APPARATUS FOR ISOCHORIC GAS COMPRESSION

Abstract

An apparatus for gas compression comprising: a container containing the gas to be compressed;a first heat exchanger exchanging heat between a high temperature thermal source and the gas, to introduce heat into the gas;a second heat exchanger exchanging heat between a low temperature thermal source and the gas, to extract heat from the gas;supply means of the gas at a supply pressure and delivery means of the gas at a delivery pressure greater than the supply pressure);gas permeable means configured to accumulate and transfer heat to the gas, andgas permeable or gas impermeable movable means dividing the container into a first section in thermal communication with the first heat exchanger and in a second section in thermal communication with the second heat exchanger and in fluid communication with said supply means and said gas delivery means.

Description

TECHNICAL BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to an apparatus for isochoric gas compression, in particular to a gas for industrial use, which uses as driving energy the thermal energy deriving from waste thermal flows preferably from the same industrial plant. The apparatus according to the invention substantially does not therefore require the need for mechanical and/or electrical energy in order to carry out the compression of the gas and consequently of a thermodynamic cycle which, starting from a thermal source, in this case waste thermal flows, makes the aforementioned mechanical and/or electric energy available for carrying out the compression.

2. Brief Description of the Prior Art

Many industries in which gas compression plants are used, have at their disposal large waste thermal flows the thermal energy of which generally can be exploited for the production of electrical or mechanical energy. For this purpose, as is known, recovery organic Rankine cycle (ORC) or Rankine cycle with water vapor, are used.

Compression plants need at the same time a driving energy and a substantial portion of their energy needs is represented by the compression of air or other gases.

According to the known art it is therefore possible to install a recovery cycle (with water vapor, gas or organic fluid), in order to produce with it electric energy for actuating the compressors.

Another known technique is to couple the turbine of the recovery cycle directly to the compressor by using the mechanical energy processed in the turbine for the actuation of the compressor.

These solutions, although having high yields, require high installation costs and involve the presence of rotating machines (turbines and compressors) that require maintenance and can reduce the reliability of the system.

The Applicant has therefore recognized the need for developing an apparatus able to directly use the thermal energy for compressing a gas, without going through a thermodynamic cycle which turns the thermal energy into mechanical and/or electric energy. In this way, an apparatus is obtained which, having a reasonable efficiency, requires low installation and maintenance costs, thanks to its simplicity of construction.

SUMMARY OF THE INVENTION

Purpose of the present invention is to provide an apparatus for isochoric compressing of gas, in particular of gas for industrial use, which uses almost exclusively thermal energy as driving energy deriving from waste thermal flows of the same industrial plant.

Traditionally, for compressing a gas using a thermal source, it is necessary to install a recovery cycle, with which electrical or mechanical energy can be produced in order to actuate the compressors, as represented in a simplified way in FIG. 1. In FIG. 1, the heat exchanger 3′ uses the high temperature thermal source for preheating, vaporizing and possibly overheating a working fluid, for example an organic fluid in an ORC. The vapor is then expanded into a turbine T, then it is condensed into a 4′ condenser; the liquid in the liquid state is then compressed into a pump P by closing the cycle. The mechanical energy produced by the turbine T is used to actuate the electric generator G; the electric energy thus produced is used for actuating the motor M connected to the compressor (any difference between the electric energy generated by G and that absorbed by M is supplied or absorbed by the network to which G and M are electrically connected). The compressed gas can optionally be cooled with a 4″ exchanger. In other configurations, the compressor can be directly coupled to the generator (the drive rotating train would in this case consist of turbine, generator and compressor), or directly to the turbine, instead of the generator.

Instead, the apparatus according to the present invention, although having a limited efficiency, does not require the use of the rotary machines, if not for some auxiliary functions as will be seen below for some embodiments, and therefore it allows low installation and maintenance costs, thanks to its simplicity of construction.

