JP2010078289A - Heat exchanger and air conditioner equipped with the same - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a heat exchanger with high heat transfer performance and an air conditioner using the heat exchanger by reducing a diameter of a heat transfer tube and optimizing a shape parameter. <P>SOLUTION: The heat exchanger includes a plurality of plate fins 1 disposed side by side at predetermined intervals and having a fluid pass through the intervals, and a plurality of circular heat transfer tubes 2 inserted in a direction orthogonal to the plate fins and carrying a coolant through interiors. In the heat transfer tube 2, an outer diameter is 4-6 millimeters, a spiral groove 4 with a depth of 0.1-0.2 millimeters is formed on a tube inner face, an angle between a straight line in parallel with a tube axial direction and an extending direction of the groove 4 in the tube inner face is 10-20 degrees when a coolant mass flow rate is 5-20 kg/h, and the angle between the straight line in parallel with the tube axial direction and the extending direction of the groove 4 in the tube inner face is 0-10 degrees when the coolant mass flow rate is 20-45 kg/h. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to a heat exchanger and an air conditioner equipped with the heat exchanger. The present invention is mainly applied to a ceiling-embedded air conditioner, but can also be applied to other air conditioners such as a wall-mounted type and a floor-mounted type air conditioner. It is.

For example, a conventional fin tube type heat exchanger is composed of a plurality of plate-like fins arranged in parallel with each other at regular intervals, and a heat transfer tube that penetrates the plate-like fins in the normal direction, Heat exchange is performed between the air flowing between the plate fins and the refrigerant flowing inside the heat transfer tube.
In recent years, from the viewpoint of preventing global warming, reduction of energy consumption of air conditioners and reduction of the amount of refrigerant used as a working fluid have been strongly demanded. Miniaturization is required.
On the other hand, since the gas passage speed is kept low from the viewpoint of suppressing noise increase in order to ensure comfort, the heat transfer on the air side remains low with respect to the heat transfer coefficient inside the heat transfer tube. Therefore, the heat transfer on the air side is improved by increasing the heat transfer area on the air side.

  In other words, due to demands for downsizing heat exchangers and installation space limitations, the number of heat exchangers installed in the air flow direction (stage direction) is increased, or the length of the heat transfer tube plate fins in the stacking direction (straight pipe) The heat exchanger is not enlarged by extending the size of the heat exchanger by extending the length of the same part) A method of increasing the heat transfer area of the heat exchanger by increasing the number of installation rows in the row direction is adopted. For example, in the past heat exchanger tube diameter was about 10 mm, fin pitch was about 1.5 mm, and heat exchangers with two rows were commercialized, but recently the heat transfer tube diameter is about 7 mm. The fin pitch is narrowed to about 1.1 mm, and heat exchangers with three or more rows have been commercialized.

And the heat transfer tube outer diameter D is in the range of 3 mm ≦ D ≦ 7.5 mm,
1.2D ≦ Lp ≦ 1.8D
2.6D ≦ Dp ≦ 3.5D
At this time, Lp: row pitch in the gas passage direction of the heat transfer tube
Dp: The heat transfer performance is improved by setting the step pitch in the direction perpendicular to the gas passage direction of the heat transfer tube (step direction), and the slit fins protruding on both surfaces of the plate fins are An invention has been disclosed in which a plurality of rows are formed by “cutting and raising” in the direction perpendicular to each other to improve heat transfer performance and promote gas mixing in the cut and raised portion (see, for example, Patent Document 1).

JP-A-63-3188 (page 2-3, FIGS. 3 and 4)

In the heat exchanger of Patent Document 1, a circular tube is used as the heat transfer tube, and by defining the heat transfer tube outer diameter D, the row pitch Lp, and the step pitch Dp as described above, the ventilation resistance in the heat transfer tube portion is reduced. Although the dead water area generated on the wake side of the heat transfer tube is suppressed, another step is required to further improve the heat transfer performance of the heat exchanger.
Conventionally, for the purpose of improving the performance of heat exchangers, the heat exchange performance has been improved by increasing the height of the groove shape in the heat transfer tube on the refrigerant side and devising the fin shape on the air side. However, in order to achieve higher performance than the current level, it is conceivable to reduce the diameter of the heat transfer tube. However, by reducing the diameter of the heat transfer tube, the heat transfer coefficient in the tube increases, but the pressure loss increases. Therefore, it is necessary to optimize them.

  Also, the refrigerant flowing through the heat transfer tube usually contains refrigeration oil, which is a lubricant for the compressor, and simply forming a groove on the inner surface of the heat transfer tube increases the pressure loss. Therefore, sufficient in-tube (condensation / evaporation) performance cannot be obtained. Such an increase in pressure loss is more remarkable as the content of refrigeration oil in the refrigerant is higher.

  The present invention has been made in order to solve the above-described problems. By reducing the diameter of the heat transfer tube and optimizing the shape parameters, the heat exchanger has high heat transfer performance, and the heat exchange. It aims at providing the air conditioner which uses a vessel.

  The heat exchanger according to the present invention includes a plurality of plate-like fins that are arranged side by side at a predetermined interval and through which the fluid passes, and a plurality of circles that are inserted in a direction orthogonal to the plate-like fin and through which the refrigerant flows. The heat transfer tube has an outer diameter of 4 to 6 mm, a spiral groove with a depth of 0.1 to 0.2 mm is formed on the inner surface of the tube, and the refrigerant mass flow rate Is 5 to 20 kg / h, the angle formed by the straight line parallel to the tube axis direction on the inner surface of the tube and the direction in which the groove extends is 10 to 20 °, and when the refrigerant mass flow rate is 20 to 45 kg / h, the tube on the inner surface of the tube The angle formed by the straight line parallel to the axial direction and the direction in which the groove extends is 0-10 °.

  In the present invention, when a refrigerant having a refrigerant mass flow rate of 5 to 20 kg / h is used, the outer diameter is 4 to 6 mm, the groove depth is 0.1 to 0.2 mm, and the pipe inner surface is in the tube axis direction. When the angle formed by the parallel straight line and the direction in which the groove extends is 10 to 20 ° and the refrigerant having the refrigerant mass flow rate of 20 to 45 kg / h is used, the outer diameter is 4 to 6 mm and the groove depth is 0. .1 to 0.2 mm, and the angle formed by the straight line parallel to the tube axis direction on the tube inner surface and the direction in which the groove extends is 0 to 10 °, so that the heat transfer rate is improved without increasing the pressure loss. Therefore, a heat exchanger and an air conditioner excellent in heat transfer performance can be obtained.

