JP2002139282A - Heat exchanger, refrigerating air conditioner and manufacturing method of heat exchanger - Google Patents

Abstract

PROBLEM TO BE SOLVED: To resolve a problem that a conventional heat exchanger employing a circular tube or a flat tube does not permit to improve remarkably a ventilating resistance or a heat transfer performance and contrive energy saving due to the restriction in manufacturing or structure. SOLUTION: A multitude of sheet type fins are arranged in parallel and flat type heat transfer tubes are inserted from the upstream side of air flow into groove holes to braze the tube to the fin whereby the heat exchanger, miniaturized in size, reduced in a ventilating resistance and excellent in heat transfer performance can be obtained.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a structure and a manufacturing method of a heat exchanger using a fin and a heat transfer tube for exchanging heat between a refrigerant and a fluid such as a gas. Related to an air-conditioning and refrigeration system.

[0002]

2. Description of the Related Art FIGS.
This will be described with reference to FIG. FIG. 22 is a partial side view showing a fin tube type heat exchanger used in the air-conditioning refrigeration apparatus disclosed in Japanese Patent Laid-Open No. 33595/1990. This heat exchanger is generally called a plate fin tube type, and is disposed at regular intervals and through which gas (air) flows (the gas passing direction is indicated by 40 in the figure).
And a heat transfer tube 2 inserted at a right angle into each of the plate-like fins and through which a refrigerant flows, and a plate-like shape between adjacent heat transfer tubes in a stepwise direction (a direction perpendicular to the gas passing direction). A slit group 50 is provided on the fin surface. The slit group is located such that the side end of the slit 50 faces the wind direction, and at this side end, an effect of updating the velocity boundary layer and the temperature boundary layer of the air flow can be expected, and heat transfer is promoted. It is said that the heat exchange capacity is increased. Also,
Legs formed at both ends of the slit 50 and having cut and raised plate-like fin surfaces are installed at an angle to the gas passing direction 40. By doing so, the heat transfer tube 2
It is considered that there is an effect of preventing a decrease in heat transfer due to a dead water region (a velocity loss region generated in a downstream portion of the heat transfer tube) generated downstream of the heat transfer tube 2 when a flow is formed along the heat transfer tube 2 and there is no slit. .

Further, as shown in FIG. 23, there is a heat exchanger using a heat transfer tube 2 having a flat cross section in the flat fin heat exchanger of FIG. By using a flat tube as the heat transfer tube, there is an advantage that ventilation resistance can be significantly reduced as compared with a circular heat transfer tube.

In order to manufacture a heat exchanger using such a flat heat transfer tube, as shown in FIG. 24, a large number of plate fins 1 stacked at appropriate intervals are fixed with a jig. ,
The plate-shaped fins 1 are inserted into the insertion holes 6 of the plate-shaped fins 1 so that the plate-shaped fins 1 are brought into close contact with the flat heat exchanger, and then are brought into close contact with a brazing material or an adhesive.

In the conventional method described above, the heat exchanger of FIG. 22 has a circular heat transfer tube,
Because the ventilation resistance in the part where the slit is installed increases,
Since the air selectively flows in a region where the ventilation resistance near the heat transfer tube where the slit does not exist is relatively small, the flow velocity of the air flowing between the slits 50 decreases, and the expected heat transfer promoting effect cannot be obtained sufficiently. . This is because there has not been enough research to optimally control the formation of the boundary layer on the slit with respect to the air flow flowing between the fins with respect to the slit width and the number of installed slits. There was a problem that the performance was not sufficiently extracted. Also,
Since the same number of slits are cut and raised on the fin 1 surface as the upstream and downstream, only the ventilation resistance increases in the downstream portion where the air temperature and the refrigerant temperature are close, despite the small heat exchange amount. There was a point. In addition, since a circular tube is used for the heat transfer tube 2, there is a problem that it is difficult to suppress the ventilation resistance in the heat transfer tube portion and the dead water area generated downstream of the heat transfer tube. In addition, the fin-
Since the pipes are joined, the contact heat resistance generated between the plate fins and the pipes is large, and there is a problem that the heat transfer rate of the air-refrigerant pipes is reduced.

In the heat exchanger of FIG. 23, when the flat tube 2 is inserted into the insertion hole 6 of the plate-like fin 1 as shown in FIG. 24, the plate-like fin 1 is bent by friction as shown in FIG. As a result, the interval between the plate-like fins 1 becomes uneven, which not only deteriorates the appearance of the heat exchanger but also increases the ventilation resistance. Therefore, in order to insert the flat heat transfer tube 2 into the insertion hole 6 of the plate-like fin 1, there is a problem that high accuracy and skill are required, assembling and manufacturing are troublesome and time consuming, and the manufacturing cost is increased. . As a technology using a flat heat transfer tube, a heat transfer tube described in Japanese Patent Application Laid-Open No. H10-89870 with a tapered shape to improve the assemblability has been proposed. The performance of heat transfer and ventilation as a heat exchanger is neglected, and such a technique has a problem that it is not possible to obtain a size reduction and energy saving effect.

SUMMARY OF THE INVENTION The present invention has been made to solve the above-described problems, and reduces the dead water area downstream of a pipe in a heat transfer pipe section of a heat exchanger, thereby reducing ventilation resistance. And the amount of heat exchange is improved. Further, the present invention is to obtain a heat exchanger and a refrigerating air conditioner having good heat transfer efficiency.
Another object of the present invention is to provide a heat exchanger and a method of assembling the heat exchanger, which have excellent adhesion between the heat transfer tubes and the plate-like fins, good heat transfer performance, and excellent assemblability.

[0007]

The heat exchanger according to claim 1 of the present invention is arranged in parallel with a plurality of plate fins through which a primary refrigerant flows, and inserted into each of the plate fins,
And a plurality of heat transfer tubes arranged in a flat shape and provided with a plurality of chambers through which the secondary refrigerant flows, and a plurality of heat transfer tubes arranged to flow the primary refrigerant flowing through the plate-like fins along the long axis direction of the flat shape.
The flat shape of the heat transfer tube reduces the length of the flat refrigerant in the short axis direction as the flow of the primary refrigerant proceeds in the leeward direction.

A heat exchanger according to a second aspect of the present invention is arranged in parallel with a plurality of plate-like fins through which a primary refrigerant flows, and inserted into each of the plate-like fins to flow a secondary refrigerant therein. Flat-shaped heat transfer tube and fin fin formed on the leeward side of the primary refrigerant from the end of the plate-like fin so that the heat transfer tube can be inserted from the end on the windward side where the primary refrigerant through which the plate-like fin flows , And the heat transfer tube is always open on the windward side from the plate-like fin.

According to a third aspect of the present invention, in the heat exchanger, the front edge of the heat transfer tube comes into contact with the plate-like fin up to near the windward end of the plate-like fin and projects from the plate-like fin at this end. It is open to the public.

In the heat exchanger according to a fourth aspect of the present invention, the heat transfer tube is inserted so as to be covered with the plate-like fin at an interval with respect to the most downstream end of the primary refrigerant of the plate-like fin.

[0011] In the heat exchanger according to claim 5 of the present invention, the outer shape of the heat transfer tube has a curved surface as a whole, and the curvature increases toward the leeward direction.

In a heat exchanger according to a sixth aspect of the present invention, a straight portion is provided in an outer tube shape.

In a heat exchanger according to a seventh aspect of the present invention, the plurality of chambers in the heat transfer tube have substantially the same sectional area.

In the heat exchanger according to claim 8 of the present invention, the flattened short axis of the heat transfer tube is da and the long axis is db, and the flatness is
H (= db / da) is 8 ≦ H.

According to a ninth aspect of the present invention, when the width of the plate-shaped fin in the direction in which the primary refrigerant flows is L and the long axis of the heat transfer tube is db, 1 ≦ L / db ≦ 1.5. is there.

