JPH037878B2 - - Google Patents

Description

Claim 1: A plate heat exchanger having a corrugated heat exchange surface, wherein the corrugations of the corrugated heat exchange surface are arranged in a plurality of rows displaced from each other by half the pitch of the corrugations, and the fins of each corrugation consists of a straight part located at the top of the waveform, a straight part located at the bottom, and an intermediate part, and these fins are arranged symmetrically and at equal angles to the axis of symmetry of the waveform to form a waveform, and the heat The intermediate portion 11 of the fin 7 of each corrugation 4 of the exchange surface 1 has two arcuate portions 12, 13, both arcuate portions 12, 13 having an equal radius of curvature.
R ₁ , both radii of curvature R ₁ are arranged facing each other, and the firm connection of the corrugations 4 with the fins 7 is such that at the junction 15 of the arcuate portions 12, 13, the thickness of the material of the corrugations 4 1. A plate-shaped heat exchanger characterized in that the heat exchanger is formed in a quadrilateral-shaped part with sides that are approximately equal to each other.

TECHNICAL FIELD The present invention relates to thermal engineering, and more particularly to planar heat exchangers.

BACKGROUND OF THE INVENTION Well-known heat exchangers use corrugated heat exchange surfaces as the heat exchange surface, the cross-section of the passages forming it having a quadrilateral and compound trapezoidal profile, which allows the flow of the heat-carrying medium to It has short fins in the direction. From a thermofluidics point of view, the most effective mechanism of heat exchange enhancement is realized in this type of corrugated surface design. When the heat-carrying medium is flowing in a state characterized by the following relationship, R _e > R _c (where R _e is the Reynolds number and R _c is the stability within the layered structure of the flow of the heat-carrying medium). is the critical value of the Reynolds number at which the first loss of . The vortex flow is dimensionally commensurate with half the thickness of the corrugated fin and is located within the wall boundary layer of the flow at a distance from the wall commensurate with half the thickness of the fin.
Therefore, the additional energy added to the flow of the heat-carrying medium to enhance the convective heat exchange in the passages of such corrugated surfaces is dissipated only for the turbulence of the wall boundary layer of the heat-carrying medium. , is not consumed to make the center of the flow turbulent.

As a result, turbulent conduction and heat flux density values are small within the narrow wall boundary layer of the passage where the main thermal resistance to thermal radiation is concentrated. However, the presence of eddies in the wall boundary layer causes the turbulent viscosity value to increase sharply, which leads to an increase in the turbulent conduction value and thus to an increase in the heat flux density value. As a result, the aforementioned passages ensure high values of the heat extraction coefficient at a moderate increase in energy consumption to enhance convective heat exchange when compared to planar passages. The above therefore makes it possible to significantly reduce the volume, mass and cost of heat exchangers using heat exchange surfaces of this type.

However, the reduction in volume and mass of plate heat exchangers is not only due to the very effective process of enhancing convective heat exchange realized within the passages of their heat exchange surfaces, but also due to the very ensured by its compact design. Known designs of heat exchange surfaces with short fins in the direction of flow of the heat-carrying medium do not allow a feature of surface compactness.

Plate heat exchangers known to those skilled in the art, e.g.
French Patent No. 1371493 (issued September 28, 1963) uses a corrugated heat exchange surface formed by a plurality of short fins arranged in the flow direction of the heat-carrying medium; are quadrilateral in shape and their corrugations have flat tops. The waveform along the flow direction of the heat-carrying medium is
It is displaced relative to the preceding waveform by half the pitch of the waveform. Waveforms arranged consecutively are
They are firmly connected to each other by a flat top.

In this way, a quadrilateral profile joint is formed with sides equal to the thickness of the corrugated material and equal to half of the base of the corrugated flat top.

In the design of the plate-shaped heat exchange surface described above,
The maximum possible compactness achieved in the surfaces formed by the quadrilateral channels cannot be ensured.

For a surface formed by a passage with a cross section in the shape of an equilateral triangle, the compactness value is approximately 2.
It's double.

