JP4897968B2 - Heat transfer tube and method of manufacturing heat transfer tube - Google Patents

Description

  The present invention relates to a heat transfer tube used in a heat exchanger such as a refrigerator or an air conditioner and a method for manufacturing the same.

  In general, heat transfer tubes used in air conditioners, refrigerators, etc., perform heat exchange with the fluid flowing outside the tubes by evaporating or condensing the refrigerant in the tubes, which improves the efficiency and energy saving of the heat exchanger. From the viewpoint, the use of internally grooved tubes is increasing.

  In this internally grooved tube, a fine triangular or trapezoidal groove is formed on the inner surface of the tube linearly or spirally with respect to the tube axis. By providing these grooves on the inner surface of the tube, the heat transfer area is increased as compared with the smooth tube, and the heat transfer performance can be improved by the stirring action of stirring the refrigerant liquid.

In recent years, there has been a strong demand for higher performance and smaller size and weight especially for heat exchangers for air conditioners, and with the revision of the Energy Saving Law, there has been a further demand for higher performance of heat transfer tubes.
However, although the number of grooves, the lead angle, the groove shape, and the like have been improved in the conventional inner surface spiral grooved tube, the performance required as described above is insufficient.

  Therefore, as a heat transfer tube that replaces these conventional spiral finned tubes, for example, in Patent Document 1, in order to promote the stirring action of the refrigerant liquid, a main groove on the inner surface of the tube and a sub-groove having a depth for dividing the fins are provided. A heat transfer tube with a cross groove formed in is disclosed.

This heat transfer tube with a cross groove is provided with a plurality of three-dimensional protrusions (3) having a substantially S shape in a plan view obtained by dividing fins by sub-grooves on the inner surface of the tube.
More specifically, the three-dimensional protrusion (3) includes a burr (3a) protruding at the front end thereof so that the refrigerant flow along the main groove can be guided toward the sub-groove, and the burr (3a) at the rear end. It has a burr (3b) protruding in the opposite direction to 3a).

  According to Patent Document 1, the refrigerant flowing through the main groove is guided to the sub-groove direction by the burr (3a), so that the stirring action by the complicated flow of the refrigerant can be obtained, and as a result, the heat transfer coefficient is obtained. It is stated that it can be done.

  However, in the heat transfer tube in Patent Document 1, the refrigerant can be stirred in the vicinity of the inner peripheral surface side of the refrigerant by the burrs (3a) and (3b) provided in the three-dimensional projection (3). The refrigerant flowing in the center side was not agitated, and as a result, the transmission performance satisfying the above-mentioned requirements could not be obtained.

JP-A-8-178574

  Therefore, an object of the present invention is to provide a heat transfer tube capable of improving the heat transfer coefficient in the tube without increasing the pressure loss, and a manufacturing method thereof.

The present invention is a heat transfer tube in which spiral fins having a predetermined angle with respect to the tube axis direction are formed on the inner surface of the tube, and the fins are divided by sub-grooves and protrude in a spiral shape on the inner surface of the tube. The first projecting piece is formed with a fin configuration portion, and protrudes between the fin configuration portion and the adjacent fin on the upstream side in the tube axis direction, and at an acute angle with respect to the fin. And projecting between the fins adjacent to each other on the downstream side in the tube axis direction, and forming parallel to the first projecting pieces formed on the other fins by forming linearly from the first projecting pieces. And a second projecting piece which is formed in parallel with the first projecting piece while projecting between the second projecting piece and the fin adjacent to the upstream side in the tube axis direction on the upstream side in the spiral direction of the fin component. With protruding pieces Characterized in that was.

  As an aspect of the present invention, the sub-grooves can be formed in a number of 4 to 60 per pipe circumference.

  As an aspect of the present invention, the first projecting piece, the second projecting piece, and the third projecting piece are formed with a projecting length of 0.1 to 0.9 with respect to the interval between the fins. be able to.

As described above, according to the present invention, on the downstream side in the spiral direction of the fin component, the first protrusion formed between the adjacent fins on the upstream side in the tube axis direction and formed at an acute angle with respect to the fin. Projecting between one piece and the fins adjacent on the downstream side in the tube axis direction, and forming linearly from the first projecting piece, it is parallel to the first projecting piece formed on the other fin. And a second projecting piece formed on the upstream side in the spiral direction of the fin component and between the fins adjacent on the upstream side in the tube axis direction and parallel to the first projecting piece. It is the structure provided with 3 protrusion pieces.

  With this configuration, a part of the refrigerant flowing between the fins collides with the first projecting piece among the projecting pieces (the first projecting piece to the third projecting piece) and is scooped up in the tube radial direction. A three-dimensional unsteady flow can be generated effectively.

