DESCRIPTION

Field of application

The present finding refers to a heat exchanger according to the preamble of the independent claim 1.

The heat exchanger according to the finding is of the so called "bar and plate" type and it is advantageously intended to be used for transferring heat between two or more fluids of which at least one of such fluids is able to undergo a change of phase during the outflow inside the exchanger.

In particular, the heat exchanger object of the finding belongs to the field of the production of air conditioner machines, in which it can be advantageously used as a cooler or condenser.

State of the art

As it is known, heat exchangers are devices that are capable of transferring thermal energy between two or more fluids at different temperatures that pass through corresponding passages of the exchanger.

In particular, heat exchangers are widely used in the field of conditioning plants, in which they are used as coolers or condensers.

More in detail, in conditioning plants heat exchangers of the so called "tube bundle and shell" type are commonly used, comprising a plurality of tubes, in which a cooling fluid flows, that are arranged inside a containment body, called a shell, which is filled with water with which the cooling fluid exchanges heat.

Operatively, in the case in which the heat exchanger is used as a cooler, the cooling fluid enters the tubes at an inlet portion thereof, in which the cooling fluid has a mass that is mostly at the liquid state (for example 90% at the liquid state and the remaining 10% is at the gaseous state), and advances in such tubes absorbing heat from the water contained in the shell which has a temperature that is greater than that of the cooling fluid. Following such a heat exchange the cooling fluid evaporates increasing its average percentage value of mass at the gaseous state gradually as it advances in the tubes, until it reaches an outlet section of the tubes in which the cooling fluid is completely vaporised.

Vice versa, in the case in which the heat exchanger operates as a condenser, the cooling fluid enters inside the tubes in the gaseous state and condensates during its movement along the tubes after yielding heat to the water contained in the shell of the exchanger (which has a temperature that is lower than that of the cooling fluid), until it comes, out completely in the liquid state.

Since the tubes of the condenser have a constant section, the progressive passage of state of the cooling fluid that flows inside them leads to the fact that the velocity of the fluid varies in the different portions of the tubes. Such a variation of velocity of the fluid, as it is known, leads to a variation of the heat exchange coefficient in the different portions of the tubes. The constant section of the tubes is normally calibrated so as to optimise the heat exchange coefficient in the central portion of the tubes, with consequent substantial inefficiency of the heat exchanger in the start and end portions of the tubes of the exchanger in which the cooling fluid has average percentage values of vapour/liquid that are considerably different with respect to the central portion.

Moreover, the exchangers of the "tube bundle and shell" type, briefly described above, have the drawback of being constructively complex and bulky.

In order to at least partially solve the aforementioned drawbacks of the exchangers of the known type described above, heat exchangers have been on the market for a long time, in which the different portions of tube have different sections as a function of the state that the cooling fluid takes up when passing through such a portion of tube, so as to regulate the velocity of the cooling fluid to optimise the heat exchange coefficient.

For example British patent GB 2069676 describes a heat exchanger, used as a cooler, in which many elongated bodies are arranged inside the tube in which the cooling fluid flows, each having a smaller width than the previous one. In such a way, in the end portion of the tube, in which the cooling fluid is in the gaseous state, the section is greater than the initial portion (in which the cooling fluid is in the liquid state) and thus defines, as it is known to a man skilled in the art, a velocity of the flow in the end portion that is slower with respect to the case in which the sections of the different portions are the same. This partially compensates for the increase in velocity due to the passage of cooling fluid in the gaseous state optimising the heat exchange coefficient along the entire tube.

However, also this last exchanger of the known type with the tubes having a variable section has been found in practice to not be without drawbacks.

One first drawback is due to the fact that the arrangement in the tubes of additional bodies for reducing the section of the tubes themselves leads to complication in the construction of the exchanger with a consequent increase in the design and production costs.

Moreover, when assembling the exchanger it is necessary to carry out specific dedicated operations for the insertion and the fixing of the elongated bodies to the tubes and consequently a longer production time.

Moreover, the exchanger described in GB 2069676 does not solve the problem of having a large bulk at all.

There are also known heat exchangers of the so called "plate" type, one example of which is described in patent application PD2007A000279. Such a "plate" exchanger, in a completely conventional manner, comprises a stack of corrugated steel plates, which define a plurality of passages between them and they are welded one to the next through brazing at certain contact points. Each plate is provided with two pairs of holes for allowing two fluids having a different temperature to enter and exit, respectively, through the passages defined between the plates themselves. More in detail, the holes of each plate are arranged aligned with the corresponding holes of the other plates of the exchanger and are connected to one another so that the passages crossed by a fluid are alternated with those crossed by the other fluid. Operatively, each plate separates the passage of one fluid from the adjacent passage of the other fluid and acts as a heat transmission means from the colder fluid to the hotter fluid.

The heat exchangers of the "plate" type described above, despite having a shape that is considerably more compact with respect to tube exchangers, are not suitable for operating in an optimal manner with fluids that change their phase during the outflow in the exchanger, since they have passages with a substantially constant section with drawbacks in terms of heat exchange inefficiency that have already been discussed for "tube bundle and shell" heat exchangers.

