JP2021134952A - Boiling type heat transfer pipe - Google Patents

Abstract

To provide a boiling type heat transfer pipe that enables high efficient heat transfer in the overall operation region by maintaining a high heat transfer rate from a heat transfer pipe to liquid refrigerant, even in a case where a heat exchanger is operated at low load.SOLUTION: A boiling type heat transfer pipe comprises a plurality of rows of fins projecting radially outward from a pipe outer peripheral surface and arranged in an annular shape or a spiral shape along a circumferential direction. The fin has a leg part erected on the pipe outer peripheral surface, and a pair of overhanging parts whose tips outside in the radial direction of the leg part extend in opposite directions in a pipe axial direction. Between adjacent fins, a cavity part continuous in the circumferential direction is defined by the overhanging part of one fin and the overhanging part of the other fin that project so as to approach each other. The pair of overhanging parts is provided with a plurality of recess parts recessed radially inward to connect the overhanging parts, along the circumferential direction.SELECTED DRAWING: Figure 3

Description

The present invention relates to a boiling water reactor used in a heat exchanger that boils a refrigerant outside the tube.

The boiling type heat transfer tube is incorporated in the evaporator of a steam compression type refrigerator such as a turbo refrigerator and a screw refrigerator, which is a heat exchanger, and is immersed in a liquid refrigerant (for example, freon, liquid nitrogen, etc.). Used to heat and boil liquid refrigerants. Various heat transfer surface shapes have been proposed as this type of boiling type heat transfer tube. For example, as in Patent Document 1, the liquid refrigerant in the cavity is promoted to boil and the liquid refrigerant on the outer surface of the tube is promoted. In addition, there are those that promote the turbulence of the vaporized medium to improve the heat transfer performance.

Japanese Unexamined Patent Publication No. 4-236097

By the way, in recent years, _{in response to environmental problems such as reduction of CO 2} emissions, inverter control that can finely control equipment to improve energy efficiency has been widely introduced. However, even when the inverter is controlled, when the load of the refrigerator is low, the heat transfer coefficient from the heat transfer tube to the liquid refrigerant tends to be significantly lowered when the liquid refrigerant of the evaporator is heated and boiled. This is because the bubble generating force decreases as the thermal driving force decreases. Therefore, when the load of the refrigerator is low, that is, when it is used in a low heat flux region, it is necessary to increase the load in order to compensate for the above-mentioned decrease in heat transfer coefficient, and it is difficult to exchange heat with high efficiency. there were.

Therefore, the present invention enables highly efficient heat transfer in the entire operating range by maintaining a high heat transfer coefficient from the heat transfer tube to the liquid refrigerant even when the heat exchanger is operated with a low load. It is an object of the present invention to provide a boiling type heat transfer tube.

According to one embodiment of the present invention, the following configurations are provided.
It has a plurality of rows of fins that protrude outward in the radial direction from the outer peripheral surface of the pipe and are arranged in an annular shape or a spiral shape along the circumferential direction.
The fin has a leg portion erected on the outer peripheral surface of the pipe and a pair of overhanging portions whose radial outer tips of the leg portion extend in opposite directions in the pipe axial direction.
Between the fins adjacent to each other in the pipe axis direction, a hollow portion continuous in the circumferential direction is defined by the overhanging portion of one of the fins and the overhanging portion of the other fins that project so as to approach each other. Being done
A boiling type heat transfer tube in which a pair of overhanging portions that define the hollow portion are provided with recesses that are recessed inward in the radial direction to connect the overhanging portions at a plurality of locations along the circumferential direction.

According to the present invention, by maintaining a high heat transfer coefficient from the heat transfer tube to the liquid refrigerant, highly efficient heat transfer is possible in the entire operating range.