Also, plate heat exchangers (e.g. 1956 10
(West German Patent No. 1104542 issued on February 2nd)
is known, which has a plate-like heat exchange surface, the corrugated fins of which are formed by upper and lower quadrilateral sections inclined at an equal angle to the axis of symmetry of the corrugations, and a middle section. It is formed. The intermediate section is arranged continuously in the flow direction of the heat-carrying medium and firmly connects the corrugations that are displaced relative to each other by half the pitch of the corrugations. The passages of the heat exchange surface have a cross-section of a compound trapezoidal profile.

However, the presence of an intermediate section on the corrugated fins, made in the form of a flat plate with a width commensurate with the corrugated pitch, does not guarantee a sufficiently high compactness of the heat exchange surface. Moreover, the presence of a flat top of the corrugations commensurate with the corrugation pitch also reduces the possibility that the same heat exchange surface of this design can be maximally compacted.

The effectiveness of a heat exchanger design is characterized not only by the thermo-hydraulic effect of the heat exchange enhancement process realized within the passages of the heat exchange surface, but also by the degree of compactness of the heat exchange surface. Therefore, the design considered here does not make it possible to ensure a high efficiency of the heat exchanger using this design.

Such a corrugated surface of a plate heat exchanger is
In heat exchanger designs that operate with counterpressure in their cavities, they have no resistance to compressive forces applied to the corrugations of the corrugated surface. This means that
The middle flat part is formed into a cantilever shape,
This is explained by the fact that it joins the upper and lower parts of the corrugation at an acute angle.

DISCLOSURE OF THE INVENTION The object of the invention is to provide a plate heat exchanger in which the structure of the fins of the plate heat exchanger ensures a reduction in its volume and mass.

This objective is achieved by a plate heat exchanger with corrugated heat exchange surfaces of the construction described below.
That is, the waveforms are arranged in a plurality of rows with half the pitch of the waveforms displaced relative to each other. Each corrugated fin has three portions, the ends of which are straight and located at the top, the bottom portions located at equal angles to the axis of symmetry of the corrugation, and the middle portions of which are rigidly connected. and the wavy fins are arranged in adjacent rows. According to the invention, the middle part of each corrugated fin of the heat exchange surface has two arcuate sections, both arcuate sections have equal radii of curvature, the radii of curvature are arranged opposite each other,
In addition, a strong connection of the wavy fins is achieved at the junction of both arcuate portions.

In order to improve the compactness of the plate heat exchanger, according to the invention, the radius of curvature of the arcuate portion of each corrugated fin is between 0.1 and 3.0 of the corrugation height.
Preferably, it is equal to twice.

In order to improve the compactness of the plate heat exchanger, the invention also provides that each corrugated fin is located at the bottom of the corrugated fin in a plane passing through the junction of the two arcuate parts of the corrugated fin. Preferably, the distance between the crests is equal to 1.05 to 3.0 times the distance between the fin crests of each corrugation located at the crests of the undulations in the same plane.

By using this proposed inventive plate heat exchanger, the volume, mass and price of the plate heat exchanger can be significantly reduced, the thermal fluid efficiency can be improved, and the continuous and Able to maintain advanced manufacturing characteristics for mass production.

[Brief explanation of the drawing]

Other objects and advantages of the present invention will become readily apparent from the following description of a plate heat exchanger with reference to the accompanying drawings.

1 is a perspective view of a plate heat exchanger with a corrugated heat exchange surface according to the invention, FIG. 2 is a top view of a corrugated heat exchange surface according to the invention, and FIG. 3 is a top view of a corrugated heat exchange surface according to the invention. FIG. 4 is a cross-sectional view of a corrugated surface according to the present invention.

BEST MODE FOR CARRYING OUT THE INVENTION As shown in FIG. 1, the corrugated surfaces 1 of the plate heat exchanger of the invention are arranged between separator plates 2 which are soldered on both sides. The crests 3 of the corrugations 4 of this corrugated surface 1 are soldered to the separating plate 2.

This proposed corrugated surface of the plate heat exchanger can be easily used for plate and fin heat exchangers of any design aimed at different applications.

