JP6575895B2 - Heat exchanger - Google Patents

Description

  The present invention relates to a heat exchanger used in an air conditioning refrigeration apparatus.

  Conventionally, in refrigeration equipment for showcases installed in stores such as supermarkets and convenience stores, the refrigerant compressed by the compressor of the refrigerator installed outside the store is radiated with a gas cooler, and then sent to the expansion valve. The expansion valve is depressurized and expanded. And it evaporates with the evaporator provided in the showcase, The inside of a showcase is cooled by evaporation of the refrigerant | coolant at this time.

  Also, due to problems in the natural environment, a refrigeration system for showcases has been developed that uses carbon dioxide, which is a natural refrigerant, as an alternative to an HFC (Hydrofluorocarbon) refrigerant (see, for example, Patent Document 1). ).

  In recent years, in conjunction with the adoption of carbon dioxide refrigerant in refrigeration equipment, higher pressure is required for refrigeration equipment, and heat transfer tubes used in evaporators are also required to be thinner and have higher strength. .

  Since the carbon dioxide refrigerant has a higher gas density and lower viscosity than the HFC refrigerant, the pressure loss due to the pipe friction is small even in a low dryness state. Therefore, the diameter of the pipe for sending gas can be reduced.

  However, since the refrigeration efficiency of the evaporator used in the refrigerator decreases as the dryness of the refrigerant increases, the amount of refrigerant circulation increases as the dryness increases. Therefore, the tube friction resistance of the heat transfer tube constituting the evaporator is increased, and the pressure loss in the evaporator is increased.

  For this reason, an evaporator having a plurality of refrigerant flow paths that flow through the evaporator has also been proposed. In this evaporator, the heat transfer tubes are divided at the inlet portion of the evaporator, and the heat transfer tubes are joined at the outlet portion of the evaporator.

JP 2013-245857 A

  In a heat exchanger used for an evaporator of a refrigeration apparatus, frost resistance is important. When the fin pitch (interval of the plate-like fins) of the heat exchanger is small, the gap between the plate-like fins is likely to be closed when frosting occurs in the evaporator, and it is necessary to frequently perform the defrosting process. In order to prevent deterioration in cooling efficiency due to frost formation, it is necessary to increase the fin pitch to some extent regardless of the diameter of the heat transfer tube.

  On the other hand, in the plate-like fin, a collar portion is provided around a through hole through which the heat transfer tube passes. However, if the heat transfer tube is reduced in diameter, it is difficult to increase the height of the collar portion. When the collar height is smaller than the desired fin pitch, a part of the heat transfer tube inserted into the through hole is exposed. If the use of the evaporator is continued in this state, the heat transfer tube is corroded by the drain water, so it is necessary to coat the heat transfer tube with a corrosion-resistant agent.

  The present invention has been made in order to solve the conventional technical problems, and an object thereof is to improve the cooling efficiency and the corrosion resistance in the evaporator.

In order to solve the above-mentioned object, the heat exchanger of the present invention is provided with a plurality of through holes in the row direction parallel to the gas flow direction and in the step direction orthogonal to the gas flow direction, and around each through hole. A plurality of plate-like fins provided with a cylindrical collar portion, and a heat transfer tube that passes through the through-hole and contacts the collar portion and through which the refrigerant passes, and the outer diameter D of the heat transfer tube is 7 mm or more and 10 mm. The fin pitch Fp of the plurality of plate-like fins is 8 mm or more and 12 mm or less, and the plurality of plate-like fins are arranged in parallel so that the collar portion of each plate-like fin protrudes in one direction. in the plate-like fins, collar portion of the end portion of one of the plate fin plate adjacent fins are both in contact with the surface opposite to the surface on which the collar portion of the other plate-like fins protrude The row pitch Lp of the heat transfer tubes in the row direction is 65 mm or less And at 75mm or less, step pitch Dp of the heat transfer tubes in the column direction is characterized in that at 25mm or less than 19 mm.

