JP5435460B2 - 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.

  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, smaller size, and lighter weight especially for heat exchangers for air conditioners, and with the revision of the Energy Conservation Law (Act on the Rational Use of Energy) There is more demand.
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 plan view in which fins are divided 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

  An object of the present invention is to provide a heat transfer tube capable of improving the heat transfer coefficient in the tube while suppressing an increase in pressure loss.

In the present invention, a helical fin having a height (H _f ) having a predetermined angle (β ₁ ) with respect to the tube axis direction is formed on the inner surface of the tube, and the outer diameter is smaller than 7 mm. A heat transfer tube for evaporation , wherein the fin is divided on the inner surface of the tube in the direction of fin formation by a spiral sub-groove having a depth (H _n ) of 0.1 mm to 0.30 mm. A plurality of fin constituent portions that project in a spiral shape are formed, and a projecting piece that projects between adjacent fins is formed on at least one of the upstream side and the downstream side in the spiral direction of the fin constituent portion. and, wherein the protruding piece is a spiral direction downstream side of the fin structure portion, the first projecting pieces out collision between the adjacent fins in the tube axial direction upstream side tube axis of the fin structure portion Adjacent on the downstream side in the direction opposite to the side having the first protruding piece A second projecting piece projecting between the fins and a third projecting piece projecting between the fins adjacent to the fins adjacent to the first projecting piece on the upstream side in the spiral direction the spiral direction upstream side of the fin structure portion, and a fourth projecting piece projecting between the adjacent fins on the opposite side to the side having the first projecting piece, before Symbol protruding pieces, between the fins the proportion of the interval _{(W 2)} projecting length to _{(W 1)} to _(W 1 / W ₂₎ is formed with 0.3 to 0.9, the fin forming portion, between the fins constituting portion in the fin-forming direction The evaporating heat transfer tube is characterized in that the interval (P _f ) is 5 mm or more and the angle (α) of the protruding piece with respect to the fin is set to 5 ° ≦ α ≦ 20 ° .

  In the claims and the specification of the present invention, the range described using the symbol “to” includes the numerical value described before the symbol and the numerical value described after the symbol.

As described above, the heat transfer tube has a fin height (H _f ) in the range of 0.1 mm to 0.30 mm, particularly at least on the downstream side in the spiral direction of the fin component having a thickness of 0.1 mm or more. Between the fins adjacent on the upstream side, a protruding piece protruding at an angle of 5 ° or more with respect to the fin forming direction is provided.

With such a configuration, a part of the refrigerant flowing between the fins collides with the protruding piece and is scooped up inward in the pipe radial direction, so that a three-dimensional unsteady flow can be effectively generated. In addition, since the refrigerant stays due to the collision with the projecting piece and the thickness of the liquid film of the refrigerant flowing in the main groove becomes non-uniform in the fin forming direction, the heat transfer performance is reduced in the thin part of the liquid film. improves.
Therefore, compared to a conventional heat transfer tube with a cross groove 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 to reduce the thickness of the liquid refrigerant, resulting in an excellent heat transfer coefficient. Can be obtained.

Further, even when the sub-groove depth (H _n ) is in the range of 0.1 mm to 0.30 mm, the angle (α) formed by the protruding piece and the main groove is less than 70 °, particularly 0.1 mm or more. Since the refrigerant flowing between the fins easily flows into the sub-groove, the refrigerant can be stirred on the entire inner peripheral surface of the pipe across the fins. Therefore, an excellent turbulent flow promoting effect can be obtained and an increase in pressure loss can be prevented.

Even if the fin height (H _f ) is in the range of 0.1 mm to 0.30 mm, the inner diameter of the small-diameter tube is not too small, and an increase in pressure loss can be suppressed if the height is particularly 0.25 mm or less. .

Further, by forming the interval (P _f ) between the fin components in the fin forming direction to be 1.5 mm or more, the refrigerant is stirred by a sufficient turning force without excessively obstructing the flow of the refrigerant along the fin. The effect can be obtained. Moreover, the fin strength which can be endured at the time of machine expansion is ensured.