According to the present invention, it is possible to compress the gas and return it up to a temperature close to the suction one, by exploiting a hot thermal source, a cold thermal source and a gas permeable means, hereinafter also called regenerator, which is capable of accumulating and giving its heat to the gas.

According to the present invention an apparatus is described for the isochoric compression of gas, the characteristics of which are set out in the appended independent claim.

Further embodiments of the aforesaid preferred and/or particularly advantageous apparatus, are described according to the characteristics set forth in the attached dependent claims.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will now be described with reference to the accompanying drawings, which illustrate some non-limiting embodiments thereof, in which:

FIG. 1 shows the scheme of a compression system with a recovery cycle according to known art,

FIG. 2 shows a scheme of an apparatus for the isochoric compression of gas according to a first embodiment comprising a container, a heat exchanger which uses the hot source, a heat exchanger which uses the cold source and a regenerator gas;

FIG. 3 shows a second embodiment of the invention, described according to some equivalent variants;

FIG. 4 shows a third embodiment of the invention, described according to some equivalent variants;

FIG. 5 provides details of the operation of the third embodiment of the invention;

FIG. 6 shows a fourth embodiment of the invention with a further regenerator, placed in parallel with the exchanger using the hot thermal source;

FIG. 7 shows the gas flows inside the isochoric compressor, according to the fourth embodiment;

FIG. 8 and FIG. 9 show a fifth embodiment of the invention with further regenerators on both hot and cold sides, in parallel with the corresponding heat exchangers;

FIGS. 10 A-G show various configurations of a sixth embodiment of the invention;

FIG. 11 shows a seventh embodiment of the invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring now to the aforesaid Figures and in particular to FIG. 2, the principle of operation of the invention according to a first embodiment thereof is illustrated.

The apparatus 1 for isochoric gas compression comprises a container 2 within which are positioned a first heat exchanger 3, preferably placed at the top, and a second heat exchanger 4, preferably placed at the bottom, respectively, in order to introduce heat into the system and to extract it. By way of example, the “hot” heat exchanger 3 could be crossed by a diathermal oil, whereas the “cold” heat exchanger 4 could use water cooled by a suitable circuit able to exchange heat with the environment. In the lower section 2b of the container an inlet gas duct 5 is present, which must be compressed and an outlet duct 6 for the compressed gas, both being equipped with a corresponding supply valve V1 and the discharge valve V2. The container 2 is traversed by a gas permeable means, called regenerator 7, movable between a lower and an upper position and able to accumulate and transfer heat to the gas (for instance, made with various overlapping metal mesh layers, which exchange heat with the gas during the crossing of the same and accumulate it within their mass). The regenerator 7 divides the container 2 into two sections, an upper section 2a, at a higher temperature, and a lower section 2b, at a lower temperature. The volumes of the two sections 2a and 2b are obviously variable according to the position of the regenerator 7. The pressures in the two sections are instead approximately the same, being the gas permeable septum.

Referring to the positions of the regenerator 1A to 1E, to the configurations of the valves V1, V2 (0=closed, 1=open) and to the corresponding graphs which, according to the aforementioned positions of the regenerator, respectively represent the regenerator stroke and the gas pressure, the apparatus) works according to the following logic:

• • position 1A: the regenerator 7, with the thermal storage mass at a temperature close to T1, is placed in the upper portion of the container, in correspondence of the first heat exchanger 3 and without the dead volume occupied by the heat exchanger itself; the pressure in the entire container 2 is close to the inlet pressure Pin. Almost all gas present in the apparatus has an inlet temperature T0, except for which remaining inside the dead volume at the top and a portion of the gas remaining inside the regenerator matrix, at a temperature between T1 and T0;
  • the regenerator 7 moves downwards due to an actuator (of the known type and not shown in FIG. 1) which does not perform any mechanical work other than that necessary to overcome the fluid dynamic and mechanical frictions; the regenerator when moving downwards, is crossed by a certain flow of gas and heats the same, due to the previously accumulated heat. The gas also receives a heat Qin from the “hot” heat exchanger 3. In this stage, the valves V1, V2 are closed and a gas is heated in a closed volume, the pressure increases, both in the (hot) portion at the top of the regenerator and in the one (cold) at the bottom of the same, which has a fluid connection with the upper one through the “permeable” matrix of the regenerator;
  • in position 1B, the gas has reached the desired pressure, therefore the valve V2 opens, causing a certain flow of gas to pass to an environment placed at a pressure Pout.