A heat exchanger according to a tenth aspect of the present invention comprises:
Assuming that a plurality of heat transfer tubes are arranged, and a stepwise pitch perpendicular to the flow of the primary refrigerant in this arrangement direction is Dp, and a flat short axis of the heat transfer tubes is da, 0.6 ≦ (Dp−da) / Dp ≤
0.95.

[0017] The heat exchanger according to claim 11 of the present invention comprises:
The distance between the windward end of the plate fin and the front edge of the heat transfer tube is L
Assuming that 1, L1 <da / 8.

A heat exchanger according to a twelfth aspect of the present invention comprises:
A slit provided on the surface of the plate-like fin and having an opening facing the flow of gas, and having a half length L of the fin width L.
The number of slits provided in the upstream portion is larger than the number of slits provided in the downstream portion with respect to the center line of / 2.

A heat exchanger according to a thirteenth aspect of the present invention comprises:
The intersection of the leg of the slit provided on the surface of the plate-shaped fin and the plate-shaped fin surface is substantially parallel to the direction of the local gas flow near the fin surface.

A heat exchanger according to a fourteenth aspect of the present invention includes:
The plate-like fin and the heat transfer tube are heated and fixed.

A heat exchanger according to a fifteenth aspect of the present invention comprises:
The same material is used for the heat transfer tube and the plate-like fin.

A refrigeration air conditioner according to a sixteenth aspect of the present invention includes a refrigeration cycle for circulating a secondary refrigerant through a compressor, a condenser heat exchanger, a throttle device, and an evaporator heat exchanger. The heat exchanger of the present invention is used for at least one of an exchanger and an evaporator heat exchanger.

A refrigeration / air-conditioning apparatus according to a seventeenth aspect of the present invention includes a blower for flowing a primary refrigerant between fins of a condenser heat exchanger or an evaporator heat exchanger. In this case, a propeller fan is used as a blower that blows air to a heat exchanger having the following.

According to a method for manufacturing a heat exchanger according to the eighteenth aspect of the present invention, a large number of plate-shaped fins are arranged in parallel and a primary refrigerant flows between the plate-shaped fins, and the secondary refrigerant is fixed to each of the plate-shaped fins. A plurality of heat transfer tubes that are formed in a flat shape and through which the primary refrigerant flowing through the plate-shaped fins flows along the flat long axis direction.
Forming a hole in the plate-shaped fin to open one end in the direction in which the primary refrigerant flows and closing the other end; and inserting a heat transfer tube into the hole in the plate-shaped fin from one end, which is always the open side. And assembling the plate-shaped fins and the heat transfer tubes, and then heating and fixing the fins and the heat transfer tubes.

In the method for manufacturing a heat exchanger according to the nineteenth aspect of the present invention, the plate-like fin and the heat transfer tube are heated and fixed, and then the plate-like fin is coated with a hydrophilic material.

[0026]

DESCRIPTION OF THE PREFERRED EMBODIMENTS Embodiment 1 FIG. 1 is a diagram showing fins and heat transfer tubes of a heat exchanger according to Embodiment 1 of the present invention, and FIG. 2 is a plan view thereof. FIG. 3 is a side view of the fins and the heat transfer tubes of the heat exchanger according to the first embodiment viewed from the AA cross section in FIG. A plate-like fin 1 and a heat transfer tube 2 inserted perpendicular to the plate-like fin 1 are arranged in parallel to the air flow, and the plate-like fin surface between the heat transfer tubes adjacent in a stepwise direction. Have slits 11 to 15 provided therein.

In this structure, the pitch Fp of the fins 1 in the stacking direction is Fp = 0.012 m, and the fin thickness Fp
t = 0.0001 m, the fin width in the air flow direction is L = 0.0254 m, and the front wind speed Uf of the heat exchanger is Uf
= 1.0 m / s, and the stage pitch Dp, which is the distance between the centers of the heat transfer tubes adjacent in the stage direction of the heat exchanger, is Dp = 0.133.
m, as the heat transfer tube goes downwind, the short-axis length x becomes smaller, the curvature of the tube side becomes larger, that is, the radius becomes smaller, and the fin collar 4 and the heat transfer tube 2 extend to the fin leading edge. The heat transfer tube is completely joined by brazing, and the heat transfer tube is not in contact with the fin collar only at the front end portion of the heat transfer tube. Further, in order to maintain the pressure resistance, the heat transfer tube is provided with nine partitions in the tube, and the inside of the tube is divided into ten chambers. Further, the cross-sectional area of the refrigerant flow path of each chamber in the direction perpendicular to the refrigerant flow direction is the same. The length of the portion where the heat transfer tube protrudes from the fin from the leading edge of the fin is L1 =
0.0003 m. In addition, the major axis diameter parallel to the air flow direction is db = 0.022 m, the minor axis diameter of the leading edge in the air flow direction is da = 0.0025 m, and the flatness is H = db / da = 8. .8. The number of rows, which is the number of series in the direction in which the air flows through the heat transfer tubes, is an example of one row. In addition, five slits are provided between the heat transfer tubes in the step direction.

In the heat exchanger configured as described above, the heat transfer tube 2 is formed by an extruded member made of an aluminum alloy, and the plate-like fin 1 is formed by a plate made of an aluminum alloy. The air flow as the primary refrigerant passes between the fins and exchanges heat with the fins and the heat transfer tubes.

FIG. 4 shows a fin collar 4 provided on the plate-like fin 1 and a fin expanding portion 5 which is a hole for removing the fin. The fin collar stands up over the entire fin extension. Further, the fin expanding portion has a wing shape, that is, a rattan shape, like the outer shape of the heat transfer tube 2. Thus, the curvature, that is, 1
/ The radius is large.

A method of assembling the heat exchanger configured as described above will be described with reference to FIGS. First, as S1, a flat hole is formed in a fin for inserting a heat transfer tube by a press. At this time, the fin expanding portion is formed at the same time. A predetermined number of plate-like fins 1 are fixed between jigs (not shown), set at regular intervals, and the holes are aligned to form a fin assembly structure.

Next, a brazing material is applied to the heat transfer tube (S2), and as shown in FIG. 5A, the heat transfer tube 2 is inserted into the expanded portion from the end of the fin (S3). ), As shown in the figure
The heat transfer tube 2 is moved until the fin collar 4 and the heat transfer tube in 3 come into close contact. Further, the heat transfer tube and the header part are temporarily assembled at the end position of the heat exchanger (S4), and are put into a vacuum continuous furnace to perform the heat joining of the brazing part (S5). After taking out from the furnace, a hydrophilic coating material is applied to the fin surface (S6). Finally, drying is performed without raising the temperature (S7).

As described above, when the brazing is performed as a method of bringing the fin collar and the heat transfer tube into close contact with each other, a conventional method (FIG. 23) is used.
When a brazing material is attached to the heat transfer tube as in the method of penetrating the heat transfer tube through the fins as in the case of -25), it is difficult to secure the clearance, and thus it is difficult to insert the heat transfer tube. In the case of the present embodiment, since the heat transfer tube is fitted from the front edge of the fin, if the brazing material is applied to the heat transfer tube, the fin and the brazing material can be completely adhered. Further, in the case of the present embodiment, since the heat transfer tube has a general shape in which the short-axis length x is reduced with respect to the leeward side, the brazing material is smaller than that of the conventional flat tube.
It is easy to smoothly spread over the entire contact portion of the fin, and it is expected that the contact heat transfer coefficient αc will increase to almost infinity. As described above, even when the fin is not formed into a wing shape or a rabbit shape but is mainly a straight shape, the manufacturing is simplified by assembling the fin from the side of the fin assembly into the entire fin. In the present invention, since the heat transfer tubes are always inserted from one side, which is on the windward side of the heat exchanger, from the state in which many fins are assembled, the manufacturing apparatus is simple and mass production is easy. Furthermore, even if the brazing method is not used, after the heat transfer tube is inserted into the fin assembly, a load is applied to the entire fin in the direction to narrow the expanded portion, thereby improving the adhesion and joining by high frequency heating. is there.

