AND EXPANSION OF GASES AND VAPOURS

Technical field

The invention refers to a compression process and a similar expansion process of gas and vapors, processes which leads to a progressive increase (respectively decrease) of the gas pressure in the working enclosure, from a p, starting value to a desired p _f value without significantly affecting its average T _m temperature. The methods revealed in the invention for the implementation of the proposed procedures use existing installations (or parts thereof) in the prior art, but also new installations, proposed by this invention. Through the rigorous application of the procedures proposed here, using only techniques and devices experienced in the prior art, a significant increase in the energy efficiency of these devices is expected, therefore, a significant reduction in the energy consumed for compression, respectively an increase in the energy supplied as a result of the expansion. The proposed invention does not stop at these results, but proposes a series of new devices, using which the exergetic performance of the compression and relaxation processes increases even more strongly, by increasing the performance of all technologies in which the compression and/or expansion of gas and vapors have an important share: transport, storage and liquefaction of gases, production of mechanical work using classical and, especially, renewable and waste sources, storage in tanks with pressurized fluids of energy from these sources, treatment and conditioning of air, etc. The invention also contains the description of some installations in these technological fields, to which the application of the described procedure involves a series of structural changes, through which new technologies are obtained, with superior results to those in the prior art.

State of the art

An ideal, perfectly isothermal compression in which the gas and its environment (solid or liquid, surrounding it, or containing it) have the same temperature (ambient temperature T _amb ) throughout the process is not possible (except in the case of a process of infinite duration). Through the compression process, the gas in the closed enclosure receives from the piston a quantity of mechanical energy which it instantly converts into thermal energy, which, if accumulated, would lead to an instanteous increase in the temperature of the gas in the entire volume of the enclosure. The exhaust of this thermal energy becomes possible for the gas with the average temperature T _g , (above the temperature of T _amb ) located in the immediate vicinity of some thermal absorbent bodies in the environment (components of the device in which the compression occurs and other elements located in the closed enclosure). Therefore, in an isothermal compression that takes place in an environment with a constant temperature T _amb , the average temperature T _lz of working gas must also be constant and higher than T _amb . As a result, gas temperatures at any point and average temperatures in different regions can suffer significant variations, even if the average global temperature remains unchanged. In this process, the energy received by the gas from the piston and instantly exhausted into its environment is greater than in the ideal case, where T _iz =T _amb . The speed at which this energy is discharged (and therefore the power of the compression plant) depends on the size of the difference between the two temperatures, on the physico-chemical characteristics of the gas and the materials from which the compressor is made (which contributes to the definition of an overall heat transfer coefficient C _GT ), on the size of the contact surfaces between the gas and its environment (which contributes to the definition of a total heat transfer area A _aT ) and on the distribution of temperatures within the gas. The size of the contact surfaces and how they vary during the compression/expansion process, as well as how temperatures are progressively distributed inside the apparatus, are constructive features. Maintaining a constant average temperature of the gas during the compression process is perfectly possible and can be achieved by maintaining the equality between the mechanical energy ceded by the piston to the gas (dependent on its velocity) and the thermal energy ceded by the gas to heat-absorbing surfaces in its environment. This equality can be achieved by changing the piston speed accordingly.

Fundamental theoretical studies, confirmed by the experimental results obtained, have concluded that from an energy point of view, the most efficient strategy by which a m _g quantity of gas with pressure p ₁ and temperature T _amb (the same in all the mass of the gas), located in an enclosure with temperature T _amb , is brought to p ₂ pressure and T _amb temperature within a time frame t _lz is a three step process (AIA cycle):

- an isentropic compression of the entire amount of gas, within a time interval D t 0, to the pressure p ₃ , corresponding to a working temperature7, _z

- an isothermal compression, within a time interval D t=t _lz , to a pressure p ₄ , which corresponds to the entropy of the gas in the state (p ₂ , T _amb ), during which the average temperature of the gas remains T _iZ , for which continuous change of the piston speed (isothermal trajectory) is required. Any other compression between the state (p ₃ , T _lz ) and the state (p ₄ , T _lz ) occurs within a time interval At>t _iz , or requires higher energy consumption. In the process, the bigger t _iz , the smaller is T _iz . At high T _iz values, high energy values are obtained - an isentropic expansion of the entire volume of gas, from the average temperature T _iz to the average temperature T _amb , from pressure p ₄ to pressure p ₂ , within a time interval D t 0.

If this sequence of operations is followed, the choice of working temperature T _iz is made according to the characteristics of each particular case and is a compromise between the amount of energy consumed in addition to the ideal compression and the duration of the compression cycle (therefore with the power of the compression plant).

For ideal gases, the thermodynamic processes in the current state of the art compressors and expanders are carried out with a polytropic index, between the isothermal and the isentropic index, which differs from one type of apparatus to another, but also varies during the process. In reality, there are few technologies whose optimal progres requires such an intermediate polytropic index. Polytropic compressors are mainly made on the basis of minimum cost and technical limitations of the respective configuration.

For a prescribed average T _jz gas temperature and a constant T _amb temperature for all surfaces of the compressor components / which are in direct contact with the working gas, if known (or approximated sufficiently precisely), at any time, the values of Aft) and hft) for each /, that is, the dimensional and physical characteristics of the material (both those of the components of the compressor and those of its ambient environment), the imposition of the condition to perform the isothermal compression leads us to a first-order differential equation (the equation of the isothermal velocity), in which the only unknown function is the v _lz (t): the time variation over a cycle of the piston speed. (Torok A., Petrescu S., Popescu G., Feidt M., Isothermal compressors and expanders, Revista Termotehnica nr. 2, 2012, http://www.aqir.rO/buletine/1684.pdf, Torok A., Petrescu S., Popescu G., Feidt M., Quasi-izothermal compressors and expanders with liquid piston, Revista Termotehnica nr. 2, 2013, http T/www.aqir. ro/buieti ne/1929. pdf ) . If there is an dependence between h,(t) and v _iz (t), the differential equation is of a higher order. After solving (analytical or numerical) the equation and creating an actuating system that imposes for the piston to move at speed v _iz (t), can be obtained a process in which the average temperature of the gas remains at the T _iz value throughout the f, _z duration of the compression phase, and the energy consumed for this isothermal compression is minimal, compared to any other relation of motion v(t) in which the duration of the polytropic compression phase is equal to t _jz .

Let us note that this type of compression is not reversible (the expansion can only occur if T _lz <T _amb ), so it occurs with losses of energie, as smal as the DT is smaller.

For a standardized temperature difference AT=T _iz -T _amb and a standardized gas (e.g. DT=10K and air), can be defined a standard isothermal start speed equal to the speed v _iz (t ₀ ) from the initial moment of the isothermal compression phase, a size which depends only on the constructive characteristics of the compressor and which can characterize any type of compressor (or expander) in terms of the heat exchange intensity between the gas and its environment.

In the state of the art, there are a series of theoretical studies and a series of installations made based on the results obtained by theoretical and experimental means, which propose “almost isothermal” transformations using variable speeds of the piston. The method starts from the mean values of åhj(t)Aj(t) or from a sequence of values of some parameters measured during piston movement (Farzad A. Shirazi, Mohsen Saadat, Bo Yan, Perry Y. Li, Terry W. Simon, Iterative Optimal and Adaptive Control of a Near Isothermal Liquid Piston Air Compressor in a Compressed Air Energy Storage System, 2013 American Control Conference Washington, DC, USA, June 17-19, 2013, Caleb J. Sancken, Perry Y. Li, Optimal efficiency-power relationship for an air motor-compressor in an energy storage and regeneration system, Proceedings of the ASME 2009 Dynamic Systems and Control Conference, DSCC2009, October 12-14, 2009, Flollywood, California, USA, Andrew T. Rice, Perry Y. Li and Caleb J. Sancken, Optimal Efficiency-Power Tradeoff for an Air Compressor/Expander, Journal of dynamic systems measurements and control, 2017, Mohsen Saadat, Perry Y. Li, Terry W. Simon, Optimal trajectories for a liquid piston compressor/expander in a Compressed Air Energy Storage system with consideration of heat transfer and friction, Conference Paper in Proceedings of the American Control Conference, June 2012). In this research, it is proposed an appropriate change in the speed of the piston according to less rigorous criteria than those proposed by this invention, by mechanical devices, and in the case of the liquid piston, by the variation of the flow of the working liquid, using pressure-swirl nozles.

There are also proposals (RO2013/128401 - Stirling engines, RO2013/128402 - Ericsson high-efficiency engines) for the development of installations to apply this process by introducing some cams in the cinematic drive chain. In the state of the art, there are no installations that rigorously apply the isothermal process.

For each type of compressor in the state of the art, the above-mentioned isothermal velocity v _iz (t) equation can be determined (at least approximately). In this way, we can have an intuitive description of how heat transfer takes place in many concrete situations. The v _iz (t) speed offers, by comparison, indications on the efficiency of the processes used at the prior art, to increase the polytropic coefficient of the gas expansion and compression activities. The state-of-the-art processes aim at increasing, through different methods, the sum of C _G7 =å _/ (the overall heat transfer coefficient): by using materials with a heat transfer coefficient between the gas and the compressor components as high as possible, by increasing the heat transfer surfaces (as in US4446698A - Izotermalizer system), by introducing additional solid or liquid components into the structure of the device such as lubricants, suspended liquid droplets, or jet (among which, US20130327033 - Forming liquid sprays in compressed-gas energy storage systems for effective heat exchange, US20110296822 - High-efficiency liquid heat exchange in compressed-gas energy storage systems, US20120297776 - Fleat exchange with compressed gas in energy-storage systems), by introducing aqueous foam (US20090301089 - System and method for rapid isothermal gas expansion and compression for energy storage), or metal inserts (US20100018196 - Open accumulator for compact liquid power energy storage), by recovering the discharged heat energy (for example, US8851043B1- Energy recovery from compressed gas), etc. The common feature of all these processes is their uneven and limited effect, both in space (having effect only in a limited area of the cylinder) and in time (only for a limited duration of the compression cycle). The overall heat transfer coefficient is dependent on the position of the piston, and the equation that allows the calculation of v _lz (t) is, in this situation, more difficult to solve.