Hereinafter, embodiments of the present invention will be specifically described with reference to the drawings.
Embodiment 1 FIG.
FIG. 1A is a cross-sectional view showing a horizontal cross section in the tube axis direction of the heat transfer tube of the fin-and-tube heat exchanger 100 according to Embodiment 1 of the present invention, and a cross section between adjacent heat transfer tubes 2 and the interval therebetween. The plate-shaped fin 1 is shown. FIG.1 (b) is side surface sectional drawing which cut | disconnected the heat exchanger 100 to the perpendicular direction, and has added the enlarged view which expanded a part of groove shape of the pipe inner surface.
The heat exchanger 100 of the present embodiment includes a plurality of plate-like fins 1 arranged side by side at a predetermined interval, and a plurality of heat transfer tubes 2 that are inserted in a direction orthogonal to the plate-like fins 1 and in which a refrigerant flows. It is configured. The plate-like fins 1 are made of a metal material such as copper, copper alloy, aluminum or aluminum alloy (the same applies to other embodiments), and a fluid such as air or water is provided between these plate-like fins 1. Pass through.
The heat transfer tube 2 is a circular tube having an outer diameter D, and is made of a metal material such as copper or a copper alloy, or aluminum or an aluminum alloy (the same applies to other embodiments). Further, the groove 4 is formed in a spiral shape on the inner surface of the heat transfer tube 2, and has an angle (referred to as a twist angle) formed by a straight line parallel to the tube axis direction on the inner surface of the tube and the direction in which the groove 4 extends. In order to reduce the pressure loss of the heat exchanger, the effect of lowering the torsion angle is greater than the effect of increasing the number of passes. By mounting the heat exchanger 100 using the heat transfer tube 2 having a small torsion angle, the ceiling is embedded. An air conditioner such as a type air conditioner is configured.

  Here, the outer diameter D of the heat transfer tube 2 with the inner surface groove constituting the heat exchanger 100 is 4 to 6 mm, and the depth of the groove 4, that is, the height h of the peak 5 is 0.1 to 0.2 mm. When the refrigerant mass flow rate of the refrigerant flowing into the heat exchanger 100 and flowing in the heat exchanger 100 is 5 to 20 kg / h, the angle formed by the straight line parallel to the tube axis direction on the inner surface of the tube and the direction in which the groove 4 extends is 10 When the refrigerant mass flow rate is 20 to 45 kg / h at -20 °, the angle formed by the straight line parallel to the tube axis direction on the tube inner surface and the direction in which the groove 4 extends is 0-10 °.

  Next, the reason for numerical limitation in the heat transfer tube 2 with the inner surface groove of the present embodiment will be described.

(1) Outer diameter D: 4-6mm
When the outer diameter D of the heat transfer tube 2 is less than 4 mm, the amount of increase in pressure loss is greater than the amount of increase in heat transfer coefficient, and as a result, the performance in the tube is degraded. On the other hand, when the outer diameter D exceeds 6 mm, the row pitch (referred to as the center interval between the heat transfer tube rows in the fluid passage direction) becomes large, the distance between the heat exchanger and the fan becomes close, and noise is generated.
Therefore, in the heat transfer tube 2 with the inner surface groove of the present embodiment, the outer diameter D is set to 4 to 6 mm.

(2) Mountain height h: 0.1 to 0.2 mm
In the heat transfer tube 2 with the inner surface groove of the present embodiment, the heat transfer coefficient increases as the depth of the groove 4, that is, the peak height h increases. However, when the peak height h exceeds 0.2 mm, the amount of increase in pressure loss is greater than the amount of increase in heat transfer coefficient, and as a result, the in-pipe performance is degraded. On the other hand, when the peak height h is less than 0.1 mm, the heat transfer coefficient is not improved.
Therefore, in the heat transfer tube 2 with the inner surface groove of the present embodiment, the peak height h is set to 0.1 to 0.2 mm.

(3) Twist angle θ: 10 to 20 ° (when the refrigerant mass flow rate is 5 to 20 kg / h) or 0 to 10 ° (when the refrigerant mass flow rate is 20 to 45 kg / h)
The influence of the twist angle θ on the in-tube performance of the inner surface grooved heat transfer tube 2 varies depending on the refrigerant mass flow rate. FIG. 2 is a diagram showing the relationship between the twist angle θ and the heat exchange rate depending on the refrigerant mass flow rate.
When the refrigerant mass flow rate is 5 to 20 kg / h, when the twist angle θ is less than 10 °, only the performance equivalent to the conventionally used grooved tube having an outer diameter of 7 mm can be obtained. On the other hand, when the twist angle θ exceeds 20 °, the refrigerant flows over the inner groove, the pressure loss in the pipe increases, and the heat exchange rate cannot be improved. When the mass flow rate of the refrigerant is 20 to 45 kg / h, when the twist angle θ exceeds 10 °, the refrigerant flows over the inner groove, the pressure loss in the pipe increases, and the heat exchange rate cannot be improved.
Therefore, the torsion angle θ in the inner surface grooved heat transfer tube 2 of the present embodiment is such that the torsion angle θ is 10 to 20 ° when the refrigerant mass flow rate is 5 to 20 kg / h, and the torsion angle θ is 20 to 45 kg / h. The twist angle θ is set to 0 to 10 °.

Embodiment 2. FIG.
FIG. 3 is a side sectional view of the heat exchanger 100 according to Embodiment 2 of the present invention cut in the vertical direction, and FIG. 4 is a diagram showing the relationship between the number of heat transfer tube inner surface grooves and the heat exchange rate. In the second embodiment, the outer diameter D of the heat transfer tube 2, the depth of the groove 4, that is, the crest height h, and the twist angle θ of the groove 4 are the same as in the first embodiment.
As shown in FIG. 4, in the internally grooved heat transfer tube 2, when the number of the grooves 4 is smaller than 40, the workability at the time of manufacturing the heat exchanger is remarkably lowered, and the heat exchange rate is lowered. On the other hand, when the number of stripes exceeds 65, the groove cross-sectional area becomes small, and the refrigerant liquid film overflows from the groove 4 and is covered by the refrigerant liquid film up to the summit, so that the heat transfer coefficient decreases.
Therefore, in the inner surface grooved heat transfer tube 2 of the second embodiment, the number of grooves 4 is preferably 40 to 65.

Embodiment 3 FIG.
FIG. 5 is a side cross-sectional view of the heat exchanger 100 according to Embodiment 3 of the present invention cut in the vertical direction, with an enlarged view enlarging a part of the groove shape on the inner surface of the pipe. In the third embodiment, the outer diameter D of the heat transfer tube 2, the depth of the groove 4, that is, the crest height h, and the twist angle θ of the groove 4 are the same as in the first embodiment.
As shown in FIG. 5, in the internally grooved heat transfer tube 2, when the tube thickness t at the bottom of the groove 4 is smaller than 0.19 mm, the workability at the time of manufacturing the heat transfer tube is remarkably lowered, and the pressure resistance of the heat transfer tube is lowered. On the other hand, when the tube thickness t at the bottom of the groove 4 exceeds 0.23 mm, the top of the peak portion is crushed and the adhesion between the heat transfer tube and the fins is lowered, so that the heat transfer rate is lowered.
Therefore, in the inner surface grooved heat transfer tube 2 of Embodiment 3, the tube thickness t of the groove bottom is set to 0.19 to 0.23 mm.

