JPS6132599B2 - - Google Patents

Description

[Detailed description of the invention]

(Industrial Application Field) The present invention improves the heat transfer coefficient by maintaining a high heat transfer coefficient on both the refrigerant and air sides, thereby reducing the size and maintaining high performance of refrigerators, air conditioners, etc. This invention relates to an air-to-air type heat exchanger. (Prior art) Conventional air-to-refrigerant heat exchangers consist of heat transfer tubes (bare tubes) with smooth inner and outer walls, and corrugated fins or flat fins with slits cut into them (slit fins). ) or fins with similar slits in the corrugated fins (wavy slit fins), which improves the performance of cross-fin coils on the air side, where the heat transfer coefficient is low. It was created as a. (Problems to be Solved by the Invention) However, as the heat transfer coefficient on the air side is improved, the cost increases, and there is a problem that increases air resistance, so there is a limit to this. The effects of improvement are not as effective as expected. Considering the case where a heat exchanger with each of the above-mentioned structures is used as an evaporator, taking an air conditioner as an example, from the condition that the air condition of the target room is in the comfortable range, the desired cooling capacity is Desired air volume is determined. Further, the area through which the coil passes must be determined so that the wind speed passing through the coil is within a practical range and satisfies the above-mentioned required air volume. Therefore, the required cooling capacity (required heat exchange capacity) is adjusted by adjusting the number of rows of coils to satisfy the required cooling capacity. Therefore, if the heat exchange performance is low, there will inevitably be a problem that the heat exchanger will be thick in the row direction, that is, it will be long in the depth direction based on the wind flow, resulting in increased air resistance and fan input and The disadvantage is that it makes a lot of noise. Further, in a heat exchanger having a thick shape in the column direction, it becomes difficult to apply a load per pass (one refrigerant flow path), which is undesirable because the design of the heat exchanger becomes complicated. Regarding the above-mentioned problem of the load burden per pass, that is, the amount of refrigerant circulation, if the amount of refrigerant circulation per pass is increased, the pressure drop of the refrigerant becomes large, and the heat transfer coefficient on the refrigerant side increases somewhat. However, the logarithmic mean temperature difference (△Tm) becomes smaller, resulting in the need to significantly increase the heat transfer area, and conversely, when the refrigerant circulation rate is reduced, the heat transfer coefficient on the refrigerant side decreases and refrigerant boiling becomes slower, which causes However, as a result, the heat transfer area has to be significantly increased, and the optimum range A' is narrow due to the relationship between the amount of refrigerant circulation per pass and the required heat transfer area as shown in Figure 1. This makes the design of the heat exchanger extremely difficult. Furthermore, when used as a condenser, there is a problem that the shape of the heat exchanger becomes large because it becomes thick in the column direction. On the other hand, as a technique for increasing the heat transfer coefficient on the refrigerant side of heat exchanger tubes, as disclosed in JP-A-50-128668, fine hole-like depressions made of intersections of cross grooves etc. are provided on the inner surface of heat exchanger tubes. There are some that increase the boiling heat transfer coefficient by creating so-called nucleate boiling, in which the hollows create bubble nuclei that serve as evaporation nuclei.
However, although nucleate boiling can occur in each depression on the inner wall surface of the tube, the essential liquid refrigerant does not spread over the entire inner wall of the tube, and nucleate boiling occurs only in the depressions at the bottom of the tube where the liquid refrigerant accumulates. However, there is a problem in that the heat transfer coefficient on the refrigerant side cannot be sufficiently improved. The present invention has been made in view of the above points, and its purpose is to increase the heat transfer coefficient on the air side of the heat exchanger tube in a heat exchanger, and to increase the heat transfer coefficient on the inside of the heat exchanger tube, that is, on the refrigerant side. The objective is to improve the heat transfer coefficient K, which comprehensively determines the performance of the air-to-air heat exchanger, by effectively increasing the heat transfer coefficient of the heat exchanger. (Means for Solving the Problems) In order to achieve the above-mentioned object, the means taken by the present invention are such that the air side uses fins such as needle-like fins, slit fins, and wave-shaped slit fins to prevent turbulence or leading edges from flowing air. The heat transfer coefficient on the air side is improved by forming fins that exhibit an action. On the other hand, in the heat exchanger tube on the refrigerant side, the aim is to make most of the liquid refrigerant in the heat exchanger tube into a spiral flow so that it spreads over the entire inner wall of the tube.