Note that, for a preset /^temperature, any compressor with positive displacement in the state of the art can behave isothermally, if the trajectory of the piston speed/angular speed of the rotor is properly established (complies with the isothermal trajectory). For large temperature differences AT=T _iz -T _amb the instantaneous piston speed can take large values with high energetic consumption, but for small temperature differences DT, the compression cycle time can be very high and the compressed gas flow (implicitly the compressor power) can be very small. The duration of the isothermal compression cycle can be considerably reduced, if the contact area between the gas and the ambient is large throughout the compression cycle, and especially when the piston approaches the top dead center TDC.

An isothermal compression of the gas in a compressor, at the average temperature T _iz , can be obtained if, at any time t of the compression process, the instantaneous thermal energy taken by the gas from the instantaneous work of the piston W(t) is equal to the instantaneous thermal energy åQ{t) taken from the gas by the compressor components (piston, lid, walls, lubricant) which are in direct contact with this gas through the surfaces of Ap, Ac, Aw, A _L respectively. These processes conforms to the first principle of thermodynamics and to Newton's law:

W(t)=åQ _i (t)=åh _i (t)A _i (t)(T _iz -T _amb ), whereQ _/ is the instantaneous heat transfer rate for component / [i=p, c, w, L ), h , (heat transfer coefficient) is a conventional coefficient, characterizing the intensity of this transfer, and W(t) is the instantaneous piston work allocated to compression for moment t (in some cases, as with devices fitted with an elastic thermal sponge, part of the work can be stored temporarily in the elastic elements). Most of the existing technological processes, which consist of gas compression/expansion phases, are either technologies in which rapid changes in the temperature of the gas are required (for which optimal performance is obtained when the polytropic index of compression/expansion is isentropic), or technologies in which it is desired exclusively a transfer of mechanical energy from/to the motor mechanism to/from the gas, without changing its temperature (transfer which is optimal when the polytropic index of the transformation is the isothermal one). In order to achieve these completely different objectives, the constructive characteristics of the compressors and expanders for the two types of processes must be adapted to the respective objective. During the processes, the apparatus for isentropic phases of the process shall have the ratio of the total volume l^of the gas to the total area A through which the gas gives thermal energy to the constituent elements of the apparatus, as high as possible, and the process shall be carried out at higher speed as possible, without creating irreversibility beyond an acceptable limit. Instead, devices intended for the isothermal phases of processes must be characterized throughout this phase by a ratio between the total volume V of gas and the total area A as small as possible and the rate of progress of the process to be the variable speed v _iz , resulting from the compromise between T _iz and f, _z . Due to these major differences in objective and constructive characteristics, for the isothermal compressors and expanders proposed by this invention, we will use further the common name of “isothermalizer”. respectively “densifier” for the isothermal compressor and “rarefier” for the isothermal expander.

Summary of the invention A first objective of the invention is to propose compression and expansion apparatus in which to be realized the increasing (respectively, decreasing) of gas pressure in a working enclosure, from a P _/ Starting value, to a desired p _f value, achieving equality between the initial average temperature and all intermediate average temperatures, including the final one, while respecting an optimal sequence of operations by which the proposed objective can be achieved, while achieving the best compromise between the energy efficiency and the power density of the apparatus. The processes described in this invention and the apparatus proposed for their realization aim to use consistently the sequence of the 3 phases, listed above, of the isothermal compression and expansion, in order to obtain constant temperature differences DT between the gas and its environment. The proposed devices shall be designed in such a way that the values of the instantaneous heat transfer from the gas to/from the constituents of the isothermalizer and from them to the cold/hot source are always as high as possible, This ensures a braking of the increasing trend of temperature difference DT and maximum efficiency for the technological installations using them. Although in many examples that we will describe in order to exemplify the application of this process we will very often refer to densifiers, the apparatus in question also works perfectly as a rarifier. For this it is necessary that the temperature difference T=T _iz -T _amb is negative and that the piston is moved in reverse, in order to take the gas expansion energy and transmit it to the device to be actuated.

A second objective of the invention is that the proposed devices for the compression and expansion of gases and vapors benefit from a standard isothermal start speed v _iz (t ₀ ), as well as instantaneous speeds at any time t of the process (for any compression/expansion ratio, that is, for the entire isothermal trajectory v _lz (t)), larger than that of similar devices in the state of the art. Their use in devices whose piston moves at the new isothermal speed v _iz (t), allows the circulation, in the same time interval, of larger gas flows, keeping the same temperature difference DT, thus an increase in the power and energy efficiency of the device.

The isotermalizers proposed in this invention use to a greater or lesser extent, according to the concrete applications served and the chosen constructive variant, one or more techniques for reducing the polytropic index of compression/expansion, from the current prior art. In addition, in all the proposed variants, between the inner face of the piston (solid or liquid) 5.2 and the cover of cylinder 5.1 , the isothermalizer has a “thermal sponge” fitted (5.3 in Fig.4). This is a deformable body, consisting of one or more solid or liquid component elements, with varying volume and/or position. The solid components of the thermal sponge have the total surface area which is in direct contact with the working gas approximately the same throughout the compression period and their degree of deformation is constantly controlled by the position of the piston (each position of the piston corresponds to a certain shape of the sponge, property ensured by the elasticity of some of its component elements, or by kinematic devices controlled by the movement of the piston). The liquid components of a thermal sponge fitted in reciprocating machines may also play the role of an transport agent for in excess thermal energy, if they are replaced by cooled components during the exhaust and suction phases, or they may take the role of a liquid piston if, during the transformation, the amount of fluid introduced is different from that exhausted. In most rotary apparatus, thermal sponges from solid elements are more difficult to make, but liquid elements of the sponge can be inserted and discharged, both from the suction phase and during the thermodynamic transformation, in the form of jet, of drops, of spray, in the form of foam, etc. When suction/exhaust during thermodynamic transformation, they can also take on both the role of liquid piston and coolant/heating agent.

A compressor fitted with a solid elastic thermal sponge in the form of a helical spring, with a liquid thermal sponge inside it for discharging excess thermal energy is revealed in WO2014005229 - Temperature management in gas compression and expansion (US20140007569). The disadvantages of this type of compressor are due to the complexity of its technical realization and the uneven power required in the isothermal regime. This compressor can be regarded as a combination of a state of the art compressor, located on the cylinder shaft, inserted into a peripheral compressor, with a bulky thermal sponge, which can be quite effective, if the piston speed were to be the isothermal one and would be correlated with the rate flow of coolant in such a way as to ensure a constant temperature of the gas in the enclosure

For the densifiers and rarifiers described in this invention, more energy-efficient solutions are proposed. The easiest to implement is a sponge in the form of a multi-alveolar system, with alveoli that communicate with each other (foam made of elastic, metallic, or non-metallic compounds, natural rubber, synthetic rubber, elastomers, elastic polymers, isoprene, etc., deformable sheets, with large, regular or irregular surfaces). A more complex thermal sponge is made of materials with controllable deformation, with high thermal capacity, preferably metal materials (they also have high thermal conductivity. The system alveoli consist of holes formed between solid or liquid elements of the sponge: various strips or metal plates, cords or elastic bars, springs, foam, membranes, woven nettings of elastic and non-elastic materials, metallic and non-metallic, various products with open alveoli (all communicating with each other), of other types of elastic or non-elastic metallic components, of inflatable bags with elastic or deformable walls, etc. Ideally, the actual volume of the Vb sponge, composed of the total volume of its solid components and that of the gas in the closed alveoli (not taking into account the volume of gas in the open alveoli), does not change significantly for any of the positions of the piston, not even after a large number of compression processes. Some small variations are acceptable, however, if they are oscillations around an average value.

When designing and making the thermal sponge, several objectives must be taken into account:

^■ the maximum thermal capacity

^■ reducing the distance between some point of the volume occupied by the gas and the nearest point of the volume occupied by a solid or liquid element, during the entire ompression process

^■ keeping for as long as possible its thermal, elastic and constructive dimensions from the unstrained state

^■ the total volume of its component elements should be as low as possible (unless the sponge is also a useful deposit of thermal energy)

^■ in its compressed form, the “dead volume” of the enclosure should be as small as possible

^■ creating irregularities that increase the heat transfer coefficients h,

^■ ensuring the smooth flow of gas and liquid throughout the isothermalizer enclosure, to create convective currents favorable to thermal exchange and to reduce friction losses

^■ the deliberate creation, in the volume of gas, of regions with different temperatures, in order to provoke convective displacement of the gas: execution of small holes in the elements of the sponge, use of plates made of metal netting with small mesh, etc.

^■ reducing the possibility of friction between the sponge components, as well as between them and the cylinder walls

The total volume of the sponge Vb is the volume closed by the sponge envelope Sb containing the sponge in its entirety). It is the result of the meeting between the volume Vs of the solid part, almost invariably when compressed and the volume Vg _ln of the gaseous part, composed of the sum of volumes of the alveoli, variable according to the piston position. Alveoli in the volume Vg _ln communicate with each other and with the gaseous medium, whose volume Vg _ex , also variable, lies between the sponge envelope and the compressor shell. If the system also contains closed gas alveoli, they are considered to be part of volume Vs (even if they undergo volume variations in the compression process and also absorb mechanical energy from the energy supplied by the piston). When the piston is at BDC (bottom dead center), Vb=Vi and Sb=Si, and at TDC (top dead center), Vb=Vf and Sb=Sf. The solid elements that make up the solid part Vs each have an individual outer surface, in contact with the Vg _in component of the sponge. The sum of all these individual surfaces is the contact surface Sc, through which the sponge absorbs, in the form of heat, some of the energy introduced by the piston during gas compression. During the compression, this surface shrinks quite a bit through deformation produced by the compression, but it can suffer important undesirable variations when two individual surfaces overlap, removing the gas between them, or when forming a closed alveole.

