HOLLOW PLATE HEAT EXCHANGERS

Abstract

One of these heat exchangers (76) consists of a stack of thin-metal walled hollow platelets (7S1-J5), 12 cm long and 5 wide. Each of these walls has a central region stiffened by alternating bosses with steep slopes, situated between two connection regions. Each wall is made by pressing then cutting an appropriate sheet of metal (aluminium 0.3 mm thick). The edges of the two fin walls form steps, symmetrically welded, the height of each step determining the internal half-thickness of a fin. Each platelet connection region ends in a narrow mouth with a cross section that has the same surface area as the embossed central region, and is welded to the edges of a slot made in an external manifold (80-82). The thickness of the internal channel of a platelet is about 0.4 mm when the fluid concerned is a liquid (water) and that of the spaces between the platelets is 7 mm when the other fluid is a gas (air). By hot pressing or thermoforming, sheets of glass or polymer may also be used but the performance is not as good. A radiator can be made of several exchangers mounted in parallel on each side of two flat main manifolds. Applications: any heat exchangers with high volumetric conduction, low weight and low pumping and ventilating power.

Description

The invention relates principally to a heat exchanger, made up of a stack of hollow plates, which has a very high level of performance, that is, very high bulk conductance combined with a small front surface area, low mechanical power requirements for the propulsion of the fluids involved and the possibility of handling liquid and/or gaseous fluids at relatively high differential pressures and temperatures.

The invention relates secondarily to heat exchangers similar to the above, overall with a lower level of performance than it has, but likely to be better suited to certain specific applications.

Hollow plate heat exchangers have much higher performance levels than the solid fin exchangers on the radiators for heat engines. In fact, for the same bulk conductance, in such a liquid/gas heat exchanger, the gap between adjacent hollow plates is much greater than the gap between solid fins. As a result, the weight of the former, their bulk, front surface area and power consumed (pumping of liquid(s) and/or blowing of gas) are significantly lower than those of the latter. And yet solid metal fin heat exchangers continue to be universally used in a number of fields. Under these circumstances, when heat engines are fitted with normal water/air radiators, the front surface area (main cross-section) of these radiators measures approximately 0.3 dm² per kW to be discharged, whilst their operation consumes mechanical power (ventilation and pumping) equal to up to 10% of the thermal power to be dissipated, or even more if the temperature differences are small. This demonstrates the advantage of hollow plate heat exchangers.

Heat exchangers made up of a one-piece stack of hollow plates made from polymer, glass or metal, are described in European Patent EP 1 579 163 B1, held by TET. The method of producing one of these exchangers consists of manufacturing, by thermoblowing of a polymer parison, an accordion-shaped blank provided with biconvex bellows, with walls embossed with steep alternating bosses, and then carrying out a controlled compression of this blank. Following this compression, these bellows take on the final form of a one-piece stack of rigid hollow plates, with a narrow internal channel, connected to two internal manifolds. Such one-piece polymer heat exchangers provide completely satisfactory results for numerous applications, as long as the bulk conductance sought remains in the mid-range (20 W/° C./dm³ at most) and the fluids handled are at moderate differential pressure (0.1 MPa at most) and not at a very high temperature (<100° C.). In fact, in a number of specific cases, their advantages in terms of weight, cost, bulk and power consumption (3 to 5% of the thermal power to be discharged) largely compensate for this limited performance, particularly when the initial temperature difference between the two fluids in question is relatively small (<60° C.).

This one-piece heat exchanger, made up of hollow plates with embossed polymer walls, has multiple advantages. Its walls combine a certain stiffness and a certain thinness, which are mutually contradictory characteristics, such that its weight, cost and bulk are low. Despite the laminar flow of the cooling liquid, its narrow internal channel allows for good thermal conductance between the liquid and the wall of the hollow plate. On the other hand, its embossed walls generate relatively significant turbulence of the flow of air between the plates, which allows for the gap between them to be increased greatly. This considerably reduces the energy needed to propel the air between the plates. In addition, this significant turbulence in the air circulating between the plates increases the apparent thermal conductivity of the air and therefore the overall thermal conductance of the exchanger.

However, experience has shown that this two-stage technique of thermoblowing and then controlled compression of the biconvex bellows of a polymer blank comes up against limited results if one seeks to increase the desired performance level and in particular the bulk conductance of the heat exchanger thus produced. Indeed, with this technique, it is impossible to completely control the two-stage manufacturing process of a one-piece stack of hollow plates, with regard to the thicknesses of the internal channel and of the walls of the plates, even though these thicknesses are decisive parameters for the bulk conductance value of the exchanger. In practice, for the internal channel of the hollow plates, this results in a thickness with an average value of around two millimetres, with a dispersion of at least thirty percent. With regard to the thickness of their walls, the average value is in the region of one millimetre and the dispersion is approximately fifty percent, this dispersion being mainly due to the uneven narrowing of the wall during the thermoblowing of the blank.

In addition to the limitation of performance attributable to these thickness problems, it must be noted that the presence of the internal manifolds of the stacked hollow plates adds another aspect to such limitation: the creation of a central channel, common to all of these hollow plates, which allows for the direct rapid flow of the liquid between these two manifolds. As a result, this relatively large central channel scarcely contributes to the desired heat exchange.