By forming two systems of helical grooves on the inner wall of the heat exchanger tube, which are oblique angles in opposite directions with respect to the tube axis and have different groove depths, most of the The purpose is to form a spiral liquid main flow that determines the flow direction of the liquid refrigerant, and to promote nucleate boiling in shallow grooves that are tributaries to this spiral liquid main flow and nucleate condensation at the tips of the protrusions. In other words, in the case of evaporation, by supplying the liquid refrigerant at the bottom of the tube along the deeper spiral groove formed in the inner wall of the tube to both sides or the upper surface of the tube, the liquid refrigerant is supplied to the entire inner wall of the tube. A liquid refrigerant is supplied to improve the refrigerant side heat transfer coefficient. Furthermore, the liquid refrigerant flowing through the deep spiral groove that is spread over the entire inner wall of the pipe rides up the shallow spiral groove and is branched and supplied to the tributary stream, promoting nucleate boiling of the liquid refrigerant within this shallow groove. , further improves the heat transfer coefficient on the refrigerant side. In addition, in the case of condensation, especially in the middle of the condensation process, nucleate condensation is promoted using the tips of many protrusions formed by the intersection of deeper spiral grooves and shallower spiral grooves as nuclei. ,
Moreover, the condensed liquid film on the heat exchange surface of the protrusion facing the shallower spiral groove is made into a thin liquid film by supplying the condensed liquid from the shallower groove to the deeper groove, In order to promote condensation, the deeper spiral grooves form a spiral liquid main flow that serves as a discharge path for the condensed liquid refrigerant and spreads over the entire inner wall of the pipe, thereby reducing the liquid level of the liquid refrigerant flowing through the grooves. The tip of the protrusion facing the deeper groove is not covered with liquid refrigerant, which also promotes condensation. In order to generate such a spiral liquid main stream and tributary streams, specifically, the angle of inclination of the spiral groove with respect to the tube axis of the deeper groove is set at 5° to 15°, and the angle of the spiral groove with the shallower groove is set at an angle of 5° to 15°. The bevel angle of the spiral groove is 3° to 45°, and the groove pitch is 0.15 to 1.5.
mm, the groove depth of the spiral groove in the depth direction is 0.1 to 0.5 mm, and the structure is formed by forming a large number of pyramid-shaped protrusion rows such as pyramids and truncated pyramids. (Function) As a result, in the present invention, a spiral liquid main stream and a tributary stream that function as described above are reliably generated,
Corresponding to the improvement in the air-side heat transfer coefficient, the refrigerant-side heat transfer coefficient is significantly improved. (Example) Hereinafter, specific embodiments of the present invention will be described in detail with reference to the drawings. The heat exchanger according to the embodiment of the present invention is an air-to-air heat exchanger comprising a large number of heat transfer tubes 1 arranged in multiple rows, and a large number of fins 2, 2... provided orthogonally to the outer surface of the heat exchanger tubes 1. The inside of the heat transfer tube 1 is used as a refrigerant passage, and the area around the fins 2, 2 outside the tube is used as an air passage, and is mainly used as an evaporator of a refrigeration system. An example of this heat exchanger is shown in FIGS. 2 to 4, in which the fins 2, 2... are formed into corrugated slit fins, and the heat exchanger tube 1 is formed into an internally treated tube. ing. Note that the material of the fins 2 is mainly aluminum, and the material of the heat exchanger tubes 1 is copper or aluminum. First, regarding the fin 2, a large number of tube fitting insertion holes 3, 3 are bored in the corrugated fin board 2-1, and cut edges parallel to the ridge lines 4, 4 are formed between the adjacent holes 3, 3. The slits 5, 5 are formed by cutting and raising the tongues 6, 6, and the slits 5, 5 after cutting and raising the tongues are such that the direction in which the ridge lines 4, 4 extend is compared to the direction perpendicular to the direction. It is formed so that it becomes long. Each cut-out tongue piece 6.6 is separated from the surface of the fin board 2-1 on the long side and floats up, and is connected on the short side. The fins 2, 2... having such a structure are attached to the heat transfer tubes 1, 1 arranged in multiple stages and rows, and a corrugated passage 7, 7 is formed between each fin 2, 2...