  According to the present invention, the cooling portion efficiency is improved by increasing the height of the collar portion so that the outer diameter D of the heat transfer tube is 7 mm or more and 10 mm or less and the fin pitch is 8 mm or more and 12 mm or less. be able to.

  In addition, the end of the collar portion of one plate fin of adjacent plate fins is in contact with the surface opposite to the surface from which the collar portion of the other plate fin protrudes. Can improve frost closure. Moreover, by setting it as such a structure, a part of heat exchanger tube inserted in the through-hole of a plate-shaped fin will not be exposed, and the corrosion resistance of an evaporator will improve.

The figure which shows an example of a structure of the heat exchanger which concerns on embodiment of this invention The figure which shows schematic structure of the plate-shaped fin shown in FIG. The figure which shows arrangement | positioning of the several plate-shaped fin in a heat exchanger Conceptual diagram of gas-liquid two-phase flow in heat transfer tubes The figure which shows the calculation result of the performance of the freezing apparatus in case of Lp = 72.8mm and Dp = 21mm The figure which shows the relationship between the pipe diameter and fin pitch in the case of Lp = 72.8mm and Dp = 21mm The figure which shows the calculation result of the performance of the freezing apparatus in case of Lp = 65mm and Dp = 19mm The figure which shows the relationship between the pipe diameter and fin pitch in case of Lp = 65mm and Dp = 19mm The figure which shows the calculation result of the performance of the freezing apparatus in case of Lp = 75mm and Dp = 25mm The figure which shows the relationship between the pipe diameter and fin pitch in case of Lp = 75mm and Dp = 25mm

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. Each embodiment described below is an example, and the present invention is not limited to this embodiment.

  FIG. 1 is a diagram illustrating an example of a configuration of a heat exchanger 1 according to an embodiment of the present invention. FIG. 2 is a view showing a schematic configuration of the plate-like fin 2 shown in FIG. FIG. 3 is a view showing the arrangement of the plurality of plate-like fins 2 in the heat exchanger 1.

  As shown in FIG. 1, the heat exchanger 1 includes a plate-like fin 2, a heat transfer tube 3, and a tube plate 4. A plurality of plate-like fins 2 are provided, and each plate-like fin 2 has a row direction parallel to the gas flow direction (the direction indicated by the solid arrow in FIG. 1 and the dotted arrow in FIG. 2), and gas. A plurality of through-holes 2a are provided in a step direction orthogonal to the flow direction, and a cylindrical collar portion 2b is provided around each through-hole 2a.

  In FIG. 1, the plate-like fins 2 are omitted between the plate-like fins 2 adjacent to the central tube plate 4 and the plate-like fins 2 adjacent to the tube plates 4 on both sides so that the heat transfer tube 3 can be seen. In reality, however, a large number of plate-like fins 2 are provided adjacent to each other.

  The plurality of plate-like fins 2 are arranged in parallel so that the collar portion 2 b of each plate-like fin 2 protrudes in one direction, and one of the plate-like fins 2 adjacent to each other is shaped like a plate. The end of the collar portion 2b of the fin 2 is in contact with the surface opposite to the surface from which the collar portion 2b of the other plate-like fin 2 protrudes (see FIG. 3).

  With such a configuration, since the heat transfer tube 3 inserted into the through hole 2a of the plate-like fin 2 is not exposed, it is possible to prevent the heat transfer tube 3 from being corroded by drain water. There is no need to coat the agent.

  The heat transfer tube 3 is a tube through which the refrigerant passes. The heat transfer tube 3 penetrates through the through holes 2 a provided in each plate-like fin 2 and comes into contact with the collar portion 2 b, and is formed by the tube plate 4 at the center and both sides of the heat exchanger 1. Retained.

