JP2005195192A - Heat transfer pipe with grooved inner face - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a heat transfer pipe with a grooved inner face using R32-R125 mixed refrigerant (R410A, e.g.) as refrigerant for improving heat transfer performance while avoiding excessive pressure loss of the mixed refrigerant even when the heat transfer pipe with the grooved inner face is longer. <P>SOLUTION: The heat transfer pipe with the grooved inner face having an outer diameter D of 6-10 mm uses the hydrofluorocarbon mixed refrigerant. It comprises spiral grooves and a plurality of fins formed between spiral grooves on the inner face of the pipe . Herein, a groove lead angle θ is over 30° but 55° or smaller, a groove depth h is 0.10-0.35 mm and a fin peak angle δ is 5-30°. A coefficient α calculated from α=S×cosθ using a cross section S of one groove and the groove lead angle θ is 0.05-0.10, and a coefficient β calculated from β=L×cosθ using a wetted perimeter L of all periphery and the groove lead angle θ is 27-40. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to an internally grooved heat transfer tube incorporated in a heat exchanger for air conditioning equipment and refrigeration equipment. In particular, the outer diameter of a pipe using a mixed refrigerant in which R32 and R125, which are hydrofluorocarbon refrigerants, are mixed is 6. The present invention relates to a heat transfer tube with an inner groove of 10 mm.

  The internally grooved heat transfer tube is used as a heat transfer tube incorporated in a heat exchanger for air conditioning equipment and refrigeration equipment. Conventionally, a hydrochlorofluorocarbon (hereinafter referred to as HCFC) -based refrigerant, typically R22, has been used as a refrigerant for the internally grooved heat transfer tube. However, due to serious environmental problems such as ozone layer destruction and global warming, instead of R22, it will be shifted to a hydrofluorocarbon (hereinafter referred to as HFC) refrigerant having an ozone layer destruction coefficient of 0. It was.

  As the HFC refrigerant, a non-azeotropic refrigerant mixture obtained by mixing a high boiling HFC refrigerant (high boiling component) and a low boiling HFC refrigerant (low boiling component) is often used. This non-azeotropic refrigerant mixture has a different dew point (liquefaction start temperature) and boiling point (liquefaction end temperature). For example, R407C, which has recently been widely used as a refrigerant for air conditioning equipment, has R32 (boiling point: −52 ° C.), R125 (boiling point: −49 ° C.), and R134a, which have greatly different boiling points (= dew points). 3 refrigerants (boiling point: −26 ° C.) mixed at a mass ratio of 23:25:52, respectively. The boiling point of the mixed refrigerant is −43.6 ° C. and the dew point is −36.7 ° C. And the difference reaches 6.9 ° C. Therefore, during the condensation and evaporation of the mixed refrigerant, a large amount of high-boiling components are condensed at the gas-liquid interface, and the low-boiling components are concentrated on the gas phase side. As a result, the concentration of each component in the mixed refrigerant becomes non-uniform, resulting in a difference in concentration. This concentration difference causes diffusion resistance and thermal resistance, and the heat transfer performance (evaporation performance and condensation performance) of the internally grooved heat transfer tube is , R32, R125, and R134a are lower than when each is used as a single refrigerant, and heat transfer performance equivalent to that when using conventional R22 (both boiling point and dew point is −40.8 ° C.) cannot be obtained. It was. For these reasons, in the inner surface grooved heat transfer tube, the heat transfer performance is improved by specifying the inner surface groove shape of the tube inner surface in order to eliminate the concentration difference in the mixed refrigerant.

  On the other hand, R410A, which has recently been widely used as an HFC refrigerant for air conditioning equipment, is a mixture of R32 (boiling point: -52 ° C) and R125 (boiling point: -49 ° C) at a mass ratio of 50:50. Since the difference in the boiling point of each component is small, the boiling point and dew point of the mixed refrigerant are almost the same (−51.6 ° C.). Thus, since R410A can be handled as a single refrigerant, it is classified as a pseudo-azeotropic refrigerant mixture that does not cause a difference in concentration of each component in the refrigerant mixture. As described above, since R410A can be thermally handled in the same manner as R22, there are many examples in which the internally grooved heat transfer tube used for R22 is used as it is.

  Below, the example of the heat transfer tube with an inner surface groove using a non-azeotropic mixed refrigerant and a pseudo azeotropic mixed refrigerant is shown. First, as a heat transfer tube with an inner surface groove for a non-azeotropic refrigerant mixture, a spiral continuous groove is provided on the inner surface of the tube, and the height of the fin formed between the grooves along the spiral direction is the peak of the groove. There has been proposed a structure in which high fins higher than the above are provided at a constant pitch. In this internally grooved heat transfer tube, by providing high fins on the tube inner surface, the flow of the mixed refrigerant is disturbed, the concentration difference of each component in the mixed refrigerant is reduced, and the effective heat transfer area is expanded. Therefore, heat transfer performance improves (for example, refer to patent documents 1).

