JP6177195B2 - Heat transfer tube for supercooled double tube heat exchanger - Google Patents

Description

  The present invention relates to a heat transfer tube that is installed on the outlet side of a condenser of an air conditioner and is incorporated in a double-pipe heat exchanger that is used to supercool the refrigerant of the condenser. For a supercooled double-tube heat exchanger suitable for heat exchange between a single-phase fluid flowing through the ring-shaped annular part and a fluid (single-phase flow or phase-changing two-phase flow) flowing in the inner tube at the center of the tube It relates to heat transfer tubes.

  An air conditioner is described in Patent Document 1, for example. In this air conditioner, as shown in FIG. 7, the refrigerant that has exited the condenser 1 flows through the main circuit F <b> 1 and the bypass circuit F <b> 2 and returns to the condenser 1. In the main circuit F1, the refrigerant exiting the condenser 1 passes through the branch point A1, and has a ring-like annular shape between the outer tube and the inner tube of the double-tube heat transfer tube incorporated in the supercooling heat exchanger 2. And then the pressure is reduced through the expansion valve 3 before entering the evaporator 4. Then, the refrigerant that has exited the evaporator 4 enters the compressor 5 through the junction A <b> 2, and the refrigerant that has exited the compressor 5 returns to the condenser 1. On the other hand, in the bypass circuit F2, the refrigerant exiting the condenser 1 is depressurized from the branch point A1 through the expansion valve 6, and then the inside of the inner tube of the double-tube heat transfer tube incorporated in the supercooling heat exchanger 2 Flowing. Thereafter, the refrigerant enters the compressor 5 through the junction A <b> 2, and the refrigerant exiting the compressor 5 returns to the condenser 1.

  In this air conditioner, the refrigerant condensed in the condenser 1 mainly flows through the main circuit F1, evaporates in the evaporator 4, and is then compressed in the compressor 5. On the other hand, a part of the refrigerant condensed in the condenser 1 flows through the bypass circuit F2 and is expanded and depressurized by the expansion valve 6 to be cooled into a liquid phase and a gas phase two-phase state. In the supercooling heat exchanger 2, the refrigerant supercools the refrigerant flowing through the main circuit F1. As a result, the liquid refrigerant on the main circuit F1 side is cooled to lower the temperature, and flows into the evaporator 4 through the expansion valve 3 of the main circuit F1. For this reason, the enthalpy difference between the refrigerant that flows into the inlet of the evaporator 4 and the refrigerant that flows out of the outlet of the evaporator 4, that is, the latent heat of vaporization increases, and the refrigerating capacity increases. Thereby, the refrigerating capacity of the air conditioner by the refrigerant flowing through the main circuit F1 can be improved. Further, by flowing a part of the refrigerant to the bypass circuit F2, the refrigerant flow rate of the main circuit F1 is reduced, and the pressure loss at the inlets of the evaporator 4 and the compressor 5 can be reduced. As described above, by using the double pipe heat exchanger to supercool the refrigerant, high performance and energy saving of the refrigeration capacity of the air conditioner are achieved.

  In the heat transfer tube used in this supercooled double tube heat exchanger, a low-viscosity fluid is used as the refrigerant flowing through the annular portion of the double tube heat transfer tube, but the conventional double tube heat transfer tube is When the flow rate of the low viscosity fluid is large, there is a problem that the heat transfer coefficient is low.

  The smooth tube has a smooth tube surface, and when a fluid flows through the tube, a velocity boundary layer and a temperature boundary layer are formed on the wall surface of the tube, and this boundary layer inhibits heat exchange between the fluids. Is low.

  In Patent Document 2, a wedge-shaped protrusion is provided on the outer surface of the tube, and a number of spiral grooves are provided on the inner surface of the tube to promote turbulent flow, thereby stirring the fluid, A heat transfer tube that suppresses the formation of a temperature boundary layer is disclosed.

  In Patent Document 3, a spiral fin is provided on the outer surface of the tube, and a number of spiral grooves are provided on the inner surface of the tube to promote turbulence. A heat transfer tube for a subcooler in which formation of a layer and a temperature boundary layer is suppressed is disclosed.

Japanese Patent Laid-Open No. 10-54616 JP-A 61-265499 JP 2013-79763 A

  However, in the heat transfer tube disclosed in Patent Document 2, the lead angle of many grooves formed in the tube is as small as 16 to 35 °. Therefore, when used in a supercooling heat exchanger, the fluid in the tube is outside the tube. Before exchanging heat with the other fluid, the fluid exits from the outlet of the heat exchanger, and there is a limit to improving the heat exchange performance. In addition, the wedge-shaped protrusion formed on the outer surface of the tube contributes to the performance improvement on the outer side of the tube, and when a low-viscosity fluid is used outside the tube (annular part), the turbulence is promoted, When the flow rate of the annular portion is large, a separation flow is generated at the front end portion of the wedge-shaped protrusion, which causes a problem that pressure loss increases and heat transfer performance is saturated. Further, when this heat transfer tube is used in a supercooling heat exchanger, there is a problem that pressure loss increases and refrigerant transport power on the refrigeration cycle side, that is, power of the compressor increases.