It must be considered that in the hypothesis of a perfect gas, the maximum obtainable pressure is derived from the law of gases P*V=M*R*T which, once applied to the specific case (To and T1, V constant, R constant, M constant) permits to obtain Pout_max=Pin*T1/T0. The valve V2 is actuated at a pressure lower than this maximum pressure, as by this Pout_max a useful outlet flow rate would not be associated (in fact, in the applied gas relation, the volume must be constant). The closer is Pout with respect to Pin (that is, the lower is the required compression ratio), the greater the extractable flow rate for each cycle of the system (at the expense of an obviously moderate Pout/Pin compression ratio).

The system delivers a pressurized gas until the regenerator has reached the lower dead point (position 1C). The supplied gas has approximately the temperature T0, due to the fact that it passes through the regenerator before passing through V2 and in the presence of the “cold” heat exchanger 4 placed in the lower portion of the container;

• • in position 1C (regenerator in the lower position) the valve V2 closes as the system is no longer able to supply gas at the required pressure;
  • the regenerator 7 moves upwards (position 1D) and during this movement cools the gas passing through it (that is, the regenerator heats up). The combined effect of the passage through the regenerator and the transfer of heat towards the “cold” exchanger reduces the pressure, until opening the valve V1 and thus allowing the reintegration of a new gas. The gas inlet continues until the regenerator 7 has reached the upper dead point (position 1E).

The beginning or the end of some phases may not coincide with the upper and lower dead points, as the pressure inside the apparatus, in a certain instant, also depends on the heat input supplied by the exchangers 3 and 4, and not only on the position of the regenerator 7.

In this way it is therefore possible to return the compressed gas to a temperature close to the suction one, by simply exploiting a hot thermal source, a cold thermal source and a gas regenerator, without the need for the addition of other energy as well as a thermal energy, apart from the small fraction needed to cyclically actuate the regenerator inside the container. Preferably, the regenerator 7 performs a complete cycle in a time ranging from 1 to 10 seconds, for example in relation to a container volume 2 of about 1000 liters and a height of about 1 m.

An alternative embodiment of the invention is shown in FIG. 3. This embodiment provides that the regenerator 7a is external to container 2. In particular, in the configuration 3A, the regenerator 7a is external to the container 2 and is fixed, whereas a disc or septum 8 impermeable to gases move within the container; the movement of this movable means 8 is allowed by an outer actuator (not shown in the Figure) and causes the gas to move from the lower section 2b to the upper one 2a or vice versa, by passing through the regenerator 7a which is in fluid-dynamic connection with both sections of the container 2.

The configuration 3B is similar to the previous one, but instead of moving the disc with a mechanical actuator, the gas is moved by one or more fans 9 capable of generating a reversible flow. The disc is made as light as possible and moves accordingly in order to equalize the pressures between the two sections 2a and 2b of the container.

In the configuration 3C also the first heat exchanger 3 and the second heat exchanger 4 are placed outside the container. Obviously, in FIG. 3 possible combinations are shown between the various ones.

A third embodiment of the invention is shown in FIG. 4.

In particular, in FIG. 4A the heat exchangers 3, 4 and the regenerator 7 are placed outside the container. Two recirculation ducts R1, R2 are also present with corresponding fans, in order to uniform the temperature within the two sections 2a and 2b of the container. The disc 8 is actuated by suitable not shown actuators.

The configuration in FIG. 4 B differs from the previous one only for the positioning of the heat exchangers 3, 4 which in this case are housed within the container 2.