It is a partial cross-sectional perspective view of the boiling type heat transfer tube which concerns on this invention. It is a partially enlarged plan view on the outer peripheral surface of the boiling type heat transfer tube of the 1st configuration example. FIG. 3 is a cross-sectional view showing a cross section taken along line III-III of FIG. An enlarged cross-sectional view of adjacent fins is shown. It is sectional drawing which is orthogonal to the tube axis direction of the boiling type heat transfer tube arranged in the liquid refrigerant. It is explanatory drawing which shows a mode that a liquid refrigerant is heated on the outer peripheral surface of a pipe, and evaporative bubbles are generated. It is explanatory drawing which shows typically the mode that the liquid refrigerant flows along the circumferential direction of a cavity part. It is a partially enlarged plan view on the outer peripheral surface of the boiling type heat transfer tube of the 2nd configuration example. It is the schematic of the test apparatus used for the evaluation of the heat transfer performance of a boiling type heat transfer tube. It is a graph which shows the relationship between the heat flux and the total heat transfer coefficient in Test Example 1 and Test Example 6. It is a graph which shows the relationship between the heat flux and the heat transfer coefficient outside the tube in Test Example 1 and Test Example 6. It is a graph which shows the relationship between the heat flux and the total heat transfer coefficient in Test Example 2 and Test Example 6. It is a graph which shows the relationship between the heat flux and the heat transfer coefficient outside the tube in Test Example 2 and Test Example 6. It is a graph which shows the relationship between the heat flux and the total heat transfer coefficient in Test Examples 3, 4 and 5 and Test Example 6. It is a graph which shows the relationship between the heat flux and the heat transfer coefficient outside the tube in Test Examples 3, 4 and 5 and Test Example 6.

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.
<First configuration example>
FIG. 1 is a partial cross-sectional perspective view of the boiling type heat transfer tube according to the present invention, FIG. 2 is a partially enlarged plan view of the outer peripheral surface of the boiling type heat transfer tube, and FIG. Is.
The boiling type heat transfer tube 100 is a metal tube material continuous in one direction, and includes a plurality of rows of fins 11 along the circumferential direction on the outer peripheral surface of the tube, and a plurality of rows of ribs 13 on the inner peripheral surface of the tube. Further, the boiling type heat transfer tube 100 may have a smooth cylindrical inner surface by omitting the rib 13 on the inner peripheral surface of the tube.

Here, in the present specification, the axial direction of the boiling type heat transfer tube 100 is also represented as the Ax direction, and the circumferential direction is also represented as the T direction.

As shown in FIGS. 1 and 2, the fins 11 are formed so as to project radially outward from the outer peripheral surface 15 of the pipe, and are arranged on the outer peripheral surface 15 of the pipe in an annular shape or a spiral shape along the circumferential direction T. When the rows of fins 11 are spiral, they may have one row or a plurality of rows.

As shown in FIG. 3, the fin 11 has a leg portion 11a and a pair of overhanging portions 11b and 11c. The leg portion 11a is erected on the outer peripheral surface 15 of the pipe, and the pair of overhanging portions 11b and 11c have bifurcated ends on the outer side in the radial direction of the leg portion 11a and extend in opposite directions of the pipe axial direction Ax. It is formed. Between the fins 11A and 11B adjacent to each other in the pipe axis direction Ax, the overhanging portion 11b of one fin 11A and the overhanging portion 11c of the other fin 11B that project so as to approach each other are continuous in the circumferential direction T. The hollow portion 17 to be formed is defined. On the other hand, a gap is formed between the tip of the overhanging portion 11b of the fin 11A and the tip of the overhanging portion 11c of the fin 11C. That is, the plurality of fins 11 have a mixture of positions where the cavity 17 is defined and positions where the cavity 17 is not defined.

FIG. 4 shows an enlarged cross-sectional view of adjacent fins 11A and 11B.
At least a part of the cavity 17 along the circumferential direction T is provided with a recess 19 in which a pair of overhanging portions 11b and 11c defining the cavity 17 are recessed inward in the radial direction. The recess 19 is formed by, for example, pressing the disc 27 to deform the overhanging portions 11b and 11c inward in the radial direction. As a result, the overhanging portions 11b and 11c are connected to each other to form a closed space as the hollow portion 17. The minimum height H1 of the recess 19 is smaller than the height H2 of the fin 11. Such recesses 19 are divided and arranged at a plurality of locations along the circumferential direction (see FIG. 2). For details on the method of forming the leg portions 11a and the overhanging portions 11b and 11c of the fin 11, refer to, for example, Japanese Patent Application Laid-Open No. 7-151485.

In the portion where the concave portion 19 of the hollow portion 17 is formed, it is preferable to form a convex portion 31 in which the pair of overhanging portions 11b and 11c protrude inward in the radial direction in the hollow portion 17. Since the overhanging portions 11b and 11c are recessed inward in the radial direction, the cross-sectional shape of the cavity portion 17 continuous in the circumferential direction in the pipe axis direction is such that the cross-sectional area of the recess 19 in the circumferential direction is the other circumferential position. It is smaller than the cross-sectional area of.

The recesses 19 are preferably arranged at regular intervals in the circumferential direction, but the present invention is not limited to this, and the recesses 19 may be arranged with a specific periodicity or may be arranged at random. The recess 19 shown in FIG. 2 is formed in an elongated shape in a plan view, and its major axis direction is parallel to the tube axis direction Ax, but the major axis direction may be formed perpendicular to the tube axis direction Ax.

