KR20130131499A - Heat transmission tube - Google Patents

Abstract

An object of the present invention is to provide a heat transfer tube capable of improving the heat transfer rate in a tube while suppressing an increase in pressure loss. In the heat exchanger tube 11 whose height H _f which has a certain angle (beta) with respect to the tube axial direction D1 in the tube inner surface 10 is smaller than the outer diameter of 7 mm in which the spiral fin 12 of 0.1 mm-0.30 mm was formed, The pin 12 has a pin configuration 12A segmented into a helical subgroove 14 having a depth Hn of 0.1 mm to 0.30 mm (Hn). 5 ° ≤α <70 ° angle with respect to the pin 12 between the pin structure 12A and the pin 12 adjacent to the pin 12 at least in the helical downstream side D2d on the axial upstream side D1u. The protrusion piece 16 which protrudes with (alpha) is provided. Pin configuration unit (12A), the clearance between the pin out portion (12A) formed in the pin direction (P _f) of more than 1.5mm, projecting on the pin 12 between the aperture (W ₂₎ of the projection piece 16. The length The ratio (W ₁ / W ₂ ) of (W ₁ ) was formed at 0.3 to 0.9.

Description

Heat transfer tube {HEAT TRANSMISSION TUBE}

The present invention relates to heat transfer tubes used in heat exchangers such as refrigerators and air conditioners.

In general, heat transfer tubes used in an air conditioner or a refrigerator are heat exchanged between a fluid flowing out of a tube by evaporating or condensing a refrigerant in the tube, and a tube having a groove on the inner surface is used in view of high efficiency of the heat exchanger or energy saving. Many times.

In this inner grooved tube, a fine triangular cross section or trapezoidal cross-sectional groove is formed on the inner surface of the tube in a straight or spiral shape with respect to the axis of the tube. By providing these grooves on the inner surface of the tube, the heat transfer area is larger than that of the flat flowing tube, and at the same time, the heat transfer performance can be improved by the stirring action of mixing the refrigerant liquid.

Recently, high performance and compact weight reduction are particularly required for heat exchangers for air conditioners, and according to the revision of the Energy Saving Act (Law on the Rationalization of Energy Use), the performance of heat transfer tubes is further required.

However, in the conventional tube provided with spiral grooves, the number of grooves, lead angles, groove shapes, and the like have been improved, but they are not sufficient for the performance requirements described above.

Accordingly, as a heat transfer tube replacing the conventional spiral finned tube, for example, Patent Document 1 has a heat transfer tube provided with a main groove on the inner surface of the tube and a cross groove formed with a sub groove having a depth for dividing the fin in order to promote the stirring action of the refrigerant liquid. Is disclosed.

This cross groove heat exchanger tube has a plurality of three-dimensional protrusions 3 having a substantially S-shape in the inner surface of the tube when the fins are divided by sub grooves and viewed in plan view.

More specifically, the three-dimensional protrusion 3 has a wing 3a that protrudes at the front end thereof to guide the refrigerant flow along the main groove in the sub-home direction, and at the same time, a wing protruding in the opposite direction to each wing 3a at the rear end ( Patent Document 1 discloses that the refrigerant flowing through the main groove is guided in the sub-groove direction by the vanes 3a to cause the stirring action due to the complicated flow of the refrigerant, and as a result, the heat transfer efficiency can be obtained. have.

However, the heat transfer tube in Patent Literature 1 allows the refrigerant to be agitated near the inner circumferential surface of the tube by the blades 3a and 3b provided in the three-dimensional protrusion 3, but the refrigerant flowing in the radial center side of the tube is stirred. As a result, it was not possible to obtain a transmission performance that satisfies the requirements described above.

Patent Document 1: Japanese Patent Application Laid-Open No. 8-178574

An object of this invention is to provide the heat exchanger tube which can aim at the improvement of the heat transfer rate in a pipe, suppressing the pressure loss increase.

The present invention is a heat transfer tube smaller than the outer diameter of 7mm with a spiral fin having a height (H _f ) of 0.1mm to 0.30mm having a predetermined angle (β ₁ ) with respect to the axial direction on the inner surface of the tube, the depth of the fin on the inner surface of the tube ( H _n ) is divided into a fin forming direction by a spiral sub groove having 0.1 mm to 0.30 mm, and a plurality of fin components are formed on the inner surface of the tube in a spiral manner, and at least spirally downstream of the fin components, A protruding piece that protrudes at an angle α of 5 ° ≦ α <70 ° with respect to the pin, between the pins adjacent to each other on the upstream side of the tube axial direction; in that a heat transfer pipe configured to form a unit ratio (W ₁ / W ₂₎ between the gap (P ₁₎ of 1.5mm or more, the projecting length (W ₁₎ for the interval between the (W ₂₎ the projecting pin convenience 0.3 to 0.9 It features.