The operating logic of the configuration of FIG. 4A or 4B is illustrated with reference to the subsequent FIG. 5.

When the disc 8 moves downwards, that is towards the cold source (FIG. 5A), the cold gas contained in the lower section 2b crosses the duct R1 with a flow equal to the sum of the mass flow m1 corresponding to the movement of the disc 8 within the container 2 (when there is no expulsion or introduction of gas through the valves V1 and V2) and of the flow rate m2 recycled by the fan. The “cold” heat exchanger 4 is crossed by a gas flow rate equal to m2, whereas the regenerator 7a is crossed only by the flow rate m1 corresponding to the movement of the disc 8 within the container 2.

During the movement downwards of the disc 8, the gas passes through the regenerator, heating up and then through the fan 9a and the “hot” heat exchanger 3″. In the heat exchanger 3 passes a flow rate equal to m1+m3, thanks to the flow rate provided by the fan. The flow rate m1 which has left the section 2b of the container is equal to the flow rate that enters the upper section 2a of the container (subtracted the amounts cumulated in the volumes of the respective components), whereas the remaining flow rate m3 cannot but be recycled upstream of the fan for the recirculation duct R2.

When the disc moves upwards towards the hot source (FIG. 5B), the regenerator 7a is always crossed by the flow rate m1, but in the opposite direction. The hot gas contained in the lower section 2a passes through the recirculation duct R2 with a flow rate equal to the sum of the flow rate m1 corresponding to the movement of the disc 8 in the container 2 and of the flow rate m3 recycled by the corresponding fan. The “hot” heat exchanger 3 is crossed by a gas flow rate equal to m3, whereas the regenerator 7a is crossed only by the flow rate m1 corresponding to the movement of the disc 8 in the container 2.

The gas then passes through the regenerator, cooling down and then through the fan 9b and the “cold” heat exchanger 4. In the heat exchanger 4 passes a flow rate equal to m1+m2, thanks to the flow rate provided by the fan. The flow rate m1 which has left the section 2a of the container is equal to the flow rate entering the lower section 2b of the container, whereas the remaining flow rate m2 cannot help but be recycled upstream of the fan for the recirculation duct R1.

The flow rates involved are established by the prevalence of the fan with respect to the load losses of the respective branches. More in detail, it is sufficient to increase the rotation speed of the fan 9a of the hot branch with respect to the rotation speed of the fan 9b of the cold branch, in order to obtain a lowering of the pressure at the node K with respect to the node H and therefore a flow from H to K and vice versa. In this way an alternating flow is obtained without using valves.

The gas enters or leaves the system through the valves V1 and V2 when, respectively, the pressure in the lower portion of the exchanger drops below the inlet pressure or rises above the outlet pressure, with the logic already described for the configuration in FIG. 2.

FIG. 6 shows a further possible configuration, in which the upper “hot” branch has a further gas permeable means, the regenerator 7b, placed in parallel with the “hot” heat exchanger 3.

In previous solutions, the regenerator 7, 7a supplies a great portion of the heat required to bring the gas to a higher temperature, and the “hot” heat exchanger 3 supplies the remaining heat portion. Therefore, in the heat exchanger 3, the heat exchange occurs with the gas being already at high temperature, as the gas has been already heated by the regenerator. Therefore also the heat carrier, for example diathermic oil, which flows into the heat exchanger 3 works with relatively low temperature differences. This phenomenon may adversely affect the ability to perform an effective heat recovery from a gaseous effluent as a low temperature difference of oil affects the ability to effectively cool down the gaseous source, just because the oil remains at relatively high temperatures.