In the heat exchanger having the above structure, the heat transfer performance (excluding the contact heat transfer coefficient between the fin and the heat transfer tube) and the ventilation resistance of the heat exchanger, and the qualitative tendency of the shape parameters described above are as follows. Will be described. In a heat exchanger having a flat heat transfer tube, the ventilation resistance decreases as the flatness ratio (= long axis da / short axis db) increases. In the case where a slit is provided, the fin efficiency increases as the flattening rate increases.

In the case of a tube heat exchanger, the distance from the heat transfer tube to the fin ends is long, and the difference between the heat transfer tube temperature and the fin end temperature is large. On the other hand, in the heat exchanger of the present embodiment, the major axis diameter of the heat transfer tube occupies 87% of the fin width,
The distance that heat is transferred from the heat transfer tube to the end of the fin is small. Therefore, the difference between the heat transfer tube temperature and the fin end temperature is small. Therefore, the fin efficiency of the heat exchanger of the present embodiment is higher than that of the tube heat exchanger. In addition, the circumferential length of the heat transfer tube 2 increases, and the heat transfer area on the outer peripheral portion of the heat transfer tube 2 increases, so that the amount of heat exchange on the air side increases. In addition, the amount of heat exchange increases as the area of the contact portion between the fin and the tube increases. In addition, since the heat transfer area increases in the pipe, the heat transfer performance on the refrigerant side is improved.

FIG. 7A shows an indoor heat exchanger and a blower when the heat exchanger of the present embodiment is used in a refrigeration cycle. The blower is an example of a once-through blower. FIG.
(b) shows the size of the dead water area when the flat tube having a straight side surface and the heat transfer tube in the present embodiment are placed in the wind tunnel. Since the dead water area becomes smaller as the flattening rate increases, the heat transfer coefficient improves, and the noise generated by the blower can be suppressed. In the case of a circular pipe, since the projected cross-sectional area is large with respect to the wind direction, the dead water area 10 (velocity loss area generated in the downstream part of the heat transfer pipe) in the downstream part of the pipe is large. In a normal flat tube, the side of the tube is desired to be horizontal to the wind direction, so that the wake of the tube is easily separated. On the other hand, the heat exchanger of the present embodiment has a high flatness and a large projected sectional area with respect to the wind direction of the tube. Further, in the heat exchanger of the present embodiment, since the side surface of the tube is streamlined, separation of the main flow at the side surface of the tube hardly occurs, and the dead water area 10 is extremely small.
When this dead water area separates from the fins and hits the blower, it cuts off the uneven wind, thereby significantly increasing the noise. Therefore, it is necessary to prevent the dead water area from extending from the fins to the outside. For this reason, it is more advantageous to increase the curvature of the outer shape of the heat transfer tube, change the flatness, or make it narrower straight to make the dead water area smaller toward the leeward side.

Although the inside of the heat transfer tube is divided into ten chambers as shown in FIG. 3 and the like, as the number of divisions increases, the deformation of the heat transfer tube due to the pressure of the refrigerant is suppressed. Also, as the number of divisions increases, the heat transfer area in the tube increases, so that the heat exchange capacity improves. Further, it is preferable that the cross-sectional area of the flow passage of each individual chamber is as constant as possible. This is because the flow rate, refrigerant pressure loss, and heat exchange amount in each chamber can be made uniform.

FIG. 8A shows the state of dew adhering during cooling in the heat transfer tube having a straight side surface and the heat transfer tube of the present embodiment. The dew 16 hangs down in the drawing under the influence of the direction of gravity g downwind of the heat transfer tube. In the case of the present embodiment, compared with a heat transfer tube having a straight side surface,
Since the curvature of the side surface of the heat transfer tube increases as it goes downwind, the surface tension acts in the windward direction, and the dew 16 is more easily attached to the side surface of the heat transfer tube as compared with a heat transfer tube having a straight side surface.
It is difficult to hold the dew 16 in the heat transfer tube. Also,
Since the fin arrangement direction of the heat exchanger is the vertical direction, FIG.
As shown in (b), on the side of the tube, the heat transfer tube of the present embodiment is less likely to peel off than the heat transfer tube having a straight side, so that the wind speed on the side of the tube is large and dew flows downstream of the tube. It is easy to flow down the fins from the part. Therefore, when the heat transfer tube of the present embodiment is used for a heat exchanger, the increase in ventilation resistance is less susceptible to the effects of dew as compared with a heat exchanger using a heat transfer tube having a straight side surface.

FIG. 8B shows a case where the heat exchanger of the present embodiment is used as a cooler. In the case of the heat exchanger of the present embodiment, since there is an interval between the rear edge of the heat transfer tube and the rear edge of the fin, the dew 16 flows near the rear edge of the fin using this interval as a flow path, and Don't jump out.
Therefore, it is necessary to provide a space between the rear edge of the heat transfer tube and the rear edge of the fin. Further, when the fin width L is reduced, the heat transfer area is reduced and the heat exchange amount is reduced, but the heat transfer coefficient is improved, the ventilation resistance is reduced, and the air volume can be increased. In the case of the elliptical heat exchanger, the ventilation resistance of the heat transfer tube portion is smaller than that of the circular tube heat exchanger, so that the reduction of the ventilation resistance by reducing the fin width increases. That is, a small-sized apparatus having good ventilation performance can be obtained.

When the distance L ₁ from the leading edge of the fin to the leading edge of the heat transfer tube protruding from the wind direction is reduced, the amount of heat exchange increases (however, the shape of the heat transfer tube is constant). This is because the area Ac [m ² ] of the contact portion between the fin and the heat transfer tube becomes large, the area where the heat transfer tube is within the fin width is large, and the temperature of the fin becomes close to the temperature of the heat transfer tube as a whole. Therefore, the fin efficiency is increased. Next, in order to quantitatively evaluate the shape parameters described above, the heat transfer performance and ventilation resistance of the heat exchanger are calculated by the method described below.

The heat transfer coefficient α [W /
(M ² K)] is generally defined by the following equation. α = Nu × λ / De Nu = C ₁ × (Re × Pr × De / L / Ln / 2) ^C ₂ Re = U × De / ν where Nu is the Nusselt number and Re is the Reynolds number. Pr is the number of prandles, λ is the thermal conductivity of air, ν is the kinematic viscosity of air, Ln is the number of rows, and Pr = 0.72 and λ = 0.0261 [W / (M
K)], and ν = 0.000016 [m ² / s]. C ₁ and C ₂ are coefficients. Here, the representative length De [m]
Is defined by the following equation. De = 4 × (L × Dp / 2−π × da × db / 4) × (F
p-Ft) / ｛2 × (L × Dp / 2−π × da × db /
4) + π × ((da ² + db ² ) / 2) ^1/2 × (Fp-F
t)) The wind speed U [m / s] based on the free passage 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 × L × Dp × Fp / ((L × Dp / 2−π / 4 ×
da × db) × (Fp−Ft)) / 2

The fin efficiency η is defined by the following equation. η = 1 / (1 + ψ × α) ψ = {(2 × L × Dp / π) ^0.5 − ((da ² + db ² ) / 2)
^1/2 } ² × (2 × L × Dp / π) ^0.5 / ((da ² + db ² ) / 2)
^{1/4 / 6 / Ft / λf ×} (1/2) 0.5 × (L / L 1/6) 0.2 λf [W / (mK)] is the thermal conductivity of the plate fin.

On the other hand, the pressure loss (ventilation resistance) ΔP [Pa] between the air and the plate-like fin is defined by the following equation. ΔP = F × L × Ln × ρ × U 2 / De F = C 3 × De / L / Ln / 2 + C 4 × Re C 5 × (De / L /
Ln / 2) ^{1 + C} ₅ where F is a friction loss coefficient, and C ₃ , C ₄ and C ₅ are coefficients. Ρ is the density of air, which is 1.2 at normal temperature and normal pressure.
[kg / m ³ ].