Moreover, in the aforementioned "plate" type exchangers, the arrangement of passages with a variable section would require excessive design complications. Indeed, since each fluid has a different phase change behaviour, it would be necessary to design plates with specific profiles that define the passages of the exchanger for each fluid used, with consequent excessive costs in terms of its design and for making the moulds.

Heat exchangers of the so called "bar and plate" type are also known, comprising a stack of plates that are arranged parallel to one another and on top of one another and that are spaced apart so that, between each pair of successive and opposite plates of the pack, a distinct passage for a fluid is defined. Each passage is moreover limited at its sides by a pair of parallel bars that are interposed between each pair of successive plates of the pack of plates.

The exchanger is equipped, moreover, with two or more manifolds, through which two or more fluids with different temperature enter and exit the exchanger. The manifolds are connected to the passages so that the passages crossed by a fluid are alternate with those crossed by another fluid. Functionally, the heat of the hottest fluid that flows in the corresponding passage is transmitted to the coldest fluid that flows in the adjacent passage, through the plate that separates the two adjacent passages.

Moreover, in each passage the heat exchanger is provided with an undulated plate which is interposed like a sandwich between the pair of opposite plates defining the passage itself. Such an undulated plate is in thermal contact with the opposite plates so as to promote the transmission of heat between the latter and the fluid flowing in the passage.

Heat exchangers of the "bar and plate" type briefly described above and currently widespread on the market do not make it possible to achieve the heat exchange between fluids changing phase with high performance of heat transmission, since the passages have constant sections leading to inefficiency in the heat exchange for reasons that have already been discussed for heat exchangers of the "tube bundle and shell" and "plate" types.

In particular, the current constructive structure of "bar and plate" heat exchangers does not allow the section of the passages to be modified inserting inside it elongate bodies of the type described in British patent GB 2069676.

Patent US 3,992,168 describes an example of a "bar and plate" heat exchanger, which comprises, in each passage, many different undulated plates that are arranged in succession one after the next and are fixed one after the other. Each shaped plate defines corresponding channels with a section that is different from that of the channels of the other shaped plates. In particular, the section of the channels of each plate is arranged as a function of the percentage of component at the gaseous/liquid state of the cooling fluid that passes through the channels themselves, so as to compensate for the velocity variation of the fluid itself due to the variation in state of the latter, so as to improve the heat exchange coefficient according to the principles that have already been previously outlined.

The main drawback of the heat exchanger described in patent US 3,992,168 is due to the fact that at the end of the channels of each shaped plate, before the cooling fluid is introduced in the channels of the following shaped plate, such a fluid has irregularities in the distribution of the liquid and of the gaseous component, since in some channels of the shaped plate the fluid has a percentage in the liquid state that is greater than in other channels, thus determining a velocity of the cooling fluid that differs from one channel to the next and therefore does not ensure an optimal heat exchange coefficient. Moreover, the fluid coming from each channel of the shaped plate upstream only flows inside the channels of the following plate substantially aligned with such a channel, with the consequence that the irregularities in the cooling fluid are kept for the entire flow along the passage of the heat exchanger, with a consequent further decrease in the heat exchange coefficient.

Patents FR 2085924, US 4,049,051, US 2005/274501 and EP 0797065 describe some examples of heat exchangers of the "bar and plate" type, which are equipped with many shaped plates with different shapes that are arranged attached to one another inside the passages passed through by the operating fluid that does not undergo a change of state (always remaining in particular in the gaseous or liquid state). The shaped plates arranged along the inlet portions of the aforementioned passages are provided with fins that are more tightly packed with respect to the following ones, so as to increase the surface of heat exchange in such an inlet section.

These last heat exchangers of the known type do not tackle in any way the problem of optimising the heat exchange coefficient as a function of the change of phase of the fluid along the passages of the exchanger itself.

Presentation of the finding Therefore, in this situation, the purpose of the present finding is to overcome the drawbacks of the mentioned prior art, by providing a heat exchanger of the "bar and plate" type that is capable of efficiently exchanging heat between two or more fluids of which at least one is in change of phase.

A further purpose of the present finding is to provide a heat exchanger that is constructively simple and cost-effective to make.

A further purpose of the present finding is that of providing a heat exchanger that is compact in size.

A further purpose of the present finding is that of providing a heat exchanger that is operatively completely safe and reliable over time.