As shown in FIG. 2, the corrugated surface 1 of the plate heat exchanger has rows 5 and 6 of corrugations 4 arranged one after the other in succession. The consecutively arranged waveforms 4 are displaced relative to each other by half the pitch of the waveforms 4. In each column 5 and 6,
A short continuous channel of length l' is formed in the direction of flow of the heat-carrying medium. As shown in Figure 3, each waveform 4
has a compound triangular shape defined by the corrugated fins 7. The fins 7 of the waveform 4 are formed by a straight line portion 8 located at the top 3 of the waveform 4 and a straight line portion 9 located at the bottom of the waveform 4.
The tip portions 8 and 9 of the fins 7 are aligned with the axis of symmetry 1 of the waveform 4.
It has an angle of inclination equal to 0. Waveform 4
The two straight portions 8 of the fins 7 are joined at the apex 3 of a circular waveform 4 having a radius R. The intermediate portion 11 is located between the straight portions 8 and 9 of the fin 7. The intermediate portion 11 of the fin 7 has an equal radius of curvature R ₁ , and the radius of curvature R ₁ is formed by two arcuate portions 1 arranged opposite to each other.
2 and 13. Fin 7 of waveform 4
The distance between the straight portions 9 located at the bottoms of the fins 7 of each corrugation 4 within the plane 14 passing through the junction 15 of the two arcuate portions 12 and 13 of the same plane 1
It is larger than the distance between the straight portions 8 located at the tops 3 of the fins 7 of the waveform 4 in the waveform 4.

As shown in FIG. 4, the strong coupling of the corrugations 4 arranged in succession is achieved by the arcuate portion 12 located on the fin 7 of the corrugation 4 at a distance equal to half the height h of the corrugation 4. , 13 junctions.

This strong bond takes place in a section having the shape of a quadrilateral with sides approximately equal to the thickness of the material of the corrugation 4.

As shown in FIG. 3, in order to improve the compactness of the corrugated surface 1, the radius R ₁ of the arc of the arcuate portions 12, 13 of the fins 7 of each corrugation 4 is
It should be 0.1 to 3.0 times the height h.

To further improve the compactness of the corrugated surface 1, two arcuate sections 12 of the fins 7 of the corrugation 4,
The distance between the tip portions 9 of each fin 7 of the waveform 4 at the bottom of the waveform 4 measured in a plane 14 passing through the junction 15 of the waveform 4 is the distance between each fin of the waveform 4 at the top of the waveform 4 in the same plane 14. Straight line part 8 of 7
1.05 to 3.0 times the distance between them.

The plate heat exchanger of the present invention described above operates in the following manner.

When two gaseous heat-carrying media are flowing through a plate heat exchanger, the corrugations 4 of the corrugated surface 1 of one of the cavities of this flat heat exchanger, as shown in FIG. The heat from the hotter heat-carrying medium passing through the fins 7 flows out to the separating plate 2 via the soldered tops 3 of the corrugations 4, as shown in FIG. This heat is transferred from this separation plate 2 to this separation plate 2
It is transferred to a relatively cool heat-carrying medium through the fins 7 of the corrugations 4, which have their tops 3 soldered on the other side.

When the heat-carrying medium flows through the passages of the corrugated heat exchange surface 1, swirls are generated on the edges of the fins 7 of the corrugations 4. This vortex exists within the wall boundary layer of the flowing heat-carrying medium, and its magnitude is commensurate with the thickness of the wall boundary layer. By the moving flow, these eddies are carried away in the flow direction along the fins 7 of the corrugations 4 and gradually dissipate on their walls. Due to the short length of the passage in waveform 4,
The vortex does not have time to decay and is generated again in the subsequent path of waveform 4. The hydrodynamic conditions across the section of waveform 4 are not similar. In the cross section of the passage of the corrugation 4 defined by the straight portions 8 of the fins 7, the generated vortices are larger than in the cross section of the corrugation 4 defined by the straight portions 9 of the fins 7. along the flow of the heat-carrying medium at a rate of attenuation. After all, the straight part 8 of the fin 7 is
This is because they are arranged closer to each other than the straight portions 9. As a result, in the cross section of the waveform 4 defined by the straight line portion 8 of the fin 7,
The flow of the heat-carrying medium is more easily transformed into a laminar state by proximity to or merging of adjacent laminar boundary layers. This causes the vortex system to decay more rapidly on the wall of the straight section 8 located at the top of the fin 7 when compared to the situation existing on the wall of the straight section 9 located at the bottom of the fin 7 .

The joining of the straight parts 8 and 9 of the fin 7 with the waveform 4 is as follows:
By virtue of what is taking place in the arcuate portions 12, 13, the hydrodynamic conditions across the section of the waveform 4 described above change gradually, rather than intermittently.