Further, in order to reduce the blower driving force when the heat exchanger in the present embodiment is used for an air conditioning refrigeration apparatus, the blower driving force is added to the performance evaluation items of the heat exchanger. The blowing drive force Pf [W] is defined by the following equation. Where ΔP
1 is a pressure difference across the air passage, which is a pressure loss. Pf = ΔP1 × Q where Q is the air flow rate passing through the heat exchanger [kg / s]
The length of the heat transfer tube in the longitudinal direction is W [m], and the number of stages is Dn.
Then, there is the following relationship with the front wind speed Uf [m / s] of the heat exchanger. Uf = Q / ρ / (W × Dp × Dn)

Hereinafter, the fin width L, the distance L ₁ from the leading edge of the fin to the leading edge of the heat transfer tube 2, the long axis da and the short axis d of the heat transfer tube
ΔP was calculated using b and the fin width L as parameters, and the air flow rate Q was determined under the condition that the blower driving force Pf was constant, and the heat exchange capacity E of the heat exchanger at this time was calculated. Note that the slit shape and arrangement are fixed. The heat exchange capacity is evaluated by the heat exchange amount E [W / K] per unit temperature difference, and is calculated by the following equation. E = Q × H × ε ε = 1-exp (-T) T = A o × K / (Q × H) K = 1 / (1 / αo + Ao / Ai / αi + Ac / Ai / αc) αo = 1 / ( Ao / (Ap + η × Af) Here, H [W / (kg · K)] is the specific heat of air, ε is the temperature efficiency, and K [W
/ (M ² K)] is the heat transfer coefficient, αc [W / (m ² K)] is the fin-
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 on the air side pipe of the heat exchanger, and Af [m ² ] is the heat transfer coefficient at the contact portion of the heat transfer tube. Air-side fin heat transfer area of heat exchanger, Ai
[m ² ] is the heat transfer area on the refrigerant side of the heat exchanger, Ac is the contact area between the fin and the heat transfer tube, a parameter depending on the shape of the heat exchanger, step pitch Dp, fin width L, fin leading edge. And the distance L ₁ from the heat transfer tube 2 to the front edge of the heat transfer tube 2, the fin pitch Fp, the fin thickness Ft, the short axis da, the long axis db of the heat transfer tube, and the fin-heat transfer tube contact heat transfer coefficient αc are calculated. is there.

Hereinafter, the relationship between the shape parameter and the heat exchange capacity E is shown in FIGS. In these figures, the heat exchange capacity E [W / K] is a value when the number of stages is one and the length W in the longitudinal direction of the heat transfer tube is a unit length. FIG. 9 shows the step pitch Dp, fin pitch Fp, fin thickness Ft, and fin width.
L, distance L ₁ from the leading edge of the fin to the leading edge of the heat transfer tube 2 _,
-The heat transfer tube contact heat transfer coefficient αc is constant within the range of the optimum value, the cross-sectional area of the heat transfer tube is constant, and the flatness H of the heat transfer tube 2 (= db)
It is the result of calculating the heat exchange capacity E using / da) as a parameter. FIG. 9 shows that the heat exchange performance increases as the flattening rate increases, that is, as the shape of the heat transfer tube increases. This is because, as the flattening rate increases, the ventilation area increases as a ratio and the ventilation resistance decreases, so that the air volume Q increases under the condition that the blower driving force Pf is constant, and as a result, the heat exchange performance E increases. If the range is 8 ≦ H, it is within 5% of the maximum value, and the heat exchange performance E is sufficiently large.

FIG. 10 shows the step pitch Dp and the fin pitch F
p, fin thickness Ft, the distance L ₁ of the fin width L, from the fin leading edge to the front edge of the heat transfer tube _2, the minor axis da of the heat transfer tube 2, the long axis d
b, Results of calculation of heat exchange capacity E using the step pitch Dp and the short axis da as parameters, with the fin-heat transfer tube contact heat transfer coefficient αc being constant within an optimum value range. FIG.
Therefore, the parameter (Dp-da) / Dp becomes maximum near 0.8, and 0.6 ≦ (Dp-da) /Dp≦0.95
Is within 5% of the maximum value, and the heat exchange performance E is sufficiently large. Note that 5% is selected as a range in which the target value can be reliably obtained. This is because when the parameter (Dp-da) / Dp is increased, (Dp-da)
In the range of da) / Dp <0.8, the ventilation resistance decreases proportionately, and the air volume Q increases under the condition that the blower driving force Pf is constant, so that the heat exchange performance E increases. In the range of 0.8 ≦ (Dp−da) / Dp, (Dp−
When da) / Dp is increased, the fin efficiency is reduced, and the extra-tube heat transfer coefficient αo is reduced, and as a result, the heat exchange performance E is reduced. The parameters described here (Dp
−da) / Dp indicates the pressure loss of the heat exchanger portion as an air passage. In other words, the parameter is in the range of 0.6 to 0.95 in the open region where the air passage loss is small. This indicates that the PQ characteristic is desirable, that is, the combination with a fan such as a propeller fan is good.

FIG. 11 shows a step pitch Dp and a fin pitch F.
p, the fin thickness Ft, the distance L ₁ from the leading edge of the fin to the leading edge of the heat transfer tube, the short axis da, the long axis db of the flat tube, and the contact heat transfer coefficient αc of the fin-heat transfer tube portion within the optimum values. This is a result of calculating the heat exchange capacity E using the fin width L as a parameter, while keeping the fin width constant. According to FIG. 11, the fin width L is larger than the long axis diameter db by L
= Maximum around 1.2db, db ≦ L ≦ 1.5db
Is within 5% of the maximum value, that is, a value at which the target value is reliably obtained, and the heat exchange performance E is sufficiently large. This is because when L is increased, db ≦ L ≦ 1.5d
In the range of b, the heat exchange capacity increases as the heat transfer area increases.
Although E is improved, when 1.5 db ≤ L, the decrease in air volume due to the increase in ventilation resistance becomes dominant to the heat exchange capacity E,
The heat exchange capacity E drops sharply.

FIG. 12 shows the step pitch Dp and the fin pitch F
p, the fin thickness Ft, the short axis da, the long axis db of the heat transfer tube, and the fin-heat transfer tube contact heat transfer coefficient αc within a range of the optimum value, and are constant from the fin front edge to the heat transfer tube 2 front edge. Distance L ₁
Is a result of calculating the heat exchange capacity E using as a parameter. From FIG. 10, it is understood that the smaller the distance L ₁ from the leading edge of the fin to the leading edge of the heat transfer tube 2, the greater the heat exchange performance E. This is because the larger the value of L1, the smaller the contact area Ac between the heat transfer tube and the fin collar, and the smaller the distance that the heat transfer tube enters the fin width, so that the fin efficiency ε decreases. However, within the range of 0 <L1 <L / 8, the maximum value is within 5%,
The heat exchange performance E is sufficiently large.

FIG. 13 shows a step pitch Dp and a fin pitch F.
p, fin thickness Ft, fin width L, distance L ₁ from the leading edge of the fin to the leading edge of the heat transfer tube 2, major axis da and short axis d of the heat transfer tube
This is a result of calculating the heat exchange capacity E using the fin-heat transfer tube contact heat transfer coefficient αc as a parameter, with b being constant within an optimum value range. FIG. 13 shows that the heat exchange performance E increases as the fin-heat transfer tube contact heat transfer coefficient αc increases. As the fin-heat transfer tube adhesion method, αc = 10000 [W / m ² K] in the expanded tube method, but αc becomes infinite by performing heat joining by brazing in a furnace. In this case, the heat exchange capacity E is improved by about 8%.

Next, the slit will be described. FIG. 14 shows the width and position of the slit with respect to the gas passage direction 40. The slit width is the same for all five slits.
Reference numeral 15 denotes a slit width e ₁ = e ₂ = e ₃ = e ₄ = e ₅ = 0.002 m.
The slit length is f1 = f2 = 0.0059m, f3 =
0.0064m, f4 = 0.0069m, f5 = 0.0074
m. When the position of the slit is indicated by the distance from the front edge of the fin, the distance between the front edge of the fin and the slit 11 is seen from the windward side.
w ₁ = 0.002 m, w ₂ = 0.
002 m, w ₃ = 0.002 m between the slits 12 and 13,
W ₄ = 0.002 m between 13 and 14, w between 14 and 15
₅ = 0.002 m. Here, the step pitch D
p, fin width L, the distance L ₁ from the fin leading edge to leading edge of the heat transfer tube _2, the minor axis da of the heat transfer tube is constant in the range of approximately optimum long axis db.