In the following description, the portion of the cavity 17 in which the recess 19 is formed is also referred to as a “narrow portion” to distinguish it from the portion in which the recess 19 is not formed. In the case of the example of FIG. 3, the position where the recess 19 is formed by the fins 11A and 11B is a narrow portion, and the other fins 11 are not narrow portions.

(Action)
FIG. 5 is a cross-sectional view of the boiling type heat transfer tube 100 arranged in the liquid refrigerant perpendicular to the tube axis direction.
The boiling type heat transfer tube 100 having the above configuration is arranged in the liquid refrigerant 21 in the evaporator with the tube axis direction horizontal, and the heated water 23 is supplied into the boiling type heat transfer tube 100. Then, the liquid refrigerant 21 is heated, and evaporation bubbles 25 are generated on the outer peripheral surface 15 of the boiling type heat transfer tube 100. As shown by the arrows in FIG. 5, the generated evaporation bubbles 25 move along the peripheral surface from the lower side to the upper side of the outer peripheral surface 15 of the pipe, and the liquid of the liquid refrigerant 21 is moved from the upper side of the boiling type heat transfer tube 100. Head to the face. Further, some of the evaporated bubbles 25 head toward the liquid surface through the gaps between the overhanging portions 11b and 11c of the fins 11 shown in FIG.

FIG. 6 is an explanatory view showing how the liquid refrigerant 21 is heated on the outer peripheral surface 15 of the pipe to generate evaporative bubbles 25.
The liquid refrigerant 21 flows into the inside of the cavity 17, and is filled with the liquid refrigerant 21. When the heating water 23 for heating is supplied into the boiling type heat transfer tube 100, the fins 11A and 11B heat the liquid refrigerant 21 in the cavity 17. At this time, the heating layer 21a of the liquid refrigerant 21 is formed along the inner side surfaces 33 of the fins 11A and 11B facing the cavity 17.

Further, when the evaporative bubbles 25 of the liquid refrigerant 21 generated by heating pass through the inner side surfaces 33 of the fins 11A and 11B, the inner side surfaces 33 are not in contact with the liquid refrigerant 21, and the liquid refrigerant passes through the inner side surfaces 33. It becomes difficult for heat to be transferred to 21. Therefore, ideally, the inner side surfaces 33 of the fins 11A and 11B are always covered with the liquid refrigerant 21 to generate the heating layer 21a, and the evaporation bubbles 25 are allowed to flow along the circumferential direction T inside the heating layer 21a. It is better to do it.

According to this embodiment, the liquid refrigerant 21 of the heating layer 21a gradually moves in the circumferential direction so as to follow the flow of the evaporated bubbles 25, and a new liquid (new liquid refrigerant 21) is constantly continuously applied to the inner surface 33. Be supplied. As a result, the heat exchange efficiency on the inner surface 33 is improved.

Then, in the narrow portion where the concave portion 19 of the hollow portion 17 is formed, the overhanging portions 11b and 11c are recessed inward in the radial direction as described above to form a convex portion 31 protruding inward in the radial direction. do. The convex portion 31 agitates the liquid refrigerant 21 flowing through the cavity 17 (arrow M in FIG. 6) to make the flow of the liquid refrigerant 21 in the heating layer 21a turbulent. As a result, the replacement of the new liquid in the heating layer 21a is promoted, and the heat exchange efficiency can be further improved.

Further, the convex portion 31 is formed on the inner surface of the cavity portion 17 on the outer side in the radial direction. Therefore, the convex portion 31 does not obstruct the flow of the evaporated bubbles 25 that move along the outer peripheral surface 15 of the boiling type heat transfer tube 100 as shown in FIG. Therefore, the flow resistance of the liquid refrigerant 21 at the fin root portion becomes small, and the cavity portion 17 does not hinder the gas-liquid two-layer flow between the evaporative foam 25 flowing from the lower side to the upper side of the boiling type heat transfer tube 100 and the liquid refrigerant 21. The liquid refrigerant 21 can be supplied in the entire circumferential direction of the above. Further, in the hollow portion 17 above the boiling type heat transfer tube 100, gas-liquid separation is performed by the buoyancy of the evaporative foam 25, and the supply of the liquid refrigerant 21 is promoted.