Also, during the compression process, in different phases of it, in the compressor cylinder (we shall consider the concept of cylinder in its broadest sense, the section perpendicular to the shaft may have some shape, not necessarily circular) a variable volume of liquid V _L may appear, bordered by a variable limiting surface S _L . In areas where the liquid part adheres to the solid elements of the piston or occupies the alveoli of the thermal sponge, the solid surface covered by the liquid is considered to be part of S _L . For example, if all the inner surfaces of the cylinder components are coated with lubricant, thermal transfer occurs between gas and liquid through the surface S _L , and the excess heat is evacuated by conductive thermal transfer between the liquid and the metal components. The volume of the liquid may be due to a single element or a large series of elements, scattered both in the Vg _in composition of the sponge and also outside it, in the volume of gas Vg _ex , between the sponge and compressor walls.

Therefore, when the piston of the densifier is located at BDC, the initial volume of the cylinder Vi _Cyi is composed of a volume of gas Vgi=Vg _in +Vg _ex , the volume of the solid part of the spongers and possibly, the initial volume of liquid V _LI . In the case of solid piston compressors, the volume V _u and the temperature 7 , of the lubricant have insignificant variations in the compression process: a quantity of lubricant equal to that introduced (continuously, or at a certain moment in the cycle) through the lubrication system inlet line is discharged through the discharge line.. Once this lubricant tranche is discharged, a part Q _L of the heat input due to piston action during that compression cycle is also discharged.

In an isothermal compression, in the case of solid piston compressors, as the piston moves toward TDC, V _L and V _s remain unchanged, instead both components Vg _in and Vg _ex of Vg shrink accordingly. A Q, amount of thermal energy equal to the instantaneous W, work done by the piston is instantly transferred to all the gas in the cylinder, which tends to increase its temperature. Gas particles in close proximity to the piston, cylinder cover and walls, lubricant particles as well as those near and inside the thermal sponge transfer some of this heat to them, so that the polytropic index of compression is lower than the adiabatic index. The distribution of temperatures inside the gas volume becomes uneven, resulting in an instantaneous average temperature T _m , of the gas. Also, solid and liquid particles that have received thermal energy from the gas in the cylinder transfers some of this heat to solid and liquid particles in their immediate vicinity, and these transmit it, by conduction, to the rest of the body (liquid or solid), that results in uneven piston, cap, walls, lubricant and sponge temperatures and the appearance of instantaneous average temperatures of them: 7p _m , _; 7c _m , Tw _mi 7 _m ; respectively T _Sm/ . It should be noted that, due in particular to the high ratio of the mass of solid and liquid elements in the process to the mass of gas in the cylinder, at not very high compression ratios (below 100), the temperature of solid and liquid elements increases very little during a single cycle. A significant increase occurs only at very high compression ratios, or after a large number of cycles. For these reasons, at low compression ratios, in explaining thermodynamic processes in the system we can consider the compression process as quasi-stationary, because the average instantaneous temperatures of Tp _ml Tc _ml, Tw _ml T _Lml T _Sml and T _mi do not change during a cycle.

The elastic thermal sponge can be made in such a way that, under the concrete temperature and pressure conditions of the respective process, the total surface area of its constituent components åA _jb (where / refers to the order of the components) remains almost constant throughout the compression period. If this area is large enough, the contribution of the other elements contributing to the overall heat transfer coefficient (e.g. sidewalls) is insignificant. If we admit the previous simplifying hypothesis of a quasi-stationary regime and if we consider the h _jb coefficients constant (as well as the corresponding total heat transfer coefficient åh _jb A _jb ) or having a variation that can be considered linear in relation to the pressure, the system equations leads, in the case of compression, to an exponential decrease in time of v _lz (t). This decrease is much less pronounced than in state-of-the-art compressors, at which the coefficient of å/?A decreases with the displacement of the piston (more pronounced at high pressures). In contrast, when the piston approaches the TBC and the gas pressure is higher, the amount of gas in the immediate vicinity of the thermal sponge is higher, as is the amount of heat transferred to it instantly.

Simple calculations help us compare two similar compressors: a common solid piston compressor, made of metal components, maintained at constant temperature T _amb , the cylinder having the length L of 30 cm and the diameter D of 20cm and a compressor identical to it, of the same length but with diameter of 200 cm (inner volume 100 times higher). Considering the contribution of the sidewall to the discharge of the excess heat, in both cases negligible, the variation curve of the piston speed for an isothermal compression is the same. For an aspirated gas temperature with pressure p _amb and temperature T _iz =T _amb +10°C, an isothermal compression until the pressure pr=7.39 p _am( , (7.39=e ² ), occurs in a 166 seconds cycle, with an exponential variation of the velocity, with the initial piston speed of 3.6 mm/s, and a final speed of 0.49 mm/s. In this way, in one hour almost 20 cycles are carried out and in the first case, about 0.2 m ³ of compressed gas are obtained and in the second, about 20 m ³ of compressed gas is obtained, consuming (and transferring to the environment) an energy 100 times higher.

If in the second compressor we insert an elastic thermal sponge, made of 100 plates of the same metal as the piston and the cover, 0.1 mm thick and about 199 cm diameter, spaced between them by means of elastic spacers, at an initial distance of 3 mm, the desired isothermal compression occurs in the first cycle at the speed v _lz (t), about 100 times higher. If an efficient process for the external discharge of excess heat is also implemented in the system, the isothermal regime shall also be maintained for subsequent cycles. In order to maintain the initial volume of the compressor, the length of the cylinder must be increased by about 1 cm to compensate for the thickness of the sponge in the total compressed state (plus a length corresponding to the implementation of a suitable cooling system), in which case about 2000 m ³ of compressed gas can be obtained in one hour of operation. In an isothermal compression, the mechanical energy transferred by the piston to the gas in the cylinder (whose temperature is T _iz ) and converted into heat is taken up entirely by the components of the compressor (which, being in contact with the environment, transmit part of this energy to it) and by the thermal sponge (which is in contact with the components of the compressor on very small portions, otherwise being in contact only with the gas in the cylinder). Therefore, if the active surface of the sponge is much larger than the active surfaces of the compressor, most of the excess thermal energy is taken over by the sponge, whose temperature gradually increases. If the thermal energy taken by the sponge is not removed, the gas temperature 7^ cannot be maintained at this value unless the piston speed is gradually reduced accordingly. Due to the very high ratio between the density of the gas to that of the solid and liquid elements of the sponge, if the equation of movement v _iz (t), in case of continued keeping, the increase of their temperature (and, consequently, of the gas in the cylinder) is slow, being necessary a significant number N of piston strokes to make this change noticeable in the increase in mechanical power absorbed by the piston actuator (N is greater as the mass of the thermal sponge is greater). If the densifier continues to operate without removing the heat accumulated by the sponge, the mechanical energy received from the outside by the piston is accumulated both as potential energy stored in the compressed gas tank and in the thermal sponge in the form of internal energy. When the temperature of the sponge reaches high values, the energy accumulated in the sponge is equivalent to the potential energy stored in a reservoir of appreciable size, containing compressed gas at an appreciable pressure. These theoretical considerations justify two different strategies for using the densifier: ■ without a sponge cooling system, which allows the accumulation of an appreciable amount of heat energy inside it and its subsequent use

^■ with a sponge cooling system, which allows it to be kept at a constant temperature T _amb , so keeping the temperature of the gas at a value T _iz .

In some applications, the two strategies can be combined. It can be noted that devices equipped with a thermal sponge with a large absorption surface and efficient cooling systems can compress/expanded, even at speeds comparable to those of polytropic devices of the state of the art, large volumes of gas without exceeding a small temperature difference. If the excess temperature is not permanently discharged, but at specified intervals, the inclusion of a compressor and an isentropic expander in the circuit allows (apart from the role of executing the isentropic steps of AIA cycle), by gradually changing the compression ratio and the expansion ratio, maintaining the temperature difference AT=T _iz -T _amb and storing an appreciable amount of thermal energy in the relatively small volume of the sponge.

It should be noted that in this type of compressor configuration and in this mode of piston movement (maintaining constant the diference D T), even if the compression ratio is very high, the division of the compression process into several stages becomes unnecessary. However, a step compression without intermediate heat exchangers, remains useful for a more efficient distribution of available power. Also, the adiabatic compressor which had the role of bringing the gas to the temperature of T _iz , can be replaced with a simpler and cheaper type of compressor (blower) from the state of the art, with the help of which to obtain the initial amount of gas with the temperature T _iz and starting pressure £ _r p ₀ [åi being the compression ratio of this step). In this way, rapid piston movements from the initial phase are avoided and the cylinder portion near the BDC is used more efficiently.

Again, we make the clarification that the previous statements made in the case of gas densifiers remain perfectly valid also in the case of rarifiers, except that in their case the gas is cooler than the thermal sponge and its components, therefore the direction of heat transfer is from them to the gas.