  Therefore, as compared with a conventional heat transfer tube in which a cross groove is formed on the inner peripheral surface of the tube, it is possible to promote turbulent flow including the inner side in the tube radial direction, and an excellent heat transfer coefficient can be obtained.

  Furthermore, among the projecting pieces (the first projecting piece to the third projecting piece), in particular, the third projecting piece makes it easier for the refrigerant flowing between the fins to flow into the sub-groove, so that a secondary flow is generated. In particular, the refrigerant stirring effect can be obtained in the vicinity of the inner peripheral surface of the pipe, and an increase in pressure loss can be prevented.

Further, according to the present invention, as described above, the first projecting piece is formed at an acute angle with respect to the fin, and the second projecting piece and the third projecting piece are formed in parallel to the first projecting piece. The refrigerant flowing between the fins easily flows between the plurality of fin constituent portions in the fin forming direction, and the heat transfer coefficient can be improved by a more excellent diffusion action.
Furthermore, since the refrigerant flow rate in the tube axis direction increases, an increase in pressure loss can be prevented.

Furthermore, if the protruding pieces (first to third protruding pieces) are formed at an acute angle with respect to the fin, the protruding pieces can be easily processed.
Further, in the present invention, by forming the sub-grooves with a number of 4 or more per pipe circumference, the interval between the fin constituent portions in the fin forming direction does not become too large, and a sufficient refrigerant stirring action can be obtained. .

  On the other hand, according to the present invention, as described above, by forming the sub-grooves with a number of 60 or less per pipe circumference, it is possible to secure a fin strength that can withstand mechanical expansion.

In particular, the sub-grooves are more preferably formed with a number of 8 to 30 per pipe circumference.
By forming the sub-grooves in such a number within the range of 8 to 30 per pipe circumference, the heat transfer performance does not change greatly, and a preferable balance between the refrigerant stirring action and the fin strength can be obtained.

  In the present invention, as described above, the projecting pieces (first projecting piece to third projecting piece) are formed with a projecting length having a ratio of 0.1 or more with respect to the interval between the fins. As a result, the refrigerant flowing between the fins can easily flow between the fin components in the fin forming direction, and the refrigerant can be promoted to the inside in the radial direction, so that a large refrigerant stirring action can be obtained. The heat transfer rate can be improved.

  Meanwhile, as described above, the protruding pieces (first to third protruding pieces) are formed with a protruding length having a ratio of 0.9 or less with respect to the interval between the fins. The protruding piece (first protruding piece to third protruding piece) does not protrude too much, and the flow of the refrigerant flowing between the fins is not excessively inhibited by the protruding piece (first protruding piece to third protruding piece). Can be prevented from increasing.

  According to the present invention, it is possible to provide a heat transfer tube capable of improving the heat transfer coefficient in the tube without increasing the pressure loss, and a manufacturing method thereof.

An embodiment of the present invention will be described below with reference to the drawings.
As shown in FIGS. 1 to 3, the heat transfer tube 11 in the present embodiment has spiral fins 12 formed on the tube inner surface 10.

  Note that the heat transfer tube 11 in the present embodiment includes the main grooves 13 between the fins 12 by forming the spiral fins 12 on the tube inner surface 10.

  Further, the heat transfer tube 11 forms a plurality of fin constituent portions 12A by dividing the fin 12 in the spiral direction D2 (fin forming direction).

  In addition, the heat transfer tube 11 in this embodiment forms a sub-groove that divides the fin 12 by the sub-groove 14 between the fin configuration portions 12A in the spiral direction D2 by forming the fin configuration portion 12A on the tube inner surface 10. A portion 15 is provided.

  FIG. 1 is a partially expanded perspective view schematically showing a state of the tube inner surface 10 of the heat transfer tube 11 of the present embodiment, and FIG. 2 is a longitudinal sectional view of the heat transfer tube 11 of the present embodiment. FIG. 3 is an enlarged plan view of the vicinity of the fin component 12A. In FIGS. 2 and 3, the ridgeline between the fin component 12 </ b> A and the pipe inner surface 10 is schematically shown.

  Further, the fin component 12A includes a protruding piece 16 (first protruding piece 16a) protruding into the main groove 13 on the upstream side D1u in the tube axis direction on the downstream side D2d in the spiral direction.

  The first projecting piece 16a is formed at an angle (α) of 12 ° which is within a range of 5 ° to 90 ° with respect to the main groove 13.

  The sub-grooves 14 are formed in the spiral fins 12 with a number of 32 stripes within a range of 4 to 60 stripes per pipe circumference.

  The first projecting piece 16a has a ratio (W1 / W2) of 0.3 within a range of 0.1 to 0.9 with respect to the width (W2) of the main groove 13 to the main groove 13. The protrusion length (W1) is used.