  In addition, as an internally grooved heat transfer tube for non-azeotropic refrigerant mixtures, a spiral crest (fin) is formed on the inner surface of the tube, and the two bottom angles of the crest (fin) are different angles. Has been. In this internally grooved heat transfer tube, by providing ridges (fins) with different bottom angles, the inclination angle of the side surfaces of the ridges (fins) differs, and the pressure loss varies depending on the refrigerant flow direction. Since the pressure loss is suppressed at the time of condensation, the heat transfer performance is improved (see, for example, Patent Document 2).

  In addition, as an internally grooved heat transfer tube for a non-azeotropic refrigerant mixture, a tube in which a spiral groove provided on the tube inner surface is arranged at a large angle of 45 ° or more with respect to the tube axis has been proposed. Moreover, what mixed HFC-32 (R32) and HFC-134a (R134a) which are the same HFC system as R407C is described as a non-azeotropic refrigerant mixture. In this internally grooved heat transfer tube, by arranging the groove at an angle of 45 ° or more, a mixed refrigerant vortex with a large vorticity is generated in the groove, and the concentration difference of each component in the mixed refrigerant is reduced. The heat transfer performance is improved (see, for example, Patent Document 3).

  Similarly to the internally grooved heat transfer tube of Patent Document 3, the spiral groove is disposed at an angle of 25 ° or more with respect to the tube axis, and preferably the depth of the groove is 0.15 mm to 0.00. An internally grooved heat transfer tube of 35 mm has also been proposed. Further, R407C is described as a non-azeotropic refrigerant (see, for example, Patent Document 4).

  In addition, as an internally grooved heat transfer tube that can use any refrigerant that is less viscous than R22 (HCFC22), for example, both R407C (non-azeotropic refrigerant mixture) and R410A (pseudo azeotropic refrigerant mixture), A spiral fin is provided, the fin pitch is 0.22 to 0.28 mm, the fin height is 0.23 to 0.28 mm, or in addition, the lead angle between the fin and the tube axis is 14 to 30 What has been proposed. In this internally grooved heat transfer tube, by setting the fin shape as described above, the heat transfer performance is improved because the refrigerant is sufficiently disturbed even when a refrigerant having a viscosity lower than R22 is used. (For example, refer to Patent Document 5).

As another example of the internally grooved heat transfer tube, a spiral groove having an inclination angle of 5 to 45 ° with respect to the tube axis is formed on the tube inner surface, and the tube outer diameter D and the tube axis right angle ( A shape parameter represented by a cross-sectional area a per groove in the (orthogonal) cross section, a groove depth d, and an area A of a circle in contact with the tip of the fin formed between the grooves around the tube axis: [2 ( d / D) ² · A / a] is specified within a range of 0.8 to 2.0. Further, as the refrigerant used, R22 is used as a refrigerant in the heat transfer performance test, but there is no example of a preferable refrigerant (for example, see Patent Document 6).

Further, a spiral trapezoidal groove is formed on the inner surface of the tube, and the apex angle of the angle protrusions (fins) formed between the trapezoidal grooves in a cross section perpendicular to the tube axis is 10 to 30 °, and the groove depth H is the tube. H / D ₁ = 0.04 to 0.05 in terms of the ratio to the inner diameter D ₁ , the sectional area S of each groove is S / H = 0.2 to 0.4 in terms of the ratio to the groove depth H, and An arcuate part having a radius of curvature R is provided between the inclined part and the groove bottom, and the arcuate part is specified as H / R = 4 to 10 in a ratio to the groove depth H. In addition, as the refrigerant used, R22 is used as a refrigerant for the heat transfer performance test, but there is no example of a preferable refrigerant (for example, see Patent Document 7).

JP-A-6-307787 (paragraph numbers [0001], [0008], [0013], FIG. 2 and FIG. 3) JP-A-8-61877 (Claim 1, paragraph numbers [0001], [0010], [0011], [0018] and FIG. 1) JP-A-8-145585 (paragraph numbers [0007], [0011], [0017], FIGS. 3 and 7) JP-A-9-42881 (paragraph numbers [0005], [0008], [0010] and FIG. 1) JP 2001-343194 (paragraph numbers [0019] to [0021], [0045], [0058], FIG. 2 and FIG. 3) JP 2001-33185 A (paragraph numbers [0010], [0054], FIGS. 2 and 3) Japanese Patent No. 2912826 (paragraph numbers [0008], [0026], FIG. 1 and FIG. 2)

  In the heat transfer tubes with inner surface grooves of Patent Documents 1 to 5, the refrigerant used is a non-azeotropic refrigerant mixture due to high fining, asymmetrical, high density, or high lead angle of the fins provided on the inner surface of the tube. In such a case, the mixed refrigerant is disturbed during the evaporation and condensation of the mixed refrigerant, the concentration difference of each component constituting the mixed refrigerant is reduced, and sufficient heat transfer performance is obtained. However, the refrigerant (for example, R410A) in which R32 and R125 used in the present invention are mixed is a pseudo-azeotropic refrigerant mixture, and has a small difference in boiling point between the components in the refrigerant mixture. Therefore, there is no difference in the concentration of each component during the evaporation and condensation of the mixed refrigerant. Therefore, the effect of improving the heat transfer performance due to the disturbance of the mixed refrigerant is small, and sufficient heat transfer performance can be obtained only by specifying the shape of the inner surface groove. There was a problem that it was not possible.