  As the refrigerant, for example, a carbon dioxide refrigerant is used. Since the carbon dioxide refrigerant has a higher gas density and lower viscosity than the HFC refrigerant, there is little pressure loss due to pipe friction even in a low dryness state. Therefore, the diameter of the pipe for sending gas can be reduced.

  In addition, for example, a high-strength copper pipe is used for the heat transfer pipe 3. As a result, the diameter of the heat transfer tube 3 can be further reduced, and the inner diameter of the heat transfer tube 3 can be increased by reducing the thickness. As a result, the degree of dryness of carbon dioxide used as the refrigerant increases, and the pressure loss in the evaporator can be suppressed even when the amount of refrigerant circulation increases accordingly.

  In the heat exchanger 1 in the present embodiment, the fin pitch Fp (see FIG. 3) in the stacking direction of the plate-like fins 2 is 10 mm, the fin thickness Ft (see FIG. 3) is 0.25 mm, and the gas in the heat exchanger 1 The row pitch Lp (see FIG. 2), which is the distance between the centers of the adjacent heat transfer tubes 3 in the direction along the passage direction, is 72.8 mm, and is adjacent in the direction perpendicular to the gas passage direction of the heat exchanger 1. The step pitch Dp (see FIG. 2) which is the distance between the centers of the heat transfer tubes 3 is 21 mm, and the outer diameter D of the heat transfer tubes 3 after expansion is 9.9 mm.

  Further, the heat transfer tube 3 is branched into a first system and a second system at the inlet of the heat exchanger 1 and merges at the outlet of the heat exchanger 1. Further, as shown in FIG. 1, the first system heat transfer tube 3 forms a meandering pipe flow path in two rows on the front side of the plate-like fin 2, and the second system heat transfer tube 3 A pipe flow path is formed in a meandering manner in two rows on the rear side of the plate-like fins 2. And the heat exchanger tube 3 of a 1st system and a 2nd system is arrange | positioned in zigzag form seeing from the gas passage direction.

  The qualitative tendency of the shape parameters described above will be described below with respect to the heat transfer performance and the ventilation resistance of the heat exchanger 1.

  In order to improve the frost resistance of the heat exchanger 1 (evaporator), it is considered effective to increase the fin pitch Fp. When the fin pitch Fp is increased, the ventilation resistance decreases. Therefore, the air volume can be increased, but the number of the plate-like fins 2 that can be accommodated is reduced, so that the heat transfer area is reduced. On the other hand, when the fin pitch Fp is reduced, the heat transfer area increases, but the ventilation resistance increases and the air volume cannot be increased.

  Similarly, when the row pitch Lp and the step pitch Dp are increased, the heat transfer coefficient on the fin surface is improved, but the fin efficiency defined by the relationship between the distance from the outer periphery of the heat transfer tube 3 to the fin end and the heat transfer Will decline. Further, since the ventilation resistance is reduced, the air volume can be increased.

  In addition, when the fin thickness Ft is increased, the fin efficiency is improved, but the ventilation resistance is increased. On the other hand, when the fin thickness Ft is reduced, the fin efficiency is reduced, but the ventilation resistance is reduced.

  As described above, each of the above-described shape parameters has an optimum value, and in order to quantitatively evaluate this, the heat transfer performance and the ventilation resistance of the heat exchanger 1 are calculated by the method described below.

In general, the heat transfer coefficient α [w / m ² · k] between air and the plate fin is defined by the following equation.
α = Nu × λ / De
Re = U × De / ν
Nu = 0.664 × Re ^1/2 × Pr ^1/3 (when Re <3.2 × 10 ⁵ )
Nu = 0.037 × Re ^0.8 × Pr ^1/3 (when Re> 3.2 × 10 ⁵ )

Here, Re is the Reynolds number, Nu is the Nusselt number, and these values are obtained by an approximate expression. Pr is the Prandtl number, λ is the thermal conductivity of air, ν is the kinematic viscosity coefficient of air, and Pr = 0.72, λ = 0.0261 [w / m · k], ν at room temperature and normal pressure, respectively. = 0.000016 [m ² / s].