Brief description of the drawings

The technical characteristics of the finding, according to the aforementioned purposes, can be clearly seen by the content of the attached claims and the advantages thereof shall become clearer from the following detailed description, given with reference to the attached drawings, which represent one embodiment, which is purely given as an example and not for limiting purposes, in which:

- figure 1 illustrates a perspective view of the heat exchanger object of the present finding;

- figure 2 illustrates a perspective view of a detail of the heat exchanger of figure 1, relating to the stack of plates defining the passages for the operating fluids;

- figure 3a illustrates a section view of the detail of the heat exchanger illustrated in figure 2 according to the line III - III of figure 2 itself, in accordance with a first embodiment of the present finding;

- figure 3b illustrates a section view of the stack of plates of the heat exchanger object of the present finding, in accordance with a second embodiment of the present finding; - figure 4a illustrates an exploded view of a detail of the heat exchanger object of the present finding, relative to one of the passages for the operating fluids, in accordance with the first embodiment of the present finding illustrated in figure 3a;

- figure 4b illustrates an exploded view of a detail of the heat exchanger object of the present finding, relative to one of the passages for the operating fluids, in accordance with the second embodiment of the present finding illustrated in figure 3b.

Detailed description of a preferred embodiment

In accordance with the attached drawings, reference numeral 1 wholly indicates the heat exchanger of the "bar and plate" type that is object of the present finding.

The heat exchanger 1 according to the finding can be advantageously used in air conditioning machines and plants, in which it is intended to operate as a cooler or as a condenser, as shall be explained in detail in the rest of the present description.

With reference to the embodiment illustrated in figure 1, the heat exchanger 1 comprises a battery 2 of metallic plates 3 that are arranged parallel to one another and spaced apart by pairs of parallel bars 4, also preferably made from metallic material.

In particular, with reference to the embodiment illustrated in figure 2, the metallic plates 3 have a flat shape and a preferably rectangular (or square) shape in plan view, and are arranged one after the other in the battery 2 according to one alignment direction Z that is perpendicular to their resting plane, in such a way preferably defining a parallelepiped shaped battery 2.

The metallic plates 3 and the bars 4 of the battery 2 define a plurality of first passages 5 for the outflow of a cooling fluid that is able to undergo, in a per se known manner, a phase change inside such first passages 5, between the liquid phase and the vapour phase. The same metallic plates 3 as well as the bars 4 of the battery 2 also define a plurality of second passages 6 for the outflow of an operating fluid like for example water or air. The first and the second passages 5 and 6 are alternated with one another in the battery 2, and are separated from one another by the metallic plates 3, which are suitable for transmitting heat between the cooling fluid and the operating fluid. Each first and second passage 5, 6, moreover, extends along a respective direction of extension X, Y that is parallel to the metallic plates 3, between a first opening 7 and a second opening 8 of the battery 2 preferably obtained on opposite sides of the battery 2 itself.

More in detail, with reference to the embodiments illustrated in the attached figures, each of the aforementioned first and second passages 5 and 6 is limited by a pair of metallic plates 3 in succession in the battery 2, and is closed at the side by one of the pairs of bars 4 that are arranged parallel with respect to the direction of extension X, Y and are interposed between the pair of metallic plates 3 in succession.

In particular, with reference to the embodiment of figure 2, each metallic plate 3, except for the first and the last of the battery 2, is interposed between one of the first passages 5 and one of the second passages 6 adjacent to the latter, so that the metallic plate 3, with one face thereof, defines the first passage 5 and with the other face thereof it defines the second passage 6.

Advantageously, in accordance with the embodiment illustrated in figure 1, the heat exchanger is provided with manifolds 9 that are connected to corresponding feeding and extraction tubes 10, 10' for connecting the first and the second passages 5, 6 to corresponding external circuits that are not illustrated, in which the cooling fluid (consisting for example of a hydrofluorocarbon HFC, of hydro fluoro olefin HFO, or of a hydrocarbon HC) and the operating fluid (consisting for example of water or air), circulate respectively.

The heat exchanger 1 comprises a plurality of turbulence means that are housed inside the first passages 5 (and advantageously foreseen also in the second passages 6 as shall be made clearer in the rest of the description), in thermal contact with the pairs of metallic plates 3 to increase the heat exchange of the cooling fluid.

Such turbulence means should be considered as heat exchange promoters. In order to enhance the heat exchange coefficient it is indeed necessary to have an increase of load losses, in particular to promote wall turbulence minimising the laminar layer, without however such a load loss excessively affecting the drop in pressure of the cooling fluid and then excessively increasing the work of the compressor.

It is therefore necessary to optimise the use of such turbulence means so as to maximise the heat exchange whilst still maintaining the load losses low. Such a purpose is achieved by the present invention as specified in the rest of the description.

The aforementioned turbulence means are made in the form of shaped plates 11, 12, 13 of metallic material arranged inside the first passages 5 and fixed to the pairs of metallic plates 3. More clearly, each shaped plate 1 1, 12, 13 is in thermal contact with the two metallic plates 3 in succession that define the corresponding first passage 5, to transmit heat between the cooling fluid that flows inside it and the two metallic plates 3 themselves that define the first passage 5 itself.

Preferably, the heat exchanger 1 also comprises further shaped plates 14 that are also positioned inside the second passages 6 made in the battery 2 where it is foreseen for there to be the passage of the operating fluid for the heat exchange with the cooling fluid that on the other hand passes in the first passages 5.