A third objective of the invention is to propose new complex installations, made by incorporating the types of densifiers and rarifiers described above. By using new installations, due to the increase in the performance of compression and expansion processes, the performance of all technologies in which the compression and/or expansion has an important weight increases: transport, storage and liquefaction of gases, mechanical work using classical sources and, in particular, of the renewable and waste ones, the storage in tanks with fluids under pressure of the energy coming from these sources, the treatment and conditioning of the air, etc. Brief description of the drawings

The description of the invention shall be made in relation to the following figures:

- fig. 1 : the principle scheme of the ideal 3-stage isothermal compressor

- fig. 2: the principle scheme of the linear motor-driven isothermal compressor controlled by a pressure regulator

- fig. 3: telescopic rod for hydraulically operated solid piston

- fig. 4: the principle scheme of the thermal sponge

- fig. 5: thermal sponge made of helical springs with rectangular section

- fig. 6: thermal sponge made of elastic cords and horizontal metal plates - fig. 7: thermal sponge made from peripheral mounted helical springs and metal plates

- fig. 7a: example of making the plate-to-spring coupling system

- fig. 7b: thermal sponge made from inside mounted helical springs and metal plates

- fig. 7c: the thermal sponge made from springs and metal plates, in the compressed position

- fig. 8: thermal sponge made of helical springs and horizontal metal plates with vertical fins - fig. 8a: the thermal sponge of fig.8 in the state of maximum compression

- fig. 8b: the thermal sponge of fig.8 in the partial compression state

- fig. 9: thermal sponge made of rigid flat metal plates and arched elastic plates

- fig. 10: thermal sponge made of flat metal plates and arched elastic plates with central storage space; cross section and flat section - fig. 11 : thermal sponge made of horizontal flat metal plates mounted on sliding supports

- fig. 12: thermal sponge made of horizontal flat metal plates mounted with bolts on the supports-harmonics

- fig. 13: the principle scheme of diaphragm densifier

- fig. 14a: the principle scheme of the densifier without dead volume - fig. 14b: minidensifier for collecting compressed gas, attached to a densifier

- fig. 15: system for collecting compressed gas by replacing it with liquid

- fig. 16: extracting a discharge box to a helical spring densifier

- fig. 17: densifier with storied liquid piston

- fig. 18: densifier with storied liquid piston and sprinklers - fig. 19: horizontal section of storied liquid piston densifier in fig.17

- fig. 20: detail of the installation of the sprinklers in the densifier with storied liquid piston

- fig. 21 : configuration resulting from the combination between a solid piston densifier with sponge mechanically deformable and a liquid piston densifier

- fig. 22: configuration resulting from the combination of the solid piston densifier of Fig.7 with a liquid piston densifier

- fig. 23: configurations resulting from the combination of a liquid piston densifier, with a group of solid piston densifiers and thermal sponge with elastic elements

- fig. 24: densifier with liquid piston introduced into the enclosure by spraying

- fig. 24a: liquid piston densifier with uneven distribution of the elements that make up the thermal sponge

- fig. 25: the principle scheme of a gas piston densifier, thermal sponge with horizontal plates, cooling with sprinklers and foam generators

- fig. 26: the principle scheme of a gas piston densifier with conveyor belt

- fig. 26A: alternative configuration for fitting the conveyor belt to the densifier in fig.25 - fig. 27: gas piston densifier with gas/liquid and liquid/medium heat exchangers

- fig. 28: gas piston densifier made by coupling two identical densifiers

- fig. 29: gas piston densifier with foam and liquid jets cooling

- fig. 30A: rotary densifier with a blade in the rotor and cooling with sprinklers

- fig. 30B: pressure step densifier, with 3 overlaid densifiers with a blade in rotor - fig. 31 : pressure step densifier, with 2 side by side densifiers with a blade in the rotor

- fig. 32: processes for the improvement of the rotary densifier with a blade in the rotor

- fig. 33: rotary densifier with rolling rotor and telescopic blade in rotor

- fig. 34: rotary densifier with a blade in the rotor and thermal sponge made of cylindrical tubes

- fig. 35: thermal sponge drive system of the rotary densifier with a blade in the rotor - fig. 36: piston-rolling densifier

- fig. 37: rotary vane densifier and cooling with external sprinklers

- fig. 38: rotary vane densifier and cooling with internal and external sprinklers

- fig. 39: rotary vane densifier, with thermal sponge and cooling with external sprinklers

- fig. 40: double-effect solid piston densifier with profiled cam drive, with horizontal plate, thermal sponge and sprinkler cooling

- fig. 41 A: gears densifier and sprinkler cooling

- fig. 41 B: cams densifier with sprinkler cooling

- fig. 42A: scroll densifier with thermal sponge from spiral plates

- fig. 42B, C: mounting variants of the thermal sponge of the screw densifier - fig. 42D:: peristaltic densifier

- fig. 42E, F: mounting variants of the thermal sponge of the peristaltic densifier

- fig. 43: improving the efficiency of internal combustion engines by implementing a densifier

- fig. 44: active thermal insulation system with recuperator fluid

- fig. 45A: T-s diagram of the processes in the ACAES energy storage system - fig. 45B: new ACAES system for energy storage

- fig. 46: T-s diagram of energy storage system processes ICAES+ACAES

- fig. 47: combined ICAES+ACAES system for energy storage

- fig. 48: ICAES system with hybrid thermodynamic cycle, for energy storage in thermal tanks

- fig. 49: ICAES system with hybrid thermodynamic cycle, for energy storage in thermal tanks and gas tanks under constant pressure

- fig. 50: system for sterilization and cooling, respectively heating the air in an enclosure, consisting of two loops, one for air and one for working gas

- fig. 51 A: system for thermodynamic sterilization of air, consisting of an adiabatic compressor and an adiabatic expander, between which is mounted a 4-way valve with large passage sections

- fig. 51 B: isothermalizer whose exhaust/inlet is made through a 3-way valve with large sections of passage, in a tank under constant pressure

- fig. 52. system for cooling or heating the air in an enclosure, similar to that in Fig.50, in which the working gas loop operates in a Stirling cycle - fig. 53: T-s diagram of the liquefier in Fig.54

- fig. 54: gas liquefaction system, which operates in a Siemens cycle, the condenser being cooled with a heat pump

Description of the best embodiments of the invention Achieving the first objective of the invention (moving the mobile organ of the device at such a speed that a constant difference DT between the average temperature of the gas and the temperature of the components bordering it is kept at all times) it results in maximum exergetic efficiency for any known type of positive displacement compressor/expander, especially the alternative ones. Simple replacement of the most commonly used drive system, the crankshaft-type one (which imposes a quasi-sine variation in piston speed), with a system driven by a variable power motor, directed by a controller designed to generate the isothermal speed v _iz (t) (in the case of piston devices, this speed is usually described mathematically by a downward exponential curve) can lead to significant reductions in energy consumption in any installation. In addition, moving the piston at this speed (in the case of reciprocating apparatus) or changing the angular speed of moving organs (in the case of rotary apparatus), allows to be kept the average temperature of the gas at all times, between two close limits and offers the possibility of harnessing the energy contained in low potential sources (the absolute temperature of which is only a few percent above the ambient temperature).

In the case of rotary devices containing, for the compression/expansion of the gas, a single working chamber (for example, compressors with a single blade in the rotor), keeping the temperature between two close limits shall be done by varying the angular speed of the rotor during each rotation, and in the case of those with several enclosures with different volumes, each in different compression phases (lobe compressors, geared compressors, scroll compressors, screw compressors, etc.), the speed can be kept constant, but in each enclosure the coolant/heating flow and the flow of liquid piston must be varied. As we have already pointed out, this first objective is achieved if the three stages of isothermal evolution are rigorously observed, in a single device, or by combining several distinct devices. The device in Fig.1 is composed of three distinct elements: isentropic expander 1 , isothermal densifier 2 and isentropic compressor 3. Depending on the direction of gas flow through the device, it performs compression or isothermal expansion. In Fig.1 A, on a T-s diagram, the corresponding temperature variation is shown, as well as the mechanical work consumed to compress the gas in a cycle.

The isentropic compression stage may be replaced for a small difference 217 by a polytropic compression, if this leads to a significant cost reduction and a significant increase in the initial pressure. Similarly, the isentropic expansion stage can be replaced for a sufficiently large difference 217, by an isobaric cooling stage, if the heat energy thus recovered can be used efficiently (e.g. for mechanical energy production). The corresponding temperature variation after replacement of this isentropic components is shown on the T-s diagram in Fig.l B. Simultaneously, the device must also fulfill the second objective of the invention: to accelerate the transfer of heat from the gas to its environment. Therefore, a thermal sponge is inserted inside the device charged with the isothermal transformation. The simplest thermal sponge and its cooling system can be the lubrication system and the lubricant cooling system with the compressors of the state of the art are fitted, if it transfers a sufficiently large amount of heat.

In order for the isothermal compression/expansion operation to be carried out under the most economical conditions, the determination of the isothermal velocity v _iz (t) must be as accurate as possible (which implies an exact determination of all the quantities involved in the differential equation describing the phenomenon and of all the correlations between these quantities), and the actuators designed to achieve this speed must be robust, have the lowest response time possible and have a sufficiently large adjustment range. The proper modification of the piston speed can be made by making suitable kinematic chains, driven by variable speed motor assemblies (preferably direct current motors, linear motors, stepper motors). In the case of the liquid piston and in the case of the hydraulically operated solid piston, the variation in the flow of the working liquid can be done using variable section nozzles, pressure-swirl nozzles.

If, for some types of installations these determinations can be made with acceptable accuracy (especially for large areas of the thermal sponge, for large diameters and small lengths of the cylinder of the device concerned, and for small temperature differences 217), other types require empirical research and a significant amount of test bench measurements. In carrying out these determinations, it should be borne in mind, that a small improvement in the operation of the prototype translates into significant total energy gains when switching to series production.