In the configuration of FIG. 5, the central regenerator 7c which is crossed by the entire gas flow, has a smaller size compared to those of the previous configurations and therefore its recovery is lower. This is compensated by the presence of the second regenerator 7b, in parallel to the heat exchanger 3. In this way, the gas which moves upwards and outwards of the central regenerator 7c will exit at the node ‘X’ at a lower temperature (with respect to the case of a single generator 7c). One portion of the gas is then heated by the second regenerator 7b and another portion is parallel heated by the heat exchanger 3 in counter-flow with the diathermic oil. In this way the introduction of heat from the diathermic oil takes place at a variable temperature and starting from a temperature at the lower node ‘X’, has a greater effectiveness then regarding the recovery of the gaseous source for what has been said previously. In the circuit there is present a non-return valve 100 which allows the gas to cross both the regenerator 7b and the exchanger 3 when the flow is substantially directed upwards (that is, when the septum within the container moves downwards) but permits to cross only the regenerator 7b when the flow is directed downwards (that is, when the septum within the container moves upwards).

In fact, when the flow is directed downwards it would be counterproductive to let the hot gas flow through the exchanger 3 which has the function of giving heat and not of absorbing it. On the other hand, the giving of heat from the heat exchanger 3 to the gas is not continuous, but takes place only for a half-cycle of operation, or in any case in an uneven manner.

FIG. 6B shows a similar solution, in which the exchangers 3 and 4 and the regenerator 7b are placed within the container 2, in order to minimize the dead volumes, that is the spaces occupied by the compressed gas which the system cannot expel. The exchanger 3 is placed in parallel with the regenerator 7b; for example, and the exchanger 3 develops on a circular crown inside which there is the regenerator 7b. The gas flow directed upwards in the heat exchanger 3 is prevented by one or more non-return valves 100, for example of the clapet type.

In FIG. 7 the gas flows within the isochoric compression apparatus are shown, depending on the movement of the disc 8 placed in the container. If the disc moves downwards, the “cold” heat exchanger 4 is crossed by a flow rate m1+m2, of which m2 is the flow rate recycled by the lower fan 9c. The gas then goes up the central regenerator 7c, heating up with a flow rate m1, generated by the central fan 9, which can generate a flow in both directions. Then the gas is divided between the “hot” heat exchanger 3 and the second regenerator 7b arriving in the upper section 2a of the container.

Another possible configuration with regenerators in parallel with the heat exchangers is shown in FIGS. 8 and 9, in which the upper “hot” branch has a further gas permeable means (the regenerator 7b), placed in parallel with the “hot” heat exchanger 3 and the “cold” lower branch has a further gas-permeable means (the regenerator 7d), placed in parallel to the “cold” heat exchanger 4. The operating logic corresponds to that already described for FIGS. 5A and 5B, with the addition of two further regenerators 7b and 7d placed on the recirculation branch and in parallel with the “hot” and “cold” exchangers.

FIG. 9 shows a possible arrangement of the regenerators and the fans according to the diagram of FIG. 8. The fans are preferably of axial or centrifugal type and receive at their inlet the outgoing gas from the central regenerator 7c and the second hot regenerator 7b, arranged around the inlet duct of the fan 9a. The flow rate expelled by the fan 9a passes the hot exchanger 3 and arrives inside the container 2. The same arrangement is adopted for the cold side.

FIG. 10A shows a further configuration, in which the septum 8 separating the hot and the cold environment within the container does not translate but rotates around an axis o-o. Its operation is exactly equal to that described for the configuration of FIG. 2. When turning the septum clockwise, this causes the gas to move from the cold to the hot side, by passing through the regenerator 7e. Preferably the septum will be characterized by a peripheral speed (at the point furthest from the axis o-o, and therefore with a greater speed), mediated on a complete working cycle (therefore with return to the initial position) included in the range between 1 and 7 m/s.

For the lower values of average peripheral velocity, a law of motion will be chosen which foresees the length of angular acceleration and deceleration concentrated towards the beginning and the end of the displacement (in order to minimize the load losses).

For higher average speeds, the motion will preferably be close to a sinusoidal motion, in order to minimize the forces of inertia generated by the motion. The hot source is distributed in the exchange pipes (or is collected by them) through suitable collectors 11; the cold source is either distributed or collected by the collectors 12. Duct 5 and duct 6 are respectively the inlet and outlet ducts of the gas to be compressed.