More specifically, the heat transfer tube 11 of the present embodiment has an outer diameter of 7.0 mm, and the spiral fins 12 are formed at an angle (β) of 30 ° with respect to the tube axis direction D1. Further, the number of fins per pipe circumference is 48, the fin height (main groove depth) is 0.25 mm, and the cross-sectional shape of the fin 12 is formed in a substantially trapezoidal shape.
The sub-groove 14 has a depth of 0.25 mm and a groove shape with an inverted triangular cross section having an inner angle of 30 degrees.

  Further, the fin component 12A includes a protruding piece 16 (second protruding piece 16b) protruding into the main groove 13 on the downstream side D1d in the tube axis direction on the downstream side D2d in the spiral direction.

  On the other hand, the fin constituting portion 12A is also provided on the upstream side D2u in the spiral direction with the projecting piece 16 (third projecting piece 16c) projecting into the main groove 13 on the upstream side D1u in the tube axis direction and on the downstream side D1d in the tube axis direction. A protruding piece 16 (fourth protruding piece 16 d) protruding in the main groove 13 is provided.

  Therefore, as shown in FIG. 3, the fin component portion 12 </ b> A is formed in a shape in which the letter H is inclined (inclined H shape) when seen in a plan view.

Then, the manufacturing apparatus 17 of the heat exchanger tube 11 mentioned above is demonstrated with FIG. 4, FIG.
FIG. 4 is a schematic view showing a cross section of the manufacturing apparatus 17 for the heat transfer tube 11 in the present embodiment, and FIG. 5 is a detailed view of the notched plug.

  On the upstream side in the drawing direction X (left side in FIG. 4), a first reduced diameter die 23 and a floating plug 24 are provided.

A first connecting rod 25 connected to the floating plug 24 on the downstream side of the floating plug 24 disposed in the pipe is rotatably connected and held in the pipe, and has a plurality of spiral shapes on the outer peripheral surface. A grooved plug 26 having a groove 26a is provided.
The grooved plug 26 is configured with a lead angle of 40 ° in the right twist direction with respect to the tube axis direction D1.

  A plurality of rolling balls 27 that revolve while pressing the copper tube 11a are provided at positions facing the grooved plug 26 on the outside of the tube.

  Further, on the downstream side of the grooved plug 26, a notch plug 29 is rotatably connected and held by a second connecting rod 28 connected to the grooved plug 26 on the downstream side.

As shown in FIGS. 5A and 5B, the notch plug 29 includes a plurality of convex notch blades 29 a on the outer peripheral surface.
The notch blade 29a has a lead angle of 25 ° in the right twist direction with respect to the tube axis direction D1. Furthermore, as shown in the enlarged view of the main part of FIG. 5A, the notch blade 29a is formed with a height of 0.25 mm and an apex angle of 30 °. The top is configured to be flat.

  Further, a second diameter reducing die 33 for further reducing the diameter of the copper pipe 11a is provided at a position facing the notch plug 29 on the outer side of the pipe.

  Here, the diameter reduction of the copper tube 11a is not limited to the use of the second diameter reduction die 33, and may be performed by rolling with a ball, for example. By reducing the diameter of the tube 11a, the cross section of the copper tube 11a can be kept circular, which is preferable.

The heat transfer tube 11 of this embodiment is manufactured by the following manufacturing method using the manufacturing apparatus 17 described above.
First, the diameter of the copper tube 11a is reduced by the first diameter reducing die 23 and the floating plug 24. Then, the copper tube 11a is further reduced in diameter by the grooved plug 26 and the rolling ball 27, and the spiral fin 12 having a predetermined lead angle with respect to the tube axis direction D1 is formed in the tube, A fin processing step for forming the main groove 13 between the fins 12 is performed.

  Subsequently, the notch plug 29 and the second reduced diameter die 33 form a plurality of the fin constituent portions 12A obtained by dividing the fin 12, and at least on the downstream side D2d in the spiral direction of the fin constituent portions 12A, the tube axis A protruding piece machining step for forming a first protruding piece 16a protruding in the main groove 13 on the upstream side D1u in the direction is performed.

  As shown in FIGS. 1 to 3, the above-described manufacturing method makes it possible to manufacture a heat transfer tube 11 having a plurality of divided fin components 12 </ b> A having an inclined H shape when viewed in plan.

The heat transfer tube 11 described above can obtain various actions and effects as follows.
As described above, the heat transfer tube 11 of the present embodiment includes the first projecting piece 16a projecting into the main groove 13 on the upstream side D1u in the tube axis direction at least on the downstream side D2d in the spiral direction of the fin component 12A. ing.