  Further, the heat transfer tube disclosed in Patent Document 3 has a small lead angle of 12 to 20 ° formed on the inner surface of the tube, so that when the tube is used in a supercooled heat exchanger, the fluid in the tube Before the heat exchange with the fluid outside the tube, the heat exchanger exits from the outlet of the heat exchanger, and there is a problem that there is a limit to the improvement of the heat exchange performance. Also, the fin formed on the pipe outer surface contributes to the performance improvement on the pipe outer surface side, and the turbulent flow is promoted when a low-viscosity fluid is used. When the separation flow occurs in the portion, the pressure loss increases and the heat transfer performance is saturated. Further, when this heat transfer tube is used in a supercooling heat exchanger, there is a problem that the power of the compressor, which is the refrigerant conveyance power on the refrigeration cycle side, increases due to an increase in pressure loss.

  The present invention has been made in view of such problems, and in the case where a low-viscosity fluid is caused to flow at a high flow rate through the annular portion of the double-pipe heat exchanger, the heat transfer coefficient is improved and the annular portion fluid is improved. An object of the present invention is to provide a heat transfer tube for a supercooled double-pipe heat exchanger that can reduce pressure loss.

The heat transfer tube for a supercooled double tube heat exchanger according to the present invention is:
A refrigerant composed of a low-viscosity single-phase flow fluid having a viscosity coefficient of 350 μPa · s or less flows through a tubular portion between the outer tube and the inner tube, and a single-phase fluid or gas-liquid two-phase flow composed of a liquid flows through the inner tube. In the heat transfer tube used for the inner tube of the supercooled double-tube heat exchanger for flowing a refrigerant consisting of fluid,
A plurality of protrusions that are formed in a rectangular shape on the outer surface of the tube and aligned in the tube axis direction and the tube circumferential direction, forming a trapezoidal trapezoidal shape;
One or more ribs spirally formed on the inner surface of the tube;
Have
The protrusion has a bottom width of the first groove between the protrusions adjacent to each other in the tube axis direction of 0.3 to 0.80 mm, and a bottom width of the second groove between the protrusions adjacent in the tube circumferential direction is 0. 0 mm. 10 to 0.30 mm,
The rib has a width of the bottom of the groove between the ribs adjacent to each other in the tube axis direction of 0.15 to 1.10 mm,
The lead angle, which is the angle formed with respect to the tube axis direction of the rib, is 40 to 65 °, and the peak angle of the rib in the tube axis direction is 55 to 110 °.

  This heat transfer tube is used as an inner tube of a double tube heat exchanger. Therefore, the fluid flowing through the ring-shaped annular portion of the double tube is in contact with the outer surface of the heat transfer tube, and usually this fluid is a single-phase flow fluid, and the double tube heat exchanger is disposed on the inner surface of the heat transfer tube. The fluid flowing in the inner pipe is in contact and is usually a single-phase flow or a two-phase flow in which the phase changes. And heat exchange is performed between the fluid which contacts the inner surface and outer surface of this heat exchanger tube. The outer tube of the double tube heat exchanger defines an annular portion for allowing fluid to flow in a ring-shaped annular portion between the outer tube and the inner tube.

In this supercooled double-tube heat exchanger heat transfer tube, preferably,
The formation pitch of the rib in the tube axis direction is 0.60 to 1.24 mm,
The height of the rib is 0.12 to 0.35 mm.

Also preferably,
The height of the protrusion is 0.18 to 0.50 mm,
The formation pitch of the projections in the tube axis direction is 0.65 to 1.15 mm,
The pitch in the tube circumferential direction of the second groove between the projections in the tube circumferential direction is 0.51 to 1.04 mm, and the peak angle in the tube axis direction of the projection is 4 to 65 °. The peak angle in the direction perpendicular to the tube axis is 25 to 75 °.

Furthermore, preferably,
The cross-sectional shapes of the first groove between the protrusions in the tube axis direction and the second groove between the protrusions in the tube circumferential direction are inverted trapezoids.

  ADVANTAGE OF THE INVENTION According to this invention, the heat transfer rate of a heat exchanger tube can be improved in the heat exchanger tube for supercooled double tube type heat exchangers which flow a low-viscosity fluid at a high flow rate to the annular part of a double tube type heat exchanger. At the same time, the pressure loss of the annular fluid can be reduced.

It is a plane exploded view which shows the protrusion 31 of the outer surface of the heat exchanger tube which concerns on embodiment of this invention. (A) is sectional drawing (partially perspective view) of the tube axis direction of the heat exchanger tube of embodiment similarly, (b) is sectional drawing of the tube axis perpendicular direction of a heat exchanger tube. It is a schematic diagram which shows the flow of the fluid of an outer surface. It is a schematic diagram which similarly shows the flow of the fluid of an outer surface. It is a figure which shows a groove shape. It is a test device for heat exchange and pressure loss. It is a circuit diagram which shows a supercooled double tube type heat exchanger.