Further, the representative length De [m] is defined by the following equation.
De = 4 × (Lp × Dp -π × D 2/4) × (Fp-Ft) / {2 × (Lp × Dp-π × D 2/4) + π × D × (Fp-Ft)}

The relationship between the free passage volume based wind speed U [m / s] between the plate fins and the front wind speed Uf [m / s] of the heat exchanger is defined by the following equation.
U = Uf × Lp × Dp × Fp / {(Lp × Dp-π × D 2/4) × (Fp-Ft)}

Further, the fin efficiency η is defined by the following equation.
η = 1 / (1 + ψ × α)
ψ = {(4 × Lp × Dp / π) ^0.5 −D} ² × (4 × Lp × Dp / π) ^0.5 / D ^0.5 / 6 / Ft / λf
Here, λf [w / m · k] is the thermal conductivity of the plate fin.

On the other hand, the ventilation resistance ΔP [Pa] between the air and the plate fin is defined by the following equation.
ΔP = 2 × F × Lp × Ln × ρ × U ² / De
F = 14.227 / Re
Here, F is a friction loss coefficient. Further, ρ is the density of air, and is about 1.2 [kg / m ³ ] in the case of normal temperature and pressure.

  Moreover, when using the heat exchanger 1 in this embodiment for an air-conditioning refrigerating apparatus, it becomes important to reduce the driving force of a fan. Therefore, the blower driving force is further considered here.

The blower driving force Pf [w] is defined by the following equation.
Pf = ΔP × Q
Here, Q is an air flow rate [kg / s] passing through the heat exchanger.

Further, assuming that the length in the longitudinal direction of the heat transfer tube is W [m] and the number of stages is Dn, the following relationship exists between these and the front wind speed Uf [m / s] of the heat exchanger.
Uf = Q / ρ / (W × Dp × Dn)

  The air flow rate Q described above is determined under the condition that ΔP is calculated using the step pitch Dp, the row pitch Lp, the fin pitch Fp, the fin thickness Ft, and the outer diameter D of the heat transfer tube as parameters, and the fan driving force Pf is constant. Can do.

In this case, the heat exchange amount E [w / k] per unit temperature difference of the heat exchanger 1 is calculated by the following equation.
E = Q × H × ε
ε = 1−exp (−T)
T = Ao × K / (Q × H)
K = 1 / (1 / αo + Ao / Ai / αi)
αo = 1 / (Ao / (Ap + η × Af) / α)
Ao = Ap + Af

Here, H [w / kg · k] is the specific heat of the air, ε is the temperature efficiency, K [w / m ² · k] is the heat passage rate, and Ao [m ² ] is the total heat transfer on the air side of the heat exchanger 1. Area, Ap [m ² ] is the heat transfer area of the air side heat transfer tube of the heat exchanger 1, Af [m ² ] is the heat transfer area of the air side fin of the heat exchanger 1, and Ai [m ² ] is the heat transfer area of the heat exchanger 1. It is a refrigerant side heat transfer area. These areas are values that can be calculated if the dimensions depending on the shape of the heat exchanger 1, the step pitch Dp, the row pitch Lp, the fin pitch Fp, the fin thickness Ft, and the outer diameter D of the heat transfer tube 3 are determined. .

Also, a paper by Lixin Cheng et al., “New flow boiling heat transfer model and flow pattern and carbon diverted in 70 minutes, which is described in“ New flow boiling heat transfer and carbon 40 ”. The heat transfer coefficient αi [w / m ² · k] of the fluid flowing in the pipe of the vessel can be obtained by the following formula according to the state of the refrigerant flowing in the pipe.