is formed, and air is blown in a direction perpendicular to each of the ridge lines 4, 4. By doing so, the air flows in a meandering manner within the wave-shaped passages 7, 7 with the ridge lines 4, 4 as turning points while becoming turbulent, and a part of the air passes through the slits 5, 5 into the adjacent air. The air flows into the corrugated passage 7, and the synergistic effect thereof further promotes turbulence in the air flow. In addition, each of the cut-out tongues 6, 6 collides with the airflow and functions as a leading edge effect of the airflow, and the above action prevents the development of a temperature boundary layer, and the air and the fins 2, 2. It becomes possible to improve the heat exchange performance between... Note that the fins 2, 2, . . . are not limited to the wave-shaped slit fins of the above example, but may of course be slit fins formed by cutting and raising slits in a flat substrate. Next, the basic aspect of the heat exchanger tube 1 is that a spiral groove 8 is provided as an inner surface treatment, but as illustrated in FIGS. The inner wall has two-way spiral grooves 8, 8 which are both clockwise and counterclockwise at an oblique angle, and these two-way spiral grooves 8, 8 allow
A large number of protrusions 81, 81 are formed in a spiral arrangement so that the tops thereof are directed toward the center of the tube. Further, the spiral grooves 8, 8 are formed into V-shaped grooves and U-shaped grooves with different groove depths, and with this structure, the portion where the upper edges of both grooves 8, 8 intersect with each other is set as the apex. A countless number of quadrangular pyramid-shaped projections 81 are formed, and countless diagonal fluid passages that intersect with each other are formed between the projections 81.
A heat exchanger tube with an increased inner surface area can be obtained. In addition, the heat exchanger tube 1 illustrated in FIGS. 5 to 7
is a structure in which both spiral grooves 88, 8 have a slight difference in groove depth, and the spiral groove 8 on the shallow groove side forming the tributary side is the spiral groove 8 on the deep groove side forming the main stream side of the flow path. The grooves 8 intersect with each other, thus forming a spiral main stream of liquid flowing along the main stream side flow path 9 and a tributary stream flowing over the tributary side passages 10, 10 spanning between the main liquid streams, The movement of liquid can be effectively utilized in two stages, and a heat transfer tube with higher heat transfer coefficient can be obtained. That is, during evaporation, the main flow of the spiral liquid, that is, the liquid refrigerant at the bottom of the tube along the deeper spiral groove 8 formed on the inner wall of the tube 1, is supplied to both sides of the tube or to the inner surface of the tube. Liquid refrigerant is supplied to the entire wall, improving the heat transfer coefficient on the refrigerant side. Furthermore, the liquid refrigerant flowing through the deep spiral groove 8 that spreads over the entire inner wall of the pipe climbs onto the shallow spiral groove 8 and is branched and supplied to the tributary stream.
Nucleate boiling of the liquid refrigerant is promoted within the refrigerant, further improving the heat transfer coefficient on the refrigerant side. Therefore, the large surface area of the protrusions 81 increases the heat exchange surface, and together with this, it is possible to significantly improve the heat transfer coefficient on the refrigerant side during evaporation. In addition, during condensation, especially in the middle of the condensation process, the tips of a large number of protrusions 81 formed by the intersection of the deeper spiral grooves 8 and the shallower spiral grooves 8 become nuclei. Condensation is promoted. Moreover, the condensed liquid film on the heat exchange surface of the protrusion 81 facing the shallower spiral groove 8 is a thin liquid film as the condensed liquid is supplied from the shallower groove to the deeper groove. This results in a condensate film, which promotes condensation. Furthermore, in the deeper spiral grooves, a spiral liquid main flow that serves as a discharge path for the pseudo-condensed liquid refrigerant is formed and spreads over the entire inner wall of the pipe, so the liquid level height of the liquid refrigerant flowing in the groove 8 can be lowered. The tip of the protrusion 81 facing the deeper groove 8 is not covered with liquid refrigerant, so condensation is promoted here as well. These promote condensation and also improve the heat transfer coefficient on the refrigerant side. In addition, each groove depth of the spiral grooves 8, 8, that is, the protrusion 81
Groove depth (mountain height) in different parts of slope form
Various experiments have shown that when a relationship ratio of h ₂ ≦ (4/5) h ₁ is set between _{h 1} and h ₂ , a heat transfer tube with low pressure loss and high heat transfer coefficient can be obtained. The result was clear. Further, both spiral grooves 8, 8 have respective twist angles θ.