In the claims and the specification of the present invention, the range described using the ˜ symbol includes the numerical value indicated before the symbol and the numerical value described after the symbol.

The heat transfer tube has fins between the fins adjacent to the tube axial upstream side at least spirally downstream of the fin configuration, in particular at least 0.1 mm, even though the fin height H _f is within the range of 0.1 mm to 0.30 mm. It is a structure provided with the protrusion which protrudes at an angle of 5 degrees or more with respect to a formation direction.

According to this configuration, a part of the refrigerant flowing between the fins may collide with the protruding pieces and move strongly inward in the radial direction of the tube to effectively generate a three-dimensional abnormal flow. In addition, since the coolant stays by colliding with the protruding pieces, the liquid film thickness of the coolant flowing in the main groove becomes nonuniform in the pin formation direction, so that the heat transfer performance is improved in the thin film thickness.

Therefore, compared with the conventional heat exchanger tube which simply provided the cross groove in the inner periphery of a tube, turbulence including the radial direction of a tube can be promoted, or liquid refrigerant thin film can be thinned, and the outstanding heat transfer rate can be obtained.

In addition, even if the sub-groove depth (Hn) is within the range of 0.1mm to 0.30mm, in particular 0.1mm or more, by forming the angle (α) formed between the protruding piece and the main groove smaller than 70 degrees, the refrigerant flowing between the fins easily flows into the sub-groove. Therefore, the stirring of the refrigerant can be achieved in the entire circumferential surface of the pipe via the pin. Therefore, an excellent turbulence promoting effect can be obtained, and an increase in pressure loss can be prevented.

In addition, even if the pin height H _f is within the range of 0.1 mm to 0.30 mm, especially when the pin height is 0.25 mm or less, the inner diameter does not become too small in the small-diameter tube, and the increase in pressure loss can be suppressed.

In addition, by forming the gap P _f between the fin components in the fin formation direction of 1.5 mm or more, the refrigerant stirring action can be obtained with sufficient turning force without excessively inhibiting the refrigerant flow along the fin. And it is possible to secure the strength of the pin obtained when expanding the machine.

In the Figure, in particular _{W 1 / W 2 = 0.4 ~} 0.9 protrusion length (W ₁₎ the ratio (W ₁ / W ₂₎ of 0.3 to 0.9 range knock is preferred. Specifically, in the small diameter tube having an outer diameter of less than 7 mm, the fin height H _f is relatively low due to the above-mentioned pressure loss problem or processing limitation. a shift ratio (W ₁ / W ₂₎ of the projecting length (W ₁₎ projecting through the pin is required to be 0.4 or more.

When the evaporation rate is 0.9 or less, supply of the liquid film between the fins is not impeded, so that the evaporation heat transfer rate decrease due to dry out can be prevented. And an outer diameter of about the pressure loss increases or decreases within the small globule scenery than 7mm influence of the fin height (H _f) or a lead angle (β _2), the gap (P _f) between the pin out portion is the dominant of the projecting length (W ₁₎ Since the influence becomes relatively small, the heat transfer promoting effect described above can be obtained without excessively increasing the pressure loss even with the above configuration.

Heat transfer pipe for condensation of the present invention is the height of the fin H _f > 0.12mm, the sub groove depth H _n > 0.12mm, the angle (α) for the protruding piece pin is 20 ° or more, the angle (β ₂ ) of the sub groove in the pipe axis direction is -10 ° to 15 °, and the gap between the pin components in the pin forming direction (P _f ) 2 to 5mm, the width (W ₃ ) of the sub groove in the pin formation direction can be formed to 0.1mm or more.

The heat transfer tube for condensation has a configuration suitable for a condensation tube having a fin height H _{f of} greater than 0.12 mm. In other words, by increasing the fin configuration, it is possible to prevent the problem that the entire heat transfer surface is immersed in the condensed water, the performance degradation occurs.

If the sub groove depth (H _n ) is larger than 0.12 mm, the sub groove lead angle (β ₂ ) is -10 to 15 °, and the width of the sub groove (W ₃ ) is 0.1 mm or more, the flow path area of the sub groove is largely secured. By reducing the lead angle β ₂ , the condensate discharge action in the axial direction of the tube is promoted and the top surface of the tube is kept in a dry state, thereby improving condensation performance.

In addition, if the angle α between the protruding piece and the main groove is 20 ° or more and the gap P _f between the pin components is 5 mm or less, the refrigerant flow flowing between the pins easily collides with the protruding piece, thereby obtaining a larger refrigerant stirring action. have.