αi = {θ _dry × α _v + (2π−θ _dry ) × α _wet }
Here, as shown in FIG. 4, θ _dry is an angle of a region where the liquid refrigerant does not exist around the entire inner wall of the tube. Α _v [w / m ² · k] is the heat transfer coefficient of the gas refrigerant, and α _wet [w / m ² · k] is the heat transfer coefficient of the liquid refrigerant.

Furthermore, the heat transfer rate α _v [w / m ² · k] of the gas refrigerant, the heat transfer rate α _wet [w / m ² · k] of the liquid refrigerant, the nucleate boiling transfer rate α _nb [w / m ² · k], forced boiling heat transfer coefficient α _cb [w / m ² · k] is given by the following equation.

α _v = 0.023 × Re _v ^0.8 × Pr _v ^0.4 × (k _v / D)
α _wet = {(α _nb ) ³ + (α _cb ) ³ } ^1/3
α _nb = 131 × Pr ^−0.0063 × ( ^−log ₁₀ Pr) ^−0.55 × M × q ^−0.58
α _cb = 0.0133 × {4G (1-x) δ / μ _L (1-β)} ^0.69 × Pr _L ^0.4 × (k _L / δ)

Here, Re _v is the Reynolds number of the gas refrigerant, Pr _v is Prandtl number of the gas _{refrigerant, k v [w / m ·} k] is the thermal conductivity of the gas refrigerant. G [kg / m ² · s] is the speed of the refrigerant two-phase flow, M [kg / kmol] is the molecular weight, q [W / m ² ] is the heat flux, and δ [m] is the liquid refrigerant liquid in the pipe. film thickness (see Figure 4), x is the dryness _{(vapor quality), μ L [} N · s / m 2] is the viscosity coefficient of the liquid refrigerant, beta sectional vapor volume ratio (cross-sectional vapor void fraction) , Pr _L is the Prandtl number of the liquid refrigerant, and k _L [w / m · k] is the thermal conductivity of the liquid refrigerant. Here, β is a parameter indicating the abundance ratio of the gas refrigerant shown in FIG.

  FIG. 5 shows that when the row pitch Lp is 72.8 mm and the step pitch Dp is 21 mm, {refrigeration capacity / (air side pressure loss × wind velocity)} calculated using the above-described formula, the pipe diameter, and It is a figure which shows the relationship between fin pitches.

  Here, the value of {refrigeration capacity / (air-side pressure loss × wind speed)} is calculated by making the front wind speed constant at 1.1 m / s and changing the pipe diameter and fin pitch as parameters. Moreover, the length of the plate-like fin 2 in the step direction was 84 mm, the length in the row direction was 330 mm, and the thickness of the plate-like fin 2 was 0.25 mm.

  From FIG. 5, it can be seen that, at each fin pitch, {refrigeration capacity / (air-side pressure loss × wind speed)} increases as the tube diameter increases, reaches a maximum at a certain tube diameter, and then decreases.

  When the fin pitch is made small, the ventilation resistance when the air flow passes between the plate-like fins 2, that is, the air-side pressure loss increases, so that the refrigeration capacity tends to decrease. On the other hand, when the fin pitch is increased, the refrigeration capacity increases up to a predetermined fin pitch. However, if the fin pitch exceeds the appropriate range, the ventilation resistance, that is, the air-side pressure loss is reduced, but the heat transfer area is reduced, so that the refrigerating capacity tends to be lowered.

  FIG. 6 is a diagram showing an appropriate range of the pipe diameter and the fin pitch. The points represented by squares in FIG. 6 correspond to the peak values of the fin pitch graphs indicated by solid arrows in FIG. For example, in the graph in FIG. 5 where the fin pitch is 10 mm, the tube diameter corresponding to the peak value is approximately 10 mm. Therefore, FIG. 6 shows a point represented by a square at a position where the fin pitch is 10 mm and the tube diameter is approximately 10 mm.