That is, as shown in FIGS.
flows in the direction of First, the physical mechanism of the flow of the heat carrier medium in the passages to the heat exchange surface 1 will now be explained.

The construction of the heat exchange surface 1 of the plate heat exchanger according to the invention, as already mentioned, actually consists of a system of rigidly connected short rows of complex triangular shapes.
Within this system, each row of short passages is offset by one-half short passage fin pitch relative to the previous row, as shown in FIGS. 2 and 4. Furthermore, the third
As shown in the figure, the inlet and outlet edges of the fins of the short passage have an acute triangular shape, and the middle part of the short passage is formed by connecting two circular arcs with equal curvature and facing in opposite directions. There is. The heat exchange surface of the invention therefore consists of a system of complex triangular passage openings. As is known, at the beginning, where the channel walls have sharp edges, the heat-carrying medium moves more than three times faster than in a vortex flow. This happens at the acute entrance edge of the passage where a three-dimensional spiral vortex is formed (in a circular passage).
This is because the heat carrier medium flows separately. After the flow of the heat-carrying medium impinges on the wall of the channel, a new adjoining thermal and hydrodynamic layer begins to form, the thickness of which layer is smaller than the thickness of the layer at the beginning of the channel. Furthermore, in the area where the heat-carrying medium separates and in the area where the thermal layer and the hydrodynamic layer merge, the degree of turbulence is higher (1.5 to 2 twice)
This is a characteristic. As a result, the value of the turbulent heat transfer increases rapidly at the beginning of the passage compared to the hydrodynamically stable flow of the heat-carrying medium in the adjoining layers forming the flow. This is the factor that causes the value of heat conduction at the beginning of the passage to increase rapidly. However, in order to increase the heat exchange efficiency mentioned above, additional power must be consumed to incorporate the heat carrier medium. but,
The heat-carrying medium can be reduced by reducing the cross-sectional dimensions of the area where the flows separate and the size of the inlet vortices that occur, and by moving these vortices as close to the walls as possible while still maintaining sufficient strength. moreover,
To obtain zero-gradient flow by reducing the plate thickness, the cross-sectional dimensions of the separation zone are reduced to reduce the vortex magnitude while simultaneously increasing the vortex strength. Additionally, the added turbulence is dispersed over a shorter length from the entire edge, and the vortex size is equal to half the thickness of the plate. In the structure of the invention, i.e., in the structure of short passages staggered in a complex triangular manner, in which turbulent flow of the heat-carrying medium occurs, a similar zero-gradient flow occurs in the wide part adjacent to the base of the cross-section of the passage. be observed. That is, the adjacent layers constituting the heat-carrying medium flow in adjacent fins of the passage do not meet in said wide portion. Thus, at the inlet of the short passage, the degree of turbulence that is sufficiently high that occurs in the wide part of the cross-section of the short passage decreases slightly during flow from the short passage to the outlet. In sufficiently adjacent narrow sections of the short passageway, adjacent fins are more adjacent to each other. This is the factor which causes the adjoining layers constituting the heat-carrying medium stream to merge in the above-mentioned specific operating conditions of turbulent flow of the heat-carrying medium.

As is known, the flow rate of the heat-carrying medium in the adjoining layers is greater than the flow rate in the wide section of the passage (in the flow core). As a result, in narrow sections of the passageway, as the heat-carrying medium flows through the fins of the passageway, separation of the heat-carrying medium flow increases the size and reduces the strength of the vortices. As a result, the degree of turbulence generated in this area is lower than the degree of turbulence in the wider section of the passage. Turbulence disappears faster and therefore over a shorter distance of the passage. This results in
Considerable changes in the degree of turbulence can be expected from the narrow section of the passageway from the inlet to the outlet. Further, the fluid pressure resistance coefficient in the narrow section of the passage is larger than in the wide section of the passage. This results in a non-uniform distribution of the flow rate of the heat-carrying medium over the cross-section of the passage,
This is why the value of the heat transfer coefficient varies around the cross section of the passage (the minimum value defines around the narrow part of the cross section of the passage). The conventional analysis of the process of passage configuration of the heat exchange surface of the present invention has been
It is based on the results of fundamental research on the process of change in heat retention in -corne-red passages).