High-performance cooling devices for various applications are described in international application WO 2006/010822, filed by TET. In these devices, the radiators are heat exchangers produced in accordance with the method in TET's European patent. For one particular application (the cooling of the exhaust gases from a diesel engine with a view to recycling them), provision is made in the application for using a one-piece heat exchanger having hollow metal plates, capable of withstanding much higher differential pressure and temperature than those to which a one-piece polymer exchanger can be subjected. To this end, the metal accordion-shaped blank for the heat exchanger had to be manufactured by hydroforming. This known technique seems promising in the field of one-piece heat exchangers having hollow metal plates but, at the moment, it has not yet been possible to implement it correctly and, moreover, it is itself limited with regard to its theoretical efficiency. Indeed, as the thermal resistivity of the cooling liquids, water or oil, is high, the thermal resistance of the layer of liquid, flowing laminarly in such hollow plates, is inevitably high, given an average thickness of at least 2 mm. This removes a large part of the advantage of the low thermal resistance that would be provided by the metal walls envisaged.

Consequently, another way of producing metal heat exchangers had to be developed for several specific applications, in particular for the application initially envisaged and, more generally, for any device involving the possibility of having very high-performance heat exchangers. To this end, these new metal heat exchangers must have weight, bulk, front surface area and mechanical power consumption that are as low as those of the one-piece exchangers described above. They must do this whilst having a much higher bulk conductance (at least 100 W/° C./dm³, for example) and, above all, the possibility of operating correctly at high differential pressures and temperatures, for example 1 MPa and 600° C. In addition, derived from these first metal exchangers, other lower-performance polymer or glass exchangers are also possible, which relate to specific particular applications, notably those that use corrosive fluids.

To this end, unlike the one-piece metal exchangers, initially envisaged for the cooling of the exhaust gases of diesel engines, the new heat exchangers, particularly envisaged for this specific use, must be fitted with hollow metal plates, provided with an internal channel as narrow and accurate as possible and walls that are both stiff and very thin. With regard to the general characteristics of such a heat exchanger, they will obviously be completely different from those of the previous heat exchangers. They will be borrowed from a heavy, bulky heat exchange device developed for cooling electric transformers in power distribution systems, described in patents U.S. Pat. No. 3,153,447 of 1964 and U.S. Pat. No. 3,849,851 of 1974. This device is made up of large hollow metal plates with embossed walls, connected by welding to two external manifolds, capable of being arranged vertically and cooled by air circulating by natural convection.

The first subject of the invention is a high-performance heat exchanger, made up of hollow plates with thin metal walls stiffened by appropriate embossing, simultaneously having a low weight, bulk and surface area, low mechanical power consumption and high bulk conductance, whilst being suitable for reliable, easy to control industrial production and, moreover, capable of handling liquid and/or gaseous fluids, at high temperatures and/or differential pressures.

The second subject of the invention relates to improved heat exchangers, similar to the previous exchanger, with lower performance than it, but better suited to given particular applications, different from those of the previous exchanger, comprising a stack of hollow plates with thin polymer or glass walls, stiffened by appropriate embossing.

The third subject of the invention is a compact radiator with a small front surface area, made from these improved heat exchangers, having high thermal conductivity and requiring very low pumping and ventilation power.

According to the invention, a heat exchanger with low weight and bulk and very high bulk conductance, capable of handling fluids at high differential pressures and temperatures, in which:

• • hollow metal plates, with a narrow internal channel, are stacked evenly spaced and connected to external manifolds;
  • these plates comprise a embossed central zone, located between two connecting zones provided with narrow openings with an area approximately equal to the area of a cross-section of the central zone;
  • the walls of these plates have been produced by stamping and cutting a metal sheet;
  • the lateral edges of the two walls of a hollow plate are welded;

is characterised in that:

• • the walls of each hollow plate are both rigid and very thin, their embossed central zone having one or more sets of alternating aligned bosses, provided with steep strain hardened faces, creating a large number of sharp edges, facing in directions oblique or perpendicular to the alignment of the bosses;
  • the gap between opposite faces is uniform, very small, exactly known and practically constant, in the range of the envisaged differential pressures;
  • the gaps separating the plates are relatively narrow.

Before a commentary is given on the advantage of these new arrangements, it will be noted that in the US patents in question, the walls of the plates do not need to be thin and their rigidity is not a particular problem, in such a way that the embossing of the central zone of the walls is not a solution to a stiffness problem which, in this case, barely exists. A sufficient thickness of walls made from a normal metal meets this need without difficulty. The embossing is simply to increase the heat exchange area of the plates without increasing their dimensions. This is achieved by longitudinal undulations that result from relatively small recesses, evenly spaced in the walls. The particular profile of these undulations is shown; it is ordinary and can hardly be characterised by any originality, as this aspect of things is of no interest in this type of exchanger. However, due to these undulations, the internal channel of the hollow plates has an undulating thickness that varies symmetrically around a relatively high average value. Furthermore, the walls of the internal channel do not comprise opposing sloping faces.