  Further, the point represented by a triangle is a point on the left side of the peak of each graph shown in FIG. 5 and is a point reduced by 15% from the peak value of each graph (point indicated by a dotted arrow in FIG. 5). Corresponding to In a range not exceeding this point, the value of the graph gradually decreases, so that the refrigeration apparatus can be operated while maintaining high performance.

  For example, in the graph with the fin pitch of 10 mm in FIG. 5, the value decreases by 15% from the peak value when the tube diameter is approximately 5.6 mm. Therefore, FIG. 6 shows a point represented by a triangle at a position where the fin pitch is 10 mm and the tube diameter is approximately 5.6 mm.

  If the pipe diameter and fin pitch in the region between the straight line connecting the points represented by the squares shown in FIG. 6 and the straight line connecting the points represented by the triangles are selected, the fin pitch is shown in FIG. The performance within 15% is obtained as compared with the performance at the peak value of the graph.

  Here, when the tube diameter is smaller than 7 mm, it is difficult to form a fin collar (a hole and a collar into which the heat transfer tube is inserted) exceeding 8 mm in height in terms of manufacturing technology. On the other hand, when the tube diameter is larger than 10 mm, it is necessary to increase the thickness of the tube in order to increase the pressure resistance, and it is difficult to bend the heat transfer tube into a hairpin shape.

  Moreover, if the fin pitch is smaller than 8 mm, the air-side pressure loss increases as described above, so that the refrigerating capacity tends to decrease.

  Moreover, in the heat exchanger 1 of this embodiment, in order to make a fin pitch 12 mm or more, it is necessary to make the height of a fin collar 12 mm or more. However, it is difficult to form a fin collar having a height exceeding 12 mm on the plate-like fin 2 in terms of manufacturing technology.

  Therefore, the range indicated by the dotted line in FIG. 6, that is, the range in which the tube diameter D is 7 mm to 10 mm and the fin pitch Fp is 8 mm to 12 mm is desirable.

  FIG. 7 is a diagram showing a relationship between {refrigeration capacity / (air side pressure loss × wind speed)}, tube diameter, and fin pitch when the row pitch Lp is 65 mm and the step pitch Dp is 19 mm. Moreover, FIG. 8 is a figure which shows the appropriate range of the pipe diameter and fin pitch in this case. Other conditions are the same as those in FIGS.

  As in the case of FIG. 6, the points represented by squares in FIG. 8 correspond to the peak values of the fin pitch graphs indicated by solid arrows in FIG. 7. Further, the point represented by a triangle is a point on the left side of the peak of each graph shown in FIG. 7 and a point reduced by 15% from the peak value of each graph (point indicated by a dotted arrow in FIG. 7). Corresponding to

  In this case, similarly to the case shown in FIG. 6, the range indicated by the dotted line in FIG. 8, that is, the range where the tube diameter D is 7 mm or more and 10 mm or less and the fin pitch Fp is 8 mm or more and 12 mm or less can be selected as a desirable range. . This is because within this range, a performance within 15% of the performance at the peak value in the graph shown in FIG. 7 is obtained.

  Note that a part of the range indicated by the dotted line in FIG. 8 is above the straight line connecting the points indicated by the squares. However, even when the pipe diameter and fin pitch in this region are selected, the refrigerant circulation By optimizing the amount and the wind speed, a performance within 15% can be obtained as compared with the performance at the peak value of the graph shown in FIG.

  FIG. 9 is a diagram showing the relationship between {refrigeration capacity / (air-side pressure loss × wind speed)}, tube diameter, and fin pitch when the row pitch Lp is 75 mm and the step pitch Dp is 25 mm. is there. Moreover, FIG. 10 is a figure which shows the appropriate range of the pipe diameter and fin pitch in this case. Other conditions are the same as those in FIGS.