  In addition, when the internal grooved heat transfer tube is used by being incorporated in a heat exchanger for an air conditioner, the length of the internal grooved heat transfer tube is increased. In such an internally grooved heat transfer tube having a long tube length, the groove sectional area cannot be increased only by specifying the inner groove shape, and the pressure loss of the mixed refrigerant increases. As a result, there was a problem that sufficient heat transfer performance could not be obtained.

  Further, only by specifying the inner surface groove shape, it is not possible to increase the wet edge length of the inner surface of the tube, which greatly affects the effective heat transfer area of the gas-liquid interface, and the effective heat transfer area cannot be sufficiently increased. As a result, there was a problem that sufficient heat transfer performance could not be obtained. In particular, since the mixed refrigerant (for example, R410A) in which R32 and R125 used in the present invention are mixed has a lower viscosity than R22 and R407C, the wet edge length has a great influence on the heat transfer performance.

  In addition, in the heat transfer tubes with inner groove in Patent Documents 6 and 7, the heat transfer performance is improved by increasing the effective heat transfer area at the gas-liquid interface by specifying the shape parameter using the groove cross-sectional area of the inner surface of the tube. ing. However, just by specifying the shape of the inner surface groove in Patent Documents 6 and 7, as in Patent Documents 1 to 5, the wet edge length cannot be increased, and the effective heat transfer area cannot be sufficiently increased. As a result, there was a problem that sufficient heat transfer performance could not be obtained.

The present invention uses a mixed refrigerant (for example, R410A) in which R32 and R125 are mixed as the refrigerant, and even if the tube length of the inner surface grooved heat transfer tube is increased, the pressure loss of the mixed refrigerant does not become excessive and is excellent. It is an object of the present invention to provide an internally grooved heat transfer tube capable of obtaining high heat transfer performance. More specifically, an example will be described. When the refrigerant mass rate is 300 kg / m ² s, the evaporation performance as the heat transfer performance is 0.054 kW / m ² K or more and the condensation performance is 0.033 kW / m ² K. Provided is an internally grooved heat transfer tube as described above.

  The inner surface grooved heat transfer tube according to the present invention is an inner surface grooved heat transfer tube having an outer diameter D of 6 mm or more and 10 mm or less using a mixed refrigerant obtained by mixing R32 and R125 which are hydrofluorocarbon refrigerants. The inner surface of the heat transfer tube has a spiral groove and a plurality of fins formed between the grooves, and the groove lead angle θ formed by the groove and the tube axis exceeds 30 ° and is 55 ° or less, orthogonal to the tube axis. The groove depth h of the groove in the cross section is 0.10 mm or more and 0.35 mm or less, the fin crest angle δ of the fin is 5 ° or more and 30 ° or less, and one groove portion in the tube axis orthogonal cross section of the tube inner surface The coefficient α calculated by α = S × cos θ using the cross-sectional area S and the groove lead angle θ is 0.05 or more and 0.10 or less, and the wet edge length L of the entire circumference in the tube axis orthogonal cross section of the tube inner surface , And β = × coefficients calculated by cos [theta] beta are those configured as a heat transfer tube with inner surface grooves is 27 or more and 40 or less.

  According to the above configuration, the groove lead angle θ, the groove depth h, the fin peak angle δ, the groove cross-sectional area S, the coefficient α calculated by the groove lead angle θ, the wet edge length L, and the groove lead angle θ are calculated. By setting the coefficient β to a predetermined range, the groove cross-sectional area is increased in the groove moldable region, and the pressure loss of the mixed refrigerant does not become excessive. In addition, the wet edge length L is increased in the groove formable region, and the effective heat transfer area at the gas-liquid interface is increased.

In such an internally grooved heat transfer tube, a mixed refrigerant (for example, R410A) in which R32 and R125 are mixed as a refrigerant is used, and the length of the internally grooved heat transfer tube is increased by incorporation into a heat exchanger for an air conditioner. Even so, the pressure loss of the mixed refrigerant does not become excessive, and excellent heat transfer performance can be obtained. Further, since the pressure loss of the mixed refrigerant does not become excessive, there is no temperature difference in the mixed refrigerant temperature at the inlet / outlet of the inner surface grooved heat transfer tube. Therefore, the compressor of the mixed refrigerant can be downsized, and the air conditioner (heat exchanger) can be downsized. In addition, the power consumption is reduced and the energy consumption efficiency COP (Coefficient) of the air conditioner (heat exchanger)
of Performance).