According to the first arrangement of the invention, it firstly involves plates with very thin rigid walls (for example 0.15 mm for certain steels) that are endowed with a particularly high hardness and limit of elasticity by their strain hardening, obtained “as a bonus” at the time of standard (cold) stamping; each face of these hollows and bosses serves as a rigid strip and, moreover, each sharp edge behaves as a beam in which these strips are embedded. These strips can therefore only take a very limited deflection, under the action of the differential pressures applied. Particularly when the overpressure is external, this deflection always remains considerably less than half of the internal thickness of the hollow plates, which thickness, measured between the faces of the bosses, is by design exactly known and particularly small (0.3 mm, for example). This prevents any contact between walls of opposite faces so that the heat exchange function between the two fluids is always correctly performed.

Under these conditions, each embossed hollow plate according to the invention owes its remarkable primary stiffness to the fact that the metal constituting its walls is strain hardened and, furthermore, that its alternating bosses significantly increase its moment of inertia. These doubly stiff very thin strips are thus able to act perfectly as an efficient heat exchanger between the two fluids circulating along their two surfaces, even if there is a high differential pressure between the fluids. The immediate characteristics of these stamped alternating bosses, which must provide this stiffness, define the basis of the invention. They take the form of steep strain hardened faces, generated by significant local elongations of the initial flat sheet, which thus create a number of very thin, very stiff strips all of the edges of which are embedded in beams formed by the sharp edges of the bosses.

The sharp edges of the dihedrons, which form between them these steep faces, have a second known effect, that of increasing the apparent thermal conductivity of the air; the edges, which are orientated obliquely and/or perpendicularly to the direction of flow of the air, have the effect of generating significant turbulence in the generally fast airflow that passes through the relatively narrow gaps separating the plates. This arrangement would be meaningless in the case of the vertically arranged undulated plates in the US patents in question, as slow airflow passes through the gaps with unspecified dimensions separating them, circulating by natural convection.

If we now, to conclude this argument, refer to TET's European patent, it can be seen that all of the causes of the performance limitations set out above are eliminated in this new heat exchanger and replaced by their opposites: the walls and the internal channel have very thin, precise and well-known thicknesses, and, as will be set out in detail below, the central channel can disappear. Conversely, all of the positive characteristics, relating to the embossed walls of the hollow plates of the one-piece polymer heat exchanger described in this European patent, are retained. These characteristics are supplemented by those arising from the strain hardening of the sheets of metal used. Due to their combination with the advantages of the exchanger described in the US patents together with the use (a priori ill-advised, in the context of the high differential pressures envisaged) of very thin walls and the creation of a particularly narrow internal channel, a new, non-obvious heat exchanger is produced. This new exchanger consequently has performance levels that greatly transcend those, already very efficient, of the one-piece polymer heat exchangers according to TET's European patent.

According to particular characteristics, supplementary to the main characteristics above,

• • each hollow plate comprises at least two rows of alternating bosses;
  • two adjacent rows are separated by a narrow, straight partition, formed by two stamped internal protrusions, assembled by welding;
  • the height of these protrusions is equal to half of the value of the internal thickness of the plates, at the crests of their bosses.

These latter arrangements, taken from a possibility envisaged in the US patents in question to improve the rigidity of the plates when they have large dimensions (m²), give two particularly advantageous results for the heat exchanger according to the invention. Firstly, under the effect of a relatively high internal overpressure, applied to the hollow plates of such a heat exchanger, the straight internal partition maintains the internal thickness of the embossed central zones at a value that is practically independent of the differential pressure to which the thin walls of the plates are subjected. The result of this is that the hollow plates, with very think walls stiffened by appropriate embossing, are able to withstand a relatively high internal overpressure without damage. Without such welded internal protrusions, the adjacent rows of highly rigid alternating bosses would be separated by a flexible zone acting as a hinge. In response to such overpressure, this would lead to a slight bulging of the plates, causing a significant reduction in the heat exchanges in the gaps between them or even rapid deterioration of the plates. However, with such a partition formed by these two welded internal protrusions, it is not necessary to systematically increase the thickness of the very thin walls of the hollow plates to enable them to withstand a temporarily high internal overpressure. This means that lighter, less costly heat exchangers can be produced.

The second advantage of these welded internal protrusions comes in the form of greater efficiency of the desired heat exchange. The internal partition formed in this way between two adjacent rows of alternating bosses constitutes a barrier for the flow of liquid entering the hollow plate. The first effect of each barrier is to prevent a significant direct flow between the two external manifolds, along a smooth wall with a small surface area and therefore inefficient for the desired heat exchange, in that this surface is not swept by a strong airflow as it is in the rear zone of the upstream manifold. On the other hand, the second effect of this barrier is to direct the incoming flow towards the two rows of alternating bosses, which have high heat exchange efficiency, and thus maximise the heat exchanges performed.

It will be noted that these two advantages are of little interest for the heat exchanger with large hollow plates with relatively thick walls described in the US patents in question. In this exchanger, the maximum differential pressure, which occurs at the foot of the large vertical plates, is the relatively low hydrostatic overpressure generated by the cooling oil. This does not concern the heat exchanger according to the invention, which can obviously be installed in any relevant position and above all can operate with very high differential pressures. Moreover, as the oil circulates from top to bottom by natural convection in hollow plates much larger than the external manifolds, the low upstream dynamic pressure, due to a low circulation speed, prevents it from being able to favour a rapid direct trajectory from one manifold to another.