The configuration of FIG. 10B is similar, except that the manifolds supplying the heat exchange tubes are arranged along the axis of the container 2, so that the manifolds 11, 12 and the pipes are fixed to one of the bases of the container. In this way the heat exchangers 3, 4 and 7 and the regenerator can be easily extracted in the axial direction.

FIG. 10C shows a further configuration for the system characterized by a rotating septum around an axis. The configuration is characterized in that the exchanger with the hot source 3 extends for a fraction x of the passage surface available upon crossing the gas pushed by the septum during its rotation about the axis o-o. The 1-x fraction is instead dedicated to a further matrix 7e′ with characteristics suitable for use as a regenerator. The fluid within the matrix can be channeled in a substantially tangential direction, thanks to the presence of non-permeable or low permeable walls NP. In a completely similar manner, as shown in FIG. 10D, a portion y of the access surface can be dedicated to a further regenerator 7e″. Also, in this case, the non-permeable walls NP can benefit the correct flow orientation. Moreover, in FIG. 10D the manifolds are arranged within the container 2. FIG. 10E shows a section of the previous Figure, in which the manifolds 11-12 are evidenced which protrude through the base flange FB, so as to allow an easy removal of the dome of container 2 for maintenance of internal components, in particular regenerators and heat exchangers, with the hot and cold sources.

In FIG. 10F a solution is shown with the heat exchange elements (heat exchangers with the sources and regenerators) which are reproduced in a mirror-like manner, and with a double septum 8 and 8′. The advantage of this configuration is to allow a better balance of the rotating masses, with the consequent possibility of strongly increasing the oscillation speed of the septum, and therefore of increasing the production of compressed gas.

In a further version of the proposed scheme, the mobile septum 8 is also made by an exchange matrix adapted to constitute a regenerator. In this case the flow rate passing through the exchangers 3 and 4 respectively and the matrices of the adjacent regenerators is pushed through said components as a consequence of the load loss which is generated during the motion of the septum. For better clarification, the flow rates are divided between the rotating septum and the fixed exchange components in relation to the pressure loss generated by the flow passing through them.

In FIG. 10G some possible details are added, such as brush seals SP, between septum and dome. The regenerator can be made with different compactness, in order to distribute the load losses and therefore the flow rate in a different way between 7e and 7e′. An eventually unidirectional valve VNR promotes the motion of the gas in a certain direction, for example by limiting its passage through the hot exchanger 3 during the gas cooling phase.

In a further embodiment of the present invention, as shown in FIG. 11, a volume is present with a front surface which is helically increased. The rotating septum and the exchangers/regenerators are helically shaped. The heat exchange elements 3, 4, the gas permeable means 7e are fixed either to the base flange or to the walls of the tank and have non-permeable walls NP and brush-like seals SP connected to the rotating hub. The helical septa 8 are fixed to the rotating hub. The elements just described are clearly visible in the enlargement of FIG. 11: the volume occupied by the tubes of the hot source is indicated with 3, the volume occupied by the tubes of the cold source is indicated with 4 and the regenerator is indicated by 7e.

This embodiment has peculiar characteristics and consequent advantages. The helical arrangement releases the frontal area of the exchangers from the sectional surface of the container. For example, in the rotating configuration with flat heat exchangers/regenerators, the section of the latter can only equal to that of the cylinder. The advantage of the helical arrangement is to permit to greatly increase the frontal area of the exchange matrices, substantially by disengaging it from the area of the axial section of the container.

Moreover, the conical shape gives the system a certain elasticity in order to cope with differential thermal expansions and with the internal pressure.

This solution makes it possible to vary, during the project, the passage surface by varying the number of threads (in the direction of windings) or the angle α.

Further advantages consist in that the masses are roughly balanced around the axis and that the cone resists to the pressure better than a flat surface.

