JPH0228798B2 - - Google Patents

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention provides a heat transfer surface channel that increases the ratio of forced convection heat transfer coefficient to working fluid driving power and also has the function of a differential user. The present invention relates to a diffuser type heat exchanger having:

[Conventional technology]

In order to improve the heat transfer coefficient of the heat transfer surface of the heat exchanger, increase the heat transfer surface area per channel volume, and suppress the increase in flow resistance during forced convection,
Up to now, a wide variety of heat transfer surface shapes and flow path configuration methods have been developed.

[Problem that the invention seeks to solve]

However, forced convection heat transfer has a deep causal relationship with flow resistance, and when the heat transfer coefficient is increased, it is inevitable that the flow resistance, that is, the power that drives the fluid, will increase. Therefore, much effort is still being put into techniques for improving heat transfer, and although there is much to be seen in this development, there has not been much improvement from the perspective of the ratio of heat transfer coefficient to power. For economical heat exchanger planning,
Currently, based on the accumulation of experience and the refinement of calculations, we are trying to optimize the design of various heat transfer surfaces and the flow paths formed by them without waste. On the other hand, if the main aim is to reduce the driving power, it is effective to increase the cross-sectional area of the flow path and reduce the average flow velocity, but this will directly increase the flow path volume, that is, the equipment, and , the idea of lowering the flow velocity has been given little consideration in the past because it results in a significant decrease in the heat transfer coefficient and requires an increase in the required heat transfer surface area, which ultimately results in a result that does not contribute much to the reduction of the driving force. I couldn't help it.

In view of the above points, it is an object of the present invention to obtain a diffuser type heat exchanger that can increase the heat exchange rate per fluid driving power.

[Means for solving problems]

In the diffuser type heat exchanger according to the present invention, the cross-sectional area of the fluid flow path increases as it goes downstream in the direction of the mainstream from the inlet of the flow path without decreasing along the way, and The length of the channel is larger than the hydraulic diameter of the inlet cross section of the channel, and the static pressure of the working fluid at the outlet is greater than the static pressure at the inlet. The peripheral wall of the flow path in a range larger than the inlet is configured as a heat transfer surface.

[Effect]

In this invention, since the fluid flow path is formed from an expanded flow path in which the average flow velocity gradually decreases, the diffuse user effect of this expanded flow path and the occurrence of separated flow in the diffuse user flow path, which was found as a result of experiments, can be avoided. By utilizing heat transfer characteristics, it is possible to increase the heat exchange rate per fluid drive power, and
It is also possible to incorporate the heat exchanger according to the invention into piping or duct systems that require a diffuser.

〔Example〕

First, the principle of the diffuser type heat exchanger according to the present invention will be explained based on FIGS. 3 and 12 to 16. FIG. 3a is a perspective view of a conventional direct flow path 1 having a rectangular cross section that does not expand in cross-sectional area, and FIG. 3b is a perspective view of the conventional direct flow path 1 having a rectangular cross section that does not expand in cross-sectional area, and FIG. 2 is a perspective view showing an enlarged flow channel 2 according to the present invention, and in each figure, b ₀ is the entrance width of the flow channel, W is its height, L is a heat transfer section, and T is one side wall forming a heat transfer surface. The heat transfer coefficient, the amount of heat transfer, and the pressure loss in the section L of the T-wall in the DC channel 1 can be easily calculated using conventional knowledge. However, if the inlet width b ₀ is increased to the outlet width b ₁ (b ₁ >b ₀ ) while maintaining the above-mentioned direct flow path shape, the heat transfer coefficient decreases due to the decrease in the average flow velocity U of the flow path.

Now, as shown in Fig. 3b, if the T-wall of the DC channel 1 is widened from the inlet width b ₀ to the outlet width b ₁ in section L to form an enlarged channel 2 in which the static pressure increases, the opening angle θ is ( It is given by the ratio of the exit width b ₁ -inlet width b ₀ ) and the section length L. In the case of a flow path with such a simple form, if the opening angle θ or the flow path enlargement ratio AR ₁ =b ₁ /b ₀ is set appropriately, there will be no separation in the fluid flow on the T wall, which is the heat transfer surface. Flow occurs, its heat transfer coefficient h _s and pressure drop rate
K _d comes to exhibit unique characteristics as shown below.