According to characteristics complementary to the previous ones:

• • the angles formed by the normals to two adjacent faces of the alternating bosses measure at least 30°, so that the sharp edges of these faces can be effective in the creation of turbulence and comparable to beams in which the faces of these bosses are embedded;
  • the maximum angle of the normals to two adjacent faces is limited by the restrictions imposed on the conditions under which the metal sheet in question is stamped.

According to a characteristic complementary to the previous ones, the opposite faces of a plate have parallel walls and the gap separating these walls is constant and of the same order of magnitude as their thickness.

According to characteristics complementary to the previous ones:

• • the alternating bosses have, of their own, two faces in the form of an isosceles trapezium, having a common longitudinal edge and, shared, two rhomboid faces;
  • the long diagonal of the rhomboid faces can measure several tens of times the thickness of the wall of the plates.

According to characteristics alternative to the previous ones:

• • the alternating bosses have, of their own, two triangular faces and, shared, two hexagonal faces, having a common transverse edge;
  • the distance between the transverse edges of the hexagonal faces can measure several tens of times the thickness of the wall of the plates.

According to a characteristic complementary to the previous ones, the embossed central zone of each hollow plate is connected to the external manifolds by two connecting zones provided with lateral edges having a significant slant and smooth walls comprising portions of truncated cones.

According to a characteristic complementary to the previous ones, the external manifolds have an aerodynamic profile capable of minimising their drag.

According to a possible characteristic complementary to the previous ones, symmetrical boss faces appear to be cut in a diamond pattern and comprise several secondary faces, provided with additional sharp edges.

As a result of these different arrangements, the bulk conductance of the heat exchanger thus produced is particularly high. There are several reasons for this: (1) the plates have metal walls that have negligible thermal resistance, (2) the thermal resistance of the very thin layer of water or oil inside the plates is low, despite the laminar flow of the layer and the relatively high thermal resistivity of these liquids and (3) the turbulence and the apparent thermal conductivity of the air circulating between the plates increase with the height of the bosses and the total number of sharp edges they comprise. With at least two rows each comprising several alternating bosses, provided with faces sloped at approximately 45°, an efficient compromise is achieved between the different parameters involved. The stamping of the bosses, the slope of the faces of which is less than approximately 50°, is a standard operation that poses no production problems, and a minimum angle of 30° between the normals to two adjacent faces ensures satisfactory turbulence in the air flow and a minimum width for each of the rows of bosses in the central zone of the plates, when the height of these hollows and bosses is fixed. Furthermore, a minimum angle of 30° between the normals to two adjacent faces gives the edge in question sufficient stiffness for it to be comparable to a beam, and the edges are then collectively comparable to a network of beams.

Moreover, with a heat exchanger formed by the stacking of a large number of such identical plates, connected to two external manifolds, the pressure drop of a liquid circulating within them at a constant flow rate and flowing laminarly, which can, if applicable, be relatively fast, can be reduced considerably. In any event, such a stack significantly reduces the power necessary to pump the liquid. In addition to the use of external manifolds with an aerodynamic profile, despite the relatively large gap separating the plates, their largest dimension, installed parallel to the flow velocity of the two intersecting fluids, leads to a significant reduction in the aerodynamic drag of the radiator and/or the power necessary to ventilate it.

With regard to the metals that can be used for the production of the walls of the hollow plates according to the invention, it will be noted that these are not numerous but are well known to specialists in stamping and that ultimately the choice (aluminium or steel, for example) will mainly be determined by the mechanical behaviour of these metals in the operating temperature range of the heat exchangers that will incorporate these plates.

As a result of these various arrangements, the industrial manufacture of the very high-performance heat exchangers according to the invention comprises a set of completely controllable operations that are relatively easy to automate, which results in an advantageous cost price for the mass production of such exchangers. These operations are as follows:

• • 1) stamping and cutting of identical plate walls from a thin metal sheet;
  • 2) turning of one wall head to foot;
  • 3) assembly of two adjacent walls, by welding of their lateral flanges and the protrusions of their internal central partition;
  • 4) mounting and fixing by welding of these hollow plates to their two external manifolds.

According to the invention, a compact radiator with very high bulk conductivity is characterised in that:

• • it comprises two identical groups of thin metal hollow plate heat exchangers, associated with two main upstream and downstream manifolds, provided with flat rectangular trapezoid surfaces, slightly separate from each other and arranged so that their square corners are opposite each other;
  • the individual upstream and downstream manifolds of the exchangers in each group are connected respectively, at constant intervals slightly larger than the width of the central zone of the exchangers, to the two surfaces of the two main upstream and downstream manifolds.

By means of these arrangements, a radiator can be constructed with very high bulk conductivity and as small a main cross-section as possible (up to 0.10 dm² per kW to be discharged). A large number of heat exchangers themselves formed by a large number of metal hollow plates stacked according to the invention can easily be assembled on either side of the two main flat manifolds. This compact radiator also requires particularly low pumping and ventilation power, around five times lower than the power required by solid fin radiators with the same thermal conductance.