In the evaporation heat exchanger tube of the present invention, the fin height H _f is 0.1 mm or more, the sub-groove depth H _n is 0.1 mm or more, and the angle α with respect to the fin of the protruding piece is 20 ° or less, the tube axis. 10 degrees or more of the lead angle (beta) _{2 of the} sub groove with respect to a direction, and the space | interval P _f between the pin component parts of a pin formation direction can be formed in 5 mm or more.

The heat transfer tube for evaporation has a configuration suitable for an evaporation tube having a fin height H _f of 0.10 mm or more. In other words, the effective heat transfer area of the inner surface of the tube is increased and the effect of thinning the liquid film near the top of the fin is sufficient to improve the evaporation performance.

In addition, if the sub-groove depth (H _n ) is 0.10 mm or more and the angle (α) formed between the protruding pieces and the main groove is 20 ° or less, the refrigerant flowing between the fins easily flows into the sub-grooves. The area is increased to obtain an evaporation promoting effect. In addition, since the refrigerant flow rate increases in the tube axial direction, it is possible to prevent an increase in pressure loss.

In addition, if the gap P _f between the fin components is 5 mm or more, an increase in pressure loss can be suppressed without excessively blocking the refrigerant flow.

In addition, when the lead angle β ₂ of the sub groove with respect to the tube axis is larger than 10 °, the liquid is retained in the tube for a long time, so that evaporation is repeated, resulting in an improved heat transfer rate.

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

1 is a partially enlarged exploded perspective view showing a heat pipe of the present embodiment;
FIG. 2 is an enlarged plan view of the inner surface of the tube having fin components in the heat transfer tube of FIG.
Figure 3 is a partially enlarged exploded perspective view showing another embodiment of the heat transfer tube of the present invention
FIG. 4 is an enlarged plan view of the inner surface of the tube with a fin configuration showing the heat transfer tube of FIG.
5 is a schematic diagram of an experimental apparatus used for heat transfer pipe performance evaluation according to the present embodiment.
Figure 6 is an enlarged plan view of the inner surface of the tube having a fin configuration showing another embodiment of the heat transfer tube of the present invention;

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

In the heat exchanger tube 11 of this embodiment, as shown in FIG. 1, FIG. 2, the helical fin 12 is formed in the tube inner surface 10. As shown in FIG.

In addition, in this embodiment, the heat exchanger tube 11 is provided with the main groove 13 between the fins 12 by forming the helical fin 12 in the tube inner surface 10.

The heat transfer pipe 11 divides the fin 12 into the main groove helical direction D2 (fin formation direction) of the fin 12 to form a plurality of fin components 12A.

Moreover, the heat exchanger tube 11 of this embodiment forms the protrusion piece 16 which protrudes spirally with respect to the sub groove spiral direction D3 on the tube inner surface 10, and is between the fin component parts 12A of the main groove spiral direction D2. The sub groove 14 includes a sub groove forming portion 15 for dividing the pin 12 by the sub groove 14.

FIG. 1 is a partially enlarged exploded perspective view schematically showing the shape of the tube inner surface 10 of the heat transfer pipe 11 of the present embodiment, and FIG. 2 is an enlarged plan view near the fin constitution 12A. 2, only the top part of the said pin structure part 12A and the said projection piece 16 is shown typically.

Moreover, the said pin structure part 12A is provided with the protrusion piece 16 and the 1st protrusion piece 16a which protrude to the said main groove 13 of the tube axial upstream side D1u in the main groove spiral direction downstream side D2d. .

The heat pipe 11 was configured with a 5mm outer diameter of the outer diameter is less than 7mm range.

The pin 12 was formed with a height of 0.15 mm with a height H _{f in the} range of 0.1-0.30 mm.

The subgroove 14 was formed to a depth of 0.15 mm having a depth H _{n in the} range of 0.1 mm to 0.30 mm.

The protruding piece 16 is formed at an angle of 34 ° with an angle α with respect to the pin 12 in a range of 5 ° ≦ α <70 °.

The pin component 12A is formed to be 2.4 mm with a gap P _f between the pin components in the pin forming direction within a range of 1.5 mm or more.

More specifically, in the heat exchanger tube 11 of this embodiment, the helical fin 12 is formed in 40 degree angle (beta) ₁ with respect to the tube axial direction D1. And the pin number of the peripheral tubes 56, the fin height (H _f) 0.15mm, the section of the pin in a trapezoid shape.

The subgroove 14 is formed at an angle β ₂ of 6 ° with respect to the axial direction D1, and is formed along the subgroove spiral direction D3. Further, in the pin forming direction, the sub groove width W ₃ is formed in an inverted triangular cross-sectional groove shape of 0.15 mm.