In all the embodiments of the invention, the regenerator must be able to accumulate a relatively large portion of the heat exchanged in the different phases, therefore it must have an adequate overall mass. In order to limit the overall dimensions and optimize the heat exchange, the regenerator for example can be realized as follows:

• • a porous matrix;
  • a set of small diameter wires or tapes which are welded or crushed with each other and are externally crossed by the gas;
  • metal wire meshes or other material suitable for the temperatures of the cycle (that is, ceramics meshes), which are tightly and overlapping arranged.

The present invention, in every one of its possible configurations, can in theory operate with any difference in temperature between the hot and cold source; obviously, the higher the temperature of the hot source and the greater the difference in temperature with the cold source, the better are the performances, in terms of efficiency and compression ratio. Regarding the cold source, this could be made of a water-cooled circuit with air cooler, therefore with water temperatures typically ranging from 10° C. to 50° C. depending on the year season and its location. The hot source could be made of waste fumes coming from an industrial process or from the exhaust of an internal combustion engine or gas turbine, therefore with temperatures typically ranging between 200° C. and 800° C.; however, as the gas/gas heat exchangers need large exchange surfaces, it is more convenient to realize the present apparatus in such a way, that an intermediate heat exchange circuit is formed between waste fumes and gas to be compressed (that is, diathermic oil or molten salt). The Author considers therefore convenient to realize the present invention so that the “hot” exchanger 3 receives at its inlet a carrier fluid at a temperature ranging between 200° C. and 450° C., as such temperatures are sufficiently high to obtain good compression ratios, but, at the same time, remain below the limits of use of the common diathermic oils present on the market.

Within these limits of temperature, the highest compression ratios attainable (Pout/Pin) are roughly comprised between 1.1 and 2.5. These are maximum values being achieved in a closed system, that is without entry or exit of gas (therefore with zero efficiency); the extraction of compressed gas leads to the achievement of lower pressures with respect to the limits set out above: the greater the required flow rate, the lower the pressure reached. According to the Author, a good compromise between discharge pressure and flow rate is obtained for compression ratios ranging between 1, 1 and 2.

The present invention therefore allows to have relatively low compression ratios, for example close to 1.3.

It is therefore particularly useful that the inlet pressure is already high, for example equal to 3 MPa, as with a ratio of 1.3, pressures close to 4 MPa can be achieved.

It is also evident that higher values of compression ratio can be achieved by arranging in series a greater number of apparatuses according to the present invention.

Due to the fact that the apparatus outputs compressed gas in a non-continuous way, the energy is related to the speed of displacement of the separation septum or regenerator (depending on the configuration considered). The Author believes that the system can preferably operate with a cycle time of between 1 and 10 seconds, for example in relation to a volume of container 2 of about 1000 liters and with a height of about 1 m. These time values consider inertias of the separation septum, thermal energies reasonably achievable from the exchangers and the regenerator, energies of fans and therefore load losses.

In addition to the embodiment of the invention, as described above, it is to be understood that numerous further variants exist. It must also be understood that such embodiments are only exemplary and limit neither the scope of the invention, nor its applications, nor its possible configurations. On the contrary, although the above description makes it possible for the skilled technician to implement the present invention at least according to an exemplary embodiment thereof, it must be understood that many variations of the described components are conceivable, without thereby departing from the scope of the invention, as defined in the attached claims, which are interpreted literally and/or according to their legal equivalents.