  For this reason, a part of the refrigerant flowing through the main groove 13 collides with the first projecting piece 16a and is scooped up in the pipe radial direction, so that a three-dimensional unsteady flow can be generated.

  Therefore, since the effect of promoting turbulent flow including the inside in the radial direction of the tube can be obtained, a conventional heat transfer tube in which a cross groove is formed on the inner surface of the tube, or a conventional heat transfer tube having a burr in the vicinity of the sub groove, In comparison, the heat transfer rate can be further improved.

  Further, as described above, the spiral fin 12 has, in particular, a third projecting piece 16c projecting toward the main groove 13 on the upstream side D1u in the tube axis direction at least on the upstream side D2u in the spiral direction of the fin component 12A. As a result, the refrigerant flowing through the main groove 13 is more easily flown into the sub-groove 14, so that a secondary flow is generated, and a stirring effect of the refrigerant can be obtained particularly in the vicinity of the inner peripheral surface of the pipe. Increase of pressure loss can be prevented.

  Further, as described above, the spiral fin 12 is formed at 30 ° which is an angle (β) of 25 ° to 50 ° with respect to the tube axis.

  As described above, by forming the fin 12 with a lead angle of 25 ° or more with respect to the tube axis, sufficient heat transfer performance can be obtained, while by forming the fin 12 with a lead angle of 50 ° or less, the fin 12 Fin processing can be easily performed in the processing step.

Moreover, the heat exchanger tube 11 of this embodiment forms the said protrusion piece 16 at the angle ((alpha)) of 12 degrees which is the range of 5 degrees-90 degrees with respect to the said main groove 13 as mentioned above.
For this reason, the flow of the refrigerant flowing through the main groove 13 is not excessively hindered, and a part of the refrigerant collides with the protruding piece 16 so that the main groove 13 can flow into the sub groove 14. The heat transfer coefficient can be improved by the refrigerant stirring action.

  Furthermore, since the refrigerant flow rate in the tube axis direction D1 increases, an increase in pressure loss can be prevented.

  In particular, as in the heat transfer tube 11 of the present embodiment, the projecting piece 16 is formed in an angle range of 5 ° to 20 ° with respect to the main groove 13, thereby further improving the heat transfer coefficient and pressure. It is possible to prevent an increase in loss.

  Further, in the heat transfer tube 11 of the present embodiment, as described above, the sub-groove 14 is formed in the spiral fin 12 with the number of strips of 32 per tube circumference, so that the notch pitch by the sub-groove 14 is increased. Since it becomes an appropriate size, it is possible to obtain a sufficient refrigerant stirring action and to obtain a fin strength that can withstand mechanical expansion.

  In addition, as described above, the heat transfer tube 11 of the present embodiment is configured so that the protruding piece 16 is moved to the main groove 13 having a ratio (W1 / W2) of 0.3 with respect to the width (W2) of the main groove 13. By forming with a protruding length (W1) of this, the scooping up of the refrigerant radially inward is promoted, an excellent refrigerant stirring action can be obtained, and as a result, the heat transfer coefficient can be improved. .

  Moreover, since the length of the refrigerant flowing in the main groove 13 is such that the flow of the refrigerant does not become too hindered, the refrigerant flowing in the main groove 13 is allowed to flow into the sub-groove 14 while preventing an increase in pressure loss. Can do.

  Moreover, it is preferable to comprise the said subgroove 14 by the depth of 30 to 100% of the depth of the main groove 13, like the heat exchanger tube 11 of this embodiment.

  By configuring the sub-groove 14 at a depth of 30% or more of the depth of the main groove 13, the refrigerant flowing through the main groove 13 can easily flow into the sub-groove 14, and an excellent heat transfer rate can be obtained. Increase of pressure loss can be prevented.

  On the other hand, by forming the sub-groove 14 at a depth of 100% or less of the depth of the main groove 13, the tube thickness does not form a notch deeper than the main groove depth. (Thickness specification) can be maintained.

  The sub-groove 14 is particularly preferably configured to have a depth of 70 to 80% of the depth of the main groove 13 in view of the balance between ease of processing and performance when the cut-out plug 29 is formed.

  Moreover, in the manufacturing method of the heat exchanger tube 11 mentioned above, the heat exchanger tube 11 provided with several fin structure part 12A which made the planar view inclination type H shape in the pipe inner surface 10 easily by performing the said protrusion piece process process especially. Can be manufactured.

  According to the manufacturing method of the heat transfer tube 11 of the present embodiment, when the sub-groove 14 is formed in the fin 12, the sub-groove 14 is formed by plastic deformation on the tube inner surface 10. There is no pollution.

  Furthermore, in the manufacturing apparatus 17 according to this embodiment, the second reduced diameter die 33 presses the notched plug 29, so that the cross section of the copper tube 11a can be kept circular.