  Embodiments of the present invention will be specifically described below with reference to the drawings. FIG. 1 is a partially enlarged cross-sectional view showing a cross-sectional shape of an internally grooved heat transfer tube when broken in the tube axis direction, FIG. 2 (a) is a cross-sectional view taken along the line AA in FIG. (A) is a partially enlarged sectional view, FIG. 3 (a) is a front view showing a heat exchanger incorporating an internally grooved heat transfer tube, and (b) is a view of the heat exchanger of (a) from the U-bend tube side. (C) is a view of the heat exchanger of (a) as viewed from the hairpin tube side, and FIG. 4 is a graph for measuring heat transfer performance and pressure loss of a heat exchanger incorporating an internally grooved heat transfer tube. FIG. 5 is a schematic view of a suction type wind tunnel to be used, and FIG. 5 is a schematic view of a refrigerant supply device that supplies a refrigerant to the suction type wind tunnel of FIG.

  As shown in FIGS. 1, 2 (a) and 2 (b), the internally grooved heat transfer tube 1 of the present invention is used as a heat transfer tube for an air conditioner. The main heat transfer tube 1 has a main flow of 6 mm or more and 10 mm or less, and the inner surface of the tube has an inner surface groove shape for improving the thermal conductivity by evaporation and condensation of the refrigerant in the tube. Moreover, the refrigerant | coolant supplied to the pipe | tube inner surface of the inner surface grooved heat exchanger tube 1 uses the mixed refrigerant | coolant which mixed R32 and R125 which are HFC type refrigerant | coolants. Moreover, as a material of the raw tube of the inner surface grooved heat transfer tube 1, copper or a copper alloy is used, for example, phosphorus deoxidized copper such as alloy numbers C1220 and C1201 defined in JISH3300. The method for forming the inner surface groove shape of the inner surface grooved heat transfer tube 1 includes a rolling method and a rolling method, but is not particularly limited.

(Pipe outer diameter D)
As described above, the pipe outer diameter D is required to be 6 mm or more and 10 mm, which is the mainstream as a heat transfer pipe for an air conditioner. When the pipe outer diameter D is less than 6 mm, a mixed refrigerant (for example, R410A) in which R32 and R125 are mixed is used as a refrigerant, and is used by being incorporated in the heat exchanger 10 for an air conditioner (see FIG. 3). In FIG. 3, the inner grooved heat transfer tube 1 is described as a hairpin tube 1a bent into a hairpin shape), and since the inner grooved heat transfer tube 1 is long, the pressure loss of the mixed refrigerant becomes excessive, and the inner groove The heat transfer performance, particularly the evaporation performance, of the attached heat transfer tube 1 is lowered. Moreover, when the pipe outer diameter D exceeds 10 mm, the inner surface grooved heat transfer tube 1 becomes heavy, and the air conditioner (the heat exchanger 10, see FIG. 3) incorporating the inner surface grooved heat transfer tube 1 cannot be reduced in weight. .

(Mixed refrigerant)
As the mixed refrigerant, R410A, which is a pseudo-azeotropic mixed refrigerant in which R32 and R125, which are two types of HFC refrigerants, are mixed at a mass ratio of 50:50 is most often used. R410A is a refrigerant that does not contain chlorine and contains hydrogen as a refrigerant that replaces R22 (HCFC-based refrigerant) containing chlorine that causes the recent destruction of the ozone layer and global warming. The boiling points of R32 and R125 are −52 (C and −49 (C, respectively), and the boiling points of the refrigerants are close to each other. Therefore, the mass ratio of R32: R125 = 50: 50 hardly changes depending on the state of the mixed refrigerant. There is no difference in density.

(Inner groove shape)
The inner surface groove shape is composed of a continuous spiral groove 2 and fins 3 formed between the grooves 2. Moreover, each groove | channel 2 is comprised from the fin peak summit curve part 3a, the fin slope straight line part 3b connected smoothly to this, and the groove bottom part 2a which connects each fin slope straight line part 3b. In addition, the groove bottom part 2a is smoothly continuous with a straight part and an arbitrary fin root radius R, and the arbitrary fin root radius R may be smoothly continuous without a straight part (not shown).

  The groove lead angle θ formed by the groove 2 and the tube axis is more than 30 ° and not more than 55 °, the groove depth h of the groove 2 in the cross section perpendicular to the tube axis is 0.10 mm or more and 0.35 mm or less, and the fin peak of the fin 3 The angle δ is 5 ° or more and 30 ° or less, and the coefficient α calculated by α = S × cos θ using the groove cross-sectional area S of one groove and the groove lead angle θ in the cross section orthogonal to the tube axis on the inner surface of the tube is 0. The coefficient β calculated from β = L × cos θ using the wet edge length L and the groove lead angle θ in the tube axis orthogonal cross section of the tube inner surface is 27 or more and 40 or less.

  Hereinafter, the grounds for limiting the numerical values in the inner groove shape will be described.