In the cross-section of three-cornered passages with heat exchange surfaces according to the invention there are two zones of heat-carrying medium flow, these two zones having a higher turbulent flow in the wider part of the cross-section. It is characterized by a higher degree of turbulence. It is also characterized by less turbulent flow and lower degree of turbulence in the narrow part of the cross section. On the boundaries of these cross-sections, unless active measures are taken, significant static pressure gradients develop that create conditions that significantly shift the heat-carrying medium flow in the normal direction at the average velocity of the heat-carrying medium. This movement of the heat-carrying medium covers most or all of the passages, which significantly increases the power consumption, since the degree of heat exchange is practically unchanged. Therefore, in order to eliminate such undesirable phenomena and reduce the static pressure gradient in the cross-section of three-cornered passages by gently connecting the narrow and wide parts of the cross-section of the passage, the third As shown in the figure, the intermediate portion 11 of the fins 7 of each corrugation 4 of the heat exchange surface 1 consists of two arcuate portions 12, 13 having equal curvature and facing in opposite directions.

As a result, the distance between the fins of each passage is minimized due to the fin structure in which adjacent passage rows are rigidly connected by a flat connecting plate at the intermediate portion 11 of the fin 7, thereby minimizing the heat exchange surface. Can be made smaller. This is the gist of the invention.

Furthermore, the physical mechanism of the heat exchange process in the second and subsequent rows of passages is analyzed.

As mentioned above, at the exit of the first row of passages,
Furthermore, the degree of turbulence at the entrance to the second row of passages varies depending on the height of the cross-section of the passages. Furthermore, the degree of turbulence of the heat-carrying medium flow at the entrance to the narrow section of the passages in the second row of passages is greater than at the entrance to the wide section of the passages. Apart from this, the degree of turbulence of the heat-carrying medium flow as it enters the wide section of the passageway in the second row of passages avoids the fins that define the narrow section in the previous first row of passages. It becomes even higher due to the adjacent vortices. From the above explanation, it follows as follows. That is, the gradient of the degree of turbulence of the heat-carrying medium at the entrance to the second row of passages is reduced over both the wide and narrow sections. Furthermore, the degree of turbulence of the heat-carrying medium flow at the entrance to the narrow section of the passageway cross-section is higher than at the entrance to the wide section. As a result, the constant velocity lines in the narrow part of the passage and the second passage row are denser and have a velocity gradient having a larger value than in the narrow part of the cross section of the first passage row. Therefore, the formation of higher strength adjoining vortices at the inlet is fundamentally desirable at the narrow section of the second passage bank than at the inlet to the narrow section of the cross section on the first passage side. Therefore, the gradient of the heat transfer coefficient around the cross-section of the second row of passages is reduced, which is a factor that increases the efficiency of the heat exchange process compared to the first row of passages. The results of studying the process of enhancing the heat exchange effect using short passages staggered from each other are as follows. That is, immediately after the second row of passages, the degree of heat exchange in each subsequent row reaches a certain high level, and the gradient of heat transfer around the passages provided after the second row of passages becomes minimum.

In addition, the fin structure of a known heat exchanger was developed in 1983.
131051) and the present invention are as follows.

That is, in Utility Model Application Publication No. 50-131051, the heat exchange surface consists of individual narrow wavy bands (fins 2a, 2
b), each band having a triangular waveform. The waveforms are positioned sequentially and without any gaps in the flow direction 9, 10 of the heat medium, and the waveform of each band is shifted by 1/2 pitch with respect to the waveform of the previous band. Each fin 2a has a waveform with a triangular cross section,
2b has a straight line portion connecting the radial portion of the top of the waveform and the base of the waveform. The design of this surface, when conceptually created as a hypothesis, is thermally and hydraulically effective. However, it is clear that this design is not practicable. This is because the heat exchange surface becomes strained and its thermal fluid pressure efficiency decreases. This is due to the following fact. This means that individual narrow profile bands are produced and subsequently transported to an assembly station.
The inability to maintain the pitch of the corrugated fins within predetermined limits in assembling the individual bands to the heat exchange surface. This occurs because generally thin sheet materials (foils) are used in the manufacture of corrugated bands. As a result of variations in the pitch of the corrugated fins during the manufacturing process of each individual corrugated band, when the corrugated bands are assembled into a heat exchanger, the corrugated fins of adjacent corrugated bands are precisely 1/2 the pitch of the adjacent fins. It is impossible to shift it. Also, the effective cross-sectional area of the corrugations of the assembled heat exchanger varies, leading to non-uniform distribution of the heat medium flow and considerably reducing the efficiency of the thermal fluid pressure of said surface.