The air flowing between the fins 1 of the heat exchanger is heated or cooled by exchanging heat with the fins. FIG. 15 shows a state of the temperature boundary layer developed on the slit. On the surface of the fin, a temperature boundary layer develops as shown in FIG. 15A, and heat transfer is performed through the boundary layer 30. Generally, as the boundary layer 30 is thinner, the amount of heat transfer per unit temperature difference between the air and the fin is larger, and as shown in FIG. 15 (b), the temperature boundary layer is updated at the windward ends of the slits 33, 34, 35, The thickness of the temperature boundary layer at the upstream end of the slits 33, 34, 35 in the air flow direction becomes extremely thin. The fins are stacked at a pitch Fp in the stacking direction,
For example, the boundary layer 30 developed from the upstream end in the air flow direction of the slit 33 interferes with the boundary layer 30 developed from the slit 34 adjacent in the stacking direction at a downstream position. Downstream from the position where the interference occurs, the thickness of the boundary layer is constant,
The heat transfer amount per unit length in the flow direction is a constant value.
On the other hand, if the thickness of the thermal boundary layer and d _t, the thickness of the temperature boundary layer at a distance y [m] in the flow direction from the air flow direction upstream end of the slit d _t [m] is represented by the following formula. dt = 5.0 × (ν × y / U) ^0.5 / Pr ^0.3 where ν is the kinematic viscosity coefficient and ν in the case of air at normal temperature and normal pressure.
= 0.000016 [m ² / s]. U is the wind speed based on the above-mentioned free passage, and U = 1.31 [m / s]. Also,
Pr is the Prandtl number and in the case of air at normal temperature and normal pressure, Pr =
0.72. Now, let fin interval Hf be Hf = Fp-F
t, and the position of the slit in the fin stacking direction is H
When f / 2, the heat transfer between the slit surface and the air is promoted because the thickness _{dt of the} temperature boundary layer at the downstream of the slit, that is, at x = e, is smaller than の of the inter-fin width Hf. Also need to be small. Here, e [m] indicates the slit width. Therefore, the slit width e is e ≦ U / ν × Pr ^0.6 × (Hf / 10) ² = 510 × U × F
It is established to satisfy the condition of p ^2. In the standard use range of the air conditioning heat exchanger, the wind speed based on the free passage volume is U = 0.
Since it is 5-2 m / s, e ≦ 255 × Hf ² -1020 × Hf ² . However, it should be noted that the unit of e and Hf is [m] in the calculation. For example, Fp =
If 0.0012 m and Ft = 0.0001 m, Hf
= 0.0011 m and e ≤ 0.00031 to 0.0
0123 m. Even when the position of the slit in the stacking direction of the fins is other than Hf / 2, if the slit width e is set based on the above concept, substantially the same effect can be obtained. By the way, at this time, the heat transfer coefficient αs [W /
m ² K] is given as follows. That is, αs = K / e × 0.664 × Re ^0.5 × Pr ^0.3 where K is the thermal conductivity of air and Pr is the Prandtl number. In the case of normal temperature and normal pressure, respectively, K = 0.0261 [W
/ MK] and Pr = 0.72 [-]. Re is a Reynolds number and is defined as follows. Re = U × e / ν Therefore, by substituting αs = 3.914 × {U / e } 0.5 e ≦ 510 × U × Hf 2, whereas αs ≧ 0.173 / Hf, flat fins when the slit is not Heat transfer coefficient αb
[W / m ² K] can be calculated approximately as follows. αb = k / (Hf × 2) × 4.3 Therefore, αb = 0.056 / Hf Assuming that the number of fin-shaped slits having a fin width L along the air flow direction is N, the effective heat transfer coefficient αeff Is the area-weighted average of the two heat transfer coefficients described above. That is, αeff = αb + (N × e / L) × (αs−αb) = 0.05
6 / Hf × {1 + N × (1274 × U × Hf ² / L)} Therefore, αeff increases as N increases.

On the other hand, the relationship between the ventilation resistance and N for giving the optimum value of the number of slits N will be described. When the number N of slits along the flow direction is large, the amount of heat transfer increases due to the effect of updating the boundary layer described above, but the ventilation resistance of the heat exchanger increases, and the blowing driving force Pf increases. It is necessary to limit the number N. Now, the pressure loss (ventilation resistance) ΔPb per unit length between fins without slits is given as follows. ΔPb = 32 / Refp × (1 / Hf) × 1/2 × (γ
/ G) × U ² where Refp is defined as follows: Refp = U × (2 × Hf) / ν Further, γ is a specific weight [N / m ³ ] of air at normal temperature and normal pressure, and g is a gravitational acceleration [m / s ² ]. On the other hand, the pressure loss (ventilation resistance) ΔPs per unit length of the slit portion is generally ΔPs = 2 × 1.328 / Rea ^0.5 × (1 / Hf) × 1 /
2 × (γ / g) × U ² Therefore, the sum ΔP ^* of pressure loss (ventilation resistance) is ΔP ^* = ｛(L−N × e) × 32 / Refp × (1 / H
f) + N × e × 2.656 / Rea ^0.5 × (1 / Hf)｝ ×
1/2 × (γ / g) × U ² = L × ΔPb + N × e × (ΔP
s−ΔPb) Therefore, the ventilation resistance ΔP is proportional to the number N of the slits.
^* Means to increase. Then, the blower driving force Pf
, And the heat exchange capacity E ^* per unit length is calculated.

FIG. 16 shows the results of calculation of the heat exchange capacity E ^* per unit length in the longitudinal direction of the heat transfer tube, with the blower driving force Pf being kept constant, using the number N of slits as a parameter. According to FIG. 16, the heat exchange capacity E ^* has a maximum value near N = 5. This is because the ventilation resistance increases linearly, but the amount of heat transfer q asymptotically approaches a constant value. Therefore, when a slit of N = 5 or more is cut and raised, the increase in pressure loss (ventilation resistance) ΔP ^* increases the amount of heat transfer q This is because the air volume Q is lower than the increment.
Therefore, in this embodiment, it is desirable that the number of slits is five, but if the number of slits is in the range of three to eight, the heat exchange capacity E ^* is sufficiently large, and the effect is exhibited. Further, in order to avoid the influence of the slit upstream of the airflow, the slits are desirably arranged at substantially equal intervals so as to provide an interval before and after the slit.

Since the temperature difference between the air and the fins is large at the windward side, the amount of heat exchange due to the leading edge effect of the slits is large. Therefore, the slits need to be cut more than at the downstream side.
In particular, in the heat exchanger of the present embodiment, since the wind speed distribution in the fin is smaller than that of the tube heat exchanger, the locally high wind speed portion does not collide with the slit leading edge, and is not attached to the slit. The increment of ventilation resistance is smaller than the tube heat exchanger. Further, since the dead water area in the downstream part of the heat transfer tube 2 is smaller than that of the circular tube heat exchanger, it is not necessary to form many slits in the downstream part and perform rectification. Therefore, many slits are provided in the upstream part with respect to the center line of the half L / 2 of the fin width L. Further, in the heat exchanger of the present embodiment, since the speed distribution is small in the fin, the slit does not have to be wider than the downstream side for rectification in the upstream portion.

