JPS6229713B2 - - Google Patents

Description

[Detailed description of the invention]

The present invention relates to a heat exchange tube for a condenser, and particularly to a heat exchange tube for a condenser, which is disposed almost horizontally in an atmosphere of condensable gas, the inside of the tube is used as a passage for a cooling fluid, and the outer periphery of the tube is provided with grooves and fins in the tube axis direction. The present invention relates to the structure of a conduction tube with fins in which fins are integrally arranged around the fins so as to be arranged alternately in a thermally conductive manner. It is well known that in this type of condenser heat exchange tube, it is effective to form a fine and uneven surface structure on the tube as a means to capture condensation heat transfer on the surface of the heat exchange tube. be. The factors that promote condensation heat transfer in this case are (a) an increase in the heat transfer area, and (b) a thin liquid film on the convex parts of the heat transfer surface and a thick liquid film on the concave parts due to the effect of surface tension. The two effects are that the effective heat transfer area can be ensured by the formation of the This is a problem. By the way, heat exchanger tubes with a loaf-in, in which the loaf-in is provided as a thermally conductive integral part around the outer periphery of the heat exchanger tube, are often used as a device that can achieve both of the effects (a) and (b) above. If the arrangement pitch of the loaf-ins is made too small for the purpose of (a) above, the grooves formed between adjacent loaf-ins will be filled with condensate, and this liquid layer will form a layer between the condensable gas layer and the heat transfer tube wall. As a result of preventing direct contact, the effective heat transfer area actually decreases. When the above-mentioned heat exchanger tube with loaf-in is placed horizontally with the fins aligned vertically, there seems to be no problem because the condensate filling the groove between the loaf-ins flows down the tube wall due to gravity and eventually drips. However, when considering the balance between gravity and surface tension, the narrower the width of the groove, the more the liquid layer formed in the groove expands from the bottom of the tube to the top, covering the tube wall. As a result, the effective heat transfer area decreases, and the effect of increasing the surface area by providing a large number of loaf-ins cannot be fully utilized, and the heat transfer coefficient cannot be expected to improve beyond a certain value. In this way, in view of the fact that there is a limit to the improvement of heat transfer performance by providing minute irregularities on the tube surface, and further improvement cannot be expected, the present invention provides a heat exchanger for condensers consisting of heat exchanger tubes with fins. The purpose of this is to achieve both of the effects (a) and (b) mentioned above by creating a unique structure in the pipe. Therefore, in order to achieve the above-mentioned object, the present invention is arranged almost horizontally in an atmosphere of condensable gas, the inside of the tube is used as a passage for cooling fluid, and grooves and fins are formed on the outer periphery of the tube in the direction of the tube axis. In a heat exchanger tube with fins, in which fins are integrally arranged around the fins for thermal conductivity so that the fins are alternately located, a notch groove is formed at the bottom of each of the fins in the longitudinal direction of the tube, reaching almost from the lower end to the bottom surface of the tube. At the same time, a porous plate capable of exhibiting capillary action is inserted into each of the notched grooves by intersecting each of the fins, and the upper edge of the porous plate is brought into substantially contact with the bottom surface of the tube. , and the lower edge portions are made to protrude below the lower end portions of the fins, so that the liquid existing in the grooves between the fins is transferred to the porous plates by the capillary action of the porous plates. It is possible to attract and drip immediately, thin the liquid film formed on the surface of the finned heat transfer tube, and minimize the heat transfer surface area covered by the thick liquid layer, increasing the effective heat transfer area. was accomplished. One embodiment of the present invention will be described in detail below. FIG. 1 is a partially cutaway perspective view of one example of the heat exchange tube of the present invention. The tube 2 made of a material with good thermal conductivity such as copper has grooves 3a in the longitudinal direction of the tube 2 on the outer periphery. For example, fins 3, 3, such as loaf fins, are provided around the fins 3, 3, etc. as a thermally conductive integral body so that the fins 3, . . In the example shown in Figure 1, bellows-like or spiral-shaped fins 3, 3, etc. are formed on the surface of a thick smooth copper tube by cutting or extrusion. It may also be a form in which a large number of separately processed disc fins are fitted and integrated by brazing or the like to ensure good heat conduction. The heat exchanger tube 1 with fins is used by being installed horizontally or almost horizontally in an atmosphere of condensable gas, using the inside of the tube as a passage for cooling fluid, such as water. At the bottom of the heat exchanger tube 1, a rectangular porous plate 4 whose width is slightly longer than its thickness is erected integrally so as to extend in the longitudinal direction of the tube 1 and intersect with the fins 3, 3... It is structured as follows. The porous plate 4 is formed by solidifying fine linear materials to form a plate member, and may have a large number of passages inside.The point is that the porous plate 4 has a capillary effect on liquid due to its porosity. It is sufficient as long as it exhibits the following. Further, the porous plate 4 is press-fitted into a notch groove 5 provided in each fin 3, and the notch groove 5 is provided at a predetermined location of each fin 3, that is, the heat exchanger tube 1 is placed almost horizontally. The porous plate 4 is cut from the lower end of the fins until it almost reaches the bottom of the tube 2 at a location that becomes the bottom when it is installed horizontally, and is press-fitted into the notched grooves 5. The edge is almost in contact with the bottom of the tube 2, and the lower edge is in contact with the fin 3.
It has a shape that protrudes slightly downward from the lower end. When the heat exchange tube having the above structure is used as a heat exchange tube arranged in multiple stages and in multiple rows in a shell end tube condenser, the surrounding condensable gas is transferred to tube 2 and fin 3.
It condenses and adheres to the surface of the groove 3a between the fins 3, 3.
It becomes a refrigerant liquid layer and fills the area. Then, it flows down in the groove due to its own weight, but at that time, the porous plate 4 existing at the bottom of the tube 1 exerts an effect of attracting the condensate in the groove part by its capillary force, and as a result, the flow exceeds its own weight. Since the condensed liquid is actively flowed down, the dripping of the condensed liquid is further promoted, and the liquid film of the refrigerant liquid existing on the surface of the heat exchange tube becomes thinner and the area of the attached part becomes smaller, thus increasing effective heat transfer. It is possible to increase the area and improve the condensing capacity. Therefore, in order to clarify that the heat exchange tube having the porous plate 4 exhibits substantially effective heat exchange performance, various performances were compared with a conventional heat exchange tube with loaf-in. When the test was conducted, the following results were obtained. However, the test tubes used in the comparison test were a smooth tube (A, conventional tube), a tube with a loaf-in (B, C,