Figure H _n 1 represents the average depth (H _n) of buhom 14 over the buhom spiral direction downstream side (D3d) in buhom spiral direction upstream side (D3u).

The protruding piece 16 is formed with a protruding length W ₁ with respect to the main groove 13 having a ratio W ₁ / W ₂ with respect to the width W _{2 of} the main groove 13.

The pin component 12A has a projection piece 16 (second projection piece 16b) protruding from the main groove helical downstream side D2d to the main groove 13 on the tube axial downstream side D1d. have.

On the other hand, the pin configuration portion 12A is a projection piece 16 (third projection 16c) and a tube which protrude into the main groove 13 on the tube axially upstream side D1u, respectively, in the main groove spiral direction upstream side D2u. The protrusion piece 16 (fourth protrusion piece 16d) which protrudes to the said main groove 13 of the axial downstream side D1d is provided.

Therefore, as shown in FIG. 2, the pin structure part 12A is formed in the shape which inclined H shape (inclined H shape) in plan view.

The heat pipe 11 described above can obtain various actions and effects as follows.

The heat-transfer tube 11 of this embodiment protrudes to the said main groove 13 of the pipe axial direction upstream D1u at least in the main groove spiral direction downstream D2d of the said fin structure 12A as described above. The 1st protrusion piece 16a is provided.

For this reason, a part of the refrigerant flowing through the main groove 13 may collide with the first protruding piece 16a and flow strongly inward in the radial direction of the tube to generate a three-dimensional abnormal flow.

Therefore, the turbulent flow promoting effect including the radially inner side of the tube can be obtained. Thus, the heat transfer rate can be improved compared with the conventional heat exchanger tube having a cross groove on the inner surface of the tube or a conventional heat transfer tube having wings near the sub-groove. .

In addition, as described above, the helical fin 12 projects, in particular, at least on the main groove helical upstream side D2u of the pin component 12A, and protrudes toward the main groove 13 side of the tube axial upstream side D1u. Since the piece 16c is provided, the refrigerant flowing through the main groove 13 is more likely to flow into the sub groove 14, so that secondary flow occurs, and in particular, the refrigerant stirring effect can be obtained near the inner circumferential surface of the tube. Increase in pressure loss can be prevented.

In addition, the heat exchanger tube 11 of this embodiment forms the said projection piece 16 by 34 degree angle (alpha) with respect to the said main groove 13 as mentioned above.

For this reason, it is possible to impinge a portion of the coolant flowing through the main groove 13 into the sub-groove 14 from the main groove 13 without impeding the flow of the refrigerant excessively. By this, the heat transfer rate can be improved.

In addition, since the refrigerant flow rate increases in the pipe axial direction D1, an increase in pressure loss can be prevented.

In addition, as described above, the heat transfer pipe 11 of the present embodiment forms a gap P _f between the fin components in the fin forming direction to 2.4 mm to obtain a sufficient refrigerant stirring action, and at the same time, the fin strength to withstand the expansion of the machine. Can be obtained.

In the heat transfer tube 11 of this embodiment is one wherein the juhom 13 is made of a 0.5 ratio (W ₁ / W ₂₎ for the projection pieces 16 in the width (W ₂₎ of the juhom 13 as described above; By forming the protrusion length W ₁ with respect to, it is possible to move the refrigerant strongly inward in the radial direction to obtain an excellent refrigerant stirring action, resulting in improved heat transfer rate.

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 pipe 11 of the present embodiment.

By configuring the sub groove 14 to 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 easily flows into the sub groove 14, thereby obtaining an excellent heat transfer rate and increasing pressure loss. You can prevent it.

On the other hand, since the sub groove 14 is formed to a depth of 100% or less of the depth of the main groove 13, the pipe thickness does not penetrate deeper than the main groove depth, thereby maintaining the quality (bottom thickness specification) of the pipe.

As mentioned above, although the heat exchanger tube 11 which is one Embodiment of this invention was demonstrated in detail, the experiment performed to verify the performance of the heat exchanger tube 11 of this invention is demonstrated.

In this experiment, nine types were produced as Examples 1-9 and nine types of the heat exchanger tubes used as a comparison object as the heat transfer tubes of this invention, and four types were produced to Comparative Examples 1-4.

Heat transfer tubes of Examples 1 to 9 are manufactured in a shape having an outer diameter, a fin, a sub groove, and a protruding piece as shown in Table 1, respectively.