Embodiment 2 FIG. 17 shows Embodiment 2 and is a refrigerant circuit diagram of a refrigeration / air-conditioning apparatus. The refrigerant circuit shown in the figure includes a compressor 21, a condensing heat exchanger 22, a throttling device 23,
It comprises an evaporating heat exchanger 24, a blower 25, and a blower motor 26. In this configuration, the chlorine-free refrigerant is discharged from the compressor 21 at a high temperature and a high pressure, condensed and radiated by the condensing heat exchanger 22, expanded by the expansion device 23 to a low pressure, and evaporated by the evaporating heat exchanger 24. The heat is absorbed and sucked into the compressor. The heat exchanger having good heat exchange performance according to the first embodiment is replaced by the condensing heat exchanger 22 or the evaporating heat exchanger 2.
By using a refrigeration cycle using a refrigerant used for 4 or both, an air conditioner such as an air conditioner with high energy efficiency and a refrigeration system such as a refrigerator or a freezing warehouse can be realized. Here, the energy efficiency is, for example, configured by the following equation in an air conditioner. Heating energy efficiency = indoor heat exchanger (condenser) capacity / all inputs Cooling energy efficiency = indoor heat exchanger (evaporator) capacity / all inputs

Each of the heat exchangers is driven by a blower motor 26 which is efficiently driven by an inverter, and air is blown between the fins 1 of each heat exchanger. By selecting the flat heat transfer tube described in the first embodiment in an appropriate arrangement and shape as shown in FIGS. 9 to 13, the slits provided in the fins between the flat heat transfer tubes are further reduced to the temperature boundary. This device improves the heat transfer performance as well as the ventilation performance by arranging the layers so that they do not interfere with each other and has low ventilation resistance. It is possible to obtain a refrigeration / air-conditioning device having a higher energy saving effect.

Referring to FIGS. 18 and 19, the configuration of the indoor unit of the evaporating heat exchanger 24 and the blower 25 of the air conditioner when the air conditioner performs cooling will be described. 27 is an indoor unit, 28 is an air outlet, 29 is a guide vane, and 36 is a drain plate for collecting dew condensation water. As shown in FIG. 19, the heat exchanger that sucks indoor air from the surroundings is configured to have a shape in which a flat heat transfer tube 2 is inserted into a fin 1 from the ventilation direction. A chamber, which is a large number of grooves through which the refrigerant flows, is provided inside the flat space, and the flat shape has a straight outer shape. Such a heat exchanger is divided into five parts so as to surround the central space, and the air is sucked from the surroundings, heat is absorbed by the refrigerant circulating in the refrigerant circuit of FIG. 17 through the heat transfer tubes and the fins, and the air is cooled. are doing.

As shown in FIG. 18, the blower 25 provided in the axial direction from the outer periphery of the heat exchanger rotates to draw air into the center, and the air in the center space of the cooled heat exchanger 24 is cooled by the guide vanes 29. The direction is changed and the air is blown out from an air outlet provided below. In the figure, a description is given in which a blower is provided on one side. However, it goes without saying that this blower may be provided on both sides, or not only one on one side but also two or more. Furthermore, a propeller fan is used as a blower, and it is installed at a slant in the direction of air flow. There are no partitions or bends that provide a large ventilation resistance for the entire ventilation path of the unit, and the ventilation resistance is low as a PQ characteristic of the blower. In this case, the effect of further reducing energy can be obtained. Further, noise can be suppressed by separating the heat exchanger and the blower so that dead water generated downstream of the heat transfer tube does not reach the rotating fan of the blower. By using a propeller fan to blow a fan to a heat exchanger with a flat heat transfer tube, it can be used in a range where the pressure of the fan is low and the air volume is large, that is, the energy used is greatly increased because the open side is used. Can be reduced to

FIG. 20 is a diagram showing the arrangement of the condensing heat exchanger 22 and the blower 25 as in an outdoor unit of an air conditioner. The fin 1 transfers heat from the refrigerant flowing through the heat transfer tube 2 of the heat exchanger 22 to the fin 1. The heated air is blown out to the outside by the rotation of the fan of the blower 25. Note that, as in the first embodiment of the present invention, the heat transfer tube 2 is opened as shown in the figure on the inflow side of the fin 1 where the air is sucked. It is to be noted that two air blowers 25 for sucking air from the heat exchanger 22 and blowing the air to the outside are provided in parallel. In this drawing, only an example of the arrangement of the heat exchanger and the blower is shown, and other components provided in the outdoor unit of the air conditioner, and a box and a ventilation circuit showing the structure of the outdoor unit are not shown. With such an arrangement, it is possible to cope with a change in capacity only by increasing or decreasing the number of heat exchangers or blowers, and to obtain a device having a large energy saving effect with a simple configuration.

As shown in FIG. 19, the heat exchanger is arranged so as to surround the central space, and each heat exchanger is provided with eight flat heat transfer tubes, and each heat transfer tube has a groove serving as a refrigerant passage. The rooms are arranged. How the refrigerant is circulated to the plurality of heat exchangers is shown in the configuration diagram of FIG. Reference numeral 37 denotes a header. Refrigerant circulating in the refrigerant circuit first enters the header 37 of one heat exchanger, from which the refrigerant flows in parallel to eight flat heat transfer tubes and flows into the air through fins to warm or cool. Tell Other heat exchangers flow through the header and circulate. Headers are provided at both ends between heat exchangers. All grooves in each heat transfer tube may be connected in parallel without providing a partition inside the header, or a partition is provided inside the header and a groove in each heat transfer tube. May be divided into a plurality of blocks, and the blocks may be connected in series to flow the refrigerant. The flat heat transfer tube allows the refrigerant to pass through a number of grooves, and the header can be formed in a cylindrical shape. Therefore, even if the pressure of the refrigerant is high, the structure is highly reliable.
For example, a highly reliable apparatus can be obtained even for a natural refrigerant such as carbon dioxide or an alternative chlorofluorocarbon-free substitute such as R410A.

The heat exchangers described in the first and second embodiments and the air-conditioning and refrigeration systems using the same have the HCFC (R22) and the HFC (R116, HFC).
R125, R134a, R14, R143a, R152
a, R227ea, R23, R236ea, R236f
a, R245ca, R245fa, R32, R41, R
C318 and several kinds of mixed refrigerants R407 of these refrigerants.
A, R407B, R407C, R407D, R407
E, R410A, R410B, R404A, R507
A, R508A, R508B, etc.), HC (butane, isobutane, ethane, propane, propylene, etc., or a mixture of several of these refrigerants), natural refrigerants (air, carbon dioxide,
The effect can be achieved by using any kind of refrigerant, such as ammonia or a mixed refrigerant of several kinds of these refrigerants) or a mixed refrigerant of several kinds of these refrigerants.

Although the working fluids of the primary and secondary refrigerants are air and refrigerant, the same effect can be obtained by using other gas, liquid, or gas-liquid mixed fluid.

Although the heat transfer tube and the fin are often made of different materials, the same material such as copper is used for the heat transfer tube and the fin, and aluminum is used for the heat transfer tube and the fin. Can be attached, the contact heat transfer coefficient between the fin portion and the heat transfer tube is dramatically improved, and the heat exchange capacity is greatly improved. In addition, since no dissimilar metal is used, troubles such as current flow can be prevented, and the recyclability at the time of disposing of the device is improved, which often contributes to environmental protection.

As a method of bringing the heat transfer tube and the fins into close contact with each other, when brazing in a furnace is performed, post-processing is performed to apply the hydrophilic material to the fins. It can prevent burn-off.

In order to increase the pressure resistance, measures such as increasing the wall thickness or increasing the number of partitions 3 inside the heat transfer tube may be taken. Increasing the thickness also increases the outer diameter of the elliptical heat transfer tube and increases the cost of the heat transfer tube.However, by adjusting the step pitch, row pitch, flattening ratio, number and shape of cut-and-raised fins, ventilation resistance and transfer If these values are appropriately set in consideration of the balance of heat promotion, the effect of the present embodiment can be sufficiently exerted.

The heat exchangers described in the first and second embodiments and the air-conditioning and refrigeration systems using the heat exchangers include mineral oils, alkylbenzene oils, ester oils, ether oils, and fluorine oils. Regardless of the refrigeration oil, regardless of whether the refrigerant and oil are soluble,
That effect can be achieved.