The differences between the heat exchange surface of the present invention and the heat exchange surface of Utility Model Application No. 50-131051 are as follows. In the present invention,
Each fin 7 of the waveform 4 is not a straight line connecting the radial part of the top of the waveform and the base of the waveform, but two fins connected to the above-mentioned radial part of the waveform at an equal angle to the axis of symmetry of the waveform. It has a straight part. These two straight parts of each fin 7 are connected by a middle part 11 of the fin 7 over a distance of 1/2 the height of the corrugation 4. heat exchange surface 1
The intermediate portion 11 of the fin of each waveform 4 has two arcuate portions 12 and 13. These two arcuate portions 12, 13 have equal curvature and are arranged such that the curvatures point towards each other. The fins 7 of adjacent rows of corrugations 4 have a rectangular cross-section and are joined at the junction 1 of the arcuate portions 12, 13 by bridges with sides approximately equal to the thickness of the material of the corrugations.
A rigidly bonded integral corrugated surface at 5 can be made from sheet material having a width equal to the width of the heat exchange surface. This design is characterized by ease of manufacture in the assembly and manufacture of the heat exchanger, compared to the previously known heat exchange surface design. Implementation of the heat exchange surface design of the present invention results in the highest thermo-hydraulic efficiency values currently known. This is because in adjacent rows of short passages, a strictly constant waveform shift value of 1/2 pitch of the waveform fin can be obtained. This allows a uniform distribution of the heating medium along the short path of the heat exchange surface.

Furthermore, all other things being equal, the heat exchange surface of the present invention is more compact than the heat exchange surface of Utility Model Application No. 131051/1983. This means that
This is explained by the fact that for the pitch of the corrugations 4 and fins 7 of equal height, the length of the fins is greater in the present invention than in the cited example. Furthermore, this fact makes it possible to improve the thermal fluid efficiency of the heat exchange surface of the present invention.
Improve the thermal fluid efficiency of the heat exchange surface of No. 131051.

The best result of convective heat exchange enhancement within the passages of the inventive heat exchange surface 1 of this proposed design is that each corrugation 4
This is achieved when the value of the radius of curvature of the arcuate portions 12, 13 of the fins 7 is equal to 0.1 to 3.0 times the height of the waveform 4.

The smaller the hydrodynamic difference in the flow across the section of the corrugation 4, the better the result, and the improvement in the compactness of the corrugated surface 1 is due to the junction of the two arcuate portions 12, 13 of the fins 7 of the corrugation 4. 15 of the straight line portion 9 of each fin 7 of the waveform 4 at the bottom of the waveform 4 and the distance of each fin 7 of the waveform 4 at the top 3 of the waveform 4 in the same plane 14. This is achieved by reducing the difference between the distances between the straight sections 8.

The best result of the enhancement of convective heat exchange in the passages of the corrugated surface 1 of the design of the invention described above and the simultaneous improvement of its compactness is the junction of the two arcuate sections 12, 13 of the fins 7 of the corrugation 4. plane 14 passing through 15
The distance between the straight portions 9 of each fin 7 of the waveform 4 located at the bottom of the waveform 4 measured within the same plane 1
This is achieved if it is equal to 1.05 to 3.0 times the distance between the straight portions 8 of the fins 7 of each corrugation located at the top 3 of the corrugation 4 in the corrugation 4.

It is often operated at Kα _nio (where K is the heat transfer coefficient of the heat exchanger and α _nio is the minimum heat extraction coefficient of one of the two heat exchange surfaces of the heat exchanger). The use of the corrugated surfaces of the present invention in plate heat exchanger designs reduces heat exchanger volume, mass, and The price can be reduced.

Industrial Applicability The plate heat exchanger of the present invention can be used for gas-to-gas,
Air cooling for condensing and vaporizing various fluids, which can be used in liquid-to-gas and liquid-to-liquid different heat exchangers, and also used in automotive, aviation, power engineering, refrigeration, chemical and other industries It can also be used in condensers and vaporizers.