Further, even when the ratio (W ₁ / W ₂ ) of the protrusion length (W ₁ ) is within the range of 0.3 to 0.9, it is particularly preferable that W ₁ / W ₂ = 0.4 to 0.9. . Specifically, in a small-diameter tube having an outer diameter of less than 7 mm, the fin height (H _f ) is relatively low due to the above-described pressure loss problem and processing limitations, so that the coolant is lifted up by the protruding piece or into the sub-groove. In order to obtain a sufficient induction effect, it is necessary to set the ratio (W ₁ / W ₂ ) of the protruding length (W ₁ ) between the fins of the protruding pieces to 0.4 or more.

Further, by setting the ratio to 0.9 or less, the supply of the liquid film between the fins is not hindered at the time of evaporation, so that it is possible to prevent a decrease in evaporation heat transfer coefficient due to dryout. In addition, regarding the increase or decrease of pressure loss in a small diameter tube having an outer diameter of less than 7 mm, the fin height (H _f ), the lead angle (β ₂ ), and the interval (P _f ) between fin components are dominant, and the protrusion length Since the influence of the thickness (W ₁ ) is relatively small, the above-described heat transfer promoting effect can be obtained without excessively increasing the pressure loss even with the above-described configuration .

Also, steam Hatsuyo heat exchanger tube of the present invention, as described above, the fin height _{(H f)} of 0.1mm or more, the sub groove depth _{(H n)} of 0.1mm or more, the projecting piece The angle (α) with respect to the fin is 20 ° or less, the lead angle (β ₂ ) of the _secondary groove with respect to the tube axis direction is 10 ° or more, and the interval (P _f ) between the fin components in the fin forming direction is 5 mm or more. be able to.

The evaporation heat transfer tube has a configuration suitable for an evaporation tube having a fin height (H _f ) of 0.10 mm or more. That is, the effective heat transfer area on the inner surface of the tube is increased, and the effect of thinning the liquid film in the vicinity of the fin top is sufficiently obtained to improve the evaporation performance.

In addition, if the sub-groove depth (H _n ) is 0.10 mm or more and the angle (α) between the protruding piece and the main groove is 20 ° or less, the refrigerant flowing between the fins easily flows into the sub-groove. The refrigerant spreads across the fins, and an evaporation promoting effect due to an increase in wetted area can be obtained. Moreover, since the refrigerant | coolant flow rate to a pipe-axis direction increases, the increase in pressure loss can be prevented.

Moreover, if the space _| interval ( _Pf ) between fin structure parts is 5 mm or more, an increase in pressure loss can be suppressed, without disturbing a refrigerant _| coolant flow excessively.
Further, when the lead angle (β ₂ ) of the secondary groove with respect to the shaft is as large as 10 ° or more, since the liquid is retained in the tube for a long time, evaporation is repeated, and as a result, the heat transfer coefficient is improved.

  According to this invention, it is possible to improve the heat transfer coefficient in the tube while suppressing an increase in pressure loss during heat exchange by the heat transfer tube.

The partial expansion expansion perspective 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 FIG. The partial expansion expansion perspective view which shows the other example of this invention heat exchanger tube. The enlarged plan view of the pipe inner surface which has a fin structure part which shows the heat exchanger tube of FIG. 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 the other example of this invention heat exchanger tube.

  An embodiment of the present invention will be described below with reference to the drawings.

  As shown in FIGS. 1 and 2, 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 divides the fin 12 in the main groove spiral direction D2 (fin formation direction) of the fin 12 to form a plurality of fin constituent portions 12A.

  In addition, the heat transfer tube 11 in this embodiment forms the protruding piece 16 which protrudes spirally toward the sub-groove spiral direction D3 on the tube inner surface 10, so that it is between the fin constituent portions 12A in the main groove spiral direction D2. Includes a sub-groove forming portion 15 that divides the fin 12 by the sub-groove 14.

  FIG. 1 is a partially enlarged 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 an enlarged plan view in the vicinity of the fin constituting portion 12A. In FIG. 2, only the fin constituent portions 12 </ b> A and the tops of the protruding pieces 16 are schematically shown.

  Furthermore, 12 A of said fin structure parts are provided with the protrusion 16 (1st protrusion 16a) which protrudes in the said main groove 13 of the pipe-axis direction upstream D1u in the main groove spiral direction downstream D2d.

The heat transfer tube 11 is formed with an outer diameter of 5 mm, which is in a range where the outer diameter is smaller than 7 mm.
The fins 12 are formed with a height (H _f ) of 0.15 mm which is in the range of 0.1 mm to 0.30 mm.