FIG. 7 is an explanatory view schematically showing how the liquid refrigerant flows along the circumferential direction of the cavity 17.
As shown in FIG. 7, in the narrow portion where the cross-sectional shape in the pipe axis direction changes along the circumferential direction and the recess 19 is formed, the cross-sectional area of the cavity 17 is larger than that in the other circumferential positions. Becomes smaller. Therefore, the flow path of the flow F of the liquid refrigerant 21 flowing through the cavity 17 is narrowed in the narrow portion, and the flow velocity decreases in the narrow portion due to the increase in the flow resistance (flow velocity V2 <V1). Therefore, the liquid refrigerant 21 of the heating layer 21a continues to be heated at the same position in the hollow portion 17 on the front side in the flow direction of the narrow portion. As a result, even when the amount of heat transferred from the inner surface 33 of the cavity 17 to the liquid refrigerant 21 per unit time is small, the heat input to the liquid refrigerant 21 of the heating layer 21a is accumulated to generate evaporation bubbles. It will be easier.

That is, by arranging a plurality of narrow portions along the circumferential direction of the cavity portion 17, a plurality of small compartments 35 separated by each narrow portion are formed in the cavity portion 17. The circumferential movement of the liquid refrigerant 21 is suppressed in each of the sub-compartments 35, and the heating of the liquid refrigerant 21 in each sub-compartment 35 is promoted. In this way, the liquid refrigerant 21 is locally heated at a plurality of locations along the circumferential direction of the cavity 17, so that evaporation bubbles can be efficiently generated.

According to the boiling type heat transfer tube 100 having the present configuration, the generation of evaporation bubbles is synergistically promoted by the stirring effect of the liquid refrigerant 21 by the convex portion 31 and the local heating effect by the subsection 35. This makes it possible to generate evaporative bubbles with high efficiency even under low heat flux conditions.

<Second configuration example>
FIG. 8 is a partially enlarged plan view of the outer peripheral surface of the boiling water reactor of the second configuration example.
The boiling type heat transfer tube 100 of the second configuration example is the same as that of the first configuration example except that the recess 19A is formed in a slender shape in a plan view and the major axis direction thereof is inclined from the tube axis direction Ax. Therefore, in the following description, the same members or the same parts are given the same reference numerals to simplify or omit the description.

The concave portion 19A formed in the hollow portion 17 is formed so that the long axis thereof is inclined from the pipe axis direction Ax and the narrow portion is inclined from the circumferential direction T. Therefore, the cross-sectional area of the tube axial cross section of the cavity 17 changes continuously or intermittently along the circumferential direction in the narrow region. Ideally, the cross-sectional area of the cavity 17 gradually shrinks along the circumferential direction and then gradually expands. In other words, the cross-sectional area of the cavity 17 increases or decreases along the circumferential direction. According to such an inclined narrow portion, the flow resistance when the liquid refrigerant 21 passes is lower than that in the case of the first configuration example, and the stirring effect of the liquid refrigerant 21 is enhanced.

The inclination angle between the long axis of the recess 19A and the axial direction Ax is 10 ° to 80 °, preferably 20 ° to 60 °, more preferably 25 ° to 40 °, and the fin size and the type of liquid refrigerant 21. It can be changed to the optimum angle as appropriate.

The boiling type heat transfer tube 100 is manufactured of copper, copper alloy, aluminum, aluminum alloy, iron, and a metal material having thermal conductivity such as stainless steel material, titanium, titanium alloy, and particularly heat such as copper or copper alloy. A material having a high conductivity is still more preferable.

Next, the results of evaluating the heat transfer performance of the boiling water reactors having the above-mentioned configurations will be described.
FIG. 9 is a schematic view of a test device used for evaluating the heat transfer performance of a boiling water reactor.
The test device is a thermosiphon type heat exchanger in which a condenser 53 and an evaporator 55 of a stainless steel shell and tube heat exchanger are connected by a pipe, and the refrigerant naturally circulates due to a temperature difference. The condenser 53 and the evaporator 55 are tanks having an inner diameter of 333 mm and a length of 974 mm. Three test tubes 54 are installed in the center of the evaporator 55, and the effective measurement length of the test tubes 54 is 974 mm. Heated water is stored in the tank 56. Here, the heated water in the tank 56 is cooled by the cooling coil 59 in the tank 56. The heated water supplied from the tank 56 is heated by the heater 57. By being heated by the heater 57, the heated water supplied to the test tube 54 is controlled to a constant temperature. This heated water is supplied to the inside of the test tube from the inlet which is one end of the test tube 54. The heated water discharged from the outlet at the other end of the test tube 54 is returned to the tank 56. The evaporator 55 is filled with a liquid refrigerant, and the test tube 54 is immersed in the liquid refrigerant in the evaporator 55. Then, the liquid refrigerant heated by the heated water inside the test tube 54 evaporates and becomes refrigerant vapor, which is supplied to the condenser 53.