Performance test tube

Lead angle


Groove depth Groove depth rain
pitch
Protrusion length
Condensation Performance Evaporation performance Heat transfer rate Pressure loss ratio Heat transfer rate Pressure loss ratio β ₁ β ₂ alpha H _f Hn Hn / H _f P _f W ₁ / W ₂ αi / α _{B A SE} ΔPi / ΔP _{BA SE} αi / α _{B A SE} ΔPi / ΔP _{BA SE} Comparative Example 1 40 6 34 0.12 0.11 0.9 1.2 0.4 95.7 119.3 102.0 127.0 Comparative Example 2 40 6 34 0.12 0.05 0.1 2.4 0.1 99.4 110.1 100.2 117.0 Comparative Example 3 40 -30 70 0.12 0.12 1.0 2.2 0.6 101.7 124.2 110.4 129.6 Comparative Example 4 35 0 35 0.20 0.10 0.5 4.3 0.2 100.9 106.1 103.5 110.2 Example 1 40 6 34 0.12 0.12 1.0 2.4 0.4 101.4 111.6 107.1 110.0 Example 2 40 6 34 0.15 0.15 1.0 2.4 0.5 108.7 113.8 112.4 112.0 Example 3 35 0 35 0.20 0.18 0.9 2.2 0.6 110.5 114.1 115.0 117.0 Example 4 35 10 25 0.20 0.19 1.0 4.2 0.8 115.6 110.7 121.9 113.6 Example 5 35 0 35 0.20 0.19 1.0 4.3 0.5 109.0 108.9 113.5 109.4 Example 6 35 0 35 0.20 0.20 1.0 4.3 0.8 116.6 109.2 116.8 112.7 Example 7 35 0 35 0.20 0.18 0.9 8.6 0.6 104.6 101.2 111.3 104.2 Example 8 30 20 10 0.20 0.19 1.0 10.5 0.7 102.3 100.2 109.2 97.4 Example 9 30 20 10 0.20 0.19 1.0 17.5 0.7 102.5 100.6 107.6 94.2

Here, the heat exchanger tube of Example 2 employ | adopts the heat exchanger tube of the same shape as the heat exchanger tube 11 of embodiment mentioned above, as is evident also in the shape of each part shown in Table 1.

In addition, the heat exchanger tubes of Examples 1, 2, 8, and 9 are all formed with a plurality of divided fin configuration portions 12A having an H shape in plan view. On the other hand, the heat exchanger tube 21 of Examples 3-7 is provided with the some fin structure part 42A of J shape (J plane in plan view), when viewed from the top, as shown to FIG. 3 and FIG.

More specifically, the divided fin configuration 42A having a J-shaped plane in the heat transfer tubes 21 of Examples 3 to 7 is provided with a first protruding piece 16a at the fin configuration 42A. It is a form provided with the 2nd protrusion piece 16b and the 3rd protrusion piece 16c.

3 is a partially enlarged exploded perspective view schematically showing the shape of the tube inner surface 10 of the heat transfer tubes 21 of Examples 3 to 7. FIG. 4 is an enlarged plan view near the divided fin of the inner surface 10 of the tube 3 to 7 of the heat transfer tubes 21. FIG. 4, only the top part of the said pin structure part 42A and the said projection piece 16 is shown typically.

Moreover, about a comparative example, the comparative examples 1-3 are heat-transfer tubes provided with the fin structure part which is inclined H shape in plan view, and comparative example 4 is planar J shape.

In this experiment, the condensation experiment to verify the condensation performance of the heat transfer tube was carried out using the condensation performance measuring device 50A as shown in Fig. 5 (a), and the evaporation experiment to verify the evaporation performance was performed in Fig. 5 ( It carried out using the in-tube evaporation performance measuring apparatus 50B as shown to b).

5 (a) and 5 (b) show schematic diagrams of the condensation performance measuring device 50A and the evaporation performance measuring device 50B in the tube, respectively. Configured by a refrigeration cycle.

Specifically, in the condensation experiment, 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 are formed, but the heat transfer tube is not formed with a sub groove (fin formation). The inner spiral grooved tube was prepared and included as a performance test tube 44 in the condenser as shown in FIG.

In this case, the ratio (α _i / α _BASE ) between the heat transfer rate α _i of the heat transfer tubes of Examples 1 to 9 and Comparative Examples 1 to 4 and the heat transfer rate α α _BASE of the heat transfer tubes having only the main grooves of the same shape, respectively. And the effect on the condensation performance of the present invention was verified. Similar comparisons were made for the pressure loss ratio (ΔP / ΔP _BASE ).

In the evaporation experiment, the heat pipes of Examples 1 to 9 and Comparative Examples 1 to 4 and the heat pipes formed on the inner surface of each of the heat pipes, but did not form a sub-groove (finned portion), that is, a conventional inner spiral A grooved tube was prepared and included as a performance test tube 44 in the evaporator as shown in Fig. 5 (b).