  Hereinafter, embodiments of the present invention will be specifically described with reference to the accompanying drawings. The heat transfer tube of the embodiment of the present invention is a heat transfer tube serving as an inner tube incorporated in a supercooled double tube heat exchanger. An outer tube is concentrically disposed outside the inner tube to form a double tube. And the fluid which flows through the ring-shaped annular part of a double tube contacts the outer surface of the heat transfer tube of the inner tube of this double tube, and this fluid is usually a single-phase flow fluid. In addition, the fluid flowing in the inner pipe of the double pipe heat exchanger is in contact with the inner surface of the heat transfer pipe of the inner pipe of the double pipe, and this fluid is usually a single-phase flow or a two-phase flow in which the phase changes. . And heat exchange is performed between the fluid which contacts the inner surface and outer surface of this heat exchanger tube. The outer tube of the double tube heat exchanger defines a ring-shaped annular portion between the outer tube and the inner tube, and the inner surface of the outer tube defines the outer surface of the annular portion through which the fluid flows. To do.

  FIG. 1 shows the outer surface of the heat transfer tube of this inner tube, and FIG. 2 is a cross-sectional view (partially perspective view) along the tube axis of this heat transfer tube and a cross-sectional view in the direction perpendicular to the tube axis. The heat transfer tube 11 of this embodiment is an inner tube of a double tube heat exchanger, and the outer tube 12 is arranged outside the heat transfer tube of the inner tube concentrically with the heat transfer tube 11. A space between the inner surface 15 of the outer tube 12 and the outer surface 13 of the inner tube (heat transfer tube 11) becomes a ring-shaped annular portion, for example, a single-phase fluid flows. The space surrounded by the inner surface 14 of the inner tube (heat transfer tube 11) is the center of the double tube, and for example, a single-phase fluid or a two-phase fluid that changes in phase flows.

  On the outer surface 13 of the heat transfer tube 11, a plurality of protrusions 31 that are rectangular in plan view are formed. The protrusion 31 has a trapezoidal trapezoidal shape with a height h1. The protrusions 31 are spirally arranged on the outer surface of the tube, and thus are formed at positions aligned in the circumferential direction and the tube axis direction. A first groove 33 is formed between the projections 31 adjacent to each other in the tube axis direction, and the first groove 33 is continuous in the tube circumferential direction. Further, a second groove 34 is formed between the protrusions 31 adjacent to each other in the pipe circumferential direction, and the second groove 34 is continuous in the pipe axis direction. However, since the protrusions 31 are spirally arranged on the outer peripheral surface of the heat transfer tube 11, the first groove 33 does not extend strictly in the tube circumferential direction, but is slightly inclined from the tube circumferential direction. . Similarly, strictly speaking, the second groove 34 does not extend in the tube axis direction, but is slightly inclined from the tube axis direction. The height of the protrusion 31 is h1, and the depth of the grooves 33 and 34 is h2. The pitch of the first groove 33 in the direction perpendicular to the extending direction of the first groove 33 is P2, and the pitch of the second groove 34 in the direction perpendicular to the extending direction of the second groove 34 is P1. However, the pitch P1 is also the pitch in the tube circumferential direction of the protrusions 31, and the pitch P2 is also the pitch in the tube axis direction of the protrusions 31. The width of the first groove 33 between the protrusions 31 is 0.3 to 0.80 mm, and the width of the second groove is 0.10 to 0.30 mm. This range is appropriate as the width of the groove through which the fluid flows.

  Further, a rib 32 extending in a spiral shape is formed on the inner surface of the heat transfer tube 11, and a groove 35 is formed between adjacent ribs 32. The height of the rib 32 is h3, the angle (lead angle) formed with respect to the tube axis direction of the rib 32 is θ, and the pitch of the rib 32 in the tube axis direction is P3. The maximum diameter of the inner surface 14 of the heat transfer tube 11 (the diameter of the peripheral surface formed from the bottom surface of the groove 35) is D. The width of the groove 35 between the ribs 32 is 0.15 to 1.10 mm. Similarly, the width of the groove through which the fluid flows is appropriate.

  Next, operation | movement of the heat exchanger tube comprised in this way is demonstrated. In a supercooled double-pipe heat exchanger, a refrigerant composed of a low-viscosity single-phase fluid flows through the tubular part, and a refrigerant composed of a gas-liquid two-layer fluid flows through the inner pipe. Heat exchange. The refrigerant flowing in the inner pipe is a gas-liquid two-phase fluid that takes the heat of the single-phase fluid flowing in the tubular portion and cools it, and the liquid refrigerant flowing in the inner pipe is evaporated by receiving heat. And vaporize. Therefore, a single-phase fluid made of liquid or a two-phase fluid made of liquid and gas flows in the inner pipe. In this case, when the heat exchange is completely performed in the double pipe, all the liquid in the inner pipe is changed to gas. Therefore, when the refrigerant that comes out of the inner pipe is not a gas in all amounts but a liquid remains and is a two-phase fluid of a gas and a liquid, heat exchange is insufficient. In the heat transfer tube according to the present invention, the protrusion 31 is provided in a spiral shape on the outer surface of the heat transfer tube 11 of the inner tube, and the rib 32 is provided in a spiral shape on the inner surface, so that the shape factors of the protrusion 31 and the rib 32 are appropriately defined. For this reason, the refrigerant flowing in the inner tube is entirely converted into gas by heat exchange with the refrigerant flowing in the tubular portion, and therefore, the liquid remains in the refrigerant that comes out of the inner tube without heat exchange. Can be avoided.

  An example of the low-viscosity fluid is a chlorofluorocarbon refrigerant. Typical examples include Freon R410A, R32, R134a, R245fa, etc. used for air conditioning. The viscosities of various liquid fluids are shown in Table 1 below.