Therefore, each shaped plate 1 1, 12, 13, 14 is interposed between a pair of metallic plates 3 in succession that define the corresponding passage 5, 6 in which it itself is housed, and extends alongside the two bars 4 that laterally define the corresponding passage 5, 6.

In such a way, functionally, the cooling fluid advancing in the first passage 5 exchanges heat, not only directly with the two metallic plates 3, but also with the shaped plate 1 1, 12, 13 inserted inside the same passage 5. Such a shaped plate has a wide surface of heat exchange to promote the transfer of heat with the cooling fluid. The shaped plate 11, 12, 13 also transmits heat through conduction to the two metallic plates 3, which transmit it to the operating fluid that flows in the second adjacent passages 6, directly and through the possible further shaped plates 14 that are inserted in such second passages 6.

Preferably, each shaped plate 1 1, 12, 13, 14 is fixed to the two metallic plates 3 that define the corresponding passage 5, 6 ensuring a high mechanical resistance of the heat exchanger 1 to the pressure exerted by the fluids that flow in the passages 5, 6.

Advantageously, the aforementioned metallic plates 3, the shaped plates 1 1, 12, 13, 14 and the bars 4 of the heat exchanger 1 are made in aluminium and are fixed to one another through brazing.

More in detail and for such a purpose, the metallic plates 3 of the battery 2 are advantageously coated on both faces with one layer of braze welding that is made up of a low melting point aluminium alloy. The aforementioned metallic plates 3 are thus locked on one another in a battery holding the lateral bars 4 and the shaped plates 1 1, 12, 13, 14 interposed between them like a sandwich so as to obtain, following oven cooking, the fusion of the low melting point layer and the formation of a single, compact, rigidly assembled exchange core.

According to the present invention each one of the aforementioned turbulence means comprises at least one first shaped plate 1 1 and at least one second shaped plate 12, which are respectively housed in a first portion 18 and in a second portion 19 of a corresponding first passage 5 that are arranged in succession one after the other according to the direction of extension X of the first passage 5 itself. The first and the second shaped plate 1 1, 12 respectively define a plurality of first parallel channels 1 Γ and a plurality of second parallel channels 12' for the outflow in progression of the cooling fluid between the first opening 7 and the second opening 8.

The aforementioned first and second shaped plate 11, 12 respectively occupy a first and a second area of bulk of the first section of the first passage 5 and respectively define a first and a second useful area for the passage of the cooling fluid. Such useful areas are subdivided into the aforementioned plurality of first and second channels 1 , 12' of the two shaped plates 1 1, 12.

The first and the second useful area respectively of the first and of the second shaped plate 1 1, 12 are bigger or smaller with respect to one another respectively at a smaller or larger amount of the component in the liquid state with respect to that of the component in the vapour state.

Since the cooling fluid changes state flowing in the first passage 5, whether it is relative to a use of the exchanger as a cooler or as a condenser, it is necessary to keep the heat exchange of the cooling fluid optimised despite the varying of its conditions or rather of its average percentage values of liquid state and vapour state that it gradually takes up flowing in the first passage 5.

This is obtained according to the present invention with turbulence means that are formed by many shaped plates 1 1, 12 with a different area of bulk that are suitable for varying, in different portions of the first passage 5, the useful area for the outflow of the cooling fluid and thus suitable for varying the ratio between the area occupied by the liquid with respect to that occupied by the vapour in the different portions of the first passage 5.

Whenever, for example, along the first passage 5 there is a high average percentage value of cooling fluid in the liquid state and a low average percentage value in the vapour state, then the percentage area of the first section occupied by vapour must advantageously be low with respect to that covered by the liquid; as the fluid advances in the first passage 5 for example of a cooler, and the liquid evaporates there must be a higher percentage of area that is free from the liquid and occupied by the vapour so as to allow the liquid to have, in any case, a high contact surface for the evaporation and to allow the vapour to not produce too much friction being able to flow in a sufficiently large area.

Analogously, the further shaped plates 14 arranged in the second passages 6 define a plurality of further channels 14' through which the operating fluid that flows out in the second passages 6 is channelled.

Therefore, according to the present invention, the heat exchanger 1 foresees that, the first shaped plate 1 1 is located in the first portion 18 of the first passage 5, in which it defines corresponding first channels 11 ' having a first overall useful area given by the sum of the useful areas of the single first channels 1 1 ', and that the second shaped plate 12 is located in a second portion 19 of the first passage 5, in which it defines corresponding second channels 12' having a second overall useful area given by the sum of the useful areas of the single second channels 12'.

The cooling fluid flowing in succession through the two portions 18, 19 of the first passage 5 (with the interposition of distribution manifolds 50 as shall become clearer in the rest of the description) passes through the first and the second channels 1 1 ', 12' substantially in the direction of extension X of the corresponding first passage 5.