The characteristics and advantages of the invention will become apparent in more detail on reading the following description of a non-limitative embodiment of the invention, given in reference to the appended drawings, in which:

FIG. 1 is a top view of a first embossed wall of a hollow plate according to the invention;

FIG. 2A is a top view of a second embossed wall of a hollow plate according to the invention and FIGS. 2B and 2C are views of particular faces of this wall;

FIG. 3 is a longitudinal cross-section of the alternating bosses on this first wall;

FIG. 4 is a longitudinal cross-section of one end of a hollow plate, welded to a manifold;

FIG. 5 is an isometric perspective view of a heat exchanger with fifteen hollow plates;

FIG. 6 is a top view of a radiator according to the invention, constructed using these heat exchangers.

FIG. 1 shows a first embodiment of a thin metal wall 10 of a hollow plate. This wall has been stamped and then cut so as to have an embossed central zone 13, arranged between two connecting zones. As an example, this wall is made from aluminium and is 0.3 mm thick and its embossed central zone is 60 mm wide and 76 mm long. This central zone 13 is made up of two adjacent identical rows 12 and 14 of alternating bosses, separated by a narrow straight zone 16, 4 mm wide. The two connecting zones 18 and 20 have smooth walls. Each row comprises two identical areas of alternating embossing, made up of bosses and hollows, that is, for rows 12-14, four bosses 22_1-2 and 24_1-2 on the one hand and four hollows 22′_1-2 and 24′_1-2 on the other, these latter being shown in grey. Each boss 22_1-2-24_1-2 or each hollow 22′_1-2-24′_1-2 is shaped like a roof with four slopes having four very steep sharp edges, namely for each alternating boss in row 12: (1) of its own, two symmetrical trapeziums 26_1-2 and 28_1-2 for the bosses and 26′_1-2 and 28′_1-2 for the hollows, all with a 19 mm large base, (2) shared with the adjacent boss in the same row, two isosceles triangles 30_1-2 and 32_1-2 for the bosses and 30′_1-2 and 32′_1-2 for the hollows, all with a 28 mm base, (3) a longitudinal crest 34_1-2 for the bosses and 34′_1-2 for the hollows, all 5 mm long, and (4) the same height of 5 mm. It will be noted that the two pairs of isosceles triangles 30₂-30′₁ and 30′₂-32₁ in row 12 (and similarly in 14), which belong to two consecutive alternations of the alternating embossing, form two flat rhombuses.

At the centre of the narrow straight zone 16, which splits in two the embossed central zone 13 of the wall 10 shown, an internal protrusion 36 2 mm wide is produced by stamping, with symmetrical sides as stiff as the stamping technology allows. Such a protrusion 36 has a height equal to half of the maximum gap separating the crests of the bosses on the two walls of the hollow plate produced (that is, 0.2 mm, as specified below). Two lines 38-40 separate the parallel external edges of the two rows 12-14 of alternating bosses on one hollow plate wall from the pair of parallel external flanges 42-44, which form part of the sealing surface of two plate walls. The lines 38-40 and the flanges 42-44 are 1 mm wide and form a small step 0.2 mm high, which determines half of the internal thickness of a plate at the crests of its bosses. These two flat lines 38-40 end in the two flat parts 46-48 of the two connecting zones 18-20 of the wall 10 and these two parallel flanges 42-44 end with the two pairs of oblique external flanges 50₁-50₂ and 52₁-52₂ of these same connecting zones; they form the other part of the sealing surface of the plate walls. Each flange 50_1-2 or 52_1-2 forms an angle of 60° with the longitudinal line of symmetry of the wall 10. The end of each connecting zone 18-20 comprises an almost flat truncated cone portion 54-56, with a half-cone angle of 87.5°. This tapered portion is delimited by two pairs of arcs 58_1-2 and 60_1-2, the latter pair being 8 mm long. Their ends are connected to each other by two steps 1.5 mm high, such that the area of each of the upstream or downstream openings, thus made for a hollow plate, measures 24 mm, i.e. approximately the area of the transverse cross-section of the internal space of the embossed central zone 13 of the plate.

FIG. 2A shows a thin metal wall 11, stamped and then cut, which constitutes a second embodiment of a hollow plate wall according to the invention. This wall 11 only differs from the previous wall 10 in its embossed central zone, which only comprises a single row of bosses 15, 26 mm wide, and in the shape of its alternating bosses. This single row comprises three bosses 22_1-3 and three hollows 22′b_1-3, the latter being shown in grey. Each boss 22b_1-3 and each hollow 22′b_1-3 is shaped like a roof with four steep slopes. For the three alternating bosses in row 15, this gives for each one: (1) of its own, pairs of symmetrical lateral triangles 25b_1-2, 27b_1-2, 29b_1-2 for the bosses, and similarly 25′b_1-2, 27′b_1-2, 29′b_1-2 for the hollows, all with a 14 mm large base and (2) shared with the adjacent boss, central hexagons 31_1-5, all with a transverse crest 18 mm long, and the same height of 5 mm.