In this case, the ratio (α i / α _BASE ) between the heat transfer rate α _i of the heat transfer tubes of Examples 1 to 9 and Comparative Examples 1 to 4 and the heat transfer rate α α _BASE of the heat transfer tubes having only the main grooves having the same shape, respectively. The effect on the evaporation performance of the present invention was verified by measuring. Moreover, the same comparison was performed about ratio of pressure loss ((DELTA ₎ P / (DELTA ₎ P _{BA SE)} .

As shown in Fig. 5 (a) and (b), the test section in the condensation performance measuring device 50A and the evaporation performance measuring device 50B in the tube is composed of a double tube heat exchanger, and the refrigerant in the performance test tube 44 is Flows the heat exchange water (hereinafter referred to as 'heat exchange water') in a direction opposite to the refrigerant flow to the inside of the annular portion 45 constituting the outer cell, thereby reducing the effective length of the performance test tube 44. Set it to 2m and perform heat exchange.

In addition, the heat exchange water in the condensation experiment flows low temperature water, and the high temperature water is used as the heat exchange water in the evaporation experiment.

In addition, as shown in Figs. 5 (a) and 5 (b), a predetermined portion of the test section is provided with a thermometer, a pressure gauge, and a wander meter. 5 (a) and (b), T denotes a thermometer, P denotes a pressure gauge, and G denotes a flowmeter.

Subsequently, as the experimental conditions for the refrigerant inlet and outlet of the performance test tube 44, the refrigerant inlet superheat degree, the refrigerant outlet cooling degree in the condensation experiment, and the refrigerant inlet dryness and the refrigerant outlet superheat degree in the evaporation experiment are shown in Table 2, respectively. Set.

condensation evaporation Refrigerant Average Saturation Temperature 48 ° C Refrigerant Average Saturation Temperature 5 Refrigerant Inlet Superheat 35 ℃ Refrigerant Inlet Dryness 0.21 Refrigerant outlet supercooling degree 5 ℃ Refrigerant outlet superheat 5 ℃ Heat pipe effective length 2 m Used refrigerant R410A

The experimental conditions in the condensation experiment and the evaporation experiment were all measured after adjusting the water temperature to be the same as the heat exchanger inlet condition of the air conditioner.

Of course, the average saturation temperature of the refrigerant at the inlet and outlet of the performance test tube 44 was set to 48 ° C. in the condensation experiment and 5 ° C. in the evaporation experiment as shown in Table 2.

As the refrigerant, R410A is used as an alternative freon. Since R410A is a mixed refrigerant, the refrigerant is collected by the refrigerant collecting unit (see FIGS. 5 (a) and (b)) installed at the outlet of the extruder during the experiment. Experiment while measuring.

In addition, the results of gas chromatograph analysis are reflected by t _s1 and t _s2 described later by calculation.

Condensing performance, the pressure loss (ΔP / Δ _BASE) and the heat transfer coefficient (α _i / α _BASE) in the tube of the performance test tube 44 that represents the evaporation performance is to be obtained as follows.

First, the pressure loss in the pipe is obtained by the pressure difference between the inlet and outlet of the performance test tube 44.

Heat transfer rate in a pipe is computed using Formula (1)-Formula (4) based on the measured value in this experiment.

In Equation (1), Q is the exchange heat capacity (kW), A is the outer surface area of the performance test tube (m ² ), t _m is the logarithmic average temperature (K), and α _o is the outside heat transfer rate (kW / m ² · K) Indicates.

In Equation (2), G is the flow rate of the heat exchange water (kg / s), Cp is the specific heat of the heat exchange water (kJ / kgK), t _W1 is the inlet temperature (K) of the heat exchange water, t _W2 is the The outlet temperature K is shown.

In formula (3) (4), t _s1 represents the refrigerant inlet saturation temperature (K), and t _S2 represents the refrigerant outlet saturation temperature (K).

In Equation (5), k denotes a thermal conductivity of the heat exchange water (kW / m · K), and D _e denotes an equivalent diameter (m) of the annular portion. D denotes the inner diameter of the cell (m), d denotes the outer diameter of the test tube (m), Re denotes the Reynolds number of the heat exchange water, and Pr denotes the number of frantelles of the heat exchange water.

That is, Q is calculated from Equation (2), tm of Equation (3) and condensation (4), and α _o from Equation (5) on the basis of measured values such as temperature and setting parameters. The heat transfer rate in the tube can be calculated by substituting the calculated value into Equation (1).

Table 3 shows the evaluation results of the condensation performance and the evaporation performance thus obtained.