[0067]

The heat exchanger according to claim 1 of the present invention has the following features.
A large number of parallel arranged, plate-like fins through which the primary refrigerant flows, and a plurality of chambers inserted into each of the plate-like fins, through which the secondary refrigerant flows, are formed in a flat shape and flow into the plate-like fins A plurality of heat transfer tubes arranged so that the primary refrigerant flows along the long axis direction of the flat plate, and the flat shape of the heat transfer tube is such that the length of the flat short-axis direction decreases as the flow of the primary refrigerant goes downwind. Since the size is reduced, a heat exchanger having a high heat exchange capacity and a small ventilation resistance can be obtained.

The heat exchanger according to the second aspect of the present invention is arranged in parallel with a plurality of plate-like fins through which the primary refrigerant flows, and inserted into each of the plate-like fins to flow the secondary refrigerant therein. Flat-shaped heat transfer tube and fin fin formed on the leeward side of the primary refrigerant from the end of the plate-like fin so that the heat transfer tube can be inserted from the end on the windward side where the primary refrigerant through which the plate-like fin flows Since the heat transfer tube is always open on the windward side from the plate-like fin, a heat exchanger having good performance and easy mass production can be obtained.

According to a third aspect of the present invention, in the heat exchanger, the front edge of the heat transfer tube comes into contact with the plate fin up to the vicinity of the windward end of the plate fin and projects from the plate fin at this end. The heat exchanger having a high heat exchange capacity can be obtained.

In the heat exchanger according to a fourth aspect of the present invention, the heat transfer tube is inserted so as to be covered by the plate fin with an interval with respect to the most downstream end of the primary refrigerant of the plate fin. As a result, a highly reliable heat exchanger against condensation can be obtained.

In the heat exchanger according to a fifth aspect of the present invention, since the outer shape of the heat transfer tube has a curved surface as a whole and the curvature increases as going in the leeward direction, the heat exchange tube has good heat exchange and heat exchange performance. A vessel is obtained.

In the heat exchanger according to the sixth aspect of the present invention, since the outer tube is provided with the straight portion, a product having good contact between the heat transfer tube and the fin can be obtained. .

In the heat exchanger according to the seventh aspect of the present invention, since the respective cross-sectional areas of the plurality of chambers in the heat transfer tube are substantially the same, the flow of the refrigerant in the tube is substantially the same, and the heat exchange performance is improved. . .

In the heat exchanger according to claim 8 of the present invention, the flattened short axis of the heat transfer tube is da and the long axis is db,
Since H (= db / da) is 8 ≦ H, a high-performance heat exchanger can be obtained.

According to a ninth aspect of the present invention, when the width of the plate-like fin in the direction in which the primary refrigerant flows is L and the long axis of the heat transfer tube is db, 1 ≦ L / db ≦ 1.5. Because there is
A small and high-performance heat exchanger can be obtained.

A heat exchanger according to a tenth aspect of the present invention comprises:
Assuming that a plurality of heat transfer tubes are arranged, and a stepwise pitch perpendicular to the flow of the primary refrigerant in this arrangement direction is Dp, and a flat short axis of the heat transfer tubes is da, 0.6 ≦ (Dp−da) / Dp ≤
Since it is 0.95, when the fan driving force is constant,
A high-performance heat exchanger having a large heat exchange amount can be obtained.

[0077] The heat exchanger according to claim 11 of the present invention comprises:
The distance between the windward end of the plate fin and the front edge of the heat transfer tube is L
Assuming that 1, L1 <da / 8, a heat exchanger having high heat exchange performance can be obtained.

A heat exchanger according to a twelfth aspect of the present invention comprises:
A slit provided on the surface of the plate-like fin and having an opening facing the flow of gas, and having a half length L of the fin width L.
Since the number of slits provided in the upstream part is larger than the number of slits provided in the downstream part with respect to the center line of / 2, a heat exchanger having high heat exchange capacity and small ventilation resistance can be obtained.

A heat exchanger according to a thirteenth aspect of the present invention comprises:
Since the intersection of the leg of the slit provided on the plate-like fin surface and the plate-like fin surface is almost parallel to the local gas flow direction near the fin surface, the heat exchange capacity is high. Thus, a heat exchanger having a low ventilation resistance can be obtained.

A heat exchanger according to a fourteenth aspect of the present invention comprises:
Since the plate-shaped fins and the heat transfer tubes are heated and fixed, a heat exchanger having a high heat exchange capacity can be obtained.

A heat exchanger according to a fifteenth aspect of the present invention comprises:
Since the same material is used for the heat transfer tube and the plate-like fins, an apparatus having high reliability and excellent environmental measures can be obtained.

A refrigeration / air-conditioning apparatus according to a sixteenth aspect of the present invention includes a refrigeration cycle for circulating a secondary refrigerant through a compressor, a condenser heat exchanger, a throttle device, and an evaporator heat exchanger. Since a heat exchanger having good heat exchange performance is used for at least one of the exchanger and the evaporator heat exchanger, a refrigeration / air-conditioning apparatus with high energy efficiency can be obtained.

A refrigeration / air-conditioning apparatus according to a seventeenth aspect of the present invention includes a blower for flowing a primary refrigerant between fins of a condenser heat exchanger or an evaporator heat exchanger. Since the propeller fan is used for the blower that blows air to the heat exchanger having the above, an apparatus having good performance and high energy saving can be obtained.

The method for manufacturing a heat exchanger according to the eighteenth aspect of the present invention is directed to a method of manufacturing a heat exchanger, comprising a plurality of plate-shaped fins which are arranged in parallel and through which a primary refrigerant flows, and which are fixed to the respective plate-shaped fins and have a secondary refrigerant therein. A plurality of heat transfer tubes that are formed in a flat shape and through which the primary refrigerant flowing through the plate-shaped fins flows along the flat long axis direction.
Forming a hole in the plate-shaped fin to open one end in the direction in which the primary refrigerant flows and closing the other end; and inserting a heat transfer tube into the hole in the plate-shaped fin from one end, which is always the open side. And a step of heating and fixing the fins and the heat transfer tubes after assembling the plate-shaped fins and the heat transfer tubes, so that a highly reliable and high-performance heat exchanger can be easily mass-produced.

In the method for manufacturing a heat exchanger according to the nineteenth aspect of the present invention, the plate-like fin and the heat transfer tube are heated and fixed, and then the plate-like fin is coated with a hydrophilic material. High heat exchanger can be manufactured.

[Brief description of the drawings]

FIG. 1 is an external view illustrating a configuration of a heat exchanger according to a first embodiment of the present invention.

FIG. 2 is a cross-sectional plan view illustrating the configuration of the heat exchanger according to the first embodiment of the present invention.

FIG. 3 is a partial front view showing the configuration of the heat exchanger according to the first embodiment of the present invention.

FIG. 4 is a partial external view showing a fin collar and a fin expanding portion according to the first embodiment of the present invention.

FIG. 5 is an explanatory view showing a method of assembling the heat exchanger according to the first embodiment of the present invention.

FIG. 6 is a flowchart showing a procedure for assembling the heat exchanger according to the first embodiment of the present invention.

FIG. 7 is a comparative diagram showing the occurrence of dead water areas 10 of the circular tube, the flat tube, and the heat transfer tube according to the first embodiment.

FIG. 8 is a comparison diagram showing how dew flows during cooling of the flat tube and the heat transfer tube of the first embodiment.

FIG. 9 is a characteristic diagram showing a relationship between a flattening rate and a heat exchange capacity of the heat transfer tube according to the first embodiment of the present invention.

FIG. 10 is a characteristic diagram showing the relationship between the step pitch, the short axis diameter of the heat transfer tube, and the heat exchange capacity according to the first embodiment of the present invention.

FIG. 11 is a characteristic diagram showing a relationship between a fin width and heat exchange capacity according to the first embodiment of the present invention.

FIG. 12 is a characteristic diagram showing the relationship between the distance between the leading edge of the fin and the leading edge of the heat transfer tube and the heat exchange capacity according to the first embodiment of the present invention.