FIGS. 2B and 2C show, as variants to the faces of the bosses in FIGS. 1 and 2A, two of the main faces in FIG. 2A having secondary faces. FIG. 2B shows a triangular lateral face 25, provided with three secondary faces 37_1-3 forming a relatively flat trihedron with three sharp edges, with a diamond point 39, located at the centre of gravity of this triangle. FIG. 2C shows a hexagonal longitudinal face 31, provided with six triangles with coplanar sides 41_1-6, with a central diamond point 43_1-6, similar to the point 39 in FIG. 2B. The height of these points is determined by the limits of the metal sheet stamping technology.

FIGS. 1 and 2A show two of the possible forms that the bosses on the embossed walls of the hollow plates according to the invention can take. FIGS. 2B and 2C show the possible variants of the main faces of these bosses, in order to improve their ability to produce turbulence in the airflow between plates.

FIG. 3 shows an enlargement of a longitudinal cross-section along an axis AA′ (see FIG. 1) of one of the ends of part of a hollow plate before it is connected to a manifold. This plate is the result of the welding of the two walls 10a and 10b, the wall 10b being the wall 10a turned head to foot, around the transverse axis of symmetry BB′ (see FIG. 1). This cross-section AA′ is made along the crests 35₂ and 35′₂ of the alternating embossing formed by the boss 24₂ and the hollow 24′₂ in row 14 and it passes through the connecting zone 18 of the wall 10a of this plate. FIG. 4 shows an enlargement of a cross-section of the same plate end, made along the longitudinal line of symmetry CC′ (see FIG. 1) of rows 12 and 14 of alternating bosses and the connecting zones 18 and 20 of the wall 10a.

In FIG. 3, the bosses and hollows of the first embodiment of the bottom wall 10b and the top wall 10a of a plate are reversed, so that parts 24₂ and 24′₂ of the top wall 10a, seen in profile in FIG. 3, appear respectively as a hollow and a boss. The boss 24′₁ and the hollow 24₁ on the wall 10b defined above are nested in the hollow and the boss respectively. The thickness of the part 62 of the internal channel of a hollow plate, located between the nested crests 34₁-35′₂ or 34′₁-35₂ of the embossed zone of the plate is 0.4 mm, and the thickness of the part 64 of the internal channel, located between the 45° slopes of the ascending or descending sides of the bosses is 0.28 mm. The thickness of the internal channel 66, between the flat parts of the connecting zones 18 and 20, is 0.4 mm.

According to FIG. 3, the right-hand part of the cross-section along the line AA′ shows (1) the start 68 of the gradual separation of the walls from the two facing conical sections 54-56 of the walls 10a-10b, which end these two connecting zones, (2) the two symmetrical steps of these walls that start with the circles 58₂ and 58₁ and (3) the two symmetrical external flanges 52₂ and 50₁ which define the sealing surface of the walls 10a and 10b.

According to FIG. 4, the cross-section shown is made along the longitudinal axis of symmetry CC′ of one hollow plate end engaged and welded with a bead 70, in the edges and the ends of a slot 72, in the form of a 120° arc, made in the connecting shell 74 of an external manifold 75, formed by two elongated shells welded to each other. The cross-section represented shows two parallel sections 16a and 16b of the narrow central zones of the walls 10a and 10b, separated by a 0.4 mm gap 66, and two other divergent sections 54 and 56 corresponding to the facing conical sections of the connecting zones of the two walls 10a and 10b of the hollow plate. The gap between the extreme edges of these two divergent sections is 3 mm and the length of the 120° arcs 60₂ and 60₁ (see FIG. 1) is 8 mm. As a result, the area of the straight cross-sections of the internal channel with embossed walls and that of the openings in the ends of the fin are approximately equal.

FIG. 5 shows an elementary heat exchanger 76 comprising fifteen thin hollow metal fins 78_1-15 with embossed walls. The ends of these hollow plates are engaged and welded as shown above in slots with circular edges, 3.5 mm wide and spaced 8 mm apart, made in the walls of the external manifolds 80-82, with an aerodynamic profile. To enable the easy production of such welds, the manifolds 80-82 are made up of two elongated shells, with a U-shaped transverse cross-section, welded to each other along a line 83. They are made from metal strips cut out of sheets identical to those used for the production of the stamped plate walls. Slots with appropriate width, length and spacing are made in half of these strips, then the two types of strip prepared in this way are turned into front closure and connecting shells 75, by means of two matching templates, with protruding and hollow profiles. Then, the openings in the various hollow plates are welded to the slots in the connecting shells. Then, two front closing shells are in turn welded to the previous two and one of their ends is sealed off, to form both the two streamlined external manifolds and the exchanger itself.