Embodiment 4 FIG.
FIG. 6 is a diagram showing a configuration of an air conditioner according to Embodiment 4 of the present invention. As shown in FIG. 6, the air conditioner has an evaporator 51 that evaporates the refrigerant and cools air or water by the heat of vaporization at that time, and compresses the refrigerant discharged from the evaporator 51 to a high temperature. The compressor 52 supplied to the condenser 53, the condenser 53 that heats air or water by the heat of the refrigerant, and the expansion that supplies the refrigerant 51 discharged to the evaporator 51 at a low temperature by expanding the refrigerant discharged from the condenser 53 The valve 54 (throttle device) is sequentially connected by a pipe to form a refrigeration cycle. In the air conditioner of the present embodiment, the heat transfer tube 2 with the inner groove shown in the first to third embodiments is incorporated in a heat exchanger that constitutes the evaporator 51 and the condenser 53, and 5 A refrigerant containing refrigerating machine oil of mass% or less is flowed.

This refrigerating machine oil is a lubricant for the compressor 52, improves the lubricity of movable parts such as bearings and cylinders, and cools the movable parts by absorbing heat generated by friction, so that the compressor 52 Is to operate well. However, if the refrigeration oil content in the refrigerant exceeds 5% by mass, the amount of refrigeration oil that moves to the summit of the groove 22 of the heat transfer tube 2 increases, and the refrigerating machine oil is covered up to the summit and the in-pipe performance is reduced. descend.
Therefore, the refrigerating machine oil content of the refrigerant to be passed through the heat transfer tube with the inner surface groove of the present embodiment is 5 mass% or less.

  In the above heat exchanger, any one of HC single refrigerant, a mixed refrigerant containing HC, R32, R410A, R407C, or tetrafluoropropene (for example, 2,3,3,3-tetrafluoropropene) is used as the refrigerant. Thus, the efficiency of heat exchange between these refrigerants and air is improved.

Embodiment 5 FIG.
7A and 7B show a method for manufacturing the heat exchanger 100 according to Embodiment 5 of the present invention, and are front sectional views of the heat exchanger 100 cut in the vertical direction. The indoor unit-side heat exchanger and the outdoor unit-side heat exchanger are both manufactured by the same procedure.
As shown in FIG. 7, the inner surface grooved heat transfer tubes 2 shown in the first to third embodiments are each bent into a hairpin shape at a predetermined bending pitch at the center portion in the longitudinal direction, and a plurality of hairpin tubes (not shown) are formed. )). Next, these hairpin tubes are inserted in a direction orthogonal to the plurality of plate-like fins 1 arranged in parallel with each other at a predetermined interval, and thereafter, the tube expansion ball 60 is pushed into the hairpin tube by the rod 61. Each plate shape is expanded by a hairpin tube by a method (see FIG. 7 (a)) or by a fluid pressure expanding method in which the expanded ball 60 is pushed into the hairpin tube by the fluid pressure of the fluid 62 (see FIG. 7 (b)). The fin 1 and the hairpin tube, that is, the heat transfer tube 2 are joined. In this way, the heat exchanger 100 is manufactured.

  In the heat exchanger 100 according to the fifth embodiment, a large number of plate-like fins 1 and hairpin tubes (heat transfer tubes 2) can be obtained simply by expanding the hairpin tubes, which are constituent members, by a mechanical tube expansion method or a hydraulic tube expansion method. Since it joins, manufacture of the heat exchanger 100 becomes easy.

Embodiment 6 FIG.
In the fifth embodiment, the case where the plate-like fin 1 and the hairpin tube (heat transfer tube 2) are joined by expanding the hairpin tube is shown. However, in the sixth embodiment, the heat transfer tube 2 of the heat exchanger 100 is further expanded. The rate is specified.
In the sixth embodiment, the expansion rate when the hairpin tube is expanded by the mechanical expansion method or the hydraulic expansion method is set to 105.5% to 106.5% in the heat transfer tube 2 of the heat exchanger 100. Thereby, the adhesiveness of the heat exchanger tube 2 and the plate fin 1 of the heat exchanger 100 can be improved, and the highly efficient heat exchanger 100 can be obtained. However, when the expansion ratio of the heat transfer tube 2 of the heat exchanger 100 exceeds 106.5%, crushing at the peak of the groove and fin color cracking occur, and the adhesion between the heat transfer tube 2 and the plate fin 1 is improved. Getting worse. On the other hand, when the expansion ratio of the heat transfer tube 2 of the heat exchanger 100 is less than 105.5%, the adhesion between the heat transfer tube 2 and the plate-like fin 1 is poor, and a high heat exchange rate cannot be obtained.
Therefore, the expansion rate of the heat transfer tube 2 when expanding the hairpin tube of Embodiment 6 is set to 105.5% to 106.5%.
By defining the tube expansion rate of the heat transfer tube 2 in this way, the product (heat exchanger) does not vary.

  In Embodiments 5 and 6, the plate-like fin 1 and the hairpin tube (heat transfer tube 2) are joined only by expanding the heat transfer tube 2. However, after joining, the plate fin 1 and the hairpin tube 2 are further joined by brazing. In this case, reliability can be further improved.

Embodiment 7 FIG.
In the seventh embodiment, the heat exchanger 100 according to any one of the first to sixth embodiments of the present invention is used for an air conditioner.
Thereby, the pressure loss in a pipe | tube does not increase and the highly efficient air conditioner using the heat exchanger excellent in heat transfer performance can be obtained.

Embodiment 8 FIG.
8 to 10 illustrate a heat exchanger according to Embodiment 8 of the present invention. FIG. 8 is a plan view showing a part of the heat exchanger, and FIG. 9 is a view of the heat exchanger shown in FIG. 10A is a cross-sectional view taken along the line AA in FIG. 8, FIG. 10B is a cross-sectional view taken along the line BB in FIG. 8, and FIG. 10C is a cross-sectional view taken along the line C- in FIG. C sectional view and FIG. 10D are HH sectional views of FIG. In the following description, reference numerals “a, b, c...” Are omitted for those showing common contents.

  8 and 9, heat exchangers (hereinafter referred to as “heat exchangers”) 100 arranged in the air conditioner are arranged in parallel with a predetermined interval, and air passes through the interval. A plurality of plate fins 1 and a plurality of heat transfer tubes 21 inserted perpendicularly to the plate fins 1, and a plurality of slit fins facing the air flow direction on the surface of the plate fins 1. 3 is formed by cutting and raising.