In the condenser 53, a heat transfer tube 52 (effective length 974 mm) in which the end of the tube is fixed by an O-ring is horizontally installed, and the refrigerant vapor supplied from the evaporator 55 is directly connected to the heat transfer tube 52 at the refrigerant vapor inlet. A baffle plate is installed to prevent it from hitting. The brine liquid supplied from the tank 51 flows into the heat transfer tube 52, and the refrigerant vapor is condensed on the outer surface of the heat transfer tube 52. This condensed liquid refrigerant returns to the evaporator 55 by gravity.

The evaporation pressure is measured from a pressure outlet provided on the upper part of the evaporator 55 using a semiconductor strain gauge type pressure transmitter (measurement error: ± 0.05% of the set span). The temperature at the inlet and outlet of the heated water is a mixture of a platinum resistance temperature detector (Pt 100Ω, JIS-A class) calibrated to ± 0.05 ° C with a quartz thermometer in advance and attached to both ends of the test tube 54. Insert into a vessel and measure the mixed average temperature. The flow rate of heated water is measured with an electromagnetic flow meter (measurement error: 1.5% of the reading value). The test conditions are shown in Table 1 below.

The heat transfer coefficient was calculated by each of the following mathematical formulas. First, the heat transfer amount Qc of heated water _{was calculated} by Equation 1.

Here, G _c is the volume flow rate of the heated water, ρ _c is the density of the heated water, c _pc is the constant pressure specific heat of the heated water, T _Cout is the hot water outlet temperature, and T _Cin is the heated water inlet temperature. As the physical property value of the heated water, the value calculated by the arithmetic mean value of the temperature measurement value of the heated water inlet / outlet using the correlation formula prepared from the physical property value table was used. The logarithmic mean temperature difference ΔT _m is defined by the following mathematical formula 2.

Here, T _s is the refrigerant saturation temperature. The refrigerant saturation temperature T _s is calculated from the measured value and the refrigerant property values of the evaporation pressure.

_{Then, the total heat transfer coefficient K 0} , which is the heat transfer coefficient based on _{the outer surface area A 0} of the fin processed portion of the test tube, was obtained by the following mathematical formula 3.

Here, the outer surface area A ₀ of the test tube fin processed portion is based on the envelope surface calculated from the _{outer diameter D 0 of the} test tube fin processed portion, as shown in the following mathematical formula 4.

Here, l is the effective heat transfer length of the test tube. Further, the heat flux q _{0 based on} the outer surface area was calculated by the following mathematical formula 5 with _{the outer surface area A 0 of the test tube fin processed portion as a reference.}

The extratube heat transfer coefficient h ₀ was calculated by the following mathematical formula 6.

Here, h _i tube side heat transfer coefficient, A _i is the inner surface area of the test試管fin processing unit, R _wall is tube wall thermal resistance, they are determined as follows.

Internal surface area A _i of the test試管fin processing unit is defined by the following equation 7.

Here, _Dimax is the maximum inner diameter of the test tube fin processed portion. Further, the tube _wall thermal resistance R wall is defined and obtained by the following mathematical formula 8.

Here, k _wall is the thermal conductivity of the tube wall. Further, the heat transfer coefficient h _i inside the tube and the _{Nusselt number Nu i} inside the tube are obtained by defining them by the following equation 9 on the assumption that the functional form is expressed by the equation of Dittus-Boelter.

Here, C _i is a coefficient determined empirically, k _CH thermal conductivity of the heating water, Pr _CH is the Prandtl number of the heated water. Further, the Reynolds number Re _CH of the heated water was determined by defining it by the following mathematical formula 10.

Here, _VCi is the average flow velocity of the heated water, and ν _C is the kinematic viscosity coefficient of the heated water. Incidentally, C _i values for determining the tube side heat transfer coefficient h _i was determined by testing in advance in advance using the Wilson plot (Wilson-plot) technique.

Table 2, Table 3 shows the dimensions and C _i values of the test試管of the prepared test examples 1-6.