The sub-groove 14 is formed with a depth of 0.15 mm, the depth (H _n ) being in the range of 0.1 mm to 0.30 mm.
The protruding piece 16 is formed at an angle of 34 °, where the angle (α) with respect to the fin 12 is in the range of 5 ° ≦ α <70 °.

12 A of said fin structure parts are formed by 2.4 mm whose space _| interval ( _Pf ) between the fin structure parts in a fin formation direction exists in the range of 1.5 mm or more.

More specifically, in the heat transfer tube 11 of the present embodiment, the spiral fins 12 are formed at an angle (β ₁ ) of 40 ° with respect to the tube axis direction D1. Further, the number of fins per pipe circumference is 56, the fin height (H _f ) is 0.15 mm, and the cross-sectional shape of the fins 12 is formed in a substantially trapezoidal shape.

The secondary groove 14 is formed at an angle (β ₂ ) of 6 ° with respect to the tube axis direction D1, and is formed along the _secondary groove spiral direction D3. Furthermore, it is formed in a groove shape having an inverted triangular cross section with a sub groove width (W ₃ ) in the fin forming direction of 0.15 mm.

The depth (H _n ) of the sub-groove 14 is formed at the same groove depth from the upstream D3u in the sub-groove spiral direction D3u in the sub-groove spiral direction D3 toward the downstream D3d in the sub-groove spiral direction.

The protruding piece 16 is formed with a protruding length (W ₁ ) to the main groove 13 having a ratio (W ₁ / W ₂ ) of 0.5 to the width (W ₂ ) of the main groove 13. ing.

  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 main groove spiral direction.

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

  Therefore, as shown in FIG. 2, the fin component portion 12 </ b> A is formed in a shape in which the letter “H” is inclined (inclined H shape) in plan view.

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 has the first protruding piece 16a that protrudes at least on the downstream side D2d in the main groove spiral direction D2d of the fin component 12A and on the main groove 13 on the upstream side D1u in the tube axis direction. It has.

  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 is, in particular, a third protruding piece that protrudes toward the main groove 13 on the upstream side D1u in the tube axis direction at least on the upstream side D2u in the main groove spiral direction of the fin component 12A. 16c, 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 near the inner peripheral surface of the pipe. In addition, it is possible to prevent an increase in pressure loss.

In the heat transfer tube 11 of the present embodiment, the protruding piece 16 is formed at an angle (α) of 34 ° with respect to the main groove 13 as described above.
For this reason, it is possible to cause a part of the refrigerant to collide with the protruding piece 16 and flow into the sub groove 14 from the main groove 13 without excessively hindering the flow of the refrigerant flowing through the main groove 13. The heat transfer coefficient can be improved by a larger 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 addition, as described above, the heat transfer tube 11 of the present embodiment can obtain a sufficient refrigerant stirring action by forming the interval (P _f ) between the fin constituent portions in the fin forming direction at 2.4 mm. The fin strength that can be endured during mechanical tube expansion can be obtained.

Further, as described above, in the heat transfer tube 11 of the present embodiment, the projecting piece 16 has a ratio (W ₁ / W ₂ ) of 0.5 to the width (W ₂ ) of the main groove 13. By forming the groove 13 with the protruding length (W ₁ ), the scooping of the refrigerant inward in the radial direction is promoted, and an excellent refrigerant stirring action can be obtained, resulting in an improvement in the heat transfer coefficient. Can be planned.

Further, the depth (H _n ) of the sub-groove 14 is preferably 0.1 mm or more and 40 to 100% of the depth of the main groove 13 as in the heat transfer tube 11 of the present embodiment.

  By configuring the sub-groove 14 with a depth of 0.1 mm or more and 40% 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 obtain an excellent heat transfer coefficient. And increase in 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.

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

In this experiment, nine types of Examples 1 to 9 were manufactured as heat transfer tubes of the present invention, and four types of heat transfer tubes used for comparison were manufactured.
The heat transfer tubes of Examples 1 to 9 are each formed into a shape having an outer diameter, fins, sub grooves, and protruding pieces as shown in Table 1.

ここで、実施例２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, the heat transfer tube of Example 2 uses a heat transfer tube having the same shape as the heat transfer tube 11 of the above-described embodiment, as is apparent from the shape of each part shown in Table 1.