  CFC refrigerant has a smaller viscosity coefficient than water, and even R245fa, which has a larger viscosity coefficient than CFC, has a viscosity coefficient of about 53% smaller than that of water. The present invention is applied to a supercooled double-pipe heat exchanger in which a low-viscosity fluid is caused to flow through a tubular portion. However, the low-viscosity fluid is not limited to chlorofluorocarbon, and is generally a fluid having a lower viscosity coefficient than water. A fluid having a flow rate of 350 μPa · s or less is used.

  As the material of the heat transfer tube, a metal material such as copper, copper alloy, aluminum, aluminum alloy, iron, or stainless steel is used. In particular, copper or a copper alloy is suitable because of its high thermal conductivity. Also, by using a high-strength copper alloy, a heat transfer tube can be constructed using a thin material, so even if the pressure on the tubular part side is high, a thin material can be used, and the use of the material The amount is reduced, the heat resistance in heat exchange is reduced, and the heat transfer performance is further improved.

“Lead angle θ of inner rib: 40 to 65 °”
In the present invention, the lead angle θ of the inner rib 32 is 40 to 65 °. When the lead angle θ of the inner surface rib 32 is less than 40 °, the length of the groove 35 on the inner surface of the tube is shortened, the amount of refrigerant in the liquid phase in the tube is reduced, and the residence time of the refrigerant is shortened, In the case of a heat transfer tube having a short length or a heat exchanger having a short length, the liquid refrigerant is not completely vaporized, and the heat exchange efficiency is lowered. When the lead angle θ of the inner surface rib is larger than 65 °, the length of the groove 35 on the inner surface of the tube becomes too long, the amount of refrigerant in the two-phase state in the tube increases, and the residence time becomes longer. For this reason, the flow rate of the single-phase flow fluid flowing through the tubular portion which is the main circuit is relatively reduced. That is, the refrigerant exiting the condenser 1 is separated into a main circuit and a bypass circuit, but when the amount of fluid refrigerant flowing through the bypass circuit in the inner pipe increases, the fluid refrigerant of the main circuit flowing relatively through the tubular portion becomes relatively large. The amount of decreases. For this reason, the lead angle θ of the inner rib 32 is set to 40 to 65 °. By setting the lead angle theta in this range, the length of the groove 35 between the ribs 32 formed on the inner surface of the heat transfer tube becomes appropriate, and the amount of refrigerant in the liquid phase state on the inner surface of the tube is sufficiently large and appropriate. In addition, the residence time of the refrigerant on the inner surface of the heat transfer tube is sufficiently long, the heat efficiency is high, the heat efficiency is high, and the liquid refrigerant is completely vaporized and heat exchange even with a short heat transfer tube or a short heat exchanger Efficiency is improved.

“Crest angle in the tube axis direction of the rib: 55 to 110 °”
In the present invention, the crest angle in the tube axis direction of the rib is 55 to 110 °. When the crest angle of the inner rib is smaller than 55 °, the groove between the ribs is narrowed and the liquid refrigerant becomes difficult to be held in the groove, and the heat transfer tube having a short length or the heat exchanger having a short length is formed in the heat transfer tube. This makes it difficult for the liquid refrigerant to completely evaporate, making it difficult to improve the heat exchange efficiency. When the crest angle of the inner surface rib is larger than 110 °, the groove portion between the ribs is widened, the liquid refrigerant is difficult to be held, and the liquid phase refrigerant in the groove portion easily flows in the longitudinal direction of the groove. If the heat transfer tube has a short length or the heat exchanger has a short length, the liquid refrigerant in the tube is not easily vaporized, and the heat exchange efficiency is difficult to improve. For this reason, the peak angle in the tube axis direction of the rib is 55 to 110 °. By setting the crest angle of the inner rib within this range, the liquid refrigerant is easily held in the groove between the ribs, and the residence time becomes longer, and the heat transfer tube having a shorter length or the heat exchange having a shorter length is used. Even in the oven, the liquid refrigerant is more completely vaporized and the heat exchange efficiency is improved.

“Formation pitch P3 of the inner rib in the tube axis direction: 0.60 to 1.24 mm”
Pipe pitch P3 of inner rib: 0.60 to 1.24 mm "
By setting the pipe pitch P3 of the inner ribs 32 in the range of 0.60 to 1.24 mm, the liquid phase refrigerant is easily held in the inner groove 35, and the residence time becomes longer. For this reason, even in a heat transfer tube having a short length (or a heat exchanger having a short length), the liquid phase refrigerant is completely vaporized in the heat transfer tube, and high heat exchange efficiency can be obtained. When the pitch P3 of the inner surface ribs 32 is larger than 1.24 mm, the grooves 35 between the ribs are likely to be widened, and the liquid refrigerant is not easily held in the grooves 35. Further, the liquid refrigerant in the groove 35 tends to flow in the longitudinal direction of the groove 35, the residence time is shortened, and the liquid refrigerant is transferred in the case of a heat transfer tube (or a heat exchanger with a short length). It becomes difficult to vaporize completely in the heat pipe, and the heat exchange efficiency is lowered. On the other hand, when the pitch P3 of the inner surface ribs 32 is less than 0.60 mm, the width of the grooves 35 between the ribs 32 tends to be narrow, and it becomes difficult to hold the liquid-phase refrigerant in the grooves 35. In the case of a heat exchanger having a short length), it becomes difficult for the liquid refrigerant to completely evaporate in the heat transfer tube, and the heat exchange efficiency is lowered. Therefore, the pitch P3 in the tube axis direction of the inner ribs 32 is preferably 0.60 to 1.24 mm.