Each test tube is provided with the fins and recesses described above on the outer peripheral surface of the tube, and the ribs shown in Table 3 are provided on the inner peripheral surface of the tube. In Test Examples 1 and 2, the shape of the recess is particularly different. Test Example 1 is a heat transfer tube of a second configuration example having a parallelogram recess shown in FIG. 8, and Test Example 2 is a heat transfer tube of the first configuration example having a rectangular recess shown in FIG. In Test Examples 3 to 5, the circumferential length of the rectangular recess 19 is changed in a plan view. In Test Example 3, the long axis of the concave portion is in the pipe axis direction, in Test Examples 4 and 5, the long axis is in the circumferential direction, and in Test Example 5, the circumferential length of the concave portion is set to be the largest. That is, the circumferential length of the recess increases in the order of Test Example 3, Test Example 4, and Test Example 5. However, in Test Examples 3 to 5, the arrangement pitch in the circumferential direction is adjusted so that the abundance ratio of the recesses to the peripheral length is equal.

Test Example 6 shows a test example of a test tube not having the configuration according to this embodiment, and is used as a comparison target with Test Examples 1 to 5. Here, JP-A-2017-20736 is used as the test conditions for Test Example 6.

The test試管Test Example 1-6 described above, the overall heat transfer coefficient K ₀ showing the heat transfer performance from the radially inner pipe circumference radially outward of the tube outer peripheral surface from the tube circumference radially outward _{The extratube heat transfer coefficient h 0} , which indicates the heat transfer performance of the above, was determined. The results are shown in FIGS. 10 to 15.

FIG. 10 is a graph showing the relationship between the _{heat flux q 0} and the overall heat transfer coefficient K ₀ in Test Example 1 and Test Example 6. FIG. 11 is a graph showing the relationship between the _{heat flux q 0} and the extratube heat transfer coefficient h ₀ in Test Example 1 and Test Example 6.

As shown in FIG. 10, the overall heat transfer coefficient of Test Example 1 was larger than the overall heat transfer coefficient of Test Example 6 over the entire area of the heat flux. Especially in a low heat flux region (for example, 20 kW / m ² or less), in Test Example 6, the degree of decrease in the overall heat transfer coefficient increases as the heat flux decreases, but in Test Example 1, the heat flux is 5 kW / m ² Even so, the total heat transfer coefficient was ^{maintained at 8.5 kW / (m 2} K) or higher, and the decrease in the total heat transfer coefficient at the time of low heat flux was suppressed.

Further, as shown in FIG. 11, for the extravascular evaporation heat transfer coefficient h _0, overall Test Example 1 is larger than the test example 6, reduction of extravascular evaporation heat transfer coefficient h ₀ at low heat flux , Test Example 1 was greatly improved as compared with Test Example 6.

FIG. 12 is a graph showing the relationship between the _{heat flux q 0} and the overall heat transfer coefficient K ₀ in Test Example 2 and Test Example 6. FIG. 13 is a graph showing the relationship between the _{heat flux q 0} and the extratube heat transfer coefficient h ₀ in Test Example 2 and Test Example 6.

As shown in FIG. 12, the overall heat transfer coefficient of Test Example 2 was larger than the overall heat transfer coefficient of Test Example 6 over the entire area of the heat flux. Especially in the low heat flux region (e.g. 20 kW / m ² or less), but the degree of decrease overall heat transfer coefficient with decreasing heat flux in Test Example 6 increases the heat flux in Test Example 2 is 5 kW / m ² Even so, the total heat transfer coefficient was ^{maintained at 8.5 kW / (m 2} K) or higher, and the decrease in the total heat transfer coefficient at the time of low heat flux was suppressed.

Further, as shown in FIG. 13, regarding the out-of-tube heat transfer coefficient h ₀ , Test Example 2 was larger than Test Example 6 as a whole, and the decrease _{in the out-of-tube heat transfer coefficient h 0 at the time of low heat flux was} , Test Example 2 was greatly improved as compared with Test Example 6.

FIG. 14 is a graph showing the relationship between the _{heat flux q 0} and the overall heat transfer coefficient K ₀ in Test Examples 3, 4, and 5. FIG. 15 is a graph showing the relationship between the _{heat flux q 0} and the extratube heat transfer coefficient h ₀ in Test Examples 3, 4, and 5.

As shown in FIG. 14, the overall heat transfer coefficient K ₀ is larger than that of Test Example 6 in all of Test Examples 3, 4 and 5, and the decrease at the time of low heat flux is suppressed.
As shown in FIG. 15, the extratube heat transfer coefficient h ₀ is lower in Test Example 5 than in Test Example 6 when the heat flux is 30 kW / m ² or more, but even in that case, it is 20 kW / (m ² K). It is maintained above. Further, in Test Examples 3, 4 and 5, the decrease at the time of low heat flux is also suppressed.