(Heat transfer tube)
8 and 9, the heat transfer tube 21 is formed of a plurality of straight tube portions 2s and a plurality of curved tube portions 2r that allow end portions of the straight tube portions 2s to communicate with each other. The straight tube portions 21a and 21b of the heat transfer tubes 21 in the first row are arranged in a direction perpendicular to the air flow direction (hereinafter referred to as “step direction”). , 21d (not shown) are arranged. Similarly, the straight tube portions 22a of the second heat transfer tube 21 and the straight tube portions 23a, 23b of the third heat transfer tube 21 are arranged in the step direction. A direction parallel to the air flow direction is referred to as a “row direction”. In the heat exchanger 100 of the present embodiment, only three rows of straight pipe portions 2s are arranged in the row direction.
The straight pipe portions 21a, 21b,..., The straight pipe portions 23a, 23b,... The “step pitch Dp” that is the interval in the direction and the “column pitch Lp” that is the interval in the row direction are “4 mm ≦ D ≦ 6 mm, 14 mm ≦ Dp ≦ 17 mm, 7 mm with respect to the outer diameter D of the heat transfer tube 21. ≦ Lp ≦ 10 mm ”. Here, it is designed as D = 5 mm, Dp = 15.3 mm, and Lp = 8.67 mm.

(Plate fin)
8 to 10, the plate-like fin 1 is a rectangular plate material, and a plurality of through-holes through which the straight pipe portion 2 s of the heat transfer tube 21 passes are formed in a staggered manner.
Further, between the straight pipe portion 21a and the straight pipe portion 21b, first slit fins 3a, 3c, 3e projecting on one surface side of the plate-like fin 1 so as to face the air flow direction, The second slit fins 3b and 3d protruding to the other surface side are alternately formed by cutting and raising.

The first slit fins 3a, 3c, and 3e are obtained by cutting and raising the plate-like fin 1 on one surface side, and are first slit fin planes 32a, 32c, and 32e and supporting leg portions that support the first slit fins. 1 slit fin slopes 31a, 31c, 31e and first slit fin slopes 33a, 33c, 33e. Therefore, the first slit fin grooves 34a, 34c, 34e are formed in the plate-like fin 1 by such cutting and raising. Also, the first slit fin inclined surfaces 31a, 31c, 31e and the first slit fin inclined surfaces 33a, 33c, 33e of the support leg are formed as inclined surfaces so as to form a substantially concentric arc shape with the straight pipe portions 21a, 21b. .
Similarly, the second slit fins 3b and 3d are formed by cutting and raising the plate-like fin 1 to the other surface side, and the second slit fin planes 32b and 32d and the first support leg portions for supporting the second slit fin planes 32b and 32d. 2 slit fin slopes 31b, 31d and second slit fin slopes 33b, 33d. Therefore, second slit fin grooves 34b and 34d are formed in the plate-like fin 1 by such cutting and raising.

10D, the first slit fin groove 34a and the second slit fin groove 34b, the second slit fin groove 34b and the first slit fin groove 34c, and the first slit fin groove 34c and the second slit. The fin groove 34d, the second slit fin groove 34d, and the first slit fin groove 34e are respectively continuous. Therefore, for example, a large groove-shaped hole is formed in a range sandwiched between the straight pipe portion 21 a and the straight pipe portion 21 b of the plate-like fin 1. Therefore, since air passing through the inside and outside of the slit fin grooves 34a to 34e flows so as to surround the straight pipe portion 2s of the heat transfer tube 21, the heat transfer performance in the slit fin 3 is improved and the heat transfer performance is improved. The dead water area generated on the downstream side of the heat pipe 21 can be reduced.
Note that the protruding height (H1) of the first slit fins 3a, 3c, 3e from one surface of the plate-like fin 1 and the protruding height of the second slit fins 3b, 3d from the other surface of the plate-like fin 1 are as follows. The height (H2) is 1/3 of the fin pitch (Fp) which is the surface interval of the plate-like fins 1, that is, “H1 = Fp / 3, H2 = Fp / 3”.

Embodiment 9 FIG.
FIG. 11 explains the concept of a ceiling-embedded air conditioner according to Embodiment 9 of the present invention, in which (a) is a perspective view and (b) is a cross-sectional view.
In FIG. 11, a heat exchanger 100 (see Embodiment 8) is arranged in a ceiling-embedded air conditioner (hereinafter referred to as “air conditioner”) 200. A motor 8 that drives the fan 7 is provided on the central top surface side of the unit housing 6 of the air conditioner 200, and the fan 7 is attached to the motor 8 with the lower side serving as a suction port.
A bell mouth 9 for introducing air into the fan 7 is disposed below the fan 7. A heat exchanger 100 is arranged in a substantially annular shape surrounding the fan 7, and a drain pan 13 is arranged at the lower part of the heat exchanger 100. Each side of the drain pan 13 is formed with an opening connecting the secondary side of the heat exchanger 100 and the room, and communicates with the opening of the decorative panel 10 to form the air outlet 11.
A vane 12 is attached to the blowout port 11 to enable adjustment of the blowout direction. Further, a front panel 14 and a filter 15 are arranged at the lower part of the fan 7 so as to be fitted in the center of the decorative panel 10.

  The air conditioner 200 configured as described above is generally called a “four-way cassette type”, and the primary side of the fan faces downward and sucks air from the room. The sucked air passes through the filter 15, dust and the like are removed, and the air is blown to the heat exchanger 100. In the heat exchanger 100, heat is exchanged between the air and the refrigerant, and the air that has received heat or has been taken away is blown out into the room through the air outlet 11.

(Heat transfer performance and ventilation resistance)
Next, the qualitative tendency of the shape parameters of the heat exchanger 100 will be described below with respect to the heat transfer performance and the ventilation resistance of the heat exchanger 100.

(Influence of step pitch Dp)
When the step pitch Dp is increased, the “fin efficiency” defined by the distance from the outer periphery of the heat transfer tube 2 to the end of the plate-like fin 1 and the tube diameter of the heat transfer tube 2 is reduced. "Rate" decreases. Further, when the step pitch Dp is increased, the “ventilation resistance” decreases, so that “increase in air volume” can be achieved.
On the other hand, when the step pitch Dp is reduced, the “fin efficiency” is increased and the “external heat transfer coefficient” is improved, but the “ventilation resistance” is increased.

(Influence of row pitch Lp)
When the row pitch Lp is increased, the “fin efficiency” decreases and the “external heat transfer coefficient” decreases, but the heat transfer area increases, so the heat transfer performance of the heat exchanger improves. In addition, “ventilation resistance” increases and the air volume decreases.
On the other hand, when the row pitch Lp is reduced, the “fin efficiency” is increased and the “external heat transfer coefficient” is improved, but the heat transfer area is reduced, so that the heat transfer performance of the heat exchanger is reduced. In addition, “ventilation resistance” decreases, and “increase in air volume” can be achieved.
As described above, each shape parameter of the heat exchanger has an optimum value, and in order to quantitatively evaluate this, the heat transfer characteristics and the ventilation resistance of the heat exchanger are calculated by the method described below.