Claims

1. An apparatus (1) for gas compression comprising: a container (2) containing the gas to be compressed;a first heat exchanger (3) exchanging heat between a high temperature thermal source and the gas, to introduce heat into the gas;a second heat exchanger (4) exchanging heat between a low temperature thermal source and the gas to extract heat from the gas;supply means (5) of the gas to a supply pressure (Pin) and delivery means (6) of the gas at a delivery pressure (Pout) greater than the supply pressure (Pin);gas permeable means (7, 7a, 7b, 7c, 7d, 7e, 7e′, 7e″) configured to accumulate and transfer heat to the gas, andgas permeable (7) or gas impermeable (8) movable means dividing the container (2) into a first section (2a) in heat communication with the first heat exchanger (3) and in a second section (2b) in thermal communication with the second heat exchanger (4) and in fluid communication with said supply means (5) and said gas delivery means (6); the apparatus (1) being characterized in that said first heat exchanger (3) and second heat exchanger (4), said gas permeable means (7, 7a, 7b, 7c, 7d, 7e, 7e′, 7e″) and said movable means (7, 8) cooperate to heat and cool the gas, causing it to cyclically flow between the first section (2a) and the second section (2b), to obtain the compression effect of the gas, which is introduced into the container, from a value equal to the supply pressure (Pin) up to a value equal to the discharge pressure (Pout).

2. The apparatus according to claim 1, wherein said gas permeable means (7, 7e, 7e′, 7e″) are internal to the container (2).

3. The apparatus according to claim 1, wherein said gas-permeable means (7a) are external to the container (2) and in fluid-dynamic communication with both the sections (2a, 2b) of the container.

4. The apparatus according to claim 1, wherein said first heat exchanger (3) and the second heat exchanger (4) are internal to the container (2).

5. The apparatus according to one of claim 1 wherein said first heat exchanger (3) and the second heat exchanger (4) are outside the container (2).

6. The apparatus according to claim 1, further comprising at least one fan (9) which moves said movable means (7, 8).

7. The apparatus according to claim 1, further comprising at least one recirculating duct (R1, R2) provided with a correspondent fan (9a, 9b, 9c).

8. The apparatus according to claim 1, further comprising an additional gas permeable means (7b), which is allocated fluid-dynamically in parallel with respect to the first heat exchanger (3).

9. The apparatus according to claim 1, further comprising an additional gas permeable means (7d), which is allocated fluid-dynamically in parallel with respect to the second heat exchanger (4).

10. The apparatus according to claim 1, wherein said container (2) has a shape corresponding to a solid of rotation and said movable means (8) rotates inside the container (2).

11. The apparatus according to claim 1, wherein said container (2) contains heat exchange tubes, arranged inside and parallel to its axis.

12. The apparatus according to claim 1, wherein said first exchanger (3) extends for a first fraction (x) of the passage surface crossed by the gases and a second fraction (1-x) comprises an additional gas permeable means (7e′).

13. The apparatus according to claim 1, wherein said second exchanger (4) extends for a first fraction (y) of the passage surface crossed by the gases and a second fraction (1-y) comprises a further gas permeable means (7e′″).

14. The apparatus according to claim 10, further comprising at least two collectors (11, 12) extending through a base flange (FB) to allow easy removal of the cap of the container (2).

15. The apparatus according to claim 10, further comprising a double movable means (8, 8′) which is impermeable to the gases and heat exchangers (3, 4) symmetrical to each other and gas permeable means (7e) symmetrical between them.

16. The apparatus according to claim 1, further comprising a one-way valve (VNR, 100) which hinders to passage of the gas in the first heat exchanger (3) during the gas cooling phase, at a greater degree than during the gas heating phase.

17. The apparatus according to claim 10, in that it comprising an exchange volume having a front surface, wherein rotating movable means (8), heat exchangers (3, 4), and gas permeable means (7e), are helical shaped.

18. The apparatus according to claim 17, wherein said heat exchangers (3, 4) and the gas permeable means (7e) are fixed to a base flange or to the cap, while the helical movable means (8) are fixed to a rotating hub.

19. The apparatus according to claim 1, having an average peripheral velocity at the farthest point from the axis, between 1 and 7 m/s.

20. The apparatus according to claim 1, wherein said gas permeable means (7, 7a, 7b, 7c, 7d, 7e, 7e′, 7e″) comprise a porous matrix or a set of wires or strips or even dense metal meshes, said elements being welded or pressed together and externally crossed by the gas.

21. The apparatus according to claim 1, wherein said first heat exchanger (3) is supplied with a carrier fluid having an inlet temperature between 200° C. and 450° C.