JPH051891A - Heat transfer tube with internal groove - Google Patents

Abstract

PURPOSE:To obtain a heat transfer tube with internal grooves, reduced in the increase of a unit weight and not so high in the hardness of manufacturing, by a method wherein a ratio of the sectional area of respective grooves to the depth of the groove is specified while an vertical angle of a crest positioned between respective grooves is also specified, in the heat transfer tube with internal grooves, in which a multitude of grooves are formed on the internal surface thereof with a specified twisting angle. CONSTITUTION:A heat transfer tube with internal grooves is provided with a multitude of spiral grooves, having a ratio of the depth Hf of the groove to the inner diameter Di of the tube or Hf/Di=0.02-0.03, a twisting angle beta with respect to the axis of the tube r beta=7 deg.-30 deg. and a truncated section, on the inner surface thereof for effecting the phase change of fluid flowing through the inside of the tube. The ratio of the sectional area S of the section vertical to the axis of respective grooves to the depth Hf of the grooves are specified so that S/Hf=0.15-0.40 while the vertical angle of a crest positioned between respective grooves in the section orthogonal to the axis of the tube is specified so as to be 30-60. The sectional area of respective grooves and the shape of crest defining the grooves are limited in such a manner whereby the reduction of weight, as well as the improvement of workability are permitted and the heat transfer tube can be provided with high performance.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a heat exchanger tube with an inner surface groove (hereinafter simply referred to as an inner surface grooved tube) suitable for use in a heat exchanger such as an air conditioner, a refrigerator, a boiler, etc., in which the fluid inside the tube undergoes a phase change Say) improvement.

[0002]

2. Description of the Related Art As shown in FIG. 10, an inner grooved tube is a metal tube such as a copper tube provided with a large number of spiral grooves on its inner surface.

Conventionally, this type of inner grooved tube has been announced with various restrictions on the groove depth, the groove shape, the twist angle of the groove, and the like. ,
It's not always the optimal one, and there is a feeling that it is just one step away from expanding its application. The biggest reason is the ratio of heat transfer performance to manufacturing cost. In other words, the inner grooved tube is difficult to obtain stable quality unless it is formed by the rolling method because of its fine inner surface structure. However, the rolling method is limited in the number of rotations of the motor, speed limit, etc. If so, there is a limit in manufacturing cost. However, since the smooth tube can be manufactured by the high-speed drawing method, considering the ratio of heat transfer performance to the manufacturing cost of the conventional inner grooved tube, the reason is that it is difficult to obtain the merit of switching from the smooth tube.

FIG. 11 shows a typical shape of an inner grooved tube which has been conventionally put to practical use.

However, the grooved tube having such an inner surface shape has a low performance / cost ratio for the reason described below.

First of all, it has been well known that the groove depth (Hf) and the performance are in a proportional relationship, but the limit to which the pressure loss greatly increases as compared with the smooth tube is the groove depth. (H
The ratio (Hf / Di) between f) and the inner diameter (Di) is 0.02 to
Despite being around 0.03, Hf / Di =
Since it is 0.018 or less, the groove depth (Hf) does not reach the limit of pressure loss. This is also due to the cost reason that increasing the groove depth (Hf) while maintaining the conventional shape immediately leads to an increase in unit weight (hereinafter referred to as unit weight).

Secondly, the factors that affect the performance are the groove shape and the mountain shape. The details will be described later, but in the conventional product shown in FIG. 11A, the groove has a small cross-sectional area (S), The apex angle (α) was large, so the performance was not sufficient. Also, FIG.
In the case of 1 (b), although the groove cross-sectional area (S) was larger than that in (a), satisfactory performance could not be obtained because the peaks were trapezoidal.

[0008]

SUMMARY OF THE INVENTION The present invention has been made in view of the above points, and it is an object of the present invention to increase the unit weight while improving the performance and to make the manufacturing difficult. Another object is to provide a practical inner grooved tube that is not relatively large.