  In addition, the heat transfer tubes of Examples 1, 2, 8, and 9 are each provided with a plurality of divided fin components 12 </ b> A having an inclined H shape in a plan view. On the other hand, as shown in FIGS. 3 and 4, the heat transfer tubes 21 of Examples 3 to 7 include a plurality of fin constituent portions 42 </ b> A having a J-shaped shape (J shape in plan view) when viewed in plan. In preparation.

  More specifically, in the heat transfer tube 21 of the third to seventh embodiments, the divided fin component 42A having a J shape in plan view includes the first protrusion 16a in the fin component 42A, and the second protrusion It is the form provided with the piece 16b and the 3rd protrusion piece 16c.

Here, FIG. 3 is a partially expanded perspective view schematically showing a state of the tube inner surface 10 of the heat transfer tube 21 of Examples 3 to 7, and FIG. 4 is a view of the heat transfer tube 21 of Examples 3 to 7. It is an enlarged plan view of the fin vicinity of the pipe inner surface divided. In FIG. 4, only the fin component 42 </ b> A and the top of the protruding piece 16 are schematically shown.
In addition, about a comparative example, the comparative examples 1 to 3 are the heat transfer tubes provided with the fin configuration portion in the plan view inclined type H shape and the comparative example 4 in the plan view J shape.

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. 5A, 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.
5 (a) and 5 (b) show 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, fins are formed on the heat transfer tubes of Examples 1 to 9 and Comparative Examples 1 to 4 and the inner surfaces of the respective heat transfer tubes, and sub-grooves (fin forming portions) are formed in the fins. As shown in FIG. 5 (a), a heat transfer tube that does not form a heat transfer tube, that is, a conventional internally grooved tube, was prepared and incorporated in a condenser as a test tube 44.

In this case, the heat transfer coefficient (α _i ) of the heat transfer tubes of Examples 1 to 9 and Comparative Examples 1 to 4 and the heat transfer coefficient (α _BASE ) of the heat transfer tube having only the main grooves of the same shape as each of them. The effect on the condensation performance of the present invention was verified by taking the ratio (α _i / α _BASE ). The same comparison was made for the pressure loss ratio (ΔP / ΔP _BASE ).

  In the evaporation experiment, fins are formed on the heat transfer tubes of Examples 1 to 9 and Comparative Examples 1 to 4 and the inner surfaces of the respective heat transfer tubes. Sub-grooves (fin forming portions) are formed on the fins. A heat transfer tube, i.e., a conventional internally spiral grooved tube, was prepared and incorporated in the evaporator as a test tube 44 as shown in FIG.

In this case, the heat transfer coefficient (α _i ) of the heat transfer tubes of Examples 1 to 9 and Comparative Examples 1 to 4 and the heat transfer coefficient (α _BASE ) of the heat transfer tube having only the main grooves of the same shape as each of them. The effect on the evaporation performance of the present invention was verified by taking the ratio (α _i / α _BASE ). The same comparison was made for the pressure loss ratio (ΔP / ΔP _BASE ).

As shown in FIGS. 5A and 5B, 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 2 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. 5 (a) and 5 (b), a thermometer, a pressure gauge, and a flow meter are arranged at each predetermined portion of the test section. 5A and 5B, 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 and the refrigerant outlet supercooling degree, and in the evaporation experiment, the refrigerant inlet dryness and the refrigerant outlet superheat degree, respectively. Settings were made as shown in Table 2.

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２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 2, 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. 5A and 5B) installed at the compressor outlet during the experiment. The sample was collected and experimented while measuring the refrigerant composition ratio by gas chromatography.
Incidentally, the analysis results of gas chromatography, reflects the t _s1 and t _s2 below by calculation.

The pressure loss (ΔP, ΔP _BASE ) and heat transfer coefficient (α _i , α _BASE ) in the test tube 44 showing the condensation performance and the evaporation performance are obtained as follows.
First, the pressure loss in the tube is obtained as the pressure difference between the inlet and outlet of the test tube 44.
The heat transfer coefficient in the tube is calculated using 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（ｍ^２）、ｔ_ｍは、対数平均温度（Ｋ）、α_Ｏは、管外熱伝達率（ｋＷ／ｍ^２・Ｋ）を示す。

Here, in Equation (1), Q is the amount of exchange heat (kW), A is the external surface area (m ² ) of the test tube, t _m is the logarithmic average temperature (K), and α _O is the external heat transfer coefficient. (KW / m ² · K).