  The notch plug 29 has a flat top portion of the notch blade 29a. By pressing such a notch blade 29a against the fin 12, the protruding piece 16 protruding toward the main groove 13 can be further formed.

  As described above, the heat transfer tube 11 according to an embodiment of the present invention has been described in detail. Subsequently, an experiment conducted for verifying the performance of the heat transfer tube of the present invention will be described.

In this experiment, five types of Examples 1 to 5 were manufactured as heat transfer tubes of the present invention, and a heat transfer tube used as a comparison target was manufactured as Comparative Example 1.
The heat transfer tubes of Examples 1 to 5 are each manufactured by the same manufacturing method as the heat transfer tube 11 of the above-described embodiment into a shape having an outer diameter, fins, sub grooves, and protruding pieces as shown in Table 1. Yes.

ここで、実施例４の伝熱管は、表１に示した各部の形状からも明らかなとおり、前述した実施形態の伝熱管１１と同じ形状の伝熱管を用いている。

Here, as is clear from the shape of each part shown in Table 1, the heat transfer tube of Example 4 uses a heat transfer tube having the same shape as the heat transfer tube 11 of the above-described embodiment.

  Further, each of the heat transfer tubes of Examples 1 to 4 is formed by including a plurality of divided fin constituent portions 12A each having an inclined H shape in plan view. On the other hand, as shown in FIGS. 6 and 7, the heat transfer tube 21 of Example 5 includes a plurality of fin constituent portions 42 </ b> A having a J-shaped shape (J shape in plan view) when viewed in plan. Forming.

More specifically, in the heat transfer tube 21 of the fifth embodiment, the divided fin structure portion 42A having a J shape in plan view includes the first protrusion piece 16a in the fin structure portion 42A and the second protrusion piece 16b. And a third protrusion piece 16c.
Here, FIG. 6 is a partially expanded perspective view schematically showing the state of the tube inner surface 10 of the heat transfer tube 21 of the fifth embodiment, and FIG. 7 is a view of the tube inner surface 10 of the heat transfer tube 21 of the fifth embodiment. It is an enlarged plan view near the divided fins. In FIG. 7, the ridgeline between the fin component 42A and the pipe inner surface 10 is schematically shown.

  The heat transfer tube of Comparative Example 1 is a conventional heat transfer tube, and fins are formed on the inner surface of the tube as shown in Table 1, but the sub-groove (fin forming portion) is not formed in the fin. It is.

In addition, the heat transfer tubes of Examples 1 to 5 and Comparative Example 1 were manufactured using a grooved plug having a shape as shown in Table 2 and a notched plug having a shape as shown in Table 3, respectively. Manufactured.
The outer diameter of the notched plug in Table 3 indicates the outer diameter including the height of the blade provided on the outer peripheral portion, and indicates the linear distance connecting the tops of the blades passing through the center of the notched plug. And

すなわち、表２に示すように、溝付プラグは、いずれの実施例においても同じ形態のものを用い、また、表３に示すように、切欠きプラグは、適宜、異なる形態のものに変更して伝熱管を製作した。

That is, as shown in Table 2, the same type of grooved plug is used in all the examples, and as shown in Table 3, the notched plug is appropriately changed to a different type. To make a heat transfer tube.

In this experiment, a condensation experiment for verifying the condensation performance in the tube of the heat transfer tube is performed using the in-tube condensation performance measuring apparatus 50A as shown in FIG. 8A, and an evaporation experiment for verifying the evaporation performance is performed in FIG. This was performed using an in-tube evaporation performance measuring apparatus 50B as shown in FIG.
8A and 8B are schematic views of the in-pipe condensing performance measuring device 50A and the in-pipe evaporating performance measuring device 50B, respectively. In any of the devices 50A and 50B, the whole is similar to a general air conditioner. Is constituted by a refrigeration cycle.

Specifically, in the condensation experiment, each of the heat transfer tubes of Examples 1 to 5 and the heat transfer tube of Comparative Example 1 was incorporated as a test tube 44 in the condenser as shown in FIG. In this case, each of the heat transfer tubes of Examples 1 to 5 was condensed by measuring the ratio of the heat transfer coefficient (heat transfer coefficient ratio) and the ratio of pressure loss (pressure loss ratio) to the heat transfer pipe of Comparative Example 1. The performance was verified.
In the evaporation experiment, each of the heat transfer tubes of Examples 1 to 5 and the heat transfer tube of Comparative Example 1 was incorporated as a test tube 44 in the evaporator as shown in FIG. In this case, by measuring the ratio of heat transfer coefficient (heat transfer coefficient ratio) and the ratio of pressure loss (pressure loss ratio) of the heat transfer pipes of Examples 1 to 5 to the heat transfer pipe of Comparative Example 1, evaporation The performance was verified.