One method that can lead to very good results is the automatic real-time adjustment of the actuator power, which avoids performing calculations, sometimes laborious, for determining the v, _z . For this, it is necessary to instantly measure the main gas state parameters using small sensors, that are enough sensitive and whose response time is fast enough. The collected signals shall be transmitted to a computing device, which produces an adequate response, transmitted to the actuator. For example, if the gas temperature is measured at enough points to determine an instantaneous average temperature T _mi , the controller (the calculating and adjusting device) compares the measured value with T _lz and transmits an acceleration, or slowing (or even stationary) command to the actuator to achieve equality T _mi = T _iz . Also, temperature sensors mounted in certain regions of the apparatus can transmit signals that, after processing, cause to be sent commands to the flow regulators on the coolant agent pipes. In this way, an isothermal evolution can be achieved, regardless of the temperature of the external environment, the temperature of the components of the compressor, of the thermal sponge or of the cooling agents. Moreover, the information collected can be used to adjust the lubricant flow, the spray coolant flow, etc.

Automatic power adjustment of the actuator does not exclude in its entirety, the calculation of the isothermal speed curve v _iz (t), or the corresponding variation of the position X _iz (t). On the contrary, knowing as accurately as possible a presumptive value allows computing devices to respond more quickly, the signal from the measurement transducers being used to correct the response provided by the theoretical knowledge of the speed (or position) of the piston.

At any given time, between the instantaneous values of volume V„ pressure p, and average temperature T _ml . there is a well-established relationship, depending on the molecular structure of the working gas, given by the law of gases. For example, if we can consider that the ideal gas model can be chosen for this gas, we will use the PiVj=nRT _mi relationship, where n is the number of gas moles in the cylinder, and Ri s the molar gas constant. Therefore, in order to achieve the equality of T _mi =T _iz =constant, there must be a direct correlation between pi and Vi. Also, for every V, there is a well-determined piston position: X,=f(V )). Consequently, a computational relationship can be inferred: Xi=f(Pi). It is preferable to choose the pressure as the control parameter, as its value is the same throughout the gas mass, and a single sensor is sufficient for its measurement. There are also simple, inexpensive, sufficiently precise transducers that give an extremely fast response (for example, piezoelectric ones). Figure 2 shows the scheme of principle for this type of installation. In this case, the processor (controller) 12.4 (DC), compares the pressure measured by the pressure sensor 12.5, with the corresponding one in an isothermal transformation, the working temperature T _iz and the position L , of the piston at that time. Depending on the result, DC sends the appropriate control to the drive system 12.3 (here, based on a linear motor), moving the piston 12.2 which moves in cylinder 12.1. The excess heat of the gas is absorbed by the thermal sponge 12.6 and the other components of the densifier.

The mode of operation of the pistons/rotors of these types of densifiers shall be chosen according to the objective pursued by the gas compression. When the introduction of the thermal sponge into the cylinder structure is aimed only for reducing the energy consumed by lowering the polytropic index, the piston is operated by one of the procedures of the prior art that are in current use. Flowever, if compression is to be as close as possible to the isothermal compression (especially when a high compression ratio is also being sought), it is necessary to use a system that instantly changes the piston speed according to its position (instantaneous compression ratio). One of the methods used for this purpose is the implementation of an appropriate cam and similar devices in the drive system. Another procedure according to this objective is that of the classic DC electric motor, or linear motor (usually mounted on the piston rod), with constant voltage and variable current. This process is suitable for isothermal compression, as it involves equality between the heat energy given by the gas to the sponge and to the densifier component elements and the mechanical energy given to the gas by the piston. In the case of a constant global heat transfer coefficient, this equality involves keeping constant current supplied to the engine.

Large compression ratios involve large differences between the instantaneous work done by the piston at the start and at the end of the compression process. For this reason, in these cases it is recommended to use a hydraulic drive, in which the force that causes the piston to move is the pressure of a fluid supplied by a hydraulic motor with variable speed (or variable flow). Therefore, the large difference in mechanical work between different stages of the process translates into large variations in fluid flow. In some state-of-the-art systems, reducing the flow gap between the different process moments is solved by the combined use of a solid and a liquid piston. In Fig.3 and Fig.3. a, a telescopic solid piston, hydraulically actuated, is shown, with the piston at the BDC position and in an intermediate position, respectively. The master piston 13.2 slides inside the master cylinder 13.1. In the configuration shown in the figure, the first part of the piston stroke is divided, by the telescopic construction attached to the piston rod 13.3, into 4 segments of equal length and a segment of variable length, but both the number of segments and their length are at the discretion of the designer. The portion 13.3 of the rod is rigidly attached to piston 13.2 in its center, and an outer ring 13.4 is attached to the opposite end, larger than the diameter of the rod. Sections 13.3i (13.3a, 13.3b, 13.3c and 13.3d), are ring cylinders, which at the top have attached an inner ring 13.5i, and at the bottom have attached an outer ring 13.4i.

These cylinders have an inner diameter equal to the outer diameter of the outer ring of the previous segment, and their outer diameter is equal to the inner diameter of the inner ring of the next segment. The outer rings of each segment slide on the inner surface of the next segment, and the inner rings slide on the outer surface of the previous segment, the seals 13.6 providing the sealing. In the configuration shown in the figure, the space between the bottom surface of the piston and the upper surfaces of the inner rings, as well as that between the outer surfaces of a segment and the inner surfaces of the next segment are vacuumed. Configurations may also be made, in which this space is occupied by a liquid or gaseous fluid at atmospheric pressure or a different one, if to this fluid is assigned an external reservoir and a series of flexible and fixed pipes for its proper circulation.

A suitable range of piston speed trajectories can also be obtained from the combination of the motion of a solid piston driven by a mechanical device, with the additional movement of extending its rod, movement due to the hydraulic power of a liquid agent.

When the piston is in the BDC , the fluid supplied by a hydraulic motor penetrates through gate 13.7 and presses on the outer ring of segment 3.3, the surface of which is much smaller than the surface of the piston and, as the gas pressure in the densifier is reduced, the piston speed will be high. As the gas pressure in the densifier increases, the piston speed decreases. When this outer ring steps on the lower surface of the inner ring of the segment 13.3a, the displacement is also transmitted to this segment, which causes the working fluid to penetrate below the lower face of its inner ring. As the active surface increases, the piston speed tends to make an upward jump, but a suitable change in the liquid flow at this time keeps it constant, to decrease again as the piston moves. The upward jump of the active surface of the piston is repeated each time an outer ring of a segment steps on the lower surface of the inner ring of the next segment. When the segment 13.3d is driven in motion, the active surface becomes equal to that of the piston and its movement continues, without jumps, at a decreasing speed, until the piston power equals that required to compress the gas to the desired pressure. The engine power may be exceeded if the telescoping continues in the same way, with ring segments with the inner surface of the outer ring larger than the diameter of the piston (and the densifier cylinder), adding an additional cylinder with the corresponding diameter.

In terms of increasing the exhaust speed of the thermal energy taken instantly by the gas at any piston movement, a huge variety of configurations can be achieved (depending on the characteristics of the application in which the need for isothermal compression occurs) through which, in the compressor construction is implemented a thermal sponge, for the purpose of increasing the power of the isothermal process or, only to decrease the polytropic coefficient and the power consumed in polytropic processes. Figures 4 to 13 show some examples of fitting a thermal sponge in solid piston densifiers (they use fixed amounts of liquid, used for lubrication, for cooling the sponge during compression and for discharging the compressed gas remaining in the cylinder when the piston reaches the TDC (without any role in gas compression). The simplest configurations are obtained by changing the configuration of the compressors from the current state of the art, by inserting in such a device a thermal sponge made according to the previous descriptions.

The principle diagram of the compressor fitted with a thermal sponge is shown in Fig.4. It consists of housing 5.1 (composed of the cover with check valves 5.5 and side walls), piston 5.2 and thermal sponge 5.3. the check valves are only one of the solutions for the suction and exhaust of the working gases.The complexity of the installations requires that, in many configurations, the circulation of gases, lubricants and cooling fluids is directed by a variety of taps, flaps, valves mechanically or electrically operated, etc., which in most drawings will be represented only schematically.

The operation of the densifier is the same as that of a sponge-free compressor: the gas suction is via the inlet valve, by moving the piston from top dead center TDC to the bottom dead center BDC, with the exhaust valve closed and compression to the desired pressure p _f , by moving it from the bottom dead center BDC to the T point, with both valves closed, during which time the heat transfer from the gas to the sponge takes place. When the piston reaches this point, the exhaust valve opens so that the gas with pressure p _f is exhausted to the desired destination, by moving the piston from T point to top dead center TDC. Most of the time, the gas is exhausted under constant pressure p _f consuming only the displacement work W _d =p _r Vg _f .

The isothermalizer in Fig.5 illustrates this type of configuration: the thermal sponge 5.4 of this isothermalizer is a helical elastic metal spring with the rectangular turn section, one end of the spring being fixed to cover 5.1 , the other to piston 5.2. With the piston in the BDC , the spring 5.4 is in a de- tensioned state (or slightly pretensioned). In the illustration, the piston is in an intermediate position.

We note that even in the situation where, with the piston in the TDC, the spring turns are stuck together (the tension is maximum), the dead volume of the isothermalizer (which includes a cylindrical space with the diameter equal to the inner diameter of the spring and the annular space between the spring and the walls of the cylinder) is quite large. It can be reduced by introducing additional springs with ever smaller diameters into this space (the inner diameter of an intermediate spring is approximately equal to the outer diameter of the next spring), increasing the heat transfer surface. Another method is that proposed in US20140007569, in which a cylindrical piece with a diameter almost equal to the inner diameter of the helical spring with the smallest diameter (Fig. 5a) is attached to piston 5.2 (or to the cylinder cover 5.1). The length of this part is chosen according to the densifier compression ratio. At the limit, when the piston is in the TDC, it can fully occupy the volume inside the fully compressed spring.