The second shaped plate 12 is made with the second channels 12' having a second overall useful area that is different from that of the first channels 1 1 ', for the outflow of the cooling fluid with a different average percentage value of gaseous phase in the first and in the second channels 1 1 ', 12'.

Advantageously, the different overall useful area of the first channels 1 1 ' with respect to the second channels 12' is obtained by arranging a number of the first channels 1 1 ' (defined by the first shaped plate 11) that is different from the number of second channels 12' (defined by the second shaped plate 12). According to the invention it is possible to arrange, in each first passage 5, many channels in succession having an overall useful area that is bigger or smaller than that of the previous channels, as a function of the increase or reduction of the average percentage value of vapour in the cooling fluid during its movement along the first passage 5.

In accordance with the embodiment illustrated in the attached figures, the first overall useful area of the first channels 1 1 ' of the first shaped plate 1 1 is smaller than the second overall useful area of the second channels 12' of the second shaped plate 12, so that the cooling fluid flows into the first channels 1 1 ' with a smaller section when it has a smaller average percentage value of its mass in the gaseous state so as to optimise the heat exchange, as explained in detail in the rest of the description.

Correspondingly, the first portion 18 of the first passage 5, in which the first shaped plate 11 is housed, is able to be passed through by the cooling fluid with an average percentage value of gaseous phase that is smaller than that present in the cooling fluid when it passes the second portion 19, in which the second shaped plate 12 is positioned.

In accordance with the embodiment illustrated in the attached figures, the first portion 18 of each first passage 5 (in which the first shaped plate 11 with the first channels 11 ' with smaller useful area is positioned) extends from the first opening 7 of the battery 2 and is preferably intended to be passed through by the cooling fluid when the latter has a mass that is mostly in the liquid state with a content that is smaller or rather with a smaller average percentage value of gaseous phase (comprised for example between 0% and 30%). In accordance with the example illustrated in the attached figures 3a,b and 4a,b, three shaped plates 11 , 12, 13 are foreseen that are housed in succession in the first passage 5. Therefore, each shaped plate 1 1, 12, 13 extends longitudinally according to the direction of extension X of the corresponding passage 5 between a first and a second transversal edge 15, 16 that are perpendicular to the direction of extension X itself. More in detail, the second portion 19 of each first passage 5 extends between the first portion 18 and a further third portion 20 that preferably ends in the second opening 8 of the battery 2 and that contains a corresponding third shaped plate 13 which defines third channels 13' in the first passage 5. More in detail, the third shaped plate 13 occupies a corresponding third area of bulk of the section of the first passage 5 and defines a respective third useful area for the passage of the cooling fluid divided into the aforementioned plurality of third channels 13'.

Preferably, in accordance with the embodiments illustrated in the attached figures, the third useful area of the third shaped plate 13 is bigger than the second useful area of the second shaped plate 12. In particular, the third channels 13' of the third shaped plate 13 have a third overall section (given by the sum of the sections of the single third channels 13') with an area that is larger than that of the second overall section of the second channels 12' of the second shaped plate 12 that is positioned in the second portion 19. Such third channels 13' are able to be passed through by the cooling fluid with a third average percentage value of gaseous phase that is greater than the second average percentage value of gaseous phase that the cooling fluid has when passing through the second channels 12'.

Of course, without for this reason departing form the scope of protection of the present patent document, each first passage 5 may also contain only two shaped plates or more than three shaped plates in succession, as a function, in particular, of the characteristics of the cooling fluid and/or of the length of the first passage 5 itself.

Each shaped plate 1 1, 12, 13 thus defines, in the first passage 5 in which it is arranged, a corresponding plurality of channels 1 1 ', 12', 13' that are parallel with one another, preferably extending according to the direction of extension X of the first passage 5, and through which the cooling fluid during its outflow through the first passage 5 itself between the first opening 7 and the second opening 8 of the battery 2, is channelled. Each shaped plate 1 1, 12, 13 preferably has a cross section with a substantially undulated profile that defines a plurality of longitudinal concavities 17 that are parallel to one another and which are defined by the metallic plates 3 of the corresponding first passage 5, thus defining the channels 1 1 ', 12', 13'.

In accordance with the embodiments illustrated in figures 3a,b and 4a,b, the first shaped plate 1 1 defines a number of first channels 1 1 ' that is larger than the number of second channels 12' defined by the second shaped plate 12 (which is preferably larger than the number of third channels 13' defined by the third shaped plate 13).

In such a way, in particular by keeping a substantially even thickness of the shaped plates 1 1, 12, 13, the cross section profile of the first shaped plate 1 1 has a greater number of sides, defining the first channels 11 ', with respect to the number that the second shaped plate 12 has, in such a way occupying a larger area than the first section of the first passage 5 and therefore leaving a smaller overall useful area of the first channels 11 ' free with respect to the second channels 12' .

Therefore, in accordance with the embodiments illustrated in the attached figures 3a,b and 4a,b, the useful area of each first channel 1 1 ' is smaller than the useful area of each second channel 12' which in turn is preferably smaller than the useful area of each third channel 13'.