(Groove lead angle θ)
The groove lead angle θ is required to be more than 30 ° and not more than 55 °. When the groove lead angle θ is 30 ° or less, the evaporation performance when the internally grooved heat transfer tube is incorporated in the heat exchanger 10 for air conditioning equipment (see FIG. 3) is less than 0.054 kW / m ² K, and The condensation performance becomes less than 0.033 kW / m ² K (refrigerant mass velocity: 300 kg / m ² s), and the heat transfer performance decreases. In addition, when the groove lead angle θ exceeds 55 °, the speed at which the groove 2 is formed on the inner surface of the tube by the rolling process is extremely reduced, and the long inner surface grooved heat transfer tube 1 is stably formed. Cannot be manufactured.

(Groove depth h)
The groove depth h is required to be 0.10 mm or more and 0.35 mm or less. When the groove depth h is less than 0.10 mm, the fins 3 formed between the grooves 2 on the inner surface of the inner surface grooved heat transfer tube 1 become lower than the condensate liquid level of the working refrigerant on the inner surface of the tube. Immersed in liquid. For this reason, the effective heat transfer area on the inner surface of the pipe is remarkably reduced, and the heat transfer performance is lowered. Further, when the groove depth h exceeds 0.35 mm, when the groove 2 is formed on the inner surface of the tube, the groove forming tool (for example, a grooved plug) is damaged, and the groove 2 is stably formed on the inner surface of the tube. Can not be molded.

(Fin peak angle δ)
The fin peak angle δ is required to be 5 ° or more and 30 ° or less. When the fin crest angle δ is less than 5 °, the width of the fin 3 is narrowed, and when the pipe is expanded when the inner surface grooved heat transfer tube 1 is incorporated into the heat exchanger 10 for air conditioning equipment (see FIG. 3). 2), the tip of the fin 3 is crushed, and the fin 3 is collapsed or distorted. Further, when the groove 2 is formed on the inner surface of the tube for forming the fin 3, the groove forming tool is damaged, and the groove 2 cannot be stably formed on the inner surface of the tube. Further, when the fin peak angle δ exceeds 30 °, the groove section sectional area S of the groove 2 is remarkably reduced, and the heat transfer performance is deteriorated. In addition, since the cross-sectional area of the fin 3 (the bottom wall thickness T of the internally grooved heat transfer tube 1) is increased and the weight of the internally grooved heat transfer tube 1 is increased, an air conditioner incorporating the internally grooved heat transfer tube 1 (heat exchange) The container 10 (see FIG. 3) cannot be reduced in weight.

(Coefficient α)
The coefficient α calculated by α = S × cos θ using the groove cross-sectional area S of one groove in the cross section perpendicular to the tube axis and the groove lead angle θ is required to be 0.05 or more and 0.10 or less. Here, the groove cross-sectional area S is calculated from the pipe outer diameter D, the bottom wall thickness T, the groove depth h, the fin tip radius r, the fin root radius R, the number of grooves, and the fin peak angle δ.

  On the other hand, when the coefficient α is less than 0.05, the groove lead angle θ becomes large, and the pressure loss of the mixed refrigerant becomes excessively large. Further, the effective heat transfer area is reduced by reducing the groove cross-sectional area S. As a result, the heat transfer performance, particularly the evaporation performance is reduced.

  On the other hand, when the coefficient α exceeds 0.10, the groove cross-sectional area S becomes excessively large, so that the number of grooves 2 becomes extremely small, or the groove depth h of the grooves 2 becomes deep. For this reason, when the groove 2 is formed on the inner surface of the tube, the groove forming tool is damaged, and the groove 2 cannot be stably formed on the inner surface of the tube.

(Coefficient β)
The coefficient β calculated by β = L × cos θ using the wet edge flow L of the entire circumference in the cross section perpendicular to the tube axis and the groove lead angle θ is required to be 27 or more and 40 or less. Here, the wet edge length L is calculated from the pipe outer diameter D, the bottom wall thickness T, the groove depth h, the fin tip radius r, the fin root radius R, the number of grooves, and the fin peak angle δ.

  On the other hand, when the coefficient β is less than 27, the wet edge length L is short, the effective heat transfer area at the gas-liquid interface is reduced, and the heat transfer performance is lowered.

  On the other hand, when the coefficient β exceeds 40, the wet edge length L becomes excessively long, the number of grooves on the inner surface of the pipe becomes extremely large, or the groove depth h of the groove 2 becomes deep. For this reason, when the groove 2 is formed on the inner surface of the tube, the groove forming tool is damaged, and the groove 2 cannot be stably formed on the inner surface of the tube.

  Next, the bottom wall thickness T, the fin tip radius r, the fin root radius R, and the number of grooves used to calculate the groove cross-sectional area S and the wet edge length L will be described. Further, the pipe outer diameter D, the groove depth h, and the fin crest angle δ used for the calculation are as described above, and thus the description thereof is omitted.

(Bottom thickness T)
The bottom wall thickness T is appropriately set in consideration of the pressure strength, weight, and the like of the internally grooved heat transfer tube 1.