That is, when the opening angle θ of the enlarged channel 2 is set to 8 degrees, the heat transfer coefficient h _s on the T wall side is shown in Fig. 3b.
When expressed as a ratio to the heat transfer coefficient _h
As shown in the graphs in FIGS. 2 and 13, the heat transfer coefficient h _s on the enlarged channel 2 side has a larger value than the heat transfer coefficient h _x calculated as the direct flow channel. AR in the figure is the local value of the channel expansion ratio in the X direction. This means that the actual heat transfer coefficient h _s does not decrease as much as predicted from the decrease in the average flow velocity due to the expansion of the flow path, and the unique phenomenon when separation occurs in the expanded flow path 2. It can be said to represent heat transfer characteristics.
In other words, experiments have shown a previously unknown fact that in the enlarged channel 2, if the opening angle θ is set appropriately, the local heat transfer coefficient h _s in the X direction is larger than the heat transfer coefficient h _x of a direct channel with the same cross-sectional area. This shows that the results have become clear.

FIG. 14 shows the measured value of the pressure recovery rate when the opening angle θ=8°. Channel inlet (Ai=
b ₀ XW), and the difference between _the ideal pressure recovery rate curve and the measured pressure recovery rate curve is shown at the bottom of the figure as the loss rate K _d of the expanded flow path 2. Note that the loss rate K _d also includes friction loss in the enlarged flow path.

When evaluating the heat transfer performance due to pump power based on the above-mentioned unique relationship between heat transfer characteristics and pressure loss, consider the ratio of the amount of heat transferred per temperature difference between the fluid and the heat transfer surface to the pump power. Opening angle θ=
If ηd is the ratio when the opening angle θ is changed with respect to the case of a DC path of 0°, there is a region where ηd>1. This means that the characteristics of the amount of heat transferred per pump power may work more favorably by enlarging the flow path than by a direct flow path.

Figure 15 shows the entrance width b ₀ =
0.03m, height W = 0.10m, section L = 1.10m, and the amount of heat transfer per 1°K temperature difference for the power calculated by changing the opening angle θ to 8°, 12°, 16°, and 20°. ,
This shows an example of calculating the ratio ηd for the case of a DC channel with an opening angle θ = 0°. When the ratio ηd > 1, it is better to use an expanded channel (θ > 0°) than a direct channel (θ = 0°). This shows that this is an advantageous area. The same figure also shows that the opening angle θ = 8° is the most advantageous, and when the opening angle θ = 8°, the Reynolds number
Calculation example in which R _e0 (the representative length is defined by the inlet width b ₀ , the hydraulic diameter D _i given by the height W, the average flow velocity U ₀ at the channel entrance, and the kinematic viscosity coefficient of the working fluid) is changed. is shown in FIG.

Therefore, in the enlarged channel 2 as shown in FIG. 3b, for example, b ₀ =0.03 m, W = 0.10 m, R _e0 =5~
By setting the length of L in the X direction to 0.4 to 0.6 m or more at 15×10 ⁴ and θ=8°, the amount of heat transferred to the power can be made more advantageous than that of a DC path.

Hereinafter, embodiments of the diffuse user type heat exchanger of the present invention will be explained based on the drawings. Fig. 1 shows the first embodiment of the diffuse user type heat exchanger that constitutes the simplest type of heat transfer surface flow path. It is something. Working fluid a flows in from the inlet of a channel whose cross-sectional area is A _i and whose hydraulic diameter is D _i in the direction of the arrow, and if the length of the line connecting the center of the cross section to the outlet is L, then the channel length is L
While flowing through the section, while conducting heat transfer with the surrounding wall, all or part of which is a heat transfer surface,
It reaches the outlet of the channel with a cross-sectional area of A _e and is discharged. Since the cross-sectional area of this flow path increases downstream from the inlet toward the outlet, naturally A _e is always larger than A _i . Also, L is always larger than D _i . The cross-sectional shape of the channel may be any shape.