As shown in FIGS. 8A and 8B, the test sections in the in-tube condensing performance measuring device 50A and the in-tube evaporating performance measuring device 50B are composed of a double-tube heat exchanger, and are provided in the test tube 44. An effective length of the test tube 44 is obtained by flowing heat exchange water (hereinafter referred to as “heat exchange water”) in a direction opposite to the refrigerant flow in the annular portion 45 constituting the outer shell. Heat was exchanged with the thickness set to 4 m.
Note that low-temperature water is flown as heat exchange water in the condensation experiment, and high-temperature water is flown as heat exchange water in the evaporation experiment.

  Further, as shown in FIGS. 8A and 8B, a thermometer, a pressure gauge, and a flow meter are arranged at each predetermined portion of the test section. 8A and 8B, T is a thermometer, P is a pressure gauge, and G is a flow meter.

  Subsequently, as the experimental conditions at the refrigerant inlet and outlet of the test tube 44, in the condensation experiment, the refrigerant inlet superheat degree, the refrigerant outlet supercooling degree, and in the evaporation experiment, the refrigerant inlet dryness, the refrigerant outlet superheat degree, respectively. Settings were made as shown in Table 4.

これら凝縮実験、蒸発実験における実験条件は、いずれも空調機の熱交換器入口条件と同一となるように、水温を調節した後に測定を行った。
さらまた、供試管４４の入口と出口における冷媒平均飽和温度は、表４に示すように凝縮実験では４８℃に設定するとともに、蒸発実験では５℃に設定した。

Measurement was performed after adjusting the water temperature so that the experimental conditions in these condensation experiments and evaporation experiments were the same as the heat exchanger inlet conditions of the air conditioner.
Furthermore, as shown in Table 4, the average refrigerant saturation temperature at the inlet and outlet of the test tube 44 was set to 48 ° C. in the condensation experiment and 5 ° C. in the evaporation experiment.

As the refrigerant, R410A is used as an alternative chlorofluorocarbon, and since this R410A is a mixed refrigerant, the refrigerant is collected at the refrigerant sampling section (see FIGS. 8A and 8B) installed at the compressor outlet during the experiment. The sample was collected and experimented while measuring the refrigerant composition ratio by gas chromatography.
The analysis result of gas chromatography is reflected in ts1 and ts2 described later by calculation.

The pressure loss ratio and the heat transfer coefficient ratio αi in the test tube 44 showing the condensation performance and the evaporation performance are obtained as follows.
First, the pressure loss ratio in the tube is obtained as the pressure difference between the inlet and outlet of the test tube 44.
The heat transfer coefficient ratio αi in the tube is calculated using the equations (1) to (4) based on the measured values in this experiment.

ここで、数式（１）中のＱは、交換熱量（ｋＷ）、Ａは、供試管外表面積（ｍ２）、ｔｍは、対数平均温度（℃）、αｏは、管外熱伝達率（ｋＷ／ｍ２Ｋ）を示す。

Here, in Equation (1), Q is the amount of exchange heat (kW), A is the external surface area (m2) of the test tube, tm is the logarithmic average temperature (° C.), αo is the external heat transfer coefficient (kW / m2K).

  In Equation (2), G is a flow rate of heat exchange water (kg / s), Cp is a specific heat (kJ / kgK) of heat exchange water, tw1 is an inlet temperature (° C.) of heat exchange water, and tw2 is Shows the outlet temperature (° C) of water for heat exchange.

  In Formula (3), ts1 indicates the refrigerant inlet saturation temperature (° C.), and ts2 indicates the refrigerant outlet saturation temperature (° C.).

  In Equation (4), k represents the thermal conductivity (kw / mK) of heat exchange water, and De represents the annular portion equivalent diameter (m). D represents the shell inner diameter (m), d represents the test tube outer diameter (m), Re represents the Reynolds number (−) of the heat exchange water, and Pr represents the Prandtl number (−) of the heat exchange water.

  That is, based on measured values such as temperature and setting parameters, Q is calculated from Formula (2), tm at the time of condensation and evaporation is calculated from Formula (3), and αo is calculated from Formula (4). By substituting in (1), the heat transfer coefficient ratio αi can be calculated.

  Table 5 shows the evaluation results of the condensation performance and evaporation performance thus obtained.

表５より明らかなように、熱伝達率に関しては、実施例１から５のいずれの伝熱管においても、凝縮性能、蒸発性能ともに、比較例１の伝熱管より高い熱伝達性能を示した。

As is clear from Table 5, regarding the heat transfer coefficient, in any of the heat transfer tubes of Examples 1 to 5, both the condensation performance and the evaporation performance showed higher heat transfer performance than the heat transfer tube of Comparative Example 1.