(Fin tip radius r)
The fin tip radius r is preferably in the range of 0.05 or more and less than 0.15 of the groove depth h. When the fin tip radius r is less than 0.05 of the groove depth h, the fin tip radius r decreases, so that the moldability of the fin 3 (groove 2) deteriorates when the fin 3 becomes high. It is difficult to obtain the fin 3 having a predetermined shape, and the groove forming tool that comes into contact with the groove 2 on the inner surface of the pipe is easily damaged. Further, when the fin tip radius r is 0.15 or more of the groove depth h, the fin tip radius r is increased, so that the cross-sectional area of the fin 3 is increased, and the weight of the internally grooved heat transfer tube 1 is increased. Become heavier.

(Fin root radius R)
The fin root radius R is preferably in the range of 1/5 or more and less than 2/3 of the groove depth h. When the fin root radius R is less than 1/5 of the groove depth h, the fin root radius R is small, so that the moldability of the fin 3 (groove 2) is deteriorated when the fin 3 is high. It is difficult to obtain the fin 3 having a predetermined shape, and the groove forming tool that comes into contact with the root of the groove 2 on the inner surface of the pipe is easily damaged. Further, when the fin root radius R is 2/3 or more of the groove depth h, the fin root radius R increases, so that the cross-sectional area of the fin 3 increases, and the bottom wall thickness T of the pipe increases. The inner grooved heat transfer tube 1 is heavier.

(Number of grooves)
The number of grooves is preferably 30 or more and 100 or less. When the number of grooves is less than 30 or more than 100, when forming the groove 2 on the inner surface of the tube, the groove formability is extremely lowered, or the groove forming tool is damaged, and the inner surface of the tube is stabilized. The groove 2 cannot be formed.

The groove bottom width w calculated from the pipe outer diameter D, groove depth h, fin peak angle δ, bottom wall thickness T, fin tip radius r, fin root radius R, number of grooves, etc. will be described.
(Groove bottom width w)
The groove bottom width w is preferably 0.18 mm or more. When it is less than 0.18 mm, a condensate pool is generated, the effective heat transfer area is reduced, and the condensation performance is lowered. The groove bottom width w is the length of the line segment connecting the inclined straight line portion 3b of two fins 3 adjacent to one groove 2 and the intersection of the extension lines of the groove bottom portion 2a.

  Next, the manufacturing method of the heat exchanger incorporating the inner surface grooved heat transfer tube of the present invention will be described by taking a plate fin tube type heat exchanger which is the majority of the heat exchanger of the air conditioner as an example. As shown in FIGS. 3A, 3B, and 3C, first, a plurality of U-shaped hairpin tubes are formed by bending an inner-grooved heat transfer tube into a hairpin shape at a center portion with a predetermined bending pitch Pa. 1a is produced. Next, a plurality of hairpin tubes 1a are inserted through a plurality of fin materials 11 made of aluminum or aluminum alloy arranged in parallel with each other at a predetermined interval (fin pitch Pb), and both are joined. . Then, a plurality of hairpin tubes 1a are joined to a plurality of U-bend tubes 12 by fitting and brazing U-bend tubes 12 which have been previously bent to the tube ends of adjacent hairpin tubes 1a. Through a plurality of stages and a plurality of rows are connected in series at a predetermined step direction pitch (same as the bending pitch Pa) and a row direction pitch Pc. Thereby, the heat exchanger 10 by which the internal heat transfer tube with a long effective heat transfer tube length is arrange | positioned between the several fin materials 11 is produced.

  Next, a method for manufacturing the inner surface grooved heat transfer tube will be described. The production of the internally grooved heat transfer tube of the present invention is performed using a conventionally known production apparatus (not shown), and the first diameter reduction processing is performed on the raw tube which is the material of the internally grooved heat transfer tube. A second step of performing a second diameter reduction process on the reduced diameter pipe obtained in the first step, and forming a spiral groove on an inner surface of the raw pipe, This includes a third step of performing a third diameter reduction processing of the raw tube in which the spiral groove is formed in the step of 2. Then, by pulling the raw tube in the drawing direction, the first step, the second step, and the third step are performed in this order, and the raw tube is processed into an internally grooved heat transfer tube. Here, a 1st process, a 2nd process, and a 3rd process are demonstrated in order.

(First step)
The raw pipe, which is the material of the inner surface grooved heat transfer pipe, is drawn out so as to pass between the reduced diameter die and the reduced diameter plug, whereby the first diameter reduction process is performed on the raw pipe.
(Second step)
The raw pipe reduced in the first step is pressed with a grooved plug inserted into the raw pipe with a plurality of rolling balls or rolls, whereby a second diameter reduction processing is performed on the raw pipe. In addition, the groove shape of the grooved plug is transferred to the inner surface of the reduced diameter pipe, and a spiral groove is formed.
(Third step)
The base tube having a spiral groove formed on the inner surface in the second step is subjected to a third diameter reduction process by a shaping die to produce an inner grooved heat transfer tube.