The heat transfer coefficient α [W / m ² K] between air and the plate fin is generally defined by the following equation.
α = Nu × λ / De ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ 1
Nu = C1 * (Re * Pr * De / Lp / Ln) ^ C2 Formula 2
Re = U × De / ν
Where Nu is the Nusselt number,
Re is the Reynolds number,
Pr is the Prandtl number,
De is the representative length
Lp is the row pitch,
Ln is the number of columns,
λ is the thermal conductivity of air,
ν is the kinematic viscosity coefficient of air,
C1 and C2 are constants.
In the case of normal temperature and pressure, Pr = 0.72, λ = 0.0261 [W / mK], and ν = 0.000016 [m ² / s].

Here, the representative length De [m] is defined by the following equation.
De = 4 × (Lp × Dp -π × D 2/4) × Fp / {2 × (Lp × Dp-π × D 2/4) + π × D × Fp} ··········・ ・ ・ ・ ・ ・ ・ ・ Formula 3
The wind speed U [m / s] based on the free passing volume between the plate fins and the front wind speed Uf [m / s] of the heat exchanger are defined by the following equations.
U = Uf × Lp × Dp × Fp / {(Lp × Dp-π × D 2/4) × Fp} ·· formula 4
Where D is the outer diameter of the heat transfer tube,
Dp is the step pitch,
Lp is the row pitch,
Fp is the fin pitch.

Further, the fin efficiency η is defined by the following equation.
η = 1 / (1 + ψ × α) ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ 5
ψ = {(4 × Lp × Dp / π) / 2−D} ² × (4 × Lp × Dp / π) / 2 / D / 2/6 / Ft / λf・ ・ ・ ・ ・ ・ ・ Formula 6
Here, λf [w / m · k] is the thermal conductivity of the plate fin.

On the other hand, the ventilation resistance “ΔP_hex [Pa]” between the air and the plate fin is defined by the following equation.
ΔP_hex = 2 × F × Lp × Ln × ρ × U ² / De (7)
F = C3 × De / Lp / Ln + C4 × Re ^ C5 × (De / Lp / Ln) ^ (1 + C5)
Here, F is a friction loss coefficient, and C3, C4, and C5 are constants. Further, ρ is the density of air, which is about 1.2 [kg / m ³ ] in the case of normal temperature and pressure.

(Blower operating force)
Further, in order to quantitatively evaluate the “blower operating force” when the heat exchanger 100 of the eighth embodiment is used in the air conditioner 200 of the ninth embodiment, the fan operating force is calculated by the following method. To do. The fan operating force Pf [W] is defined by the following equation.
Pf = ΔP_all × Q (1) in Equation 9
= (ΔP_hex + ΔP_etc) × Q (2 in Equation 9)
Here, Q is an air flow rate.

Hereinafter, “ΔP_hex” was calculated using the step pitch Dp and the row pitch Lp as parameters. Moreover, the heat passage rate K of the heat exchanger was calculated by the following formula.
K = 1 / (1 / αo + Ao / Ai / αi)
αo = 1 / (Ao / (Ap + η × Af) / α) Equation 11
Ao = Ap + Af Equation 12
Here, K [W / m ² K] is the total heat transmission rate of the heat exchanger,
Ao [m ² ] is the total heat transfer area on the air side of the heat exchanger,
Ap [m ² ] is the heat transfer area of the air side pipe of the heat exchanger,
Af [m ² ] is the air side fin heat transfer area of the heat exchanger,
Ai [m ² ] is the refrigerant side heat transfer area of the heat exchanger,
Equations 10 to 12 are values that can be calculated if the dimensions depending on the shape of the heat exchanger, that is, the step pitch Dp, the row pitch Lp, the fin pitch Fp, and the outer diameter D of the heat transfer tube are determined. It is assumed that the heat transfer coefficient αi [W / m ² K] of the fluid flowing in the pipe of the heat exchanger is constant.

In general, the energy consumption efficiency COP of an air conditioner is defined by the ratio of the heat exchange amount and the total input. By reducing the total input, the COP is improved, leading to energy saving.
Next, the total input is the sum of the compressor input and the fan operating force Pf. The compressor manpower decreases as AoK increases, and the blower operating force pf decreases as ΔP_hex decreases.
Here, a heat exchanger performance index “AoK / ΔP ^ n” is defined as a constant n. The constant n is “n = 1” when the ratio of the ventilation resistance “ΔP_hex” to the total ventilation resistance is 100%, and in the heat exchanger 100 of the air conditioner 200, the ratio of the total ventilation resistance is about half. Therefore, when ΔP_hex is doubled, tripled, or quadrupled, the overall ventilation resistance is 1.5 times, 2.0 times, or 2.5 times, respectively, and “n = 0. 59 ".
Therefore, in the heat exchanger 100 of the air conditioner 200, the heat exchanger tube diameter D and the step pitch Dp are set as the heat exchanger performance index “AoK / ΔP ^ 0.59” when the front wind speed Uf = 1 [m / s]. The relationship of the row pitch Lp was evaluated. In other air conditioners (referring to air conditioners other than ceiling-embedded air conditioners), for example, in the case of a room air conditioner indoor unit, ΔP_hex occupies about 80% of the overall ventilation resistance. ≈0.85 ”.
The larger the value of n, the larger the value of the air conditioner, ΔP_hex has a greater influence on the heat exchanger performance index “AoK / ΔP ^ n”. In the heat exchanger 100 of the ceiling-embedded air conditioner 200, other air It is characterized in that the influence of ΔP_hex is small compared to the harmonic machine.

FIGS. 13 to 16 show the effect on the heat exchanger performance index “AoK / ΔP ^ 0.59” in the heat exchanger arranged in the ceiling-embedded air conditioner according to Embodiment 8 of the present invention. 13 is a heat transfer tube diameter D, FIG. 14 is a step pitch Dp, FIG. 15 is a row pitch Lp, and FIG. 16 is a correlation diagram with a fin pitch Fp.
FIG. 13 shows the heat exchanger performance index “AoK” with the step pitch Dp = 15.3 mm, the row pitch Lp = 8.67 mm, the front wind speed Uf = 1 [m / s], and the heat transfer tube diameter D as a parameter. /ΔP^0.59 ".
When the heat transfer tube diameter D is less than 4 mm in terms of manufacturing technology, the work efficiency is significantly reduced in the step of inserting the tube expansion rod into the heat transfer tube and bringing it into close contact with the plate fin. On the other hand, when the heat transfer tube diameter D exceeds 6 mm, “AoK / ΔP ^ 0.59” is remarkably reduced. However, if D ≦ 6 mm, the heat transfer tube diameter D is 4 mm. Since the reduction is 3% or less, a heat exchanger with sufficiently high heat transfer performance can be supplied.
Therefore, by setting the range of “4 mm ≦ D ≦ 6 mm”, it is possible to supply the heat exchanger 100 with sufficiently high heat transfer performance without lowering the production efficiency.