[0009]

That is, the gist of the present invention is to set the groove depth as large as possible so as not to increase the pressure loss of the internal fluid, as described in the claims, and , The thickness of the liquid film and the surface area inside the tube 2
This is because the restrictions were added in consideration of the points, and the shape of the ridges was also restricted by comprehensively judging the internal surface area of the pipe, the unit weight of the pipe, and the workability during manufacturing.

[0010]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS FIG. 1 shows a partially enlarged cross-sectional view of an embodiment of an inner grooved tube according to the present invention. In this embodiment, the outer diameter (O
D) Grooves were formed on the inner surface of a 9.52 mm copper tube with a twist angle (β) of 18 ° with respect to the tube axis so that triangular peaks with a crest number (n) of 60 could be formed at equal intervals. The effective wall thickness (Tw) is 0.30 mm.

The reasons for limiting the numerical values of the present invention will be described below in comparison with conventional products.

All the data shown below were measured using Freon: R-22 as the fluid in the pipe, and the steam pressure was 4 kg / cm ²
G, average dryness = 0.6, heat flow rate = 10 Kw / m ² , refrigerant flow rate = 200 kg / m ² S, condensing pressure = 14.6 kg / cm
^{The conditions are 2} G, superheat at the inlet = 50 ° C., supercooling at the outlet = 5 ° C., and the inner surface area is based on the minimum inner diameter (Di).

First, the groove depth (Hf) will be described.

FIG. 2 shows a general copper pipe with internal groove (OD = φ
9.52, Di = φ8.52, β = 18 °),
The horizontal axis represents the groove depth (Hf) as a ratio with the minimum pipe diameter (Di), and the vertical axis shows the heat transfer coefficient in the pipe and the pressure loss of the fluid in the pipe as a ratio with the smooth pipe. According to this figure, the heat transfer coefficient increases proportionally as the groove depth (Hf) becomes deeper, but the amount of increase from around Hf / Di = 0.02 to 0.03 becomes slow. The pressure loss ratio is Hf /
Up to Di = about 0.03, there is not much difference from the smooth tube, but after that, it increases sharply. Therefore, in order to select the highest possible range within the range where the pressure loss is not much different from that of the smooth tube, Hf
The range of /Di=0.02 to 0.03 is desirable.

Next, the twist angle (β) of the groove will be described.

FIG. 3 shows a copper pipe with internal groove (OD = φ9.52,
For Di = φ8.52, Hf = 0.20), the horizontal axis represents the twist angle (β) of the groove with respect to the tube axis, and the vertical axis represents the ratio of the heat transfer coefficient in the tube to that of the smooth tube. According to this figure, during evaporation, there is a slight peak around 7 to 20 °,
At the time of condensation, the performance tends to gradually increase as the twist angle (β) increases. On the other hand, considering that the increase of the twist angle (β) leads to the deterioration of the workability at the time of manufacturing the pipe, the optimum twist angle (β) is 7 ° where the performances of evaporation and condensation sharply rise. It is desirable to keep the range from the vicinity to about 30 °, and it can be considered that there is no great difference in performance within these ranges.

Next, the groove sectional area (S) will be described.

Generally, the improvement of the performance of the inner grooved tube in the phase change state is considered to be the effect of stirring the fluid due to the unevenness of the inner surface, the effect of increasing the inner surface area, and the effect of the liquid film fluctuation in the uneven portion. Regarding the first stirring effect, it is clear that the groove depth (Hf) is dominant and that the larger it is, the more it contributes to the performance improvement, but it is closely related to the third liquid film behavior. That is, when the fluid flows in the tube at a certain speed or more, the liquid tends to rise up in the spiral groove by the capillary action of the fine groove and the drag force resulting from the fluid velocity to wet the entire circumference of the tube, that is, a so-called annular flow. This situation is shown in FIG.
It shows in (b).