FIG. 2 shows a second embodiment of a diffuser type heat exchanger in which the inlet cross section and the outlet cross section are not similar, and the inflow direction and outflow direction of the working fluid constitute a heat transfer surface flow path that changes angles. It is something. The cross-sectional area of the flow path monotonically increases from the inlet to the outlet,
The length L of the channel centerline is greater than the inlet hydraulic diameter D _i .

FIG. 4 shows a third embodiment of a diffuser type heat exchanger having a more complicated heat transfer surface flow passage cross-sectional shape. Although the inlet cross section is particularly shown with hatching, it satisfies the conditions for the flow path shown in the description of the first embodiment, and a part or all of its peripheral wall is used as a heat transfer surface.

FIG. 5 shows a fourth embodiment of a diffuser type heat exchanger using a corrugated heat transfer plate as part of the peripheral wall of the flow path. As shown by the imaginary line in the figure, the cross-sectional area of the imaginary flow path formed by the flat side wall and the plane connecting the peaks of the corrugated plate gradually increases in the fluid flow direction. L is larger than the hydraulic diameter of the flow path inlet cross section hatched.

FIG. 6 shows a fifth embodiment of a diffuser type heat exchanger having a plurality of fins in the flow path in the mainstream direction. A virtual flow path formed by a plane connecting the tips of the fins is in the shape of a conical diffuser.

FIG. 7 shows a sixth embodiment of a diffuser type heat exchanger in which three single heat transfer surface passages each having a rectangular cross section are provided in parallel, and each passage is similar to the first embodiment. The conditions shown in the example description are met. All or part of the partition walls, including the partition walls of the flow path, can be used as heat transfer surfaces.

FIG. 8 is a perspective sectional view showing a seventh embodiment of a diffuser type heat exchanger in which heat transfer surface channels having rectangular cross sections are arranged in series and in parallel. The channel inlet portion of the k-th heat transfer surface channel among the individual channel elements is indicated by hatching. Naturally, this flow path also satisfies the conditions shown in the description of the first embodiment.
This embodiment also allows the partition wall and the peripheral wall to serve as heat transfer surfaces.

FIG. 9 shows an eighth embodiment of a diffuser type heat exchanger in which two working fluids a and b having different temperatures flow oppositely and can exchange heat via a heat transfer surface wall H. . The channels a and b pass through are each enlarged channels, and L ₁ and L ₂ are respectively
It is larger than D _i1 and D _i2 .

Figure 10 shows two working fluids a and b with different temperatures.
This shows a ninth embodiment of a diffuser type heat exchanger in which the two flow in opposite directions to exchange heat. fluid a
flows through a plurality of cylindrical enlarged channels in a direction in which the cross-sectional area increases, and exchanges heat with the fluid b flowing outside using the circumferential wall H of the channels as a heat transfer surface. The section length L is
It is larger than the largest one among the hydraulic diameters D _io (assuming there are n pieces) of the cross section of the inlet of the channel through which a flows.

FIG. 11 is a partial perspective view showing a tenth embodiment of a diffuser type heat exchanger equipped with a finned plate heat exchanger plate. By overlapping the flat plate heat transfer surface shown by the imaginary line on the portion of the finned plate shown in the figure, a heat transfer surface flow path similar to that of the eighth embodiment is formed.

Therefore, when using the diffuser type heat exchanger according to the present invention, the heat transfer of the flow in the diffuser caused by the intermittent and steady separated flow generated by the expanded flow path is Due to its unique properties, the cross-sectional area of the flow path increases, so even if the cross-sectional average flow velocity decreases,
The degree of decrease in heat transfer coefficient can be kept small. On the other hand, as the flow velocity decreases, the tube friction resistance decreases significantly, so the heat transfer coefficient relative to the flow resistance increases, and this, combined with the gradual and continuous increase in channel volume of the diff user, increases the heat transfer surface flow. It is possible to significantly improve the ratio of heat transfer coefficient to driving power while minimizing the increase in volume of the passage. In addition, in most conventional heat transfer surface channels, the static pressure of the working fluid decreases during the passage through the channel, but in the heat transfer surface channel of the present invention, the static pressure is increased. The plan can be designed. That is, since pressure recovery, in other words, conversion of dynamic pressure to static pressure (effect as a diff user) can be achieved by the expanded flow path, the heat exchanger according to the present invention can be used in piping that requires a diff user. Alternatively, by incorporating it into a duct system, it becomes possible to perform heat exchange while also playing the role of a differential user, which in turn contributes to making the entire system more compact and improves economic efficiency.