₁ , θ ₂ are in the range of θ ₁ = 5 to 15 degrees, θ ₂ = 3 to 45 degrees, and the groove pitches P ₁ and P ₂ are in the range of 0.15 to 1.5 mm.
The groove depth _h1 of the spiral groove 8 on the deep groove side is 0.1 to 0.5 m/m.
By specifying the range of Figures 8 to 11,
As seen in the measurement results during evaporation in the figure,
Heat exchanger tubes with various excellent performances can be obtained, and the range outside the above range has advantages and disadvantages and is not a preferable range. However, the conditions related to the above measurement results are as follows. Refrigerant used…Freon R-22 Boiling liquid pressure…4.2Kg/cm ² G Condensed liquid pressure…18.8Kg/cm ² G Refrigerant circulation amount…68Kg/h Heat transfer tube diameter…φ9.52×t1.0m/m FIG. 8 shows the heat transfer coefficient ratio αn/αs between the heat exchanger tube 1 and the bare tube whose inner surface has not been treated according to the embodiment of the present invention on the vertical axis, and the depth _h1 of the groove in the heat exchanger tube 1 on the horizontal axis. This is a diagram for the case where _h1 is around 0.3m/m,
This shows that about three times the performance can be obtained compared to the bare tube. On the other hand, Fig. 9 shows the case where the vertical axis is the pressure loss ratio Pn/Ps and _the groove depth _h1 is the horizontal axis. The same pressure loss can be obtained, and in terms of both the heat transfer coefficient and pressure loss, the groove depth h ₁ is practically 0.1.
This indicates that the preferred range is ~0.5 m/m.
In addition, FIG. 10 shows that by changing the twist angle θ ₁ ,
The helix angle θ ₁ is 30°, and the groove depth h ₁ =
0.25 mm, h ₂ = 0.20 mm, and shows the heat transfer coefficient for the twist angle θ _1. The heat transfer coefficient α changes greatly depending on the twist angle, and reaches its maximum at about 10°. Furthermore, when θ ₂ is changed, the heat transfer coefficient α reaches its maximum when θ ₁ is between 5° and 15°, and θ ₁ is in the practical range of 5° to 15°. I found out something. On the other hand, in FIG. 11, the twist angle θ ₂ is changed, the twist angle θ ₁ is set to 7°, and the groove depth h ₁
= 0.25 mm, h ₂ = 0.20 mm. It shows the heat transfer coefficient with respect to the twist angle θ ₂ , and the heat transfer coefficient α reaches its maximum when θ ₂ is approximately 10°. As shown in the figure, when θ ₁ is further varied,
It has been found that the heat transfer coefficient α reaches its maximum when θ ₂ is between 3° and 45°, and a preferable range of θ ₂ is between 3° and 45°. Here, the twist angle θ1 is a small angle and has a narrow range, and _O2 has a wide range up to a considerably large angle.The reason is that _θ1 has a deep groove depth and is a mainstream spiral flow, so the flow velocity can be increased to achieve a large flow rate. Since θ ₂ has a shallow groove depth and is a tributary stream, the main flow is branched and supplied to promote nucleate boiling, etc., so it can be selected from a wide range. Next, adjust the groove pitch between adjacent grooves to 0.15 to 1.5 mm.