  Thereby, a part of the refrigerant flowing through the main groove collides with the first projecting piece and is scooped up in the pipe radial direction, so that significant turbulence including the inner side in the pipe radial direction can be promoted. . Further, it can be demonstrated that a part of the refrigerant flowing through the main groove flows into the sub-groove, so that the refrigerant diffuses over the entire inner peripheral surface of the pipe, and as a result, the heat transfer coefficient can be improved. It was.

  In addition, regarding the condensation performance, the heat transfer tubes of Examples 1, 2, and 5 having a large angle (α) formed by the main groove and the sub groove with respect to the fins have high performance. The heat transfer tubes of Examples 3 and 4 having a small angle (α) made with

  In particular, in the heat transfer tubes of Examples 1 and 2, as shown in Table 3, the lead angle of the notch plug is small (0 °), and the secondary groove is substantially parallel to the tube axis. For this reason, the fact that condensed refrigerant quickly flows in the direction of the tube axis and is eliminated is also considered as a factor for improving the condensation performance.

  Conversely, in the heat transfer tubes of Examples 3 and 4, the lead angle of the notched plug is larger (25 °) than the lead angle (0 °) of the heat transfer tubes of Examples 1 and 2, so that the refrigerant is long in the tube. It is thought that the evaporation performance was improved as a result of holding the time and repeating the evaporation.

  Further, as is apparent from Table 5, regarding the pressure loss, the heat transfer tubes of Examples 3 and 4 were less than or equal to the heat transfer tube of Comparative Example 1.

  Accordingly, the pressure loss is reduced by setting the angle (α) formed by the projecting piece with respect to the fin and the length of the projecting length of the projecting piece to the main groove side (W1 / W2) with respect to the main groove width. It was confirmed that high performance could be achieved without increasing.

  In particular, the heat transfer tube of Example 5 has a larger angle (α) formed by the projecting pieces with respect to the fins (44 °) than the heat transfer tubes of the other examples and the comparative example, so that the pressure loss increases. (123.6).

  From this, it has been confirmed that the angle (α) formed by the protruding piece with respect to the fin is preferably approximately 30 ° or less in order to suppress an increase in pressure loss.

  On the other hand, the heat transfer tubes of Examples 1, 2, and 5 both improved the total condensation performance and evaporation performance including the heat transfer coefficient ratio as compared with the heat transfer tubes of Comparative Example 1, but in particular, Example 2 The pressure loss of the heat transfer tube increased by 10% or more during both condensation and evaporation.

  As shown in Table 3, the top angle of the notch plug in the heat transfer tube of Example 2 is only 30 °, whereas the top angle of the notch plug in the heat transfer tube other than Example 2 is 30 °. A factor of 60 ° is considered.

  That is, by manufacturing the heat transfer tube of Example 2 using a notch plug having a large apex angle (60 °), the protrusion length (W1) of the protrusion piece to the main groove side with respect to the main groove width (W2) is reduced. A possible reason is that the ratio (W1 / W2) is as large as 0.63.

  Further, in the heat transfer tube of Example 1, the ratio (W1 / W2) of the protruding length (W1) of the protruding piece to the main groove side with respect to the main groove width (W2) is the ratio (W1 / W2) in Examples 3 and 4, respectively. It is formed with a larger value (0.50) than W2).

  From the above, the ratio (W1 / W2) of the protruding length (W1) of the protruding piece to the main groove side with respect to the main groove width (W2) is preferably smaller than 0.5 in order to suppress an increase in pressure loss. Was confirmed.

  Further, as is apparent from the results in Table 5, the heat transfer coefficient of the heat transfer tube of Example 4 was higher than that of the heat transfer tube of Example 3 in both the condensation experiment and the evaporation experiment.

  This is because the heat transfer tube of Example 3 has a larger number of notches, that is, the number of sub-grooves than the number of sub-grooves (8) of the heat transfer tube of Example 4 compared to the heat transfer tube of Example 4 ( 32) However, since all the other forms are the same, the heat transfer tube of Example 3 includes the inner side in the tube radial direction because the number of sub-grooves is larger than that of the heat transfer tube of Example 4. It is considered that the turbulent flow effect in the entire pipe could be achieved.

  In addition, the present invention is not limited to the heat transfer tubes of Embodiments 1 to 5 described above or the heat transfer tube manufacturing method described above, and can be configured in various forms and modes.

  For example, the heat transfer tube of the present invention includes at least a first protruding piece on the fin constituent portions 12A and 42A, like the heat transfer tube 11 of the above-described embodiment (Example 4) or the heat transfer tube 21 of Example 5. 16a can be formed.

