturing heat exchanger, and heat exchanger manufactured by this manufacturing method |
JP2008309361A (en) * | 2007-06-12 | 2008-12-25 | Panasonic Corp | Refrigerating cycle device |
JP2009041880A (en) * | 2007-08-10 | 2009-02-26 | Sumitomo Light Metal Ind Ltd | Water heat exchanger for water heater |
JP4819765B2 (en) * | 2007-08-22 | 2011-11-24 | 三菱電機株式会社 | Method for manufacturing twisted tube heat exchanger |
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JP5171280B2 (en) * | 2008-01-18 | 2013-03-27 | 日立アプライアンス株式会社 | Heat exchanger and heat pump type water heater using the same |
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JP2013088045A (en) * | 2011-10-19 | 2013-05-13 | Hitachi Appliances Inc | Heat exchanger and heat pump type water heater using the same |
JP5794952B2 (en) * | 2012-06-13 | 2015-10-14 | 三菱電機株式会社 | Twisted tube heat exchanger |
-
2013
- 2013-06-13 EP EP13886812.0A patent/EP3009767B1/en active Active
- 2013-06-13 WO PCT/JP2013/066313 patent/WO2014199479A1/en active Application Filing
- 2013-06-13 JP JP2015522338A patent/JP6075451B2/en active Active
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JPWO2014199479A1 (en) | 2017-02-23 |
WO2014199479A1 (en) | 2014-12-18 |
EP3009767A4 (en) | 2017-01-25 |
EP3009767A1 (en) | 2016-04-20 |
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