Another constructive variant is represented by the isothermalizer in Fig.6 and the one in Fig.7, where the main components of the thermal sponge are horizontal, parallel plates (preferably metal), 5.11. These plates shall be attached to a vertical elastic rope system 5.7a, inserted into the elastic bellows 5.7b, respectively to a vertically mounted helical spring system 5.12, with an outer diameter smaller than that of the isothermalizer springs in Fig.5 (where they were the main component of the thermal sponge). The space inside the elastic helical springs 5.12 (circular, rectangular, etc. section) can be occupied by a sequence of springs 5.13, with appropriately decreasing diameters, and the space inside the spring with the smallest diameter can be partially occupied by a cylinder similar to cylinder 5.2a of Fig.5. In Fig. 7 a guide rod system 5.7c attached to the piston is proposed, which pierce the cover through holes fitted with sealing gaskets 5.8 (or vice versa). If tubes are installed instead of rods, through which cooling fluids circulate, the problem of removing the excess heat can be solved completely or partially. These tubes may also be fitted with sprinklers to spray droplets of coolant, or aqueous foam. A process that achieves the same objective is the use of thermal tubes with vapor at the saturation limit. These rods also eliminate the possibility of side movements of the thermal sponge and the possibility of it coming into contact with the cylinder walls.

In order to keep the elastic properties of springs 5.12 unchanged, it is most often necessary to fix plates 5.11 by means of another deformable material, which can be rigidly fixed to horizontal plates 5.12a as shown in Fig. 7a. This coating can be continuous, over the entire length of the turns, or it can be made of rings 5.12a, mounted one for each turnof the elastic spring. The fitting must be made in such a way that the horizontal plates are not subjected to mechanical stresses when, due to the compression of the spring, the outer diameter of the spring undergoes slight changes. In Fig. 7a are represented two sections that illustrate the evolution of movement between the turn 5.12 and the support 5.12a rigidly fixed to the corresponding horizontal plate. The fixing of each plate 5.11 on each of its support springs 5.12 is done at a single point. To allow the free movement of the plates, at each spring, crossing windows are practiced in the plate. The windows that provide the largest absorbent surface of the board are circular, with a diameter equal to the maximum diameter of the support spring. In Fig. 7, the fastening system is illustrated when the springs are located at the edge of the cylinder, and in Fig. 7b the fastening system when the springs are inside them. The number of plates mounted in these systems can be very large, thus ensuring a large heat transfer surface and, consequently, a high piston speed, or a very small difference between the temperature of the gas and that of the thermal sponge.

At the limit, when the piston is in the TDC, the horizontal plates can be very close to each other, even bonded (Fig. 7c). Even when the horizontal plates come into contact with each other, many spaces remain filled with compressed working gas. For this reason, especially in the case of large compression ratios, it is appropriate to execute horizontal channels and vertical holes so that, during the exhaust phase, these spaces can be filled with a liquid agent located inside (liquid component of the thermal sponge), or introduced specifically for the removal of gas. In this way it is possible to achieve a total exhaust of gas from these spaces with minimal energy consumption. To ensure maximum plate density, the number of elastic cords 5.7a (Fig.6) and the support- springs 5.12 (Fig.7) can be multiplied, these forming distinct sets, and the corresponding sets of horizontal plates 5.11 are mounted interspersed on them.

Figure 8 shows an isothermalizer configuration in which, starting from the configuration in Figure 7, the surface area through which heat absorption, by the thermal sponge, from the gas being compressed is considerably increased by the installation of fins, or other vertical elements 5.10. This leads to very high compression ratios without the excessive increase in the working gas temperature. The density, the arrangement on the support plate, the thickness of the plates, etc. may differ from one horizontal plate to another. There is a very high degree of freedom in the shape of these fins, their dimensions (a large thickness ensures a slower increase in the temperature of the sponge, a smaller distance between the fins ensures better cooling of the gas, a larger width of these reduces the required number of horizontal plates, a larger diameter of the holes practiced in both vertical and horizontal plates ensures efficient convective circulation and less friction losses in the liquid to be inserted at the end of the compression, for the exhaust of compressed gas, etc.). This variety also provides multiple options for choosing the isothermal trajectory v, _z . The smaller number of horizontal plates ensures that the horizontal stability of the thermal sponge is ensured by guide rollers, with little friction in the bearings. In Fig.8, the thermal sponge of such an isothermalizer is shown, with the piston in an intermediate position, and in Fig.8a is the same sponge with the piston in the TDC. Vertical fins can be made with variable length by telescoping with elastic elements, and those mounted on the lower face of the plates by gravimetric telescoping (Fig.8b). In this configuration, the density of heat- absorbing vertical plates varies depending on the compression ratio, and when the piston approaches the TDC the size of the gas regions is greatly reduced.

Configurations may be made in which vertical fins are walls separating laterally distinct areas of the cylinder (in a horizontal section, they are a sequence of concentric circles, or rectangles with increasingly smaller sides, or other geometric figures placed in each other, or side by side). In this case, the horizontal plates have the raised edges along the entire contour (5.11c, figure 8), like trays. This allows a liquid layer to accumulate permanently or periodically on the horizontal plates (5.11 d, figure 8b). The height of the tray edges will determine the size of the liquid fraction of the sponge 5.11a. The vertical movement of the piston causes the vertical fins to move, penetrate the liquid layers on the horizontal plates and discharge liquid into the liquid layer formed on the top surface of the piston. In this way, the liquid is circulated among the plates, favoring the transfer of heat.

In another configuration of the isothermalizer in Figure 8 the thermal sponge is made without the elastic springs between the horizontal plates. In this way, the action of the piston is transmitted from one horizontal plate to another successively, not simultaneously. In this configuration, mechanical or electromechanical locking-unlocking mechanisms of horizontal plate movement can be installed in the side walls of the densifier, allowing a diversification for the shape of the curves of the isothermal velocities.

If the liquid fraction of the thermal sponge remains, during a cycle, at the same value, holes for gas circulation can be applied in the vertical plates. If a permanent flow of the liquid fraction is ensured and the size of this fraction can be changed and controlled by changing the flow rate of fluid introduced and discharged from the cylinder, holes can be also applied in the horizontal plates,. This results in the change of both the speed variation curve and the time and space distribution of the heat energy accumulated by the thermal sponge.

The isothermalizers in Fig. 9 and 10 also have thermal sponges made of elastic and non elastic metal components which, when the piston is in the TDC, occupy almost all the inner volume of the isothermalizer. The one in Fig. 9 is constructed by alternating flat plates 5.11 , which slide on an equal number of arched plates 5.14, the whole assembly being stabilized by a rod 5.7, which has one end fixed to the cylinder cover, and the other end pierces the piston through a hole made in the piston and sealed with the sealing gasket 5.8. Rod 5.7, in all the configurations in which it is used (e.g. Fig.6, Fig.11 ) has a non-circular section and is located in the cylinder axis, preventing the stabilized plates from rotating. If a larger number of stabilizing rods is required and some of the plates are deformable, they may be stabilized by rectangular slots with a width equal to that of the rod and a length equal to the magnitude of the voluntary movement. The isothermalizer shown in Figure 10 is similar, but the solid sponge plates are made by the welded connection 5.15a between the rigid planar plates 5.11 and a series of arched elastic bands 5.15, among which they can be mounted (on a single level, if in the TDC the arched plates become flat, or on several levels if in the TDC the arched plates maintains adequate curves) horizontal plates with corresponding widths, with two slip points (bolts mounted on the elastic bands, sliding into grooves executed in the thickness of the horizontal plates). The compressed gas is collected in the space between the horizontal plates, as well as in an inner parallelepiped space 5.2b. The dimensions of the collection space are set by the width of the arched strips 5.15, and the height of the inner parallelepiped collection space is adjusted by the dimensions of the piece 5.2a, fixed on the moving piston. In section 1 -1 , an inner, top view of the system is shown.

The isothermalizer in Fig.10a is similar to that of Fig.9, but arched plates with a different number of curves with different radii of curvature, are mounted between the horizontal plates. As in previous configurations, the highest absorption power is obtained when all the plates have the same surface, close to the section surface through the cylinder, and in the TDC they perfectly overlap.

This type of thermal sponge can be used to reduce the energy consumption of state-of-the-art compressors having a superunitary polytropic index, compressors for which the main objective is not to achieve an isothermal compression, but to obtain a large volume of compressed gas in as short time as possible. This objective can be achieved in a more economical way than at the state of the art (where the desired compression ratio is obtained by staged compression, intercaling some heat exchangers between these stages), by inserting a thermal sponge with a maximum absorption surface into the compressor, obtained with heat-accumulator elements having a minimum volume, associated with a continuous flow lubrication system, which also takes over the sponge cooling function and reduces the dead volume as much as possible when the piston is in the TDC. In addition, the introduction of a piston actuator system which (at a compression cycle time equal to that of a conventional compressor) introduces a variable piston speed, higher in the exhaust, in the suction and in the first part of the active piston stroke, and smaller toward the end of the compression process, it further reduces energy consumption and also makes the cooling system more efficient.

To any compressor with a superunit polytropic index in the prior art, the energy consumption necessary to obtain a given compression ratio may be reduced, in a given time, if a properly sized thermal sponge is inserted inside its cylinder. Most of the time, this involves some constructive modifications of the original compressor, necessary to shorten the piston stroke (or elongate the useful part of the cylinder) by a G _b , value, equal to the thickness of the sponge in the fully compressed state and for the adaptation of lubrication system to the new requirements. Given the evolution of energy prices and the objectives of reducing thermal pollution, the expenses necessary for these adaptations will be rewarded.