In Equation (2), G is the flow rate of heat exchange water (kg / s), _CP is the specific heat of heat exchange water (kJ / kg · K), and t _W1 is the inlet temperature of heat exchange water (K). , T _W2 indicates the outlet temperature (K) of the water for heat exchange.

In formula (3), t _S1 represents the refrigerant inlet saturation temperature (K), and t _S2 represents the refrigerant outlet saturation temperature (K).

In Equation (4), k represents the thermal conductivity (kW / m · K) of heat exchange water, and _De represents the annular portion equivalent diameter (m). D represents a shell inside diameter (m), d is subjected試管outer diameter (m), R _e is the Reynolds number of the heat exchange water, P _r 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 Equation (2), t _m at the time of condensation is calculated from Equation (3), α _O is calculated from Equation (4), and α _O is calculated from Equation (4). By substituting into equation (1), the in-tube heat transfer coefficient can be calculated.
Table 3 shows the evaluation results of the condensation performance and evaporation performance thus obtained.

表３より明らかなように、熱伝達率に関しては、実施例１乃至９のいずれの伝熱管においても、凝縮性能、蒸発性能ともに、従来の螺旋溝付管より高い熱伝達性能を示した。

As is clear from Table 3, with regard to the heat transfer coefficient, in any of the heat transfer tubes of Examples 1 to 9, both the condensation performance and the evaporation performance showed higher heat transfer performance than the conventional spiral grooved tube.

  As a result, a part of the refrigerant flowing through the main groove collides with the first projecting piece and is scooped inwardly in the radial direction of the pipe. It was possible to demonstrate the effect of the diffusion of the refrigerant on the entire inner peripheral surface of the pipe caused by a part of the refrigerant flowing through the main groove flowing into the sub-groove.

Further, from the results of Example 1 and Comparative Example 2, it is necessary to improve the performance by sub-groove depth (H _n ) of 0.1 mm or more. In order to improve the performance, it is necessary that the sub-groove depth (H _n ) is larger than 0.12 mm, and it can be seen that the heat transfer coefficient ratio increases as the depth increases. This is considered to be caused by a shortage of the amount of refrigerant flowing into the sub-groove when the sub-groove depth is shallow.

Further, from the results of Examples 1, 5, 7, 8, 9 and Comparative Example 1, the heat transfer coefficient is obtained when the interval (P _f ) between the fin constituent portions in the fin forming direction is 1.5 mm or more for both condensation and evaporation. It can be seen that the ratio is improved and peaks at around 4 mm. On the other hand, the pressure loss ratio decreases as the distance between the fin components (P _f ) increases.

In particular, Examples 8 and 9 have an evaporation pressure loss ratio of 100 or less, and can be said to exhibit excellent performance when used as an evaporation tube in which pressure loss is regarded as important. This is because the gap between the fin components (P _f ) is large, and the angle difference (α) between the main groove and the protruding piece is as small as 10 °, so that the refrigerant can easily flow into the sub groove, The reason is thought to be an increase in the refrigerant flow rate in the tube axis direction.

  Moreover, it can be seen from the results of Example 1 and Comparative Example 3 that the pressure loss is extremely increased when the angle (α) between the main groove and the protruding piece is 70 ° or more. This is because the refrigerant flowing in the main groove needs to change the direction to flow into the sub-groove, and in order to improve the heat transfer rate while suppressing the increase in pressure loss, the main groove protrudes from the main groove. It can be said that it is preferable that the angle difference (α) of the pieces is less than 70 °.

From the results of Examples 5 and 6 and Comparative Example 4, as the ratio (W ₁ / W ₂ ) of the protruding length (W ₁ ) between the fins of the protruding piece increases, the improvement in the heat transfer coefficient ratio increases, and W ₁ It can be seen that there is almost no effect when / W ₂ = 0.2.

Since the heat transfer coefficient is improved in Example 1, the ratio of the protrusion length (W ₁ ) needs to be 0.3 or more in order to obtain a sufficient effect, and the pressure loss ratio is about Since the protrusion length (W ₁ ) was not increased so much, it was confirmed that it is preferable to set the protrusion length (W ₁ ) large in a small-diameter tube having an outer diameter of less than 7 mm.