Examples of the present invention will be specifically described below.
(First embodiment)
Pipe outer diameter D, groove lead angle θ, groove depth h, fin crest angle δ, coefficient α = groove cross section S × cos θ (groove lead angle) and coefficient β = wet edge flow L × Test tubes (Examples 1 to 4) in which all of cos θ (groove lead angle) satisfy the claims of the present invention (Examples 1 to 4), and test tubes (Comparative Examples 1 to 4) in which at least one of the shape parameters does not satisfy the claims of the present invention. 5) was produced. The test tube is prepared by first melting, casting, hot-extrusion, cold-rolling, cold drawing, and cold drawing, melting and dephosphorating copper of alloy number C1220 defined in JISH3300. Was made. Next, a first diameter reduction process is performed on the element pipe, a second diameter reduction process is performed while forming the inner surface groove-shaped spiral groove on the diameter-reduced element pipe, and the element in which the spiral groove is formed. The tube was subjected to a third diameter reduction process to prepare a test tube having an outer diameter (tube outer diameter D) of 7 mm. Table 1 shows values of shape parameters of Examples 1 to 4 and Comparative Examples 1 to 4.

Next, the plate fin tube type heat exchanger 10 shown in FIGS. 3A, 3B, and 3C was manufactured using each of the test tubes. The specifications of the heat exchanger 10 were as follows.
(Heat exchanger 10)
The outer shape was 250 mm high × 250 mm long × 25.4 mm wide.
(Hairpin tube 1a)
Made using the test tube, and arranged in two rows and 12 steps (bending pitch Pa21 mm, row direction pitch Pc 12.7 mm) (effective heat transfer tube length was about 6.7 m).
(Fin material 11)
A plate material made of aluminum having an alloy number of 1N30 specified in JISH4000, and the surface of the plate material is coated with a resin. The thickness of the fin material 11 was 100 μm. Then, 200 fin materials 11 were arranged in parallel at a fin pitch Pb of 1.25 mm.

  Using this heat exchanger 10, heat transfer performance (evaporation performance, condensation performance) and pressure loss were measured, and the results are shown in Table 2. Here, the evaporation performance and the condensation performance are described by measuring the overall heat transfer coefficient. The pressure loss was described by measuring the evaporation pressure loss per unit length of the test tube (hairpin 1a).

  Moreover, the schematic diagram of the measuring apparatus which measures heat-transfer performance and pressure loss in FIG. 4 is shown. As shown in FIG. 4, the measuring device includes a suction type wind tunnel 100 with a constant temperature and humidity function, a refrigerant supply device 110 (see FIG. 5), and an air conditioner (not shown). In the suction type wind tunnel 100, the heat exchanger 10 is arranged in the flow path of the air that flows in from the air inlet 108 and is discharged from the air outlet 109, and upstream and downstream of the heat exchanger 10, respectively. Air samplers 101 and 102 are arranged. The air samplers 101 and 102 are connected to temperature and humidity measuring boxes 103 and 104, respectively. The temperature and humidity measuring boxes 103 and 104 measure the temperature and humidity of the air by measuring the dry bulb temperature and the wet bulb temperature of the air collected by the air samplers 101 and 102, respectively. An induction fan 105 is provided on the downstream side of the air sampler 102 and discharges air to the air discharge port 109. Rectifiers 106 and 106 for rectifying the air that has passed through the heat exchanger 10 are provided between the heat exchanger 10 and the air sampler 102 and between the air sampler 102 and the induction fan 105.

  FIG. 5 shows a schematic diagram of the refrigerant supply device 110. In FIG. 5, 107 is a refrigerant pipe, 111 is a sight glass, 112 is a heat exchanger for heating and cooling liquid (refrigerant), 113 is a dryer, 114 is a liquid receiver (refrigerant), 115 is a plug, and 116 is a condenser. 117 is an oil separator, 118 is a compressor, 119 is an accumulator, 120 is an evaporator, 121 is an expansion valve, and 122 is a flow meter. And the refrigerant | coolant which adjusted the pressure and temperature is supplied into the inside of the hairpin pipe | tube 1a (refer FIG. 3) of the heat exchanger 10 with which the suction type wind tunnel 100 was equipped through the refrigerant | coolant piping 107. FIG. In addition, a pressure gauge 123 (the temperature is a saturation temperature corresponding to the measured pressure) is provided at the inlet and outlet of the heat exchanger 10 to measure the temperature and pressure of the refrigerant. Further, the air conditioner (not shown) supplies air with controlled temperature and humidity to the air inlet 108 of the suction type wind tunnel 100.

The measurement conditions were as shown in Table 3, R410A was used as the refrigerant, and the refrigerant mass rate was 300 kg / m ² s (refrigerant flow rate 35 kg / h). Further, the flow of the refrigerant at the time of measuring the evaporation performance and the flow of the refrigerant at the time of measuring the condensation performance are different from each other (the flow direction of the refrigerant in the refrigerant supply device 110 shown in FIG. Shows the flow direction of the refrigerant.