FIG. 6 shows a top view of a compact radiator 81. Six identical heat exchangers 76_1-6 can be mounted in parallel on either side of two flat main manifolds 84-86, in the form of rectangular trapeziums arranged head to foot, in order to form a compact radiator with appropriate overall thermal conductance. These flat manifolds 84-86 have parallel sides 88_1-2 and 90_1-2 and a thickness approximately equal to the maximum dimension of the straight cross-sections of the external manifolds 80_1-2. Two adjacent exchangers are mounted so that the lateral edges of their plates are practically abutting or slightly interleaved. In the first case, the feet of the upstream 80_1-6 and downstream 82_1-6 external manifolds are engaged to the same depth in appropriate circular openings, 94_1-6 and 96_1-6, made at constant intervals along the long sides 92-93 of the faces of the main manifolds, and they are then welded. In the second case, the depths of insertion of the manifolds are different for the exchangers in the odd and even rows. The length of the longest 88₂-90₂ of the parallel sides of the two main manifolds 84-86 is determined by the number of heat exchangers 76 to be mounted. The short sides of the two main manifolds 84-86 have lengths determined by the spacing of the external manifolds 80-82 and by the gap 100 (typically 5 mm) that separates their oblique sides.

Such an assembly of heat exchangers formed by stacks of thin hollow metal plates, with very thin walls stiffened by embossing, allows for the formation of a compact radiator that is particularly advantageous for the cooling of high power thermal engines (>100 kW). They have a very small main cross-section, very high thermal conductance, low pumping and ventilation power consumption and limited bulk and weight. It is also suitable for the processing of diesel engine exhaust gases, used cooled to improve their operation at low speeds. More generally, any heat exchange between two fluids, particularly between a liquid and a gas, having a high temperature and/or differential pressure (up to around 600° C. and 1 MPa) can be carried out efficiently by means of such a compact metal assembly.

The invention is not limited to the examples described. The length and width of the hollow plates can be significantly greater than those shown in FIG. 1 and measure several decimeters. The same can be said of the number of alternating bosses in each row and the number of rows on each plate. The maximum dimensions of a plate are in practice determined by those of the table on the stamping press available. With regard to the number of hollow plates in a heat exchanger, this can rise to several tens. The same can be said of the total number of exchangers assembled in a compact radiator.

It will also be noted that it is possible to produce a hollow plate according to the invention using two appropriate embossed walls that are similar but not identical due to their different lateral edges. Instead of two identical walls, with lateral flanges comprising a small step defining the half-thickness of the internal channel of the central zone, it would be possible to have one wall with flanges having a step twice as high as the previous step and another wall without any step. This would require the use of two different pairs of stamping moulds but would have little economic impact when the production rate is high.

The figures above show hollow plates for a liquid/gas heat exchanger. The liquid circulates in these metal plates with a very narrow internal channel (0.3 mm). In the case of a gas/gas heat exchanger, the thickness of this internal channel is obviously much greater (typically >1 mm) and the gap between plates is generally smaller than that on the exchanger shown. This is so that the mass flows and speeds of the two gases are comparable on either side of the walls of the hollow plates.

Moreover, for particular applications, notably in chemistry and any other field in which corrosive fluids are used, it is often desirable and sometimes necessary to have access to high-performance glass heat exchangers perfectly suited to their conditions of use. To this end, these glass heat exchangers will be provided with high bulk conductance, but half way between the conductances given above for hollow plate exchangers made either from a one-piece polymer or from metal of the type according to the invention (20 or 100 W/° C./dm³). With regard to the maximum temperatures and differential pressures that can be applied to these glass heat exchangers, they will be lower than those withstood by the metal exchangers according to the present invention and higher than those relating to the one-piece polymer exchangers according to TET's European patent. For this same type of application, it can also be advantageous to have access to polymer heat exchangers with a bulk conductance approximately 50% higher than that of these one-piece exchangers, whilst retaining their differential pressure and temperature ranges.

To this end, the new technology for metal heat exchangers according to the invention can be adopted and adapted and, instead of a metal sheet, a sheet of glass or polymer can simply be used and processed by hot stamping or thermoforming. The manufacturing methods used in these two sheet forming techniques are similar to each other: the first uses mechanical pressure and two matching moulds comprising hollows and/or protrusions, and the second uses pneumatic pressure and a single mould with hollows and/or protrusions; both use appropriate heating. However, no strain hardening is produced.

The thicknesses of the walls and internal channels of such a glass or polymer hollow plate heat exchanger with embossed walls and external manifolds will inevitably be increased in accordance with the specific mechanical properties of the type of glass or polymer used. Their performance will be derived directly from them, as explained above.

Claims

1-1. (canceled)

12. A Heat exchanger with low weight and high bulk conductance, capable of handling fluids at high differential pressure and temperatures, in which: hollow metal plates with a narrow internal channel, are stacked evenly spaced and connected to external manifolds;these plates comprise an embossed central zone, located between two connecting zones provided with narrow openings with a surface area approximately equal to the area of a cross-section of the central zone;the walls of these plates have been produced by stamping and cutting a metal sheet;the lateral edges of the two walls of a hollow plate are welded;

13. The Heat exchanger, derived from the exchanger according to claim 12, wherein: it is made up of glass or polymer hollow plates, with a thin internal channel, stacked with constant spacing and connected to external manifolds;the walls of these hollow plates have been produced by hot stamping or thermoforming and then cutting from a sheet of glass or polymer;these plates comprise an embossed central zone, located between two connecting zones provided with narrow openings with an area approximately equal to the area of a transverse cross-section of the central zone;the lateral edges of the two walls of a hollow plate are welded;the central zone of the plates has one or more sets of aligned alternating bosses, provided with steep faces, creating a large number of sharp edges, orientated obliquely or perpendicularly to the alignment of the bosses;the gap between opposite faces is uniform, small, exactly known and practically constant, in the range of the envisaged differential pressures;the gaps separating the plates are relatively narrow.