FIG. 13 is a characteristic diagram showing a relationship between the contact heat transfer coefficient between the fins and the heat transfer tubes and the heat exchange capacity according to the first embodiment of the present invention.

FIG. 14 is a plan sectional view showing details of a slit shape and a shape according to the first embodiment of the present invention.

FIG. 15 is an explanatory diagram showing a relationship between a plate-like fin and a development state of a temperature boundary layer on a slit.

FIG. 16 is a characteristic diagram showing a relationship between the number of slits and heat exchange capacity according to the first embodiment of the present invention.

FIG. 17 is a refrigerant circuit diagram illustrating a configuration of a refrigeration / air-conditioning apparatus according to Embodiment 2 of the present invention.

FIG. 18 is an explanatory view showing a heat exchanger and a blower of a refrigeration / air-conditioning apparatus according to Embodiment 2 of the present invention.

FIG. 19 is an explanatory diagram showing a configuration of a heat exchanger of a refrigeration / air-conditioning apparatus according to Embodiment 2 of the present invention.

FIG. 20 is an explanatory diagram showing a heat exchanger and a blower of the refrigeration / air-conditioning apparatus according to Embodiment 2 of the present invention.

FIG. 21 is an explanatory diagram showing a configuration of a heat exchanger of a refrigeration / air-conditioning apparatus according to Embodiment 2 of the present invention.

FIG. 22 is a plan sectional view and a partial side view showing the configuration of a conventional circular tube heat exchanger.

FIG. 23 is an external view showing a configuration of a conventional flat tube heat exchanger.

FIG. 24 is an explanatory view showing a method of assembling a conventional flat tube heat exchanger.

FIG. 25 is an explanatory view showing a phenomenon in assembling a conventional flat tube heat exchanger.

[Explanation of symbols]

1 plate-like fin, 2 heat transfer tube, 3 inner wall, 4
Fin collar, 5 fin expansion section, 6 insertion hole,
10 dead water area, 11-15 slits, 16 dew,
Reference Signs List 21 compressor, 22 condenser heat exchanger, 23 expansion device, 24 evaporator heat exchanger, 25 blower, 2
6 motor for blower, 30 temperature boundary layer, 31 to 3
3 slits, da short axis diameter of heat transfer tube, db long axis diameter of heat transfer tube, Dp step pitch, e1 to e5 slit width, f1 to f5 slit length, L plate fin 1 fin width, L1 plate fin 1 Distance between the front edge of the heat transfer tube 2 and the front edge of the heat transfer tube 2, w1 to w5 distance between slits.

 ──────────────────────────────────────────────────の Continued on the front page (72) Kunihiko Kaga, 2-3-2 Marunouchi, Chiyoda-ku, Tokyo Mitsui Electric Co., Ltd. (72) Kenichi Yamada 2-3-2, Marunouchi 3-chome, Chiyoda-ku, Tokyo F term (reference) 3L103 AA01 AA37 BB38 BB42 CC22 CC28 DD08 DD33

Claims (19)

[Claims] 1. A plurality of plate-like fins arranged in parallel, between which a primary refrigerant flows, and inserted into each of the plate-like fins,
A plurality of chambers in which the secondary refrigerant flows, and a plurality of heat transfer tubes which are formed in a flat shape and are arranged so as to flow the primary refrigerant flowing in the plate-like fins along the long axis direction of the flat. The heat exchanger according to claim 1, wherein the flat shape of the heat transfer tube reduces the length of the flat refrigerant in the short axis direction as the flow of the primary refrigerant goes downwind. 2. A plurality of plate-like fins arranged in parallel, through which a primary refrigerant flows, and inserted into each of the plate-like fins,
A heat transfer tube formed into a flat shape with a secondary refrigerant flowing therein, and the heat transfer tube is inserted from the end of the plate-like fin so as to be inserted from the windward end of the plate-like fin through which the primary refrigerant flows. A fin hole formed on the lee side of the primary refrigerant, wherein the heat transfer tube is always open on the windward side from the plate-like fin. 3. The heat transfer tube according to claim 1, wherein a front edge of the heat transfer tube contacts the plate-shaped fin up to an end near the windward side of the plate-shaped fin, and projects from the plate-shaped fin at this end to be opened. The heat exchanger according to claim 1. 4. The heat transfer tube according to claim 1, wherein the heat transfer tube is inserted so as to be covered with the plate-like fin at an interval with respect to the most downstream end of the primary refrigerant of the plate-like fin. The heat exchanger according to any one of the above. 5. The heat exchanger according to claim 1, wherein the outer shape of the heat transfer tube has a curved surface as a whole and a curvature increases toward the leeward direction. 6. The heat exchanger according to claim 1, wherein a straight portion is provided in an outer tube shape. 7. The heat exchanger according to claim 1, wherein each of the plurality of chambers in the heat transfer tube has substantially the same sectional area. 8. The flat short axis of the heat transfer tube is da and the long axis is db.
The heat exchanger according to any one of claims 1 to 7, wherein the flattening factor H (= db / da) satisfies 8 ≦ H. 9. The apparatus according to claim 1, wherein 1 ≦ L / db ≦ 1.5, where L is the width of the plate-like fin in the direction in which the primary refrigerant flows, and db is the long axis of the heat transfer tube. A heat exchanger as described in Crab. 10. A plurality of heat transfer tubes are arranged, and the stepwise pitch in a direction perpendicular to the direction of the primary refrigerant in the arrangement direction is D.
The heat exchanger according to any one of claims 1 to 9, wherein, where p is a flat short axis of the heat transfer tube and da is 0.6 ≦ (Dp−da) /Dp≦0.95. 11. The apparatus according to claim 1, wherein L1 <da / 8, where L1 is a distance between the windward end of the plate-like fin and the front edge of the heat transfer tube. Heat exchanger. 12. A slit provided on the surface of the plate-like fin and having an opening facing the flow of gas, provided at a downstream portion with respect to a center line of a half L / 2 of the fin width L. The heat exchanger according to any one of claims 1 to 11, wherein the number of slits provided in the upstream portion is larger than the number of slits provided. 13. An intersection line between a leg portion of a slit provided on a surface of the plate-like fin and the plate-like fin surface is substantially parallel to a local gas flow direction near the fin surface. The heat exchanger according to claim 1, wherein: 14. The heat exchanger according to claim 1, wherein the plate-like fin and the heat transfer tube are heated and fixed. 15. The heat exchanger according to claim 1, wherein the same material is used for the heat transfer tube and the plate-like fin. 16. The secondary refrigerant is supplied to a compressor, a condenser heat exchanger,
And a refrigerating cycle for circulating an evaporator heat exchanger. The heat exchanger according to any one of claims 1 to 15, wherein at least one of the condenser heat exchanger and the evaporator heat exchanger is used. A refrigeration / air-conditioning system characterized by 17. The method according to claim 17, wherein the primary refrigerant is a condenser heat exchanger or
A blower flowing between the fins of the evaporator heat exchanger,
17. The refrigeration / air-conditioning apparatus according to claim 16, wherein a propeller fan is used as a blower for blowing air to the heat exchanger having the flat heat transfer tubes. 18. A plurality of plate-shaped fins which are arranged in parallel and through which a primary refrigerant flows, and which are fixed to the respective plate-shaped fins, are formed in a flat shape in which a secondary refrigerant flows therein, and are formed on the plate-shaped fins. A plurality of heat transfer tubes arranged so that the flowing primary refrigerant flows along the flat major axis direction, wherein one end of the primary refrigerant in the direction in which the primary refrigerant flows is opened to the plate-like fins and the other end side is opened. Forming a hole to close the plate, assembling the heat transfer tube by inserting the heat transfer tube into the hole of the plate fin from one end which is always open, and assembling the plate fin and the heat transfer tube. Heating the fins and the heat transfer tube to secure the heat transfer tubes after the heat transfer. 19. The method for manufacturing a heat exchanger according to claim 18, wherein the plate-like fin and the heat transfer tube are heated and fixed, and then the plate-like fin is coated with a hydrophilic material.