In the eighth embodiment, the partition wall between the two working fluid passages is configured as a heat transfer surface wall H, but it is also possible to add, for example, the opposite wall surface or side wall surface as a heat transfer surface. For example, since the heat transfer coefficients on these wall surfaces are considered to be larger than the heat transfer coefficient h _s on the heat transfer surface wall H side, there is an advantage that the heat transfer efficiency can be further improved.

〔Effect of the invention〕

As described above, according to the present invention, the relationship between the amount of heat transfer and the pump power is determined by the characteristic heat transfer characteristics and hydraulic characteristics (pressure drop characteristics) when separated flow occurs due to the enlarged flow path. This has the effect of producing a gain that is better than previously thought and can additionally provide a differential user effect.

[Brief explanation of drawings]

Fig. 1 is a perspective view showing a first embodiment of this invention, Fig. 2 is a perspective view showing a second embodiment of this invention, and Figs. 3 a and b both explain the principle of this invention. 3A is a perspective view showing a direct flow path, FIG. 3B is a perspective view showing an enlarged flow path, and FIG. 4 is a view equivalent to FIG. 1 showing a third embodiment of the present invention. ,
FIG. 5 is a view corresponding to FIG. 1 showing a fourth embodiment of the invention having a corrugated heat transfer plate, and FIG. 6 is a view corresponding to FIG. 1 showing a fifth embodiment of the invention comprising a finned tube. FIG. 7 is a perspective view showing a sixth embodiment of the present invention having a plurality of enlarged flow channels, and FIG. 8 is a perspective view showing a seventh embodiment of the present invention having a plurality of enlarged flow channels. 9 is a perspective view showing an eighth embodiment of the present invention having a plurality of enlarged flow channels, and FIG. 10 is a perspective view showing a ninth embodiment of the present invention having a plurality of enlarged flow channels. FIG. 11 is a perspective view showing a tenth embodiment of the present invention having a plurality of enlarged flow channels;
2 to 16 are explanatory diagrams for explaining the heat transfer characteristics in the diffuser flow path. A _i ...Heat transfer surface channel inlet area, A _e ...Heat transfer surface channel outlet area, D _i ...Hydraulic diameter at heat transfer surface channel inlet,
H, T...Wall (heat transfer surface), L...Length of the line connecting the center points of the cross section of the heat transfer surface flow path (length of the flow path in the main flow direction), a...Working fluid, b...Action A working fluid with a different temperature from fluid a. In each figure, the same reference numerals indicate the same or equivalent parts.

Claims (1)

[Scope of Claims] 1 The cross-sectional area of the fluid flow path increases as it goes downstream in the direction of the mainstream from the inlet of the flow path, without decreasing along the way, and the cross-sectional area of the flow path in the mainstream direction increases. the length is greater than the hydraulic diameter of the inlet cross section of the channel;
It is formed from an enlarged flow path in which the static pressure of the working fluid at the outlet is greater than the static pressure at the inlet, and the peripheral wall of the flow path in the range where the static pressure of the enlarged flow path is greater than that at the inlet is configured as a heat transfer surface. This is a diffuse user type heat exchanger. 2. The diffuser type heating device according to claim 1, wherein the expanded flow path is formed by an imaginary surface obtained by connecting tips of a plurality of convex portions formed on the inner surface in the mainstream direction. exchanger. 3. The diffuser type heat exchanger according to claim 2, wherein the convex portion is formed of a corrugated plate. 4. The diffuser type heat exchanger according to claim 2, wherein the convex portion is formed by a fin. 5. Claim 1, wherein the fluid passage is branched or divided in parallel within the fluid passage, and each passage formed after the branching or division constitutes an enlarged passage. 5. A diffuser type heat exchanger according to any one of items 1 to 4.