“Inner rib height h3: 0.12 to 0.35 mm”
By setting the height h3 of the inner rib 32 in the range of 0.12 to 0.35 mm, the liquid phase refrigerant can be easily held in the groove 35, and the heat exchanger tube (or the heat exchanger having a short length) has a short length. ), The liquid refrigerant is more easily vaporized and the heat exchange efficiency is increased. When the height h3 of the inner rib 32 is smaller than 0.12 mm, the cross-sectional area of the groove 35 is reduced, it becomes difficult to hold the liquid-phase refrigerant in the groove 35, and the residence time is shortened. In the case of a heat pipe (or a heat exchanger having a short length), the liquid refrigerant is not easily vaporized, and the heat exchange efficiency is lowered. On the other hand, if the inner surface rib height h3 is higher than 0.35 mm, the amount of liquid phase refrigerant flowing in the pipe becomes excessive, the retained film thickness of the liquid phase refrigerant on the pipe inner surface is increased, and resistance to heat conduction is increased. Therefore, the amount of heat exchange is reduced. Therefore, the height h3 of the inner surface rib 32 is preferably 0.12 to 0.35 mm.

“Height of outer protrusion h1: 0.18 to 0.50 mm”
As shown in FIG. 3, the fluid F1 flows through the second groove 34 toward the downstream side. At this time, the refrigerant tends to flow from the second groove 34 between the upstream protrusions 31 to the second groove 34 between the downstream protrusions 31, but flows through the second groove 34 between the upstream protrusions 31. A part of the refrigerant flows into the second groove 34 between the protrusions 31 on the downstream side as a fluid F1, and the remaining part flows into the first groove 33 as fluids F2 and F3. On the other hand, part of the fluids F4 and F5 of the refrigerant flowing between the first grooves 33 flows between the second grooves 34 between the protrusions 31 on the downstream side. Therefore, the fluids F1, F4, and F5 flow through the second groove 34 between the protrusions 31 on the downstream side. The refrigerant flowing through the tubular portion between the outer surface of the heat transfer tube (inner tube) 11 and the outer tube 12 flows in the first groove 33 and the second groove 34 between the protrusions 31 in this way, and passes through the heat transfer tube 11. Heat exchange is performed with the flowing refrigerant.

  At this time, by setting the height h1 of the outer surface protrusion 31 of the heat transfer tube 11 to 0.18 to 0.50 mm, the fluidity of the fluid flowing as described above is improved. Since the fluidity of the refrigerant fluid is high, the heat exchange performance of the heat transfer tube 11 is improved and the pressure loss of the refrigerant is reduced.

  When the projection height h1 is higher than 0.50 mm, when a single-phase flow is caused to flow parallel to the tube axis direction, the flow velocity of the refrigerant flowing in the second groove 34 parallel to the tube axis direction further increases. When the fluid in the groove 33 flows into the second groove 34 on the downstream side, the second groove 34 becomes a resistance, and as a result, the pressure loss increases. That is, in the second groove 34, when the height of the protrusion 31 is high when crossing the first groove 33 from the portion between the upstream protrusions 31 and moving to the portion between the downstream protrusions 31, 31 becomes resistance and pressure loss increases. On the other hand, when the height h1 of the outer surface protrusion 31 is lower than 0.18 mm, the flow velocity of the fluid flowing through the second groove 34 is reduced and the upstream protrusion is caused when a single-phase flow is made parallel to the tube axis direction. The fluid flow input when the fluid flows from the second groove 34 between 31 to the first groove 33 and the second groove 34 between the protrusions 31 on the downstream side decreases. As a result, the fluidity of the refrigerant decreases, so that the heat exchange performance decreases. Accordingly, the projection height h1 is preferably 0.18 to 0.50 mm.

“Pitch P2 in the tube axis direction of the outer protrusion: 0.65 to 1.15 mm”
By setting the formation pitch P2 of the outer surface projections 31 in the tube axis direction to a range of 0.65 to 1.15 mm, when a single-phase flow is caused to flow in parallel with the tube axis direction, the second pitch between the upstream projections 31 is increased. When the fluid flows from the groove 34 to the second groove 34 between the first groove 33 and the downstream projection 31, the inflow of refrigerant from the first groove 33 to the second groove 34 between the downstream projection 31 is promoted. Thus, the stagnation of the fluid in the first groove 33 is suppressed. Thereby, the fluidity of the refrigerant is improved, so that an increase in pressure loss is suppressed, and heat exchange is further promoted by the inflow and outflow of the fluid.