  As described above, by providing the first projecting piece 16a, a part of the refrigerant flowing through the main groove 13 collides with the first projecting piece 16a and is scooped up in the pipe radial direction, so that a three-dimensional unsteady flow is achieved. This is because the heat transfer rate can be improved by further promoting turbulence.

Specifically, for example, as shown in FIG. 9, the heat transfer tube of the present invention includes a plurality of fin components 52 </ b> A having a Z-shaped shape (planar view inclined Z shape) when viewed in plan. Alternatively, the heat transfer tube 31 can be formed.
FIG. 9 is an enlarged plan view of the inner surface of the tube having the fin configuration portion 52A having a Z shape in plan view. However, in FIG. 9, the ridgeline between the fin component 52A and the inner surface of the tube is schematically shown.

  The divided fin structure 52A having a Z shape in plan view is a form in which the fin structure 52A includes the first protrusion 16a and the fourth protrusion 16d.

  Furthermore, even if the heat transfer tube 31 flows the refrigerant with the upstream side and the downstream side in the tube axis direction D1 opposite to those shown in FIG. 9, the fourth projecting piece 16d remains in the spiral direction of the fin component 52A. In the downstream side D2d, the main groove 13 on the upstream side D1u in the tube axis direction protrudes (see the arrow in parentheses in FIG. 9).

  That is, in the heat transfer tube 31 having the above-described configuration, the first projecting piece 16a and the fourth projecting member are not affected even if the refrigerant flows through the tube to the downstream side with either one of the tube axis direction D1 and the other side as the upstream side. Either one of the pieces 16d necessarily protrudes to the main groove 13 on the upstream side D1u in the pipe axis direction D2d on the downstream side D2d in the spiral direction of the fin component 52A.

Therefore, the heat transfer tube having the above configuration can ensure the above-described excellent performance regardless of the mounting direction to the heat exchanger, and can be easily attached to the heat exchanger without being aware of the mounting direction. Can.
The present invention can be formed by various configurations as in the above-described embodiment, but is not limited to the above-described configuration, and many embodiments can be obtained.
In the correspondence between the above-described embodiment and the configuration of the present invention, the notch blade 29a of this embodiment corresponds to the blade of the present invention, and the main groove 13 of this embodiment is between the fins of the present invention. It shall correspond.

The partial expansion expansion perspective view which shows the heat exchanger tube of this embodiment. The longitudinal cross-sectional view which shows the heat exchanger tube of this embodiment. The enlarged plan view of the pipe inner surface which has the fin structure part in the heat exchanger tube of this embodiment. The schematic diagram which shows the manufacturing apparatus of the heat exchanger tube in this embodiment. Schematic of the notch plug used for manufacture of the heat exchanger tube of this embodiment. The partial expansion expansion perspective view which shows an example of this invention heat exchanger tube. The enlarged plan view of the pipe inner surface which has a fin structure part which shows an example of the heat exchanger tube of this invention. Schematic of the experimental apparatus used for the performance evaluation of the heat exchanger tube of this embodiment. The enlarged plan view of the pipe inner surface which has a fin structure part which shows an example of the heat exchanger tube of this invention.

11, 21, 31 ... Heat transfer tube 10 ... Tube inner surface 12, 42, 52 ... Fin 12A, 42A, 52A ... Fin component 14 ... Sub groove 16 ... Projection piece 16a ... First projection piece 26 ... Slotted plug 29 ... Cutting Notched plug 33 ... second reduced diameter die D1 ... tube axis direction D2 ... spiral direction

Claims (3)

A heat transfer tube in which a spiral fin having a predetermined angle with respect to the tube axis direction is formed on the tube inner surface,
The fin is divided by a sub-groove and is formed by a plurality of fin components that spirally project on the inner surface of the tube,
On the downstream side of the fin component in the spiral direction,
A first protruding piece that protrudes between the fins adjacent on the upstream side in the tube axis direction and is formed at an acute angle with respect to the fin;
Projecting between the adjacent fins on the downstream side in the tube axis direction and forming a straight line from the first projecting piece makes it parallel to the first projecting piece formed on the other fin . 2 protruding pieces,
On the upstream side of the fin component in the spiral direction,
A heat transfer tube including a third protruding piece that protrudes between the fins adjacent to each other on the upstream side in the tube axis direction and is formed in parallel to the first protruding piece. The minor groove,
The heat transfer tube according to claim 1, wherein the number is 4 to 60 per tube circumference. The said 1st protrusion piece, the said 2nd protrusion piece, and the said 3rd protrusion piece were formed in the protrusion length which consists of the ratio of 0.1-0.9 with respect to the space | interval between the said fins. The heat transfer tube described.

Images