For isothermalizers can also be made configurations without elastic components. The isothermalizer in Figure 11 (horizontal section through a vertical cylinder with a rectangular section) consists of a thermal sponge made of metal plates 5.11 , made with a thickness as small as possible (if a high power of the isothermalizer is desired), but large enough for the plates not to be subjected (due to their own weight, or too sudden movements) to some residual deformations. To ensure that the plates have a stable horizontal position and a variable value spacing (dictated by the position of the piston), a sufficient number of movable carry-supports 5.19 are fitted inside the cylinder, located in close proximity to the cylinder walls, in such a way that the movements of either of them would not impede the movements of the others, nor the movement of the plane plates. In Figure 11 , section 1-1 is a vertical section through the cylinder, executed in the area where the carry-supports are mounted. Each carry-support is made in the form of blades, or narrow rods, on whose inner side (facing the inside of the cylinder) are mounted, (by welding, riveting, embossing, etc.) supports 5.20 of the plates 5.11 , made of sheet metal, wire, pieces processed by machining, etc. A number of supports equal to the total number of plates, or equal to the number of plates in a set, shall be mounted on each carry-support at different levels (usually equally spaced), if the interlaced plate technique is used. One end of the carry-support shall be secured, by means of a movable joint, to a fishplate 5.18 rigidly attached to the piston. On the other end of the blade, a short swivel arm is attached, also via a movable joint, which has a guide roller 5.16 attached, which can run on a rail, or in a channel 5.17 of the cylinder cover.

The horizontal plates 5.11 are rectangular, occupying almost the entire horizontal section area, but they have practiced in the corners a series of cuts to avoid collision with the carry-supports and supports on the neighboring levels, as well as to create the consoles 5.21 that are laying on the supports on that level. When the piston is at the BDC , the carry-supports make the minimum angle (almost 0°) with the vertical axis, and the distance between the plates is maximum. As the piston moves, the angle made by the longitudinal axis of carry-supports whith the vertical axis increases, and the distance between the plates decreases. When the piston is at the TDC , the carry-supports make the maximum angle (almost 90°) with the vertical axis, and the distance between the plates is minimal. When carefully processing components, the plates can perfectly overlap without intermediate spaces, ensuring a small dead volume and easy circulation for the fluid intended to replace this gas.

Figure 12 shows a horizontal section through the cylinder of an isothermalizer which also consists of a thermal sponge made of very thin metal plates 5.11 layed on a carry-supports system. Here, the carry-supports are made of a sequence of pairs of fishplates 5.23 and 5.24, placed superimposed in the same vertical plane. Both fishplates of these pair have a centrally located hole through which a pin passes, around which both fishplates can rotate. At the same time, this pin, having a corresponding length, can be the support for one of the horizontal plates 5.11. In another configuration, the length of the pin is approximately equal to the thickness of the two fishplates and is an empty cylinder, which constitutes a bearing for the support attached to the plate 5.11 by a rigid joint. The ends of the fishplates are coupled by moving joints, with two other pairs of fishplates (one lower and one upper). The extreme fishplate pairs are shorter and are coupled by moving support joints 5.22, one fixed to the piston, the other to the cylinder cover. In the “magnifying glass” of Fig.12A a front view of the carry-supports system is shown in the position corresponding to the piston at the TDC, and in section 1-1 a front view of the entire sponge corresponding to the piston in an intermediate position. In this configuration, a high density of horizontal plates 5.11 can be ensured by increasing the number of carry-supports, associated with a process of interspersed mounting of horizontal plates.

State-of-the-art diaphragm compressors, in any of the constructive variants, can become isothermal by easy-to-make modifications, by inserting a thermal sponge according to the invention into the working chamber and by adding an cooling/heating system based on a heat transfer fluid. Fig.13 shows a diaphragm densifier consisting of the upper housing 5.1 , the lower housing 5.1b and the elastic diaphragm 5.33. In the configuration shown, the densifier is operated directly by the piston 5.2, but can also be operated by means of a volume of hydraulic oil, in which case the housing 5.1b has perforations for the oil circulation. The shape of the two housings is modified, the enclosure between them having a shape close to that of a rectangular parallelepiped, with "softening" the edges, which allows aspiration of a larger volume of gas and offers more choice for the type of thermal sponge. In Fig.13 we chose a sponge composed mainly of flat metal plates 5.11 , supported on supports mounted on the carry-supports of harmonic type, composed of flashplates 5.22, 5.23 and 5.24, described in Fig.12. The thermal sponge can also have a permanent liquid component, with the role of avoiding the formation of a dead volume, and an itinerant liquid component, that with the help of 6.9b sprinklers, cools the gas, subject to compression. This component can also be used as a liquid piston, with flow rate adjusted in such a way as to obtain an isothermal speed for the compression. In order to better control the volume changes of the enclosure, the diaphragm 5.33 is mounted between two metal plates, 5.30 and 5.32, and is rigidly fixed to the two plates, by means of the plates 5.34, along a median axis, the outer edges of the the diaphragm being rigidly fixed between the two halves of the housing. The free part of the elastic diaphragm is extensible under the action of the piston, and the part between the two plates can slide on some rollers 5.31.

When the piston of an isothermalizer is at the TDC, the compressed gas is unevenly distributed in the volume of the cylinder. Significant volumes of compressed gas (so-called “dead volume”) may remain inside the sponge and in the space between the sponge and the cylinder walls, which will drop to the p _amb presure once the piston is moved in reverse, before the inlet valve is opened. This fact, like the situation with the compressors in the state of the art, it leads to a decrease in the final flow of compressed gas, a decrease as much important as the compression ratio of the compressor is greater. Careful selection of the shape and dimensions of the thermal sponge components, with taking into account their modifications during the processes in the densifier, must lead to the achieving of the most regular shapes of the volumes in which the compressed gas is found, when the piston is in position T and a minimum dead volume when the piston is in the TDC position. The gas is now exhausted by moving the piston from position T to the TDC position, where the dead volume reaches the lowest value. A total elimination of the dead volume can be achieved, as in the principle diagram in Fig.14a, by introducing into cylinder 5.1a, as early as the initial phase, a liquid phase 5.3b of the 5.3a thermal sponge, consisting of an appropriate amount of lubricant, or a heat transfer liquid, the volume of which equals the dead volume. At this densifier, when the 5.2a piston reaches the T position, the 5.5a output valve opens (due to the valve adjustment, or due to a command received from the control system), located at the highest enclosure elevation. From this point on, the piston movement causes the compressed gas to transfer into the discharge line, which is completed when the piston reaches the TDC position, where only the thermal sponge (solid and liquid phase) remains inside the cylinder.

Another possible configuration for the exhaust of compressed gas is shown in Figure 14b, in which 5.1a is a small densifier, whose inlet window 5.6a is at the same time the discharge window for a larger densifier 5.1 , with which it has a common wall. This mini-densifier is equipped with the 5.2a piston and a thermal sponge made of flat plates 5.11a. Moving the piston 5.2 from the TDC to the BDC leads to the inlet of the working gas at p, pressure in both cylinders. The first phase of compression is achieved by moving piston 5.2 from the BDC to the point T, interval during which the volume of gas in the densifier 5.1a does not change, but the gas in this cylinder is compressed in the same relation to the gas in cylinder 5.1 , and its sponge contributes to the accumulation of excess thermal energy. The compression conditions in the two densifiers being different, will be different throughout the compression period, also the temperatures of the gas and of thermal sponges they contain. The liquid fraction of the sponge in the densifier 5.1 can be chosen so that when piston 5.2 reaches the TDC , it completely occupies the volume of the cylinder not occupied by the solid fraction, without entering in cylinder 5.1a at all. At this point, all the initial volume of gas in the two cylinders is transferred to cylinder 5.1a, and its pressure reaches the final p _f value. In cylinder 5.1a, the gas can be subjected to a new isothermal compression stage, or it can be exhausted into the storage tank by moving the 5.2a piston from the BDC to the TDC. In applications where severe purity conditions are imposed on the compressed gas, the thermal sponge of the densifier 5.1a shall be carried out only with a solid fraction, so that the dead volume is as low as possible. In cylinder 5.1a, the gas can be subjected to a new isothermal compression stage, or it can be exhausted into the storage tank by moving the 5.2a piston from the BDC to the TDC . In applications where severe purity conditions are imposed on the compressed gas, the thermal sponge of the densifier 5.1a shall be carried out only with a solid fraction, so that the dead volume is as low as possible. In Fig.15, the compressed gas exhaust operation is carried out, with the piston in position T, through the window 5.6r and the pipe 5.6c, which connects directly to the storage tank located at a higher elevation (or with another useful destination) and which contains, at its base, liquid from associated hydraulic circuit, with pressure p _f . If liquid is found in the pipe, the 5.6a window serves only for gas suction. In this case, when the piston, in his movement to the TDC reaches the point T, opening of the window 5.6r allows liquid from the 5.6c pipe to enter the cylinder and replace the compressed gas at the pressure p,. This, due to archimedic forces, reaches the top of the storage tank, being replaced by an equal volume of liquid. This volume of fluid is then exhausted into the associated hydraulic circuit, for example through a valve fitted in the piston, or is exhausted back into the pipe, by moving the piston from point T to the TDC . Acest volum de lichid este apoi evacuat in circuitul hidraulic asociat, de exemplu printr-o supapa montata in piston, sau este refulat inapoi in conducts, prin deplasarea pistonului din punctul T in TDC.

When the piston reaches the TDC and the exhaust window closes at the control system command, the exact amount of liquid required to remove the dead volume remains in the cylinder. This amount of liquid remains permanently in the cylinder as a liquid fraction of the thermal sponge. This process can also be applied to solid piston compressors of the state of the art. Both valves 5.5 in Figure 4 have been replaced by the wide windows 5.6a and 5.6r created in the side walls of the cylinder (in the case of a cylinder with a rectangular section, the window width may be equal to the thickness of the sponge when the piston is in the TDC, and its length can be equal to the width of the wall), which allows for rapid circulation with reduced losses of exergy. The figure shown the piston in TDC , with the suction valve 5.6a open. The valve remains open until the piston reaches the BDC .