  As in the case of FIG. 6, the points represented by squares in FIG. 10 correspond to the peak values of the graphs of the fin pitches indicated by solid arrows in FIG. Further, the point represented by a triangle is a point on the left side of the peak of each graph shown in FIG. 9 and a point reduced by 15% from the peak value of each graph (a point indicated by a dotted arrow in FIG. 9). Corresponding to

  Also in this case, similarly to the case shown in FIG. 6, the range indicated by the dotted line in FIG. 10, that is, the range where the tube diameter D is 7 mm or more and 10 mm or less and the fin pitch Fp is 8 mm or more and 12 mm or less can be selected as a desirable range. .

  Note that a part of the range indicated by the dotted line in FIG. 10 is below the straight line connecting the points indicated by the triangles. However, even when the pipe diameter and fin pitch in this region are selected, the refrigerant By optimizing the circulation amount and the wind speed, a performance within 15% can be obtained as compared with the performance at the peak value of the graph shown in FIG.

  Further, it is desirable that the row pitch Lp is 75 mm or less and the step pitch Dp is 25 mm or less. When the row pitch is 75 mm and the step pitch Dp is more than 25 mm, when the arrangement volume of the heat exchanger is considered to be constant, the outer dimensions of the plate-like fins 2 are limited, so the number of collar portions 2b (through holes 2a Need to be reduced. However, if the number of the collar portions 2b is reduced, it is necessary to increase the circulation amount of the refrigerant flowing inside the heat transfer tube 3. As a result, the in-tube pressure loss of the heat transfer tube 3 increases, and the performance of the heat exchanger 1 is deteriorated. This in-tube pressure loss increases as the tube diameter decreases.

  On the other hand, if the step pitch Dp is less than 19 mm and the row pitch Lp is less than 65 mm, it is difficult to form the collar portion 2b having a height of 8 mm or more on the plate-like fin 2 due to manufacturing technology. For this reason, it is desirable that the step pitch Dp is 19 mm or more and the row pitch Lp is 65 mm or more.

  As described above, according to the present invention, by increasing the height of the collar portion 2b so that the outer diameter D of the heat transfer tube 3 is in the range of 7 mm to 10 mm and the fin pitch is 8 mm to 12 mm, The cooling efficiency in the heat exchanger 1 can be improved.

  Further, the end portion of the collar portion 2b of one plate-like fin 2 of the adjacent plate-like fins 2 is in contact with the surface opposite to the surface from which the collar portion 2b of the other plate-like fin 2 projects. The frost resistance of the heat exchanger 1 can be improved. Moreover, by setting it as such a structure, a part of heat exchanger tube 3 inserted in the through hole 2a of the plate-like fin 2 is not exposed, and the corrosion resistance of the heat exchanger 1 is improved.

  The heat exchanger according to the present invention is suitable for use in an air conditioning refrigeration apparatus.

DESCRIPTION OF SYMBOLS 1 Heat exchanger 2 Plate-like fin 2a Through-hole 2b Collar part 3 Heat transfer tube 4 Tube sheet

Claims (2)

Plural plate-like fins in which a plurality of through holes are provided in a row direction parallel to the gas flow direction and a step direction orthogonal to the gas flow direction, and a cylindrical collar portion is provided around each through hole When,
A heat transfer tube through which the refrigerant passes through the through-hole and in contact with the collar portion;
The outer diameter D of the heat transfer tube is 7 mm or more and 10 mm or less, and the fin pitch Fp of the plurality of plate-like fins is 8 mm or more and 12 mm or less,
The plurality of plate fins are arranged in parallel so that the collar portion of each plate fin protrudes in one direction, and the color of one plate fin of the adjacent plate fins among the plurality of plate fins. The end of each part is in contact with the surface opposite to the surface from which the collar portion of the other plate-like fin protrudes ,
The heat exchanger according to claim 1 , wherein a row pitch Lp of the heat transfer tubes in the row direction is 65 mm or more and 75 mm or less, and a step pitch Dp of the heat transfer tubes in the step direction is 19 mm or more and 25 mm or less . The heat exchanger according to claim 1 , wherein the refrigerant is a carbon dioxide refrigerant.