According to the comparison between Examples 4 and 6, the larger the lead angle (β ₂ ) of the sub-groove and the smaller the angle difference (α) between the main groove and the protruding piece, the greater the effect of improving the evaporation heat transfer coefficient can be obtained. It could be confirmed.

This is because, in Examples 8 and 9, the improvement in the evaporative heat transfer ratio is large for the size of the interval (P _f ) between the fin components, and therefore, the lead angle (β ₂ ) of the relatively large secondary groove. It was confirmed that the liquid retention due to the liquid and the expansion of the liquid film across the fins due to the small angle difference between the main groove and the protruding piece lead to evaporation promotion.

The present invention is not limited to the heat transfer tubes of the first to ninth embodiments described above, and can be configured in various forms and modes.
For example, the heat transfer tube of the present invention is similar to the heat transfer tube 11 of the above-described embodiment (Examples 1, 2, 8, 9) or the heat transfer tube 21 of Examples 3 to 7, and the fin components 12A, 42A can be formed with a configuration including at least the first protruding piece 16a.

  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. 6, the heat transfer tube of the present invention includes a plurality of fin components 52 </ b> A having a Z-shaped shape (planar inclined Z shape) when viewed in plan. Alternatively, the heat transfer tube 31 can be formed.
FIG. 6 is an enlarged plan view of the inner surface of the tube having the fin constituting portion 52A having a Z shape in plan view. However, in FIG. 6, only the top part of the said fin structure part 52A and the said protrusion piece 16 is shown typically.

  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.

  Further, even if the heat transfer tube 31 has the upstream side and the downstream side in the tube axis direction D1 opposite to the directions shown in FIG. 6 and the refrigerant flows, the fourth projecting piece 16d remains in the main groove of the fin component 52A. In the spiral direction downstream side D2d, the main groove 13 is projected on the upstream side in the tube axis direction D1u (see the arrow in parentheses in FIG. 6).

  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 spiral direction downstream side D2d of the fin component 52A and protrudes to the main groove 13 on the upstream side D1u in the tube axis direction.

  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 mounted on the heat exchanger without being aware of the mounting direction. Can do.

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 main groove 13 of this embodiment corresponds to between the fins of the present invention.

DESCRIPTION OF SYMBOLS 10 ... Tube inner surface 11, 21, 31 ... Heat-transfer tube 12, 42, 52 ... Fin 12A, 42A, 52A ... Fin component 13 ... Main groove 14 ... Sub groove 15 ... Sub groove forming part 16 ... Projection piece D1 ... Tube axis Direction D2 ... Main groove spiral direction D3 ... Sub groove spiral direction

Claims (1)

A heat transfer tube for evaporation smaller than an outer diameter of 7 mm in which a spiral fin having a height (H _f ) of 0.1 mm to 0.30 mm having a predetermined angle (β ₁ ) with respect to the tube axis direction is formed on the inner surface of the tube. Because
It said inner surface, a plurality of fins that protrude while being divided in the fin forming direction by the fin depth (H _n) is spiral minor groove of 0.1Mm～0.30Mm, helically tube inner surface Forming the component,
Forming a projecting piece projecting between the adjacent fins on at least one of the upstream side and the downstream side in the spiral direction of the fin component,
In the protruding piece,
A first projecting pieces out collision between a spiral direction downstream side of the fin structure part, and the fins adjacent to each other in the tube axial direction upstream side,
A second projecting piece projecting between the fins adjacent to the fin on the downstream side in the tube axis direction on the opposite side of the first projecting piece;
A third projecting piece projecting between the fins adjacent to the fin on the upstream side in the spiral direction on the side of the first projecting piece;
A fourth projecting piece projecting between the fins adjacent to the fin on the upstream side in the spiral direction on the opposite side to the side having the first projecting piece ;
Before Symbol protruding piece protruding length to the spacing between the fins _{(W 2)} the ratio of _{(W 1) (W 1 /} W 2) is formed at 0.3 to 0.9,
The fin component portion is formed with a spacing (P _f ) between the fin component portions in the fin formation direction of 5 mm or more,
The angle (α) of the protruding piece with respect to the fin was set to 5 ° ≦ α ≦ 20 °.
Heat transfer tube for evaporation .

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