In accordance with one embodiment that is not illustrated, the different overall useful area of the first channels 1 1 ' with respect to that of the second channels 12' is obtained by providing the first shaped plate 1 1 (defining the first channels 1 Γ) with a thickness that is different from the second shaped plate 12 (defining the second channels 12'), that in turn preferably has a thickness that is different from that of the third shaped plate 13 (defining the third channels 13'). For example, the first shaped plate 11 has a thickness that is greater than the second shaped plate 12, and therefore, by keeping the number of channels substantially the same, it occupies a greater area of the first section of the first passage 5 with respect to the second shaped plate 12, leaving free a first overall useful area of the first channels 1 Γ that is smaller than the second overall useful area of the second channels 12'.

Of course, the different overall useful area of the channels defined by the different shaped plates can be obtained by providing a different number of channels of the shaped plates and, at the same time, by a different thickness of the latter, or also by a different shape of the section of the single channels.

Operatively, for example, in the case in which the heat exchanger 1 is used as a cooler, the cooling fluid is introduced in the liquid state (by means of the corresponding manifold 9) in the first passages 5 through the corresponding first openings 7 of the battery 2. The cooling fluid therefore advances through the first passages 5 absorbing heat, through the metallic plates 3 and the shaped plates 11, 12, 13, from the operating fluid that flows out in the second passages 6 and that has a greater temperature than the cooling fluid. Following such a heat exchange, the cooling fluid evaporates by progressively increasing its average percentage value of mass at the gaseous state as it advances through the first passages 5, until it reaches the second openings 8 completely vaporised.

More in detail, the cooling fluid has a mass that is mainly in the liquid state when it flows out into the first portion 18 of the first passages 5, whereas in the following portions 19, 20 it has an average percentage value in the gaseous state that gradually increases, for example between 30% and 70% in the second portion 19 and between 70% and 100% in the third portion 19.

In accordance with the laws of fluid dynamics known to a man skilled in the art, the increase of the gaseous phase of the cooling fluid tends to cause an increase in the outflow velocity of the cooling fluid itself as it advances through the first passages 5 from the first portion 18 to the third portion 19.

The arrangement of shaped plates 1 1, 12, 13 in succession in the first passages 5 and defining first, second and third channels 1 Γ, 12', 13' with an overall useful area that increases from the first openings 7 to the second openings 8, defines an outflow velocity of the cooling fluid in the second and third portion 19 and 20 that is slower with respect to the case in which the overall useful areas of the second and third channels 12' and 13' remain the same to that of the first channels 1 Γ. This, basically, compensates at least partially for the increase in velocity of the cooling fluid due to the increase in the gaseous phase and, in the same way, it contains the increase of load losses of the cooling fluid itself.

The arrangement of the channels 1 1 ', 12', 13' according to the finding avoids, therefore, important velocity variations of the cooling fluid in the different portions 18, 19, 20 of the first passages 5, and makes it possible to optimise the coefficient of heat exchange in the first passages 5 themselves.

In the case in which the heat exchanger 1 is used as a condenser, the cooling fluid is introduced in the gaseous state in the first passages 5 through the corresponding second openings 8 of the battery 2, and advances along the first passages 5 progressively condensing following the yielding of heat to the operating fluid that flows in the second passages 6 and that has a temperature that is lower than that of the cooling fluid. The cooling fluid comes out from the first passages 5 in the liquid state through the corresponding first openings 7 of the battery 2.

The increase of the liquid phase in the cooling fluid tends to cause a reduction of the velocity of outflow of the same cooling fluid as it advances in the first passages 5 from the third portion 19 to the first portion 18.

The prearrangement in the first passages 5 of shaped plates 11, 12, 13 placed in succession to define first, second and third channels 11 ', 12' and 13' with a decreasing overall useful area from the second openings 8 to the first opening 7, defines a greater velocity of the cooling fluid with respect to the case in which the sections of the first channels 1 1 ' remain the same to those of the second and third channels 12' and 13'. This prevents, also in this case, important variations of the velocity of the cooling fluid in the different portions of the first passages 5 optimising the heat exchange in the first passages 5 themselves.

As illustrated in the embodiment of figure 3a,b and 4a,b, the shaped plates 1 1, 12, 13 contained in the same first passage 5 are positioned in sequence along the direction of extension X of the first passage 5 itself.

In accordance with the idea forming the basis of the present invention, the heat exchanger 1 comprises at least one distribution manifold 50 that is positioned between the first shaped plate 1 1 and the second shaped plate 12, and is suitable for evenly distributing the flow of cooling fluid received by the channels of a shaped plate 1 1 , 12 positioned upstream (according to the direction of outflow of the cooling fluid in the corresponding first passage 5) with respect to the channels of a shaped plate 12, 11 positioned downstream. In particular, with reference to the embodiments illustrate in the figures 3a and 3b, in the passage between each shaped plate 11 , 12, 13 and the next one in the direction of extension of the first passage 5 an aforementioned distribution manifold 50 is interposed, the task of which is to homogenise the flow of cooling fluid received from the channels of the shaped plate upstream before re-introducing it into the channels of the following shaped plate downstream.