When, for the purpose for which the isothermalizer is used, it is useful to recover a larger fraction of the energy supplied to the system by means of the piston, may be implemented compressed gas exhaust devices, with the recovery of the thermal energy accumulated by the thermal sponge. In Fig.16 is shown such a process applied to the isothermalizer in Fig.5. The thermal sponge of this isothermalizer is composed of a solid fraction (a rectangular section helical spring) and a liquid fraction that completely eliminates the dead volume of the cylinder when the piston is in the TDC. Gas suction and exhaust are made through valves 5.5 in the cover of the isothermalizer (Fig.5), or through the 5.6c windows in the side walls. As we mentioned above, in the case of a densifier that is not equipped with a cooling circuit, the temperature of the gas (and implicitly, the mechanical work required to compress it during a cycle) and of the thermal sponge increase progressively with the number N of compression cycles. The heat the gas receives is accumulated, along with the mechanical energy received, in the compressed gas storage tank. After a sufficiently large number N of compression cycles, when the temperature of the sponge reaches a convenient value, the densifier sponge can be extracted entirely from the densifier, stored in an isolated enclosure and replaced with an identical sponge having the temperature of T _amb . This is possible, for example, if the cylinder has a rectangular section, the 5.6c side windows and the caps closing them have the width equal to the compressed sponge and the length equal to the side wall and if, at the moment immediately before extraction, the side caps 5.6c and the edge plates of the sponge 5.4a shall be mechanically coupled together so that they can be translated, sliding on the surface of the piston then on the outer rails (for example, pushed by a 5.2d piston, or by towing).

Another way to extract the “overheated” sponge is to extract the piston fully, coupled with the sponge, through the casing of the device. After extraction, the piston-sponge assembly is cooled in a fast cycle, or is stored and replaced with another identical one, having the temperature T _amb .

These processes for the reuse of excess thermal energy can be used in temporary storage systems for the renewable energy. In other types of uses, thermal sponge cooling is one of the most important problems of building high-performance densifiers. The thermal energy absorbed by the thermal sponge can be eliminated, depending on the characteristics of the densifier, by any of the prior art procedures listed in the previous paragraphs.

For compression/expansion systems in simple installations, when the compression/expansion ratio is low, a lubrication system combined with a lubricant cooling system leads, after a transitional regime, as in polytropic compressors, to a stabilization of the sponge temperature at a T _b value, higher or lower depending on the coolant flow and the coolant temperature. The presence of solid thermal sponge due to the additional heat absorption surface and the transfer to the lubricant (in this case the total heat transfer coefficient between the solid sponge and the liquid sponge is much higher than that between gas and lubricant) results in a lower value of T _b temperature and, implicitly, in a lower instantaneous work. In the case of large compression ratios, the temperature value of T _b can be maintained at a low value if, e.g., the system operation is supervised by an controller which, at predetermined intervals (or dictated by a feed-back adjustment system) stops the piston movement and leaves running, for a short time, only the cooling system and the lubrication system. During this time, the thermal sponge exchanges heat with the liquid in motion, and the densifier turns into a heat exchanger. To avoid too frequent stops, the on-off mode can be replaced by a multi-speed mode.

Similar to the system used in screw compressors, a liquid can be used for lubrication, which at Ti _Z temperature to generate a high concentration of in suspension particles. These particles have a great ability to limit the temperature rise above the working temperature of T _iz When certain purity performance of the compressed gas is required, or for reasons of simplification, the lubrication function is taken over by a coolant circulating through densifier with constant flow or, it is inserted intermittently into the cylinder (once for a number of N cycles). During the cooling operation, the piston can improve the efficiency of this operation by means of short back and forth movements. Also, in densifiers where the gas in the apparatus is exhausted by replacing it with the hydraulic agent in the compressed gas storage tanks, this agent can be included in a cooling circuit, continuously or intermittently and can also take over the cooling function of the thermal sponge. For example, for the densifier in Fig.15, the elimination of excess heat is done by replacing the compressed gas with colder liquid during the exhaust operation (the warmer liquid being subject to upward forces). The stopping time of the liquid agent in the compressor can be prolonged (periodically 1000 or every cycle) by the commands sent to the piston by the controller. Another possibility of replacing this fraction, increasing the flow rate of the gas, is to remove the remaining fluid in the cylinder, during the fresh gas suction operation, by opening a valve located in the piston (with the liquid spilling into the densifier housing), or by absorbing it through a component pipe of a cooling circuit equipped with a suitable heat exchanger. If the temperature of the liquid is lower than the temperature of the sponge, it 1005 will take up and discharge some of the thermal energy accumulated by the sponge, the higher the longer it stays in the cylinder.

With the imposition of higher performance criteria for the isothermalizer, the performance required for the thermal sponge cooling system increases. This is achieved by implementing more efficient cooling systems. The state of the art proposes a wide range of such procedures: by 1010 continuously introducing a coolant aqueous foam, with the elimination of excess fluid in each cycle, by continuous or intermittent spraying (toward the end of the compression cycle) of a coolant, or by any other method which advantageously combines the action of the solid piston with that of a liquid piston. The liquid piston appears whenever the instantaneous flow of coolant inserted is greater than that exhausted. If the flow rate of coolant circulating through this circuit is correspondingly correlated with 1015 the instantaneous gas pressure in the cylinder and the instantaneous piston speed, the compression carried out is almost isothermal.

As an example, in the densifier in Fig.5, in the empty space in the center of the helical spring 5.4 a tubular helical spring can be mounted, coupled by flexible tubes to an external cooling circuit with heat exchanger, through which, under the effect of a hydraulic pump, a coolant circulates. The 1020 hollow spring is fitted with horizontal spray nozzles. It introduces suspended particles between the metal plates, which considerably increases the cooling speed of the gas and of the thermal sponge. Excess fluid accumulated above the piston is eliminated in continuous flow. This system can also be applied to densifiers in Fig.6, Fig.7, Fig.9 and Fig.10, if the horizontal plates 5.11 provide, by design, overlapping circular holes in such a way as to create the space necessary to mount the tubular spring 1025 with spray nozzles.

If the horizontal plates 5.11 and vertical plates 5.10 of the isothermalizer in Fig.8 are fitted with judiciously placed perforations, this type of isothermalizer can be cooled by inserting the coolant through the top of the apparatus, at a pressure equal to that of the apparatus. This type of densifier is very suitable for cooling with aqueous foam. Foam regeneration can be done by introducing gas at a 1030 pressure equal to instantaneous pressure in trays 5.11c.

At the isothermalizers shown in Fig.11 and Fig.12, the additional coolant introduction system may be mounted on the skeleton supporting the horizontal plate system, inside it, or on similar independent structures.

In liquid piston compressors from the state of the art, the piston compressor is a liquid agent 1035 supplied by a hydraulic motor, the liquid having remarkable properties: lubrication and sealing, minimizarea of dead volume and a two-way transmission agent of heat and mechanical energy. The disadvantage of these systems is that, if thermal sponges made up of solid elements only are inserted into the cylinder, at some point they will be covered by the liquid whose level increases, so the heat absorbing surface of the sponge shrinks (just when its temperature rises faster).

1040 One way to overcome this inconvenience is to make a solid piston compressor, in which a thermal sponge with N fixed horizontal plates is mounted, with the surface slightly smaller than the horizontal section through the cylinder, that multiplies about N times the absorption surface of the apparatus. The plates are arranged in such a way that, after the introduction of liquid into the cylinder, all the plates are covered with a layer of liquid, which becomes a liquid component of the thermal 1045 sponge, as it takes from the upper surface of the solid plate the role of absorbing the excess thermal energy. Moreover, on the periphery of horizontal plates, low-height stops can be installed that permanently retain a thin layer of liquid, thus sacrificing part of the useful volume of the isothermalizer in favor of a more efficient thermal sponge. Further penetration of liquid agent increases the thickness of liquid layers, without significantly affecting the absorption surfaces, but further increases the 1050 pressure in the cylinder (the liquid agent continues its role as a piston).

An example is the liquid piston isothermalizer whose longitudinal section is shown in Fig.17 and the transverse one in Fig.19. The isothermalizer is built like a liquid compressor of the state of the art, with a cylindrical housing 7.1 and a liquid piston 7I driven by solid piston 7.5, in turn driven by a speed regulation system, whose imperative is to provide the isothermal speed v _iz . The solid piston is 1055 used in combination with a variable speed electric drive assembly. In the case of a variable fluid flow hydraulic drive, cylinder 7.2 connects directly to the liquid line.

The thermal sponge has a solid and a liquid component. The solid component consists of the horizontal plates 7.3a provided, on most of the circular sector, with the peripheral skirts 7.3b, intended for the division of the liquid piston. The other side of the circular sector, separated from the first by the 1060 vertical walls 7.6b (Fig.19), has the skirts facing toward the upper part. The liquid component consists of the liquid layers that form on each of these solid horizontal plates. The liquid component of the thermal sponge has mainly the role of a piston and acts simultaneously in the N elementary compressors formed by splitting the main compressor. In cylindrical-shaped compressors, the gas contained in each of these elementary compressors yields heat, mainly to two circular surfaces 7.3a, 1065 which have a diameter almost equal to the diameter of the master cylinder. If the horizontal plates 7.3a were missing, cylinder 7.1 , cap 7.2 and the liquid piston would constitute a liquid piston compressor, with an initial volume approximately equal to the sum of