However, if the diameter is 0.15 mm or less, the grooves will be covered with a liquid film along the pipe wall, and secondary flow (like a vortex) will not occur easily in the grooves, and the angle of the peak will not be sharp during manufacturing. If the diameter is 1.5 mm or more, the increase rate of the inner surface on the tributary side will decrease, the peak angle of the crest will become large, making it difficult for secondary flow to occur in the groove, and on the main stream side, due to the formation of a spiral flow. This is based on manufacturing constraints such that the groove depth must be increased, but the pipe wall thickness must also be increased, and after all, the preferred range for the groove pitch is 0.15 to 1.5 mm. As mentioned above, as an example different from the heat exchangers shown in FIGS. 2 to 7, a heat exchanger in which needle fins 2 are attached to a heat exchanger tube 1 whose inner surface has been treated in the manner described above is shown. It is shown in FIGS. 12 and 13. This also exerts a turbulent flow and leading edge effect on the air, and can similarly improve the heat transfer coefficient on the air side. The needle-like fins 2 are formed by bending a series of flat plates serving as fin substrates into a substantially U-shape so that the longitudinal center line is at the center, so that the base portion in contact with the heat exchanger tube 1 is closer to the opposing leg portion. The fin material has a collapsed U-shape in cross section so that it can bulge out slightly laterally, and fine pitch cuts are made in both legs. By spirally winding the fins and brazing them, it is possible to easily construct a heat exchanger in which the needle fins are radially clustered as shown in the figure. Next, referring to the characteristic diagrams in FIGS. 14 to 16, it will be explained that the heat exchanger of the present invention, whose structure is shown in each of the above examples, can exhibit excellent characteristics in improving heat exchange performance. I will explain. (a) Characteristics of heat exchanger tubes with inner surface treatment The relationship between the heat transfer coefficient inside the heat exchanger tube and the heat flow rate (heat flow rate per unit area) is α _R = f (q・G・di) …(a) However, α _R : Inner pipe heat transfer coefficient q: Heat flow rate G Refrigerant circulation amount di: Pipe inner diameter. As shown in Fig. 14, the relationship between α _R and q is based on the applicable range of q in the conventional bare tube cross-fin heat exchanger, that is, 6000 to 9000 kcal/
While the heat transfer coefficient is small at m ² hr,
Compared to conventional products, the heat transfer coefficient is 1.6 to 1.8 times higher due to the larger value of q, and the variation in the heat transfer coefficient is smaller with respect to the variation of q. is clearly indicated. Note that the pressure loss per unit length of the heat exchanger tube of the present invention is almost the same as that of a conventional bare tube. (b) Characteristics of a combination of fins and internally treated heat exchanger tubes The heat transfer coefficient K in this case is obtained by the following equation. 1/K=SHF/[{1-(Rf/R) (1-Ef)}×α _A ]+R×[1/α _R +γ]
...(b) However, SHF: Sensible heat ratio Rf: Inside/outside ratio of the fin section (Rf=R-1) α _A : Air side heat transfer coefficient, Ef: Fin efficiency, R: Outside/outside ratio γ: Contact heat of the tube The resistance and the capacity Q of the heat exchanger can be obtained from the following formula. Q=K・A・△Tm...(c) However, A: Air side heat transfer area △Tm: Logarithmic mean temperature difference.The relationship between the front wind speed and air side heat transfer coefficient for each type of fin is As shown in Figure 15,
As a result of the measurement, it was clear that the shapes were arranged in the order of wavy slit shape, slit shape, needle shape, and flat plate shape from the highest to the lowest. Although specific numerical values are not shown in FIG. 15, both the air side heat transfer coefficient (α _A ) and the front wind speed are shown on a logarithmic scale. A comprehensive comparison of equations (b) and (c) and the measurement results is shown in the table below. However, the table below shows the values when each item is set as 100 for the combination of flat fins and smooth tubes (bare tubes), and the heat exchange amount is set as 100 for all the various combinations.

【table】

[Table] As is clear from the above table, if the air side fin is a special fin other than a flat plate fin, such as the above-mentioned slit fin, and at the same time the refrigerant side pipe inner surface is machined with the above special groove, the heat transfer coefficient α _A on the air side is 1.25. ~
1.5 times, the heat transfer coefficient α _R on the refrigerant side (inside the tube) is 1.7
~2 times, and the heat transfer coefficient K is approximately 1.3 times. Therefore, it is shown that for the same amount of heat exchange Q, the required heat transfer area A can be reduced by about 25%. (c) In the case of a heat exchanger that combines an internally treated tube and a corrugated fin with slits As shown in Figure 16, compared to a conventional combination of flat fins and bare tubes (indicated by broken lines). Therefore, it is clear that the required heat transfer area to obtain the same capacity is small, and the change in area when changing the load (refrigerant flow rate) per pass is small, making it easier to design the heat exchanger. be. This is because when the inner surface treatment is used, as is clear from the table above, α _R is improved, the heat transmission coefficient K becomes better, and the length of the heat exchanger tube can be shortened, so one pass Even if the load is increased, the pressure drop will not be so large, so △
This is because the decrease in Tm is not so large and the increase in the required heat transfer area is small.On the other hand, even if the load is reduced, the refrigerant boils well in the pipes, so the degree of decrease in α _R is small and the increase in the required heat transfer area is small. This is because the decrease in the overall heat transfer coefficient K is also small, and the increase in the required heat transfer area is also small. The theoretical and experimental considerations based on the above experiments are for evaporators, but in the case of condensers as well, the internal heat transfer coefficient of internally treated tubes is within the range of normal use. (compared to bare tube)
The improvement is 1.4 to 1.8 times, and the improvement in heat transfer coefficient when combined with fins is slightly lower than when using an evaporator, but it is considerably improved, so the heat exchange ability is improved overall. (Effects of the Invention) As detailed above, the heat exchanger of the present invention has the following effects. That is, on the air side, the fins 2, 2... are shaped to exhibit turbulent flow or leading edge action on the circulating air, so it is possible to prevent the development of a temperature boundary layer and improve the heat transfer coefficient on the air side. can. Regarding the refrigerant side (a) There are two systems of spiral grooves 8 on the inner wall of the heat exchanger tube 1 that are obliquely angled in opposite directions and have different groove depths.