Operatively, at the end of the channels of a shaped plate, before the flow of cooling fluid is introduced in the channels of the following shaped plate, irregularities can be created between the layers of cooling fluid itself (due in particular to the fact that in some channels of the shaped plate upstream the cooling fluid has a percentage of component in the liquid state that is different from that in the other channels of the same shaped plate). Such irregularities are advantageously compensated with the aforementioned distribution manifold 50 that remixes the dis-homogeneous layers of the cooling fluid coming out from the channels of the shaped plate upstream with one another, so as to homogenise the cooling fluid on the entire section of the first passage 5.

Advantageously, the distribution manifold 50 is arranged between the second transversal edge 16 of the first shaped plate 1 1 and the first transversal edge 15 of the second shaped plate 12. Preferably, with reference to the embodiments illustrated in figures 3a and 3b, a further distribution manifold 50' is foreseen arranged between the second transversal edge 16 of the second shaped plate 12 and the first transversal edge 15 of the third shaped plate 13.

Advantageously, the distribution manifold 50, 50' defines, for at least one portion of the first passage 5 comprised between the shaped plate upstream and that downstream, a useful section for the passage of the cooling fluid that is greater than the useful areas defined by the shaped plates themselves.

More in detail, the distribution manifold 50 defines, for at least one portion of the first passage 5, comprised between the first and the second shaped plate 11, 12, a useful section that is greater than the first and second useful areas defined by the first shaped plate 1 1 and by the second shaped plate 12 themselves, respectively.

In particular, a very large useful section defined by the distribution manifold 50, 50' results in a smaller load loss compared to the considerable benefit of homogenising the flow and thus determining an even distribution in the different channels.

Advantageously, the distribution manifold 50 comprises a distribution space 51 that extends between the second transversal edge 16 of the first shaped plate 11 and the first transversal edge 15 of the second shaped plate 12.

In accordance with a first embodiment of the present invention illustrated in figures 3a and 4a, the distribution space 51 of the distribution manifold 50 is completely free, and preferably extends for around 2-5 mm according to the direction of extension X. In particular, the aforementioned extension of the distribution space 51 makes it possible to redistribute the cooling fluid without substantially penalizing the mechanical strength of the heat exchanger 1.

In accordance with a second embodiment of the present invention illustrated in figures 3b and 4b, the distribution manifold 50 comprises a group of flow deviators 52 that are arranged in the distribution space 51 preferably positioned spaced away from one another. More in detail, such flow deviators 52 have a mainly transversal development with respect to the direction of extension X of the corresponding first passage 5 for interfering with the direction of the flow of the cooling fluid, in particular by meaning with "mainly transversal extension" the fact that the flow deviators 52 are oriented according to a direction that with the direction of extension X forms an angle that is not smaller than around 45°.

In accordance with the embodiment illustrated in figures 3b and 4b, the flow deviators 52 are arranged perpendicular to the direction of extension X of the corresponding first passage 5, and are preferably obtained for example with the metallic fins.

In particular, again in accordance with the second embodiment illustrated in figures 3b and 4b, the flow deviators 52 are preferably organised in many transversal rows with respect to the direction of extension X, with the deviators 52 of each row that are offset with respect to those of the row next to it.

Operationally, the flow of cooling fluid hits against each deviator 52 and is divided into two semi-flows that move along towards the two opposite lateral edges of the deviator 52, to then move forward according to the direction of extension X once they have passed the corresponding lateral edge of the deviator 52, becoming mixed with the semi-flow of the liquid coming from the deviator 52 beside it (as illustrated for example in figure 3b where the broken lines schematically represent the flow of cooling fluid in one part of the distribution manifold 50). In such a way, the distorted layers of the cooling fluid become mixed with one another making the flow of fluid itself homogeneous before it enters the channels of the shaped plate downstream.

Moreover, the flow deviators 52 are fixed, preferably though brazing, to the two metallic plates 3 (that define the corresponding first passage 5 between them) so as to make the structure of the heat exchanger 1, also at the distribution manifold 50, particularly resistant to the pressures exerted by the fluids that pass through the passages 5, 6 of the heat exchanger 1 itself.

In particular, during the operation of the heat exchanger 1, the metallic plates 3, usually having a thickness of around 0.5-0.6 mm, undergo pressure also of some tens of bar (for example 30-45 bar). The flow deviators 52 fixed to the plates 3 prevent the deformation of the plates 3 undergoing the action of the pressure exerted by the fluids that pass through the passages 5, 6 of the heat exchanger 1, giving a high mechanical strength to the exchanger 1 itself at the distribution manifold 50.