[Table] Each of these tubes is installed in a sealed glass cylinder (inner diameter 102 mm, length 300 mm) filled with fluorocarbon refrigerant R-113 vapor (superheated steam at a predetermined temperature). The following measurements were carried out under operating conditions in which adjustable cooling water was distributed. ◎ Observation of the liquid film on the heat transfer tube. Install a camera and strobe on the outside of the glass cylinder to measure the tube center angle (φ) from the top (see Figure 2).
The state of the liquid film was observed by photographing with front light and backlight at 45° intervals between 0° and 90°, and at 15° intervals between 90° and 180°. ◎Measurement of wall temperature: The wall temperature was measured at 3 or 4 points on 3 or 4 cross sections in the length direction of the heat transfer portion of each heat transfer tube using a thermocouple. Note that the heat transfer surface was polished with a metal abrasive and then cleaned with acetone. Before the experiment, a vacuum pump was used to extract the air inside the glass cylinder, and during the experiment, air was removed to maintain the pressure inside the glass cylinder at over 1.4 atmospheres. The experimental range in this case is shown in Table 2 below.

[Table] ◎ Experimental results, (a) Circumferential distribution of liquid film thickness in the groove, Height of loaf-in (h) with the central angle (φ) formed between the observation point and the top of the tube as the horizontal axis As is clear from the figures in Figure 3 A to C, which plot the ratio of the liquid film thickness (δ) (distance from the bottom of the loaf-in to the lowest part of the liquid film in the groove) on the vertical axis, In all pipes B, C, and D, the liquid film thickness rapidly increases from the point where the central angle exceeds approximately 110°, and therefore, at the bottom, at the portion of 80° on each side, totaling approximately 160°, the condensed liquid layer forms between the loaf-ins. When the area becomes full, the heat transfer performance of this area is significantly impaired. On the other hand, for the heat exchange tubes BP, CP, and DP according to the present invention, the rate of increase in the liquid film thickness from the point where the central angle exceeds about 110° is slower than that of the corresponding tubes B, C, and D, respectively. especially heat exchange tubes.
In CP and DP, the increase in liquid film thickness is extremely small, and this is thought to be due to the active flow promoting action of the porous plate 4 working effectively. In addition, in FIG. 3 A to C, two solid lines parallel to the vertical axis indicate the position where the liquid fills the groove, and
The wavy line indicates the formula () below. φ1 = π-cos ^-1 (1-2y/do) = π-cos ^-1 (1-2δ/ρgr ₁ do) ...... () However, r ₁ = b/2 where δ: surface tension ρ: liquid Density r ₁ : Radius of curvature of the liquid film surface generated in the groove between the fins do : Outer diameter of the fin b : Width of the groove between the fins g : Represents the acceleration of gravity. As shown in FIG. 2, the above-mentioned _φ1 is an angle determined by the static balance between gravity and surface tension acting on the liquid filling the groove, and this value agrees well with the actually measured value. According to this equation (), the surface tension (δ)
In the case of a condensate with a large property, or in the case of a condensate with a small density (ρ), φ ₁
However, by providing the porous plate 4, the _φ1 can be increased. According to the above results, in conventional finned heat exchanger tubes B, C, and D that do not have porous plates 4, the range in which the grooves are filled with liquid increases as the fin pitch (P) decreases. It was revealed that the above range became smaller by providing the porous plate 4, indicating that the effect of the porous plate 4 was excellent. (b) Comparison of heat transfer coefficients on the condensing side, actual area-based heat transfer coefficient (α) on the condensing side
The explanation will be based on the actual measurement results shown in Fig. 4 A to C showing the relationship between the temperature and the degree of supercooling of the heat transfer surface (ΔTs, the difference between the refrigerant gas saturation temperature and the temperature at the base of the loaf-in). However, the solid line in Figure 4 represents the Nusselt equation () Nv = 0.728 (GrPr/H) = ...... () However, the saturation temperature is 50°C. As is clear from the above figure, the experimental values for test tube A agree with the solid line within a range of about ±3%. On the other hand, in test tube B which does not have the porous plate 4, the heat transfer coefficient increases by 75 to 100% in test tube B, approximately 150% in test tube C, and 80 to 110% in test tube D compared to the tube A. %. Furthermore, test tubes BP, CP, and
When comparing DP with pipes B, C, and D, the increase in heat transfer coefficient is 6 to 20%, 20 to 40%, and 10 to 50% for test pipes BP, CP, and DP, respectively. Therefore, the effect of providing the porous plate 4 is considered to be large on the low temperature difference side and small on the high temperature difference side, and this is due to the influence of the flow resistance within the porous plate 4. be judged. (c) Comparison of thermal conductance (value indicating ease of heat transfer); Figure 5 shows the results of measuring the relationship between thermal conductance C per unit length and degree of supercooling of heat transfer surface (ΔTs) for each pipe. From the figure, when ΔTs=5~5.5〓, the thermal conductance (C _F ) of smooth tube A
The results of calculating the ratio of the thermal conductance of each tube to the thermal conductance are shown in Table 3 below.