In the case of the smooth tube of FIG. 4 (a), the upper dry portion does not contribute to evaporation, whereas in the case of the grooved tube of FIG. 4 (b), evaporation is promoted over the entire surface. However, even in the grooved pipe 1, when the total cross-sectional area of the groove is different, the thickness of the liquid film has different appearances as shown in FIG. 5 when the total amount of liquid is constant. That is, when the groove cross-sectional area is large (c), the liquid film 2 is too thin and the tip of the crest is exposed,
No evaporation takes place here. Further, in the pipe (a) having a small groove cross-sectional area, the liquid film 2 is too thick, the thermal resistance between the fluid gas and the pipe wall increases, and the performance also deteriorates. Therefore, there is an optimum groove cross-sectional area such that the entire wall surface is covered with the thinnest possible liquid film and the state shown in FIG. However, when the shapes of the ridges that separate the grooves are the same, the inner surface area is inversely proportional to the groove cross-sectional area, so (c) tends to deteriorate in surface area compared to (b), but (a) shows Compared to (b), the surface area tends to improve the performance. Therefore, it is expected that the comprehensively optimum groove sectional area S (to be exact, S / Hf) is located at an intermediate position between (a) and (b) of FIG.

FIG. 6 shows an example in which the cross-sectional shape of the crests is changed with respect to the constant optimum groove cross-sectional area (S). In the figure, (a) has a larger peak apex angle (α) than (b), so that although it has excellent workability during pipe manufacturing, it has a large peak cross-sectional area, which leads to an increase in unit weight. At the same time, the inner surface area is reduced, which tends to reduce the performance. In the case of the rectangular or trapezoidal crests as in (c), the decrease of the total inner surface area due to the increase of the unit weight and the crest pitch is similar. on the other hand,
A mountain shape with a small apex angle (α) as in (d) increases the internal surface area without increasing the unit weight compared to (b), but a slim mountain with a significantly smaller apex angle significantly reduces workability. Therefore, the manufacturing cost is likely to increase significantly.

FIGS. 7 to 9 show qualitative influences of the above-mentioned groove and mountain shapes on the performance as data.

In FIG. 7, OD = φ9.52 and Di = φ8.5.
2 shows the relationship between the peak shape and the heat transfer coefficient in the inner grooved copper tube with Hf = 0.20, β = 18, and n = 60.
From this figure, it can be seen that the smaller the apex angle (α) is, the better the performance is, and the triangular mountain (B) is better than the trapezoidal mountain (A) in both evaporation and condensation. However, as the apex angle (α) becomes smaller, the workability at the time of manufacturing the pipe remarkably lowers, leading to an increase in cost. Therefore, the apex angle (α) is practically 30 to 60.
It can be said that the range of ° is desirable.

FIG. 8 is a view showing the relationship with the groove section cross-sectional area S (more accurately, S / Hf), with the apex angle being constant with triangular peaks of α = 50 °. According to this figure, the evaporation performance is S / Hf =
Although there is a peak in the vicinity of 0.3, in the range of 0.3 or less, the performance deterioration is gradual, compared with the sharp decrease in the range of 0.3 or more. On the other hand, the condensation performance sharply increases as S / Hf becomes smaller, and a slight peak is observed near S / Hf = 0.2.

From these tendencies, it can be said that the smaller the S / Hf is, the more stable the performance is, but on the other hand, the increase in the unit weight due to the increase in the number of peaks may be inversely proportional to the S / Hf. Don't forget. That is, when the specifications other than the number of peaks that define the groove are constant, a decrease in S / Hf means an increase in the number of peaks (n) and, in turn, an increase in unit weight.
Looking at this in relation to S / Hf, it shows an inversely proportional tendency. Therefore, it is necessary to comprehensively judge these and obtain the optimum specifications.

An example of trial calculation for considering the overall cost merit, which is one of the objects of the present invention, is shown below.