Performance test tube
Outer diameter Floor thickness Home number
Lead each Groove depth Groove depth rain pitch Protrusion length Do t N1 N2 β ₁ β ₂ alpha H _f Hn Hn / H _f P _f W ₁ / W ₂ Comparative Example 1 5.00


0.23


56 22 40 6 34 0.12 0.11 0.9 1.2 0.4 Comparative Example 2 56 11 40 6 34 0.12 0.05 0.4 2.4 0.1 Comparative Example 3 56 6 40 -30 70 0.12 0.12 1.0 2.2 0.6 Comparative Example 4 50 6 35 0 35 0.20 0.10 0.5 4.3 0.2 Example 1
5.00







0.23






56 11 40 6 34 0.12 0.12 1.0 2.4 0.4 Example 2 56 11 40 6 34 0.15 0.15 1.0 2.4 0.5 Example 3 50 12 35 0 35 0.20 0.18 0.9 2.2 0.6 Example 4 50 8 35 10 25 0.20 0.19 1.0 4.2 0.8 Example 5 50 6 35 0 35 0.20 0.19 1.0 4.3 0.5 Example 6 50 6 35 0 35 0.20 0.20 1.0 4.3 0.8 Example 7 50 3 35 0 35 0.20 0.18 0.9 8.6 0.6 Example 8 52 10 30 20 10 0.20 0.19 1.0 10.5 0.7 Example 9 52 6 30 20 10 0.20 0.19 1.0 17.5 0.7

As can be clearly seen in Table 3, the heat transfer rate is higher than that of the conventional spiral grooved tube based on the condensation performance and the evaporation performance in any of the heat transfer tubes of Examples 1 to 9.

As a result, a part of the refrigerant flowing through the main groove collides with the first protruding piece, and flows strongly inwardly in the radial direction of the tube, thereby remarkably promoting turbulence and thinning of the liquid refrigerant including the radially inner side of the tube, and part of the refrigerant flowing through the main groove. The effect of the refrigerant diffusion on the entire inner circumferential surface of the pipe by flowing into the sub groove could be demonstrated.

In addition, from the results of Example 1 and Comparative Example 2, the sub groove depth Hn is required to be 0.1 mm or more for improving the performance, and the sub groove depth Hn is 0.12 for the condensation performance improvement by the comparison of Examples 1, 2 and 3. It should be larger than mm, and the deeper the depth, the better the heat transfer rate. This is considered to be due to the inflow of refrigerant into the sub grooves when the sub groove depth is shallow.

In addition, from the results of Examples 1, 5, 7, 8, 9 and Comparative Example 1, the heat transfer ratio is improved when the gap Pf between the fin components is 1.5 mm or more in the pin formation direction for both condensation and evaporation. It can be seen that it forms. On the other hand, the pressure loss ratio decreases as the distance P _f between the pin components increases.

In particular, in Examples 8 and 9, the evaporation pressure loss ratio is 100 or less, and it can be said that the tube exhibits excellent performance when used as an evaporation tube in which pressure loss is important. This is due to the large gap (P _f ) between the fin components and the small angle difference (α) between the main groove and the protruding piece, which is small by 10 °, so that the refrigerant flows easily into the secondary groove and the refrigerant flow rate increases in the tube axis direction. .

In addition, it can be seen from the results of Example 1 and Comparative Example 3 that the pressure loss becomes extremely large when the angle α forming the main groove and the protruding piece is 70 ° or more. This is because the refrigerant flowing in the main groove needs to be largely changed and flowed into the sub groove. Therefore, in order to suppress the increase in pressure loss and improve the heat transfer rate, it is preferable to form the angle difference α between the main groove and the protruding piece to be less than 70 °. Can be.

From the results of Examples 5, 6 and Comparative Example 4, the larger the protruding length W ₁ ratio (W ₁ / W ₂ ) to the main groove 13 width W ₂ of the protruding piece, the greater the improvement in the heat transfer rate ratio, W _{In the case of 1} / W ₂ = 0, it can be seen that there is almost no effect with the protruding pieces.

In Example 1, the ratio of the protrusion length W ₁ needs to be 0.3 or more in order to obtain an effect sufficient for the improvement of the heat transfer rate, and the pressure loss ratio does not increase much even when the protrusion length W ₁ is lengthened. in the small-diameter tube having an outer diameter less than 7mm has been confirmed that it is desirable that the projection length is set large (W _1).

It was confirmed from the comparison of Examples 4 and 6 that 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 evaporative heat transfer rate improvement effect.

Which Examples 8 and 9 the pin configuration unit cracks in point greater is the evaporation heat transfer ratio improved as compared with the size of the (P _f), the difference is relatively large buhom lead each hold liquid due to the (β _2), juhom and projecting convenience angle less It was confirmed that the diffusion of the liquid film across the fins and the fins led to the promotion of evaporation.