8, a large number of protrusions 81 rows are formed on the inner wall of the pipe, increasing the heat exchange surface. Form a main flow to spread the liquid refrigerant all over the inner wall of the pipe,
Furthermore, it is possible to promote nucleate boiling in the shallow grooves that are tributaries to the main spiral liquid flow and nucleate condensation at the tips of the protrusions, dramatically increasing the heat transfer coefficient on the refrigerant side during evaporation and condensation. can be improved. (b) By limiting the oblique angle of both helical grooves 8, 8, the groove pitch, and the groove depth of the deeper helical groove within the range of the predetermined dimensions, the main flow of the helical liquid and the tributary stream can be well connected. Moreover, it can be generated reliably, and it is possible to realize the effect that the refrigerant side heat transfer coefficient is approximately 1.7 to 2 times higher than that of the conventional method. As described above, by improving the heat transfer coefficient on both the air side and the refrigerant side, the heat exchange performance is improved overall, and as a result, a heat exchanger that is thin in the column direction and has low air resistance and noise is obtained.
Even if the load per pass is increased, the required heat transfer area only increases by a small amount, making it possible to widen the optimum range, simplifying the design of the heat exchanger, and making it even more compact. As a result, a heat exchanger having a heat transfer coefficient approximately 1.3 times that of a heat exchanger with flat fins and bare tubes can be obtained.

[Brief explanation of the drawing]

FIG. 1 is a characteristic diagram of a conventional heat exchanger. FIG. 2 is a partial perspective view of a heat exchanger according to an embodiment of the present invention, FIG. 3 is a partially sectional front view, FIG. Figure 5 is a partially cutaway external view of the heat exchanger tube related to the heat exchanger shown in Figure 2, and Figure 6 is an enlarged view of the main part of the heat exchanger tube shown in Figure 5, showing the axis and the center of one spiral groove. 7 is an enlarged view showing the cut portion of FIG. 6, FIGS. 8 to 11 are characteristic diagrams of the heat transfer tube, and FIGS. 12 and 13 are diagrams showing the characteristics of the heat transfer tube according to the present invention. FIGS. 14 to 16 are front and side views showing essential parts of one example of a heat exchanger according to another embodiment, and FIGS. 14 to 16 are characteristic diagrams of the heat exchanger of the present invention. 1...Heat transfer tube, 2...Fin, 2-1...Fin substrate, 4...Ridge line, 5...Slit, 8...Spiral groove, 81...Protrusion.

Claims (1)

[Claims] 1 A heat exchanger having a large number of fins 2, 2... on the outer surface of the heat transfer tube 1, with the inside of the heat transfer tube 1 serving as a refrigerant passage and the outside of the tube serving as an air passage, wherein the fins 2, 2... are needle-like fins, fins formed by cutting and raising slits on a flat plate, and corrugated fins having ridge lines 4, 4 extending orthogonally to the air flow direction. The heat exchanger tube 1 is formed into a fin that exhibits a turbulent flow or a leading edge effect on the circulating air, such as a fin having a large number of slits 5, 5 formed by cutting and raising.
The inner wall has two systems of helical grooves 8 that are oblique angles in opposite directions with respect to the tube axis and have different groove depths, and the spiral groove 8 with the deeper groove depth has an oblique angle with respect to the tube axis. 5° to 15°, the oblique angle of the shallower spiral groove 8 is 3° to 45°, the groove pitch is 0.15 to 1.5 mm, and the groove depth of the deeper spiral groove 8 is 0.1. A heat exchanger characterized by forming a large number of 81 rows of protrusions in the shape of pyramids, truncated pyramids, and equilateral pyramids each having a diameter of 0.5 mm.