  When the pitch P2 of the outer protrusions 31 is larger than 1.15 mm, when a single-phase flow is caused to flow in parallel with the tube axis direction, the first grooves 33 and the downstream grooves are separated from the second grooves 34 between the upstream protrusions 31. When the fluid flows in the second groove 34 between the protrusions 31, the refrigerant flow input from the first groove 33 to the second groove 34 between the protrusions 31 on the downstream side decreases, and the fluid in the first groove 33 stagnates. By doing so, the fluidity of the refrigerant is lowered and heat exchange is hindered. On the other hand, when the pitch P2 of the outer surface protrusions 31 is smaller than 0.65 mm, the first groove 33 and the downstream are formed from the second groove 34 between the upstream protrusions 31 when a single-phase flow is made parallel to the tube axis direction. When the fluid flows in the second groove 34 between the projections 31 on the side, the inflow of the refrigerant from the first groove 33 to the second groove 34 between the projections 31 on the downstream side is inhibited, and the fluid in the groove is The refrigerant flows in the outer peripheral direction, and the fluidity of the fluid in the groove decreases, so that the pressure loss of the refrigerant increases. Accordingly, the protrusion pitch P2 is preferably set to 0.65 to 1.15 mm.

"Pitch P1: 0.51 to 1.04 mm in the circumferential direction of the outer protrusions"
By setting the formation pitch P1 of the outer circumferential protrusions 31 in the pipe circumferential direction in the range of 0.51 to 1.04 mm, when a single-phase flow is caused to flow in parallel with the pipe axis direction, the second pitch between the upstream projections 31 is increased. When the fluid flows from the groove 34 to the second groove 34 between the first groove 33 and the downstream projection 31, the inflow of refrigerant from the first groove 33 to the second groove 34 between the downstream projection 31 is promoted. Thus, the stagnation of the fluid in the first groove 33 is suppressed. Thereby, the fluidity of the refrigerant is improved, so that an increase in pressure loss is suppressed, and heat exchange is further promoted by the inflow and outflow of the fluid.

  When the pitch P1 of the outer protrusions 31 is larger than 1.04 mm, when a single-phase flow is caused to flow in parallel with the tube axis direction, the first grooves 33 and the downstream grooves are separated from the second grooves 34 between the upstream protrusions 31. When the fluid flows in the second groove 34 between the protrusions 31, the refrigerant flow input from the first groove 33 to the second groove 34 between the protrusions 31 on the downstream side decreases, and the fluid in the first groove 33 stagnates. By doing so, the fluidity of the refrigerant is lowered and heat exchange is hindered. On the other hand, when the pitch P1 of the outer surface protrusions 31 is smaller than 0.51 mm, the flowability of the refrigerant in the tube axis direction is improved when a single-phase flow is made parallel to the tube axis direction, but the inside of the first groove 33 is increased. Before the fluid contributes to heat exchange, it tends to flow downstream, so heat exchange is hindered. Accordingly, the protrusion pitch P1 is preferably 0.51 to 1.04 mm.

“Crest angle in the tube axis direction of the external projection: 4 to 65 °”
By setting the peak angle of the protrusion in the tube axis direction within this range, when a single-phase flow flows parallel to the tube axis direction, the single-phase flow flows substantially parallel to the groove 2 and the fluid reaches the groove 1. When the fluid flows into the groove portion 1 from the groove portion 2, the fluid easily flows into the groove bottom portion. As a result, the stagnation of the fluid is suppressed and the fluidity of the fluid is improved, so that an increase in pressure loss is suppressed and heat exchange is further promoted by the inflow and outflow of the fluid.

  When the peak angle _ in the tube axis direction of the projection is smaller than 4 °, the fluid does not easily flow into the groove 1, the fluid easily flows in the outer circumferential direction of the tube, and the fluid in the groove 1 stagnates. When the exchange is hindered and incorporated in the double pipe, the fluid flows to the inner peripheral side of the outer pipe of the annular part, and thereafter, the fluid on the inner peripheral side of the outer pipe tends to flow into the groove part 1 with the protrusion on the outer surface of the inner pipe as a base point. As a result, the fluidity of the fluid is hindered and the pressure loss increases.

  When the peak angle_ of the protrusion in the tube axis direction is larger than 65 °, when a single-phase flow is caused to flow parallel to the tube axis direction, the fluid flows into the groove portion 1 from the upstream groove portion 2, and the downstream groove portion. When the fluid flows through 2, the flow input of the fluid in the groove portion 1 to the downstream groove portion 2 decreases, the fluid in the groove portion 1 stagnates, the fluidity of the fluid decreases, and heat exchange is inhibited. Therefore, it is preferable that the crest angle in the tube axis direction of the outer protrusion is 4 to 65 °.

“Crest angle in the direction perpendicular to the tube axis of the outer protrusion: 25 to 75 °”
By setting the crest angle in the direction perpendicular to the tube axis of the outer protrusion in this range, when a single-phase flow flows in parallel to the tube axis direction, the single-phase flow flows substantially parallel to the groove 2 and the fluid reaches the groove 1. When the fluid flows into the groove portion 1 from the groove portion 2, the fluid easily flows into the groove bottom portion. As a result, the stagnation of the fluid is suppressed, the fluidity of the fluid is improved, the increase in pressure loss is suppressed, and the heat exchange is further promoted by the inflow and outflow of the fluid.