FIG. 14 shows the heat exchanger performance index “AoK / ΔP with the step pitch Dp as a parameter, with the heat transfer tube diameter D = 5 mm, the row pitch Lp = 8.67 mm, and the front wind speed Uf = 1 [m / s]. It is the result of calculating “^ 0.59”.
In the vicinity of the step pitch Dp = 15 mm, the heat exchanger performance index “AoK / ΔP ^ 0.59” shows a maximum value, which is a decrease of 10% or less from the maximum value in the range of “14 mm ≦ Dp ≦ 17 mm”. When the step pitch Dp is less than 14 mm, in the process of bending the heat transfer tube into a hairpin shape, since the bending pitch is small, there is a possibility that the heat transfer tube becomes flat and the appearance is deteriorated or the pressure loss in the tube is increased. is there.
On the other hand, when the step pitch Dp exceeds 17 mm, it is necessary to reduce the number of paths between the heat transfer tubes when the arrangement volume of the heat exchanger is considered to be constant, but when the number of paths is decreased, the pressure loss in the pipe increases. Reduce the performance of the heat exchanger. In particular, as the heat transfer tube diameter decreases, the pressure loss inside the heat transfer tube tends to increase. Accordingly, it is desirable that the step pitch Dp is “14 mm ≦ Dp ≦ 17 mm”.

FIG. 15 shows the heat exchanger performance index “AoK / ΔP with the row pitch Lp as a parameter, with the pipe diameter D = 5 mm, the step pitch Dp = 15.3 mm, and the front wind speed Uf = 1 [m / s]. It is the result of calculating “^ 0.59”.
When the row pitch Lp is around 8 mm, the heat exchanger performance index “AoK / ΔP ^ 0.59” shows the maximum value, and in the range of “7 mm ≦ Lp ≦ 10 mm”, it is 10% or less lower than the maximum value. The heat exchanger 100 has high heat transfer performance.
If the row pitch Lp is smaller than 7 mm, it is difficult to form fin collars (holes and collar portions for inserting heat transfer tubes) in the plate-like fins in terms of manufacturing technology.
On the other hand, when the row pitch Lp exceeds 10 mm, the heat passing rate K decreases due to the fin efficiency decreasing, and in addition, the ventilation resistance ΔP increases, so that the heat exchanger performance index “AoK / ΔP ^ 0.59”. Is significantly reduced. Therefore, the row pitch Lp is preferably “7 mm ≦ Lp ≦ 10 mm”.

FIG. 16 shows the heat transfer tube diameter D = 5 mm. The step pitch Dp = 15.3 mm, the row pitch Lp = 8.67 mm, the front wind speed Uf = 1 m [m / s], and the ratio “Hl / Fp” between the height Hl of the cut and raised and the fin pitch Fp is a parameter. As a result of calculating the heat exchanger performance index “AoK / ΔP ^ 0.59”.
The ratio of the height H1 of the cut and raised to the pitch of the fin pitch Fp “Hl / Fp = 1/3”, air flow paths can be formed at equal intervals between the board portion and the raised portion of the plate-like fin for the highest efficiency. Since heat transfer can be improved, the heat exchanger performance index “AoK / ΔP ^ 0.59” indicates the maximum value, and the heat exchanger 100 has sufficiently high heat transfer performance.

Embodiment 10 FIG.
17 and 18 illustrate a heat exchanger disposed in a ceiling-embedded air conditioner according to Embodiment 10 of the present invention. FIG. 17 is a plan view showing a part of the heat exchanger, FIG. 18 is a front sectional view of the heat exchanger. The same parts as those in the eighth embodiment are denoted by the same reference numerals and described.

(Plate fin)
17 and 18, the plate-like fin 301 is a rectangular plate material, and a plurality of through holes through which the straight pipe portion 2s of the heat transfer tube 21 passes are formed in a staggered manner.
Further, first slit fins 3a, 3c, and 3e are formed between the straight pipe portion 21a and the straight pipe portion 21b, respectively, protruding to one surface side (one surface side). That is, the plate-like fin 301 is the same as the plate-like fin 1 of the eighth embodiment in which the second slit fins 3b and 3d are removed (not cut and raised).

Therefore, between the first slit fin 3a and the first slit fin 3c, a flat plate-shaped fin strip portion 35b which is a part of the plate-shaped fin 301 is provided between the first slit fin 3c and the first slit fin 3e. Between them, flat plate-shaped fin strips 35d, which are part of the plate-shaped fins 301, are respectively present. That is, the first slit fins 3a, 3c, and 3e are formed on every other side of the plate-like fin 301 in such a manner that the flat plate-like fin strip portions 35b and 35d are sandwiched, respectively.
The first slit fins 3a, 3c, and 3e have the same width in the air flow direction (referred to as “Wa” for convenience), and the plate fin strips 35b and 35d have the same width in the air flow direction (for convenience). , Referred to as “Wb”).
Thus, even when the three first slit fins 3a, 3c, and 3e are cut and raised in the row direction, the effect of the present invention can be obtained as in the eighth embodiment.

Embodiment 11 FIG.
19 and 20 illustrate a heat exchanger disposed in a ceiling-embedded air conditioner according to Embodiment 11 of the present invention. FIG. 19 is a plan view showing a part of the heat exchanger, FIG. 20 is a front sectional view of the heat exchanger. Note that the same reference numerals are used for the same parts as those in the eighth embodiment. Unless otherwise specified, the reference numerals “a, b, c... The description is omitted.

(Plate fin)
19 and 20, the plate-like fin 401 is the same as the plate-like fin 301 of the tenth embodiment in which the first slit fin 3c is removed (not cut up).

Therefore, two first slit fins 3a and 3e are formed between the straight pipe portion 21a and the straight pipe portion 21b in the row direction protruding to one surface side. And between the 1st slit fin 3a and the 1st slit fin 3e, the flat plate-shaped fin strip part 35c which is a part of plate-shaped fin 401 exists.
The width of the first slit fins 3a and 3e in the air flow direction is the same (referred to as “Wa” for convenience), and the width of the plate-shaped fin strip portion 35c in the air flow direction is referred to as “Wb” for convenience.
Thus, even when the two first slit fins 3a and 3e are cut and raised in the row direction, the effect of the present invention can be obtained as in the eighth embodiment.

[Effect of slit fins]
FIGS. 21 and 22 are correlation diagrams for explaining the effect of the slit fins in the heat exchangers shown in FIGS. 17 to 20.
In FIG. 21, the horizontal axis is the ratio “wa / wb” between the width wa in the row direction of the slit fins 3 a and the like and the width wb in the row direction of the plate-like fin strips 35 b existing between the slit fins. The vertical axis represents the heat exchanger performance index “AoK / ΔP_hex ^ 0.59”, which is the result of calculation using the former as a parameter.
From FIG. 21, when the ratio “wa / wb” is 1, that is, when “Wa: Wb = 1: 1, Wa = Wb”, the heat exchanger performance index “AoK / ΔP_hex ^ 0.59” is sufficiently large. It becomes a heat exchanger.