From the results in Table 2, Examples 1 to 4 of the present invention have an evaporation performance exceeding the target value of 0.054 kw / m ² K, and the condensation performance is also excellent as 0.033 kw / m ² K or more, which is the target value. The heat transfer performance was. In addition, the pressure loss was not about a large pressure loss value for a heat transfer tube having a tube outer diameter of about 7 mm and an effective heat transfer tube length of about 6.7 m at about 0.8 kPa / m.

Further, in Comparative Example 1, the coefficient α exceeds the upper limit value of the claims of the present invention, so the number of grooves is extremely reduced, groove forming failure occurs, and mass production of the test tube (inner grooved heat transfer tube) cannot be performed. It was in a state. In Comparative Example 2, since the coefficient α is less than the lower limit value of the claims, the number of grooves is large, the groove lead angle θ is large, the evaporation performance is less than the target value of 0.054 kw / m ² K, the pressure loss Was larger than Examples 1-4. In Comparative Example 3, the groove lead angle θ exceeded the upper limit value of the claims, so that a groove forming defect occurred as in Comparative Example 1. In Comparative Example 4, since the coefficient α and the coefficient β are less than the lower limit value of the claims, the groove depth h is small, the evaporation performance is less than the target value of 0.054 kw / m ² K, and the condensation performance is also the target value. This was below 0.033 kw / m ² K. In Comparative Example 5, since the coefficient β exceeds the upper limit value of the claims, the number of grooves is extremely increased, and defective groove formation occurs.

(Second embodiment)
The test tube was the same as the first example except that 6.35 mm (Example 5, Comparative Example 6) and 5 mm (Comparative Example 7) were used as the outer diameter D of the test tube. Table 4 shows the values of the shape parameters of the test tubes, and Table 5 shows the measurement results of heat transfer performance (evaporation performance, condensation performance) and pressure loss.

From the results of Table 5, Example 5 of the present invention has an excellent transmission performance in which the evaporation performance exceeds the target value of 0.054 kw / m ² K and the condensation performance also exceeds the target value of 0.033 kw / m ² K. It was thermal performance. Also, the pressure loss was not a large pressure loss value for a heat transfer tube having a tube outer diameter of 6.35 mm and an effective heat transfer tube length of about 6.7 m at a pressure loss of about 1.3 kPa / m.

In Comparative Example 6, since the groove lead angle θ and the coefficient β are less than the lower limit values of the claims, the groove depth h is small, and the evaporation performance is less than the target value of 0.054 kw / m ² K. The performance was also lower than the target value of 0.033 kw / m ² K. In Comparative Example 7, since the pipe outer diameter D, the groove lead angle θ, the coefficient α, and the coefficient β are less than the lower limit values of the claims, the groove depth h is small and the evaporation performance is a target value of 0.054 kw / m. ^The pressure loss was less than ² K, and the pressure loss was about 1.5 kPa / m, the outer diameter of the tube was 5 mm, and the effective heat transfer tube length was about 6.7 m.

It is a partially expanded sectional view when it fractures | ruptures in the tube-axis direction which shows the cross-sectional shape of the heat transfer tube with an inner surface groove | channel based on this invention. (A) is sectional drawing in the AA of FIG. 1, (b) is a partially expanded sectional view of (a). (A) is the front view which shows the example which incorporated the heat transfer tube with an inner surface groove | channel which concerns on this invention in the heat exchanger, (b) is the figure which looked at the heat exchanger of (a) from the U bend pipe side, (c) FIG. 3 is a view of the heat exchanger of (a) as viewed from the hairpin tube side. It is a schematic diagram of the suction type wind tunnel used when measuring the heat transfer performance and pressure loss of the heat exchanger incorporating the internally grooved heat transfer tube according to the present invention. It is a schematic diagram of the refrigerant | coolant supply apparatus which supplies a refrigerant | coolant to the suction type wind tunnel of FIG.

Explanation of symbols

1 Heat transfer tube with inner groove 2 Groove 3 Fin D Tube outer diameter h Groove depth L Wet edge length S Groove cross section δ Fin peak angle θ Groove lead angle

Claims (1)

In an internally grooved heat transfer tube having an outer diameter D of 6 mm or more and 10 mm or less using a mixed refrigerant obtained by mixing R32 and R125, which are hydrofluorocarbon refrigerants,
A plurality of fins formed between the groove and the spiral groove on the inner surface of the inner surface grooved heat transfer tube, and a groove lead angle θ formed by the groove and the tube axis is more than 30 ° and not more than 55 °; The groove depth h of the groove in the cross section perpendicular to the tube axis is 0.10 mm or more and 0.35 mm or less, and the fin crest angle δ of the fin is 5 ° or more and 30 ° or less,
A coefficient α calculated by α = S × cos θ using a groove cross-sectional area S of one groove in the tube axis orthogonal cross section of the tube inner surface and the groove lead angle θ is 0.05 or more and 0.10 or less,
An inner surface groove characterized in that a coefficient β calculated by β = L × cos θ using the wet edge length L of the entire circumference in the tube axis orthogonal cross section of the tube inner surface and the groove lead angle θ is 27 or more and 40 or less. Heat transfer tube.

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