From the above, the overall heat transfer coefficient K ₀ and the extratube heat transfer coefficient h ₀ of Test Examples 1 to 5 are maintained at values substantially equal to or higher than the case where the heat flux of Test Example 6 is 30 kW / m ^2. Therefore, even when the heat flux is low, it is possible to suppress a decrease in _{the overall heat transfer coefficient K 0} and the extratube heat transfer coefficient h _0. Therefore, even when the heat exchanger using the boiling type heat transfer tube of the present invention is operated with a low load, the heat transfer coefficient from the heat transfer tube to the liquid refrigerant is maintained high, so that the heat transfer coefficient is high in the entire operating range. Efficient heat transfer is possible.

As described above, the present invention is not limited to the above-described embodiment, and can be modified or applied by those skilled in the art based on the combination of the configurations of the embodiments with each other, the description of the specification, and the well-known technique. This is also the subject of the present invention and is included in the scope for which protection is sought.

As described above, the following matters are disclosed in this specification.
(1) A plurality of rows of fins are provided, which protrude outward in the radial direction from the outer peripheral surface of the pipe and are arranged in an annular shape or a spiral shape along the circumferential direction.
The fin has a leg portion erected on the outer peripheral surface of the pipe and a pair of overhanging portions whose radial outer tips of the leg portion extend in opposite directions in the pipe axial direction.
Between the fins adjacent to each other in the pipe axis direction, a hollow portion continuous in the circumferential direction is defined by the overhanging portion of one of the fins and the overhanging portion of the other fins that project so as to approach each other. Being done
A boiling type heat transfer tube in which a pair of overhanging portions that define the hollow portion are provided with recesses that are recessed inward in the radial direction to connect the overhanging portions at a plurality of locations along the circumferential direction.
According to this boiling type heat transfer tube, in the cavity defined by a pair of overhangs, recesses that are recessed inward in the radial direction to connect the overhangs are narrowed at a plurality of locations along the circumferential direction. Formed and placed. Then, the heating of the liquid refrigerant in each of the small compartments is promoted by suppressing the movement of the liquid refrigerant in the circumferential direction in each of the small compartments divided by each of the narrow portions of the hollow portion. In this way, the liquid refrigerant is locally heated at a plurality of locations along the circumferential direction in the cavity, and evaporation bubbles are efficiently generated.

(2) The boiling type heat transfer tube according to (1), wherein the pair of overhanging portions that define the cavity portion form a convex portion that projects the cavity portion inward in the radial direction.
According to this boiling type heat transfer tube, the liquid refrigerant whose convex portion flows through the hollow portion can be agitated to make the flow of the liquid refrigerant turbulent. As a result, it is possible to promote the replacement of new liquid of the liquid refrigerant in the cavity and further improve the heat exchange efficiency.

(3) The boiling type heat transfer tube according to (1) or (2), wherein the recess is formed in a slender shape in a plan view, and the major axis direction thereof is parallel or perpendicular to the tube axis direction.
According to this boiling type heat transfer tube, the flow velocity of the liquid refrigerant flowing through the cavity can be efficiently reduced, and the heating of the liquid refrigerant can be promoted between the narrow portions formed by the recesses.

(4) The boiling type heat transfer tube according to (1) or (2), wherein the recess is formed in a slender shape in a plan view, and the major axis direction thereof is inclined from the tube axis direction.
According to this boiling type heat transfer tube, the flow resistance of the liquid refrigerant flowing through the cavity in the recess is reduced, the liquid refrigerant is heated between the narrow portions formed by the recess, and the liquid refrigerant flows in the cavity. Can be well balanced.

(5) The boiling according to any one of (1) to (4), wherein the pair of overhanging portions that define the hollow portion form a closed space in a cross section in the pipe axis direction at the position where the recess is formed. Type heat transfer tube.
According to this boiling type heat transfer tube, the cross-sectional area of the cross section in the axial direction of the tube becomes smaller than the other peripheral positions at the position where the recess is formed due to the formation of the closed space, and the flow of the liquid refrigerant is narrowed. As a result, the flow of the liquid refrigerant is suppressed and the liquid refrigerant is easily heated at the same position, and the generation of evaporative bubbles is promoted.

(6) The boiling type heat transfer tube according to any one of (1) to (5), wherein the recesses are arranged at regular intervals in the circumferential direction.
According to this boiling type heat transfer tube, the small sections formed by the recesses between the narrow portions become uniform in the circumferential direction, and the liquid refrigerant is heated under the same conditions. As a result, the liquid refrigerant is uniformly heated in the circumferential direction, and heat exchange with less unevenness can be performed.