  If the crest angle in the direction perpendicular to the tube axis of the outer projection is smaller than 25 °, the fluid into the groove 2 is difficult to flow in, the fluid tends to flow toward the outer periphery of the tube, and the fluid in the groove 2 stagnates. When the exchange is hindered and incorporated in the double pipe, the fluid flows to the inner peripheral side of the outer tube of the annular portion, and then the fluid on the inner peripheral side of the outer tube tends to flow into the groove portion 1 with the protrusion on the outer surface of the inner tube as a base point. Thus, the fluidity of the fluid is hindered and the pressure loss increases.

  When the crest angle in the direction perpendicular to the tube axis of the outer protrusion is greater than 75 °, when a single-phase flow is caused to flow parallel to the tube axis direction, the fluid flows into the groove 1 from the upstream groove 2, When the fluid flows in the downstream groove 2, the flow input of the fluid in the groove 1 decreases to the downstream groove 2, and the fluidity of the fluid decreases due to stagnation of the fluid in the groove 1. Be inhibited. Therefore, it is preferable that the crest angle in the direction perpendicular to the tube axis of the outer protrusion is 25 to 75 °.

“Groove shape of first and second grooves: inverted trapezoid”
The cross-sectional shapes of the first groove 33 and the second groove 34 are preferably inverted trapezoids. By making the first groove 33 and the second groove 34 into inverted trapezoidal shapes, when a single-phase flow is made parallel to the tube axis direction, the fluid flows almost in parallel with the second groove 34, and the first When the fluid reaches the groove 33, the fluid flows from the second groove 34 into the first groove 33. At that time, if the groove shape is an inverted trapezoidal shape, the fluid easily flows into the bottom of the groove, and as a result, the stagnation of the fluid in the groove is suppressed and the fluidity of the fluid is improved. Thereby, an increase in pressure loss of the fluid is suppressed, and heat exchange is further promoted by the inflow and outflow of the fluid.

  As shown in FIG. 4A, when the cross-sectional shape of the first groove 33 is an inverted trapezoid, the fluidity of the fluid is improved as described above because there is a flat surface at the bottom of the groove. However, as shown in FIG. 4B, when the shape of the groove 43 between the protrusions 41 is an inverted triangle, the fluid flow input to the root portion of the groove 43 is small, and the fluid in the direction away from the outer surface of the tube Easy to flow. That is, in the case of an inverted triangular cross-section having no flat portion at the groove bottom other than the trapezoidal groove, when a single-phase flow is caused to flow parallel to the tube axis direction, the second groove 43 between the upstream protrusions 41 is substantially parallel. When the refrigerant flows into the first groove and the fluid flows into the first groove, the fluid flows from the second groove 43 into the first groove. However, since the groove is not trapezoidal, it is difficult for the fluid to flow into the groove bottom, The fluid easily flows in the direction, that is, the direction away from the surface of the protrusion 41. Further, the fluid tends to stay at the bottom of the groove, and the heat exchange property of the fluid is hindered. Further, in the case of the groove shape having a triangular cross section, the cross sectional area of the groove portion tends to be smaller than in the case of the trapezoidal cross section. Furthermore, when it is incorporated in the double pipe structure, the fluid first flows to the inner peripheral side of the outer tube of the annular portion, and then the fluid on the inner peripheral surface side of the outer tube starts from the outer surface protrusion of the inner tube into the groove on the base 1. Inflow. For this reason, the fluidity | liquidity of a fluid is inhibited and a pressure loss increases.

  As described above, the cross-sectional shape of the groove is preferably the inverted trapezoid shown in FIG. 5 (d). However, the groove has a bottom shown in FIG. 5 (b) in addition to the inverted triangle shown in FIG. 5 (a). A curved groove having a center, or a curved groove having a bottom surface shown in FIG. Further, as shown in FIG. 5E, a trapezoidal groove having a convexly curved bottom surface may be used, and as shown in FIG. 5F, a trapezoidal groove having a convexly curved bottom surface may be used.

"Other"
In addition, the diameter of the upper surface of the protrusion on the outer surface of the heat transfer tube 11 is, for example, 7.95 to 13.5 mm. Moreover, the internal diameter D in the protrusion part of the heat exchanger tube 11 is 6.00 to 11.5 mm. Further, the number of ribs 32 formed on the inner surface of the heat transfer tube 11 can be calculated from the pitch of the inner ribs 32 and the lead angle of the inner ribs 32.

  Next, examples of the present invention will be described in comparison with comparative examples that are out of the scope of the present invention. FIG. 6 shows a test apparatus for heat exchange performance and pressure loss. This test apparatus constitutes an independent refrigeration cycle and a liquid refrigerant cycle. The test part that becomes the test part uses a double-pipe heat exchanger, arranges a smooth tube on the outer tube, and arranges the heat transfer tubes of the examples and comparative examples of the present invention on the inner tube. A comparative test was conducted.

  The inside of the inner pipe was connected to a circuit constituting a refrigeration cycle, and the refrigerant liquefied by the condenser was supplied to the inlet of the inner pipe by reducing the pressure by the expansion valve. The refrigerant supplied into the pipe of the inner pipe is in a gas-liquid two-phase flow state (a state where gas and liquid refrigerant are mixed). On the other hand, the liquid refrigerant heated to the evaluation condition is supplied to the annular portion between the inner surface of the outer tube and the outer surface of the inner tube to exchange heat with the fluid in the inner tube in a counter flow. Table 2 below shows the evaluation conditions of the examples and comparative examples.