14. The Heat exchanger according to claim 12, wherein: each hollow plate comprises at least two rows of alternating bosses;two adjacent rows are separated by a narrow, straight partition, formed by two internal stamped or thermoformed protrusions, assembled by welding;the height of these protrusions is equal to half of the maximum value of the internal thickness of these hollow plates.

15. The Heat exchanger according to claim 13, wherein: each hollow plate comprises at least two rows of alternating bosses;two adjacent rows are separated by a narrow, straight partition, formed by two internal stamped or thermoformed protrusions, assembled by welding;the height of these protrusions is equal to half of the maximum value of the internal thickness of these hollow plates.

16. The Heat exchanger according to claim 12, wherein: the angles formed by the normals to two adjacent faces of the alternating bosses measure at least 30°, so that the sharp edges of these faces can be effective in the creation of turbulence and in withstanding the pressure differences between the fluids;the maximum angle of the normals to two adjacent faces is limited by the restrictions imposed on the conditions under which the material in question is stamped or thermoformed.

17. The Heat exchanger according to claim 13, wherein: the angles formed by the normals to two adjacent faces of the alternating bosses measure at least 30°, so that the sharp edges of these faces can be effective in the creation of turbulence and in withstanding the pressure differences between the fluids;the maximum angle of the normals to two adjacent faces is limited by the restrictions imposed on the conditions under which the material in question is stamped or thermoformed.

18. The Heat exchanger according to claim 12, wherein: the alternating bosses have, of their own, two lateral faces in the form of an isosceles trapezium, having a common longitudinal edge and, shared, two central rhomboid faces;the long diagonal of the rhomboid faces can measure several tens of times the thickness of the wall of the plates.

19. Heat exchanger according to claim 13, wherein: the alternating bosses have, of their own, two lateral faces in the form of an isosceles trapezium, having a common longitudinal edge and, shared, two central rhomboid faces;the long diagonal of the rhomboid faces can measure several tens of times the thickness of the wall of the plates.

20. The Heat exchanger according to claim 12, wherein: the alternating bosses have, of their own, two lateral faces in the form of an isosceles triangle for the bosses and for the hollows and, shared, two central hexagonal faces for the bosses and for the hollows, these hexagonal faces having a common transverse edge;the gap between the transverse edges of the hexagonal faces can measure several tens of times the thickness of the wall of the plates.

21. Heat exchanger according to claim 13, wherein: the alternating bosses have, of their own, two lateral faces in the form of an isosceles triangle for the bosses and for the hollows and, shared, two central hexagonal faces for the bosses and for the hollows, these hexagonal faces having a common transverse edge;the gap between the transverse edges of the hexagonal faces can measure several tens of times the thickness of the wall of the plates.

22. Heat exchanger according to claim 12, wherein the embossed central zone of each hollow plate is connected to the external manifolds by two connecting zones provided with lateral edges having a significant slant and smooth walls comprising portions of truncated cones.

23. Heat exchanger according to claim 13, wherein the embossed central zone of each hollow plate is connected to the external manifolds by two connecting zones provided with lateral edges having a significant slant and smooth walls comprising portions of truncated cones.

24. Heat exchanger according to claim 12, wherein the opposite faces of a hollow plate have parallel walls and the gap separating these walls is constant and of the same order of magnitude as their thickness.

25. Heat exchanger according to claim 13, wherein the opposite faces of a hollow plate have parallel walls and the gap separating these walls is constant and of the same order of magnitude as their thickness.

26. Heat exchanger according to claim 12, wherein symmetrical boss faces appear to be cut in a diamond pattern and comprise several secondary faces and are provided with complementary sharp edges.

27. Heat exchanger according to claim 13, wherein symmetrical boss faces appear to be cut in a diamond pattern and comprise several secondary faces and are provided with complementary sharp edges.

28. The Heat exchanger according to claim 12, wherein: the external manifolds of the hollow plates have an aerodynamic profile capable of minimising the drag of the exchanger;each manifold is made up of two elongated shells, one for connection to the plates and the other for front closure, their transverse cross-section is U-shaped and they are fixed to each other by a weld line.

29. The Heat exchanger according to claim 13, wherein: the external manifolds of the hollow plates have an aerodynamic profile capable of minimising the drag of the exchanger;each manifold is made up of two elongated shells, one for connection to the plates and the other for front closure, their transverse cross-section is U-shaped and they are fixed to each other by a weld line.

30. A Compact, light radiator with high or very high thermal conductivity, wherein: it comprises two identical groups of heat exchangers with hollow plates made from metal, glass or polymer, according to claim 18,these two groups are associated with two thin main upstream and downstream manifolds, provided with flat rectangular trapezoid surfaces, slightly separate from each other and arranged so that their square corners are opposite each other;the individual upstream and downstream manifolds of the exchangers in each group are connected respectively, at constant intervals slightly larger than the width of the central zone of the exchangers, to two homologous surfaces of the two main upstream and downstream manifolds.