Advantageously, the distribution manifold 50 is provided with a first free volume of the distribution space 51, arranged between the group of flow deviators 52 and the first shaped plate 11 so as to promote the remixing of the cooling fluid in outlet from the first channels 11 ' of the first shaped plate 11 , and of the second free volume of the distribution space 51 , arranged between the group of flow deviators 52 and the second shaped plate 12.

In particular, the aforementioned first and second free volume extend according to the direction of extension X for around 1-5 mm, and preferably for around 2 mm, so as to not jeopardise the mechanical strength of the heat exchanger 1.

Advantageously, the distribution manifold 50, arranged between the first and the second shaped plate 11, 12, has flow deviators 52 that define, between them, a useful section (for the passage of cooling fluid) that is smaller or greater with respect to the useful section defined by the flow deviators 52' of the further distribution manifold 50' arranged between the second and the third shaped plate 12, 13, respectively at a larger or smaller amount of the component in the liquid state with respect to that of the component in the vapour state of the cooling fluid.

In accordance with the embodiment illustrated in the attached figures 3b and 4b, the distribution manifold 50 arranged between the first and the second shaped plate 1 1, 12 (and intended to be passed through by the cooling fluid with a low vapour content) has flow deviators 52 that define, between them, a smaller useful section with respect to that which is defined by the deviators 52' of the further distribution manifold 50' arranged between the second and the third shaped plate 12, 13 (and intended to be passed through by the cooling fluid with a higher vapour content).

Preferably, again with reference to the embodiment illustrated in figures 3b and 4b, the flow deviators 52 of the distribution manifold 50 (arranged between the first and the second shaped plate 11, 12) are organised more tightly packed with respect to the deviators 52' of the further distribution manifold 50' (arranged between the second and the third shaped plate 12, 13).

In accordance with different embodiments of the present invention which are not illustrated, the smaller or bigger useful section of each distribution manifold 50, 50' is obtained by providing flow deviators 52, 52' with a different shape, or rather with different dimensions.

In accordance with the embodiment illustrated in figure 3a, the first shaped plate 1 1 has its second transversal edge 16 facing the first transversal edge 15 of the second shaped plate 12, which in turn has its second transversal edge 16 facing against the first transversal edge 15 of the third shaped plate 13. Advantageously, the first channels 1 1 ', defined by the first shaped plate 1 1, have an longitudinal extension that is rectilinear and preferably parallel to the direction of extension X of the corresponding first passage 5.

Advantageously, the second and the third channels 12' and 13' have a longitudinal extension with an undulated progression. More in detail, the second channels 12' each comprise a series of longitudinal portions that are transversally offset one with the following one, by around half of the section of the second channel 12'. The third channels 13' preferably have a longitudinal extension having a progression with sinusoidal undulations.

The arrangement, according to the present finding, of channels 1 1 ', 12', 13' with overall useful areas that are different (and preferably also with a different longitudinal extension) in each first passage 5 of the heat exchanger 1 , makes it possible to define the velocity of the cooling fluid and at the same time define a different amount of turbulence of the cooling fluid in each portion 18, 19, 20 of the first passage 5, being able in particular to differentiate the turbulence of the cooling fluid as a function of the specific percentage of vapour/liquid present in the fluid when it passes through each portion 18, 19, 20 of the first passage 5.

In particular, the arrangement of the second and of the third channels 12' and 13' with a section that is greater and with an undulated progression makes it possible to generate a greater turbulence of the cooling fluid when this has an average percentage value of its mass at the gaseous state that is greater with respect to when it flows in the first portion 18 of the first passage 5. This advantageously makes it possible, in each portion 18, 19, 20 of the first passages 5, to determine the conditions of turbulence of the cooling fluid that make it possible to optimise the heat exchange.

The shaped plates 1 1, 12, 13 arranged in the first passages 5 are preferably produced by means of shearing operations of a continuous metallic band, in particular obtained by means of a die-cutting press, and subsequent mechanical deformation operations so as to obtain, by means of a succession of folds, the plate turbulence means of the desired shape. Such a turbulence means in general made with a length and a shape that are defined as a function of the portion of the first passage 5 in which it is intended to be located. In particular, the shape and the length of the plates (and therefore the section and the length of the channels) are selected as a function of the characteristics of the cooling fluid, and in particular of the average percentage values of liquid/gaseous phase of the mass of cooling fluid when passing through the different portions of the first passages 5.

The arrangement of many shaped plates 1 1, 12, 13 that are different in each first passage 5, to optimise the heat exchange between the operating fluids, obtains a considerable saving of cost in making the heat exchanger 1, since it is possible to stabilise the section and the shape of the channels as a function of the characteristics of the operating fluids without having to design and make specific shaped plates for each operating fluid used each time. In other words, the present invention makes it possible to make efficient turbulence means in an extremely versatile manner making it possible to satisfy the most varied application requirements through the composition of many shaped plates having different characteristics (or rather many ready-to-use base modules of turbulence means) associating them to one another in a manner that has a simple design as a function of the cooling fluid and, more in general, with the characteristics of the exchanger that is desired to be made. The finding described thus achieves the predetermined purposes.