[Table] As is clear from the table above, a maximum value of 12 times was obtained, which is because the increase in surface area due to loaf-in is due to the flow-promoting effect of the porous plate 4, resulting in an increase in the effective heat transfer surface. It's nothing but a result. Next, the effects of the present invention are as follows. (a) Porous plate 4 provided at the bottom of the finned heat exchanger tube 1
actively promotes the flow of the condensate between the fins through its capillary action, resulting in a thinner liquid film and a smaller area filled with liquid, expanding the effective heat transfer area. By increasing the heat transfer coefficient (50% increase in the case of refrigerant R-113 compared to a finned heat exchanger tube without porous plate 4),
It is possible to improve the condensing capacity. (b) By achieving the effect of (a) above, there is an economical advantage that the overall shape of the heat exchanger can be made smaller. (c) In particular, the present invention has the porous plate 4 cross-contacted with the fins 3 at the bottom of the finned heat exchanger tube 1, and has its upper edge almost in contact with the bottom of the tube 2, and its lower edge Since the fins 3 protrude downward from the lower end, the grooves 3 between the fins
The condensate in the portion a is effectively drawn into the porous plate 4 up to the surface of the tube 2 due to the large capillary force, and therefore the surface of the tube 2 and the surface of the fin 3 in the lower half of the heat exchanger tube 1 are The condensate film becomes thinner, and the surface of this area contributes fully to heat transfer.

[Brief explanation of the drawing]

FIG. 1 is a partially cutaway perspective view of an example of the heat exchange tube of the present invention, FIG. 2 is an explanatory diagram showing the form of a static liquid film filling the groove between the fins, and FIGS.
C is a comparative diagram showing the relationship between the thickness of the liquid film filling the groove between the fins and the position of the heat exchanger tube in the circumferential direction, comparing each example of the heat exchanger tube of the present invention with a conventional heat exchanger tube. Figures A, B, and C are comparison diagrams showing the relationship between the degree of supercooling of the heat transfer surface and the heat transfer coefficient between each example of the heat exchange tube of the present invention and a conventional heat exchange tube. FIG. 3 is a comparison diagram showing the relationship between thermal surface supercooling degree and thermal conductance. 1... Heat exchanger tube with fins, 2... Tube, 3... Fins, 3a... Groove, 4... Porous plate, 5... Notch groove.

Claims (1)

[Claims] 1 Disposed almost horizontally in a condensable gas atmosphere, the inside of the pipe 2 is used as a passage for cooling fluid, and the pipe 2
Grooves 3a... and fins 3... are provided on the outer periphery of the tube 2 in the longitudinal direction.
In the heat exchanger tube 1 with fins, the fins 3 are integrally arranged around the fins 3 for thermal conductivity so that the fins 3 are arranged alternately,
The tube 2 is connected to the bottom of each fin 3 from the lower end.
Notch grooves 5 reaching almost the bottom surface of the tube 2 are cut aligned in the longitudinal direction of the tube 2, and porous plates 4 capable of exhibiting capillary action are interposed in the notch grooves 5 by intersecting the fins 3. A heat condenser for a condenser, characterized in that the upper edge of the porous plate 4 is brought into almost contact with the bottom surface of the tube 2, and the lower edge of the porous plate 4 is made to protrude below the lower end of the fins 3. exchange tube.