The present invention is not limited to the heat transfer tubes of Examples 1 to 9 described above, and can be configured in various forms and shapes.

For example, the heat transfer tube of the present invention is the same as the heat transfer tube 11 of the above-described embodiments (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 in the structure provided with the at least 1st protrusion 16a.

In this way, by providing the first protruding piece 16a, a part of the refrigerant flowing through the main groove 13 impinges on the first protruding piece 16a and flows strongly in the radial direction of the tube to generate a three-dimensional abnormal flow and promote new turbulence. This is because the heat transfer rate can be improved.

Specifically, the heat exchanger tube of the present invention is, for example, as shown in Fig. 6, a heat transfer tube 31 having a plurality of fin configuration portions 52A which are inclined in a Z shape (flat inclined Z shape in plan view). It can be formed as.

6 is an enlarged plan view of the inner surface of the tube having the pin configuration portion 52A that is Z-shaped in plan view. 6, only the top part of the said pin structure part 52A and the said projection piece 16 is shown typically.

The planar Z-shaped segmented fin structure 52A is provided with the first protrusion piece 16a in the pin structure part 52A, in addition to the fourth protrusion piece 16d.

The heat transfer pipe 31 has the upstream side and the downstream side in the pipe axial direction D1 opposite to the direction shown in Fig. 6, and the fourth projection piece 16d has the fin configuration portion 52A even when the coolant flows. The main groove in the spiral direction downstream side (D2d) of the configuration is protruding to the main groove 13 of the tube axial upstream side (D1u) (see the arrow shown in parentheses in Fig. 6).

In other words, the heat transfer pipe 31 having the above-described configuration has the first projecting piece 16a and the fourth projecting piece 16 even though one side of the tube axial direction D1 is upstream and the refrigerant flows in the downstream direction. Either 16d) necessarily projects to the main groove spirally downstream side D2d of the pin configuration portion 52A and to the main groove 13 on the tube axially upstream side D1u.

Therefore, the heat pipe of the above configuration can secure the excellent performance described above irrespective of the direction in which the heat exchanger is mounted and can be easily mounted in the heat exchanger without being aware of the mounting direction.

Although the present invention can be formed by various configurations as in the present embodiment described above, many embodiments can be obtained without being limited to the configurations described above.

In addition, the main groove 13 of this embodiment shall correspond between the pins of this invention in correspondence with embodiment mentioned above and the structure of this invention.

The present invention can be used as a heat transfer tube included in a heat exchanger such as a refrigerator or an air conditioner.

10: inside the tube
11, 21, 31: heat pipe
12, 42, 52: pin
12A, 42A, 52A: Pin Configuration
13: main home
14: sub-home
15: sub groove forming part
16: protrusion
D1: pipe axial direction
D2: Main groove spiral direction
D3: Sub groove spiral direction

Claims (3)

A heat transfer tube having a height H _f having a predetermined angle (β ₁ ) with respect to the tube axial direction on the inner surface of the tube smaller than an outer diameter of 7 mm with a spiral fin of 0.1 mm to 0.30 mm,
Forming a plurality of fin components on the inner surface of the tube by a spiral sub groove having a depth H _n of 0.1 mm to 0.30 mm in the fin forming direction and protruding spirally on the inner surface of the tube;
At least on a spiral downstream side of the fin configuration portion, with a protruding piece projecting at an angle α of 5 ° ≦ α <70 ° with respect to the fin, between the pins adjacent to the tube axially upstream side;
The pin constitutes a proportion of the gap between the pin configuration on the part of the pin forming direction (P _f) of more than 1.5mm, projecting to the gap (W ₂₎ between the projecting convenience pin length _{(W 1) (W 1 /} W 2 ) Heat transfer tube formed 0.3 ~ 0.9. The method of claim 1,
The height of the fin is H _f > 0.12 mm, the depth of the sub groove is H _n > 0.12 mm, the angle α with respect to the pin of the protruding piece is 20 ° or more, and the angle of the sub groove (β ₂ ) with respect to the pipe axis direction. 10 ° to 15 °, the condensation heat exchanger tube having a gap P _f between the fin components in the fin formation direction of 2 to 5 mm and a sub groove width W ₃ in the fin formation direction of 0.1 mm or more. The method of claim 1,
The pin height H _f is 0.1 mm or more, the sub groove depth H _n is 0.1 mm or more, the angle α with respect to the pin of the protruding piece is 20 ° or less, and the angle of the sub groove with respect to the tube axis direction β _The evaporation heat exchanger tube which formed ₂ ) more than 10 degrees, and the space | interval P _f in the said pin formation direction (P _f ) is 5 mm or more.

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