Assuming a fin coil type heat exchanger for a room air conditioner, which is one of the typical heat exchangers, the tube outside heat resistance including slit type aluminum fins and the tube inside heat resistance when using a conventional product Assume that the ratio of and is 75%: 25%. At this time, only the inside of the pipe is changed from the conventional product shown in FIG. 11A to an embodiment product of the present invention (OD = 9.52, Hf = 0.2, β = 1.
8B, α = 50 °, n = 60), the heat transfer coefficient improvement rate is converted into a heat passage rate increase rate, which is represented by B in FIG. When the same comparison is performed for the unit weight, FIG.
It becomes the graph of A.

In this case, the specifications of the conventional product (FIG. 11A) are OD = φ9.52, Hf = 0.15, β = 25.
°, α = 90 °, n = 65.

If the tube length is shortened by the amount corresponding to the improvement of the heat transmission rate, this amount will be the cost merit as it is, and the reduction of the unit weight must be taken into consideration the deterioration of the workability at the time of manufacturing. , And the cost merit is close to this. Therefore, A + B in FIG. 9 is the total merit on the tube purchase side.

Actually, if the air-conditioning capacity and the efficiency are improved in adopting the grooved pipe, the cost merit is reduced accordingly, and if the workability is deteriorated, the merit is further reduced. The merit conversion shown in FIG. 9 is merely a guide, but as a result of studying the present invention with an emphasis on reducing the unit weight as well as improving the performance, a sufficient merit is obtained even in the region where the S / Hf is small and the performance is small. What can be read can be read from FIG. Therefore, S / Hf is 0.1
A range of 5 to 0.40 is effective.

[0030]

As described above, according to the present invention, the unit weight can be reduced and the workability can be improved by limiting the cross-sectional area of each groove and the shape of the peak defining the groove. However, it has high performance, and its practical value is great.

[Brief description of drawings]

FIG. 1 is an enlarged cross-sectional view showing an embodiment of the present invention.

FIG. 2 is a graph showing the relationship between groove depth, heat transfer performance, and pressure loss.

FIG. 3 is a graph showing a relationship between a groove twist angle and heat transfer performance.

FIG. 4 is a diagram schematically showing the flow of fluid flowing in a pipe.

FIG. 5 is a diagram schematically showing the relationship between groove dimensions and liquid film thickness.

FIG. 6 is a diagram showing the relationship between the groove size and the peak size.

FIG. 7 is a graph showing the relationship between the peak angle and heat transfer performance.

FIG. 8 is a graph showing the relationship between groove cross-sectional area, heat transfer performance, and unit weight.

FIG. 9 is a graph showing an example of trial calculation of groove cross-sectional area, heat transfer performance, unit weight reduction rate, and merits compared with conventional products.

10A and 10B are a horizontal cross-sectional view and a vertical cross-sectional view schematically showing an inner grooved tube.

FIG. 11 is an enlarged cross-sectional view of a conventional product and a diagram showing symbols for the dimensions of each part.

[Explanation of symbols]

 1 Internal grooved tube 2 Liquid film

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Makoto Hori 3550 Kidayomachi, Tsuchiura City, Ibaraki Prefecture Hitachi Cable Co., Ltd. Tsuchiura Plant

Claims (1)

Claim: What is claimed is: 1. A heat transfer tube in which a fluid flowing inside the tube undergoes a phase change, and a groove depth (Hf) is formed on the inner surface of the tube.
Hf / Di = 0.02 to 0.03 in ratio with the groove and the spiral angle with respect to the tube axis of the groove is 7 ° to 30 °, and has a plurality of trapezoidal grooves with a trapezoidal cross section. The ratio of the cross-sectional area (S) perpendicular to the axis to the groove depth (Hf) is S / Hf =
An inner grooved heat transfer tube, characterized in that the apex angle is 0.15 to 0.40, and the apex angle of the crests located between the grooves in a cross section perpendicular to the axis is 30 ° to 60 °.