  For the evaluation, after stabilizing the apparatus to the evaluation conditions, the flow rate of the annular part fluid, the temperature of the entrance and exit of the test part, and the pressure difference of the entrance and exit of the test part annular part were measured. Thereafter, the exchange heat quantity QR was calculated according to the following formula 1, using the flow rate of the fluid in the annular part and the temperature at the entrance and exit of the test part.

  The following Table 3-1, Table 3-2, Table 4-1, and Table 4-2 show the protrusions and rib shapes of the protrusion outer surface and the protrusion inner surface of the heat transfer tubes of Examples and Comparative Examples of the present invention. Table 5 below shows the heat exchange performance (exchange heat quantity ratio and pressure loss).

  As shown in Table 3-1, Table 3-2, Table 4-1, and Table 4-2, Examples 1 to 22 satisfy Claim 1. However, the projection height of Example 1 deviates from Claim 3, the projection pitch of Example 2 deviates from Claim 3, and the pitch of the second groove between the projections in the circumferential direction of Example 3 deviates from Claim 3. In Example 4, the peak angle of the projection in the tube axis direction deviates from Claim 3, and in Example 5, the peak angle of the projection in the direction perpendicular to the tube axis deviates from Claim 3. Further, in Examples 19 and 20, the formation pitch of the ribs in the tube axis direction and the height of the ribs deviate from Claim 2, and in Example 21, the formation pitch of the ribs in the tube axis direction deviates from Claim 2. No. 22 has a rib height deviating from the second aspect. On the other hand, Examples 6-18 satisfy | fill all the requirements of Claim 1, Claim 2, and Claim 3. On the other hand, Comparative Examples 17 and 18 are different from those in Claim 1 only. Further, Comparative Examples 1, 2, 3, 8, 9, 10, 11, 12, 13, and 14 are all out of the requirements of claims 1, 2, and 3. In Comparative Examples 4, 5, 6, 7, 15, and 16, the requirements of claims 1 and 3 are not satisfied.

  Table 5 shows the heat exchange ratio and pressure loss ratio of heat exchange in Examples 1 to 22 and Comparative Examples 1 to 18 as relative values with respect to Comparative Example 1. Examples 1 to 22 satisfy Claim 1 of the present invention, and both the exchange heat amount ratio and the pressure loss ratio of heat exchange were excellent. In particular, Examples 6 to 18 satisfied Claims 2 and 3, so that the heat exchange ratio of heat exchange was superior to Examples 1 to 5 and Examples 19 to 22. On the other hand, in Comparative Examples 1 to 13, at least one of the heat exchange ratio and the pressure loss ratio was inferior to that of the example.

1: Condenser 2: Supercooling heat exchanger 3, 5: Expansion valve 4: Condenser 5: Compressor 11: Heat transfer tube 12: Outer tube 22: Annular portion 31: Projection 32: Rib 33: First groove 34: Second groove 35: groove

Claims (4)

A refrigerant composed of a low-viscosity single-phase flow fluid having a viscosity coefficient of 350 μPa · s or less flows through the tubular portion between the outer tube and the inner tube, and a single-phase fluid or gas-liquid two-phase flow composed of liquid flows through the inner tube. In the heat transfer tube used for the inner tube of the supercooled double-tube heat exchanger for flowing a refrigerant consisting of fluid,
A plurality of protrusions that are formed in a rectangular shape on the outer surface of the tube and aligned in the tube axis direction and the tube circumferential direction, forming a trapezoidal trapezoidal shape;
One or more ribs spirally formed on the inner surface of the tube;
Have
The protrusion has a bottom width of the first groove between the protrusions adjacent to each other in the tube axis direction of 0.3 to 0.80 mm, and a bottom width of the second groove between the protrusions adjacent in the tube circumferential direction is 0. 0 mm. 10 to 0.30 mm,
The rib has a width of the bottom of the groove between the ribs adjacent to each other in the tube axis direction of 0.15 to 1.10 mm,
The supercooling double characterized in that a lead angle, which is an angle formed with respect to the tube axis direction of the rib, is 40 to 65 °, and a peak angle of the rib in the tube axis direction is 55 to 110 °. Heat transfer tube for tube heat exchanger. The formation pitch of the rib in the tube axis direction is 0.60 to 1.24 mm,
The heat transfer tube for a supercooled double-tube heat exchanger according to claim 1, wherein a height of the rib is 0.12 to 0.35 mm. The height of the protrusion is 0.18 to 0.50 mm,
The formation pitch of the projections in the tube axis direction is 0.65 to 1.15 mm,
The pitch in the tube circumferential direction of the second groove between the projections in the tube circumferential direction is 0.51 to 1.04 mm, and the peak angle in the tube axis direction of the projection is 4 to 65 °. The heat transfer tube for a supercooled double-tube heat exchanger according to claim 1 or 2, wherein a peak angle in a direction perpendicular to the tube axis is 25 to 75 °. The cross-sectional shape of the 1st groove | channel between the said protrusions of a pipe-axis direction and the 2nd groove | channel between the said protrusions of a pipe | tube circumferential direction is an inverted trapezoid, The one of Claim 1 thru | or 3 characterized by the above-mentioned. Heat transfer tube for supercooled double tube heat exchanger.

Images