(7) 7. Boiling type heat transfer tube.
According to this boiling type heat transfer tube, by providing ribs on the inner peripheral surface of the tube, the contact area with the liquid flowing in the tube is increased, and the heat transfer efficiency can be improved.

(8) A plurality of rows of fins are provided, which protrude outward in the radial direction from the outer peripheral surface of the pipe and are arranged in an annular shape or a spiral shape along the circumferential direction.
The fin has a leg portion erected on the outer peripheral surface of the pipe and a pair of overhanging portions whose radial outer tips of the leg portion extend in opposite directions in the pipe axial direction.
Between the fins adjacent to each other in the pipe axis direction, a hollow portion continuous in the circumferential direction is defined by the overhanging portion of one of the fins and the overhanging portion of the other fins that project so as to approach each other. Being done
The cavity is a boiling water reactor defined so that the cross-sectional area increases or decreases in the circumferential direction.
According to this boiling type heat transfer tube, the hollow portion defined by the pair of overhanging portions is defined so as to be continuous in the circumferential direction and the cross-sectional area in the circumferential direction increases or decreases. Then, as the cross-sectional area in the circumferential direction increases or decreases, the cavity is composed of a narrow portion and other portions, and the circumferential movement of the liquid refrigerant is suppressed in each of the subsections separated by each narrow portion. This promotes the heating of the liquid refrigerant in each subsection. In this way, the liquid refrigerant is locally heated at a plurality of locations along the circumferential direction in the cavity, and evaporation bubbles are efficiently generated.

11, 11A, 11B Fins 11a Legs 11b, 11c Overhangs 13 Ribs 15 Tube outer circumference 17 Cavities 19, 19A Concave 21 Liquid refrigerant 23 Heating water 25 Evaporative bubbles 31 Convex 33 Inner side 35 Subsection 52 Heat transfer tube 53 Condenser 54 Test tube 55 Evaporator

Claims (8)

It has a plurality of rows of fins that protrude outward in the radial direction from the outer peripheral surface of the pipe and are arranged in an annular shape or a spiral shape along the circumferential direction.
The fin has a leg portion erected on the outer peripheral surface of the pipe and a pair of overhanging portions whose radial outer tips of the leg portion extend in opposite directions in the pipe axial direction.
Between the fins adjacent to each other in the pipe axis direction, a hollow portion continuous in the circumferential direction is defined by the overhanging portion of one of the fins and the overhanging portion of the other fins that project so as to approach each other. Being done
A boiling type heat transfer tube in which a pair of overhanging portions that define the hollow portion are provided with recesses that are recessed inward in the radial direction to connect the overhanging portions at a plurality of locations along the circumferential direction. The boiling type heat transfer tube according to claim 1, wherein the pair of overhanging portions that define the cavity portion form a convex portion that projects the cavity portion inward in the radial direction. The boiling type heat transfer tube according to claim 1 or 2, wherein the recess is formed in a slender shape in a plan view, and the major axis direction thereof is parallel or perpendicular to the tube axis direction. The boiling type heat transfer tube according to claim 1 or 2, wherein the recess is formed in a slender shape in a plan view, and the major axis direction thereof is inclined from the tube axis direction. The boiling type heat transfer tube according to any one of claims 1 to 4, wherein the pair of overhanging portions that define the cavity portion form a closed space in a cross section in the tube axial direction at the position where the recess is formed. The boiling type heat transfer tube according to any one of claims 1 to 5, wherein the recesses are arranged at regular intervals in the circumferential direction. The boiling type heat transfer tube according to any one of claims 1 to 6, which is formed so as to project radially inward from the inner peripheral surface of the tube and includes a plurality of rows of ribs arranged in an annular shape or a spiral shape along the circumferential direction. It has a plurality of rows of fins that protrude outward in the radial direction from the outer peripheral surface of the pipe and are arranged in an annular shape or a spiral shape along the circumferential direction.
The fin has a leg portion erected on the outer peripheral surface of the pipe and a pair of overhanging portions whose radial outer tips of the leg portion extend in opposite directions in the pipe axial direction.
Between the fins adjacent to each other in the pipe axis direction, a hollow portion continuous in the circumferential direction is defined by the overhanging portion of one of the fins and the overhanging portion of the other fins that project so as to approach each other. Being done
The cavity is a boiling water reactor defined so that the cross-sectional area increases or decreases in the circumferential direction.

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