  In FIG. 22, the horizontal axis is “H2 / Fp” in which the height H2 of the slit fins 3b and the like is made dimensionless by the fin pitch Fp, and the vertical axis is the heat exchanger performance index “AoK / ΔP_hex ^ 0.59”. In this case, the calculation result is obtained using the former as a parameter. From FIG. 22, when the slit fin height H2 is ½ of the fin pitch Fp, the heat exchanger performance index “AoK / ΔP_hex ^ 0.59” is a sufficiently large heat exchanger.

Embodiment 12 FIG.
23 and 24 illustrate the concept of a ceiling-embedded air conditioner according to Embodiment 12 of the present invention. FIG. 23 is a bottom view, and FIG. 24 is a heat exchanger part of the air conditioner. It is sectional drawing. In addition, the same code | symbol is attached | subjected and demonstrated to the same part as FIG. 11 (Embodiment 9). Moreover, the heat exchanger in any one of Embodiment 1-11 shown above is used for the heat exchanger 100 of this embodiment.

In FIG. 23, a fan 7 is attached to the central top surface side of the unit housing 6 of the air conditioner 500 with the lower side as a suction port. Two heat exchangers 100 bent in an L shape so as to surround the fan 7 are arranged in a substantially annular shape.
In this way, by arranging two L-shaped heat exchangers 100 in a substantially annular shape, the refrigerant can be used as a heat transfer tube 2 as compared with the case where only one R-shaped heat exchanger is arranged in a substantially annular shape. Since the length passing through the inside can be reduced and the number of passes is doubled, the pressure loss of the refrigerant in the pipe can be reduced. This is an extremely effective means when the diameter of the heat transfer tube 2 is reduced.

  Therefore, when the heat exchanger 100 is used as an evaporator, it flows in 16 passes from the evaporator refrigerant inlet direction shown in FIG. 24, and the T-shaped three-way between the second and third rows in the air flow direction. The pipe distributes to 32 passes and exits to the outlet.

  Generally, when a refrigerant flows through the heat transfer tube of the heat exchanger of the evaporator, the state of the refrigerant changes in the order of the two-phase region and the superheated gas. At that time, the pressure loss “ΔP_ref” of the refrigerant is larger in the superheated gas than in the two-phase region. In the present embodiment, the pressure loss “ΔP_ref” of the refrigerant can be greatly reduced by the effect of increasing the number of passes from 16 to 32 passes between the second and third rows near the evaporator outlet. This is a very effective means for reducing the diameter of the heat transfer tube 2.

  When the heat exchanger 100 is used as a condenser, it flows in 32 passes from the condenser refrigerant inlet direction shown in FIG. 24, and is formed by a T-shaped three-way pipe having two rows and three rows in the air flow direction. , Merged into 16 passes and discharged to the exit.

Examples of the present invention will be described below in comparison with comparative examples that are out of the scope of the present invention.
As shown in Table 1, as Example 1 and Example 2, when the refrigerant mass flow rate is 5 to 20 kg / h, the outer diameter is 5 mm, the tube thickness of the groove bottom is 0.23 mm, and the peak height is 0.00. A heat exchanger 100 having a diameter of 15 mm and a twist angle of 10 degrees and 20 degrees was produced.
Further, as Comparative Example 1 and Comparative Example 2, a heat exchanger 100 having an outer diameter of 5 mm, a groove bottom tube thickness of 0.23 mm, a peak height of 0.15 mm, and a twist angle of 0 degrees and 30 degrees. As Comparative Example 3, a heat exchanger 100 having an outer diameter of 7 mm, a groove wall thickness of 0.23 mm, a peak height of 0.15 mm, and a twist angle of 35 degrees was produced.

  As is clear from Table 1, when the refrigerant mass flow rate is 5 to 20 kg / h, the heat exchangers 100 of Example 1 and Example 2 are both heated in comparison with the heat exchangers of Comparative Examples 1 to 3. The exchange rate was improving.

Next, as shown in Table 2, as Example 3 and Example 4, when the refrigerant mass flow rate is 20 to 45 kg / h, the outer diameter is 5 mm, the tube thickness of the groove bottom is 0.23 mm, and the peak height. Was 0.15 mm, and a heat exchanger 100 with twist angles of 0 degrees and 10 degrees was produced.
Further, as Comparative Example 4 and Comparative Example 5, a heat exchanger 100 having an outer diameter of 5 mm, a groove bottom tube thickness of 0.23 mm, a peak height of 0.15 mm, and a twist angle of 0 degrees and 30 degrees. As Comparative Example 6, a heat exchanger 100 having an outer diameter of 7 mm, a groove wall thickness of 0.23 mm, a peak height of 0.15 mm, and a twist angle of 35 degrees was produced.

  As is apparent from Table 2, when the refrigerant mass flow rate is 20 to 45 kg / h, the heat exchanger 100 of Example 3 and Example 4 is more heat-resistant than the heat exchangers of Comparative Example 4 to Comparative Example 6. The exchange rate was improving.

Next, as shown in Table 3, in Examples 5 to 7, when the refrigerant mass flow rate is 5 to 20 kg / h, the outer diameter is 5 mm, the tube thickness of the groove bottom is 0.23 mm, and the peak height. Was 0.15 mm, the twist angle was 20 degrees, and the heat exchanger 100 having 40, 50, and 65 strands was produced.
Further, as Comparative Example 7 and Comparative Example 8, a heat exchanger having an outer diameter of 5 mm, a groove wall thickness of 0.23 mm, a peak height of 0.15 mm, a twist angle of 20 degrees, and a number of strands of 30 and 75 As a comparative example 9, a heat exchanger 100 having an outer diameter of 7 mm, a groove bottom tube thickness of 0.23 mm, a peak height of 0.15 mm, a helix angle of 35 degrees, and a number of threads of 52 was produced.

  As is clear from Table 3, the heat exchangers 100 of Examples 5 to 7 all improved the heat exchange rate as compared to the heat exchangers of Comparative Examples 7 to 9.

  The present invention is not limited to the above-described embodiment and the illustrated configuration. Since the heat transfer performance is high as described above, the present invention can be widely used as various internal heat exchangers and various air conditioners equipped with the same. The present invention can also be used in a freezer dedicated to freezing, a dryer dedicated to generating hot air, and the like.

Sectional drawing which shows the cross section horizontal in the pipe-axis direction in the heat exchanger tube of the fin and tube type heat exchanger which concerns on Embodiment 1 of this invention, and side sectional drawing which cut | disconnected the same heat exchanger in the perpendicular direction. The diagram which shows the relationship between the twist angle of a heat exchanger tube inner surface groove | channel, and a heat exchange rate by the refrigerant | coolant mass flow rate in a heat exchanger. Side surface sectional drawing which cut | disconnected the heat exchanger which concerns on Embodiment 2 of this invention to the perpendicular direction. The diagram which shows the relationship between the number of strips of a heat-transfer tube inner surface groove | channel, and a heat exchange rate. Side surface sectional drawing which cut | disconnected the heat exchanger which concerns on Embodiment 3 of this invention to the perpendicular direction. The block diagram of the air conditioner which concerns on Embodiment 4 of this invention. The front sectional view which shows the manufacturing method of the heat exchanger which concerns on Embodiment 5 of this invention, and cut | disconnected the heat exchanger in the perpendicular direction. The top view which shows a part of heat exchanger which concerns on Embodiment 8 of this invention. Sectional drawing of the front view of the heat exchanger shown in FIG. Sectional drawing of each of AA of a heat exchanger shown in FIG. 8, BB, CC, and HH. The perspective view explaining the concept of the ceiling embedded type air conditioner which concerns on Embodiment 9 of this invention. Sectional drawing of the ceiling embedded type air conditioner shown in FIG. The correlation diagram which shows the influence of the heat exchanger tube diameter D which acts on the heat exchanger performance parameter | index in the heat exchanger arrange | positioned at the ceiling embedded type air conditioner shown in FIG. The correlation diagram which shows the influence of the stage pitch Dp which acts on the heat exchanger performance parameter | index in the heat exchanger arrange | positioned at the ceiling embedded type air conditioner shown in FIG. The correlation diagram which shows the influence of the row | line | column pitch Lp on the heat exchanger performance parameter | index in the heat exchanger arrange | positioned at the ceiling embedded type air conditioner shown in FIG. The correlation figure which shows the influence of the fin pitch Fp which acts on the heat exchanger performance parameter | index in the heat exchanger arrange | positioned at the ceiling embedded type air conditioner shown in FIG. The top view which shows a part of heat exchanger which concerns on Embodiment 10 of this invention. Sectional drawing of the front view of the heat exchanger shown in FIG. The top view which shows a part of heat exchanger which concerns on Embodiment 11 of this invention. Sectional drawing of the front view of the heat exchanger shown in FIG. The correlation diagram explaining the effect of the slit fin in the heat exchanger shown in FIGS. FIG. 14 is a correlation diagram for explaining the effect of slit fins in the heat exchanger shown in FIG. 13 and the like. The bottom view explaining the concept of the ceiling embedded type air conditioner concerning Embodiment 12 of this invention. FIG. 24 is a cross-sectional view of a heat exchanger part of the ceiling-embedded air conditioner shown in FIG. 23.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 Plate-like fin, 2 Heat transfer tube, 2r Curved pipe part, 2s Straight pipe part, 3 Slit fin, 3a, 3c, 3e 1st slit fin, 3b, 3d 2nd slit fin, 4 groove | channel, 5 crest, 21 Heat transfer tube 21a, 21b, 21c Straight pipe part, 22a Straight pipe part, 23a, 23b Straight pipe part, 31a, 31c, 31e First slit fin slope, 31b, 31d Second slit fin slope, 32a, 32c, 32e First slit Fin plane, 32b, 32d Second slit fin plane, 33a, 33c, 33e First slit fin slope, 33b, 33d Second slit fin slope, 34a, 34c, 34e First slit fin groove, 34b, 34d Second slit fin Groove, 35b, 35c, 35d Plate-shaped fin strip, 51 Evaporator, 52 Compressor, 53 Condenser, 5 Expansion valve, 100 heat exchanger, 200 air conditioner, 301 plate fin, 401 plate fin, 500 air conditioner.

Claims (14)

A plurality of plate-like fins arranged side by side at a predetermined interval, and a plurality of plate-like fins through which the fluid passes, and a plurality of circular heat transfer tubes that are inserted in a direction perpendicular to the plate-like fin and through which the refrigerant flows,
The heat transfer tube has an outer diameter of 4 to 6 mm, a spiral groove having a depth of 0.1 to 0.2 mm is formed on the inner surface of the tube, and a refrigerant mass flow rate of 5 to 20 kg / h. When the angle formed by the straight line parallel to the pipe axis direction on the pipe inner surface and the direction in which the groove extends is 10 to 20 ° and the refrigerant mass flow rate is 20 to 45 kg / h, the straight line parallel to the pipe axis direction on the pipe inner face is An angle formed by a direction in which the groove extends is 0 to 10 °.   The heat exchanger according to claim 1, wherein the number of grooves is 40 to 65.   The heat exchanger according to claim 1 or 2, wherein a tube thickness of the groove bottom is 0.19 to 0.23 mm.   The heat exchanger according to any one of claims 1 to 3, wherein a step pitch of the heat transfer tubes arranged in a step direction perpendicular to the flow direction of the fluid is 14 to 17 mm.   The heat exchanger according to any one of claims 1 to 4, wherein a row pitch of the heat transfer tubes arranged in a row direction parallel to the fluid flow direction is 7 to 10 mm.   In the plate-like fins, a plurality of slit fins facing each other in the fluid flow direction are formed between the heat transfer tubes, and the inclined surfaces so that the support legs of the slit fins are substantially concentric arcs with the heat transfer tubes. The heat exchanger according to claim 1, wherein the heat exchanger is formed of   The slit fins are alternately formed by cutting and raising on both sides of the plate-like fin, and the ratio between the height of the slit fin raising and the fin pitch of the plate-like fin is about 1/3. The heat exchanger according to claim 6.   The slit fins are formed by cutting and raising every other side so that a flat fin strip portion exists on one side of the plate-like fin, and the slit fin raising height and the fin pitch of the plate-like fin are The heat exchanger of claim 6 wherein the ratio is about 1/2.   9. The heat exchanger according to claim 8, wherein a width in the row direction of the plate-like fins is equal to a width in the row direction of the fin strips.   The heat exchanger according to any one of claims 1 to 9, wherein a tube expansion rate at the time of joining by expanding the heat transfer tube to the plate fin is 105.5 to 106.5%.   The heat exchanger according to any one of claims 1 to 10, wherein a refrigerant containing 5% by mass or less of refrigeration oil flows in the heat transfer tube.   In a refrigeration cycle in which a compressor, a condenser, a throttling device, and an evaporator are sequentially connected by piping and a refrigerant is used as a working fluid, the heat exchanger according to any one of claims 1 to 11 is replaced with the evaporator or the condenser. An air conditioner characterized by being used as   The heat exchanger according to any one of claims 1 to 11, wherein the heat exchanger is formed in an L shape, and the heat exchanger is formed in a substantially annular shape using two of the L shape heat exchangers. Embedded ceiling type air conditioner.   The air conditioner according to claim 12 or 13, wherein any one of HC single refrigerant, a mixed refrigerant containing HC, R32, R410A, R407C, tetrafluoropropene, and carbon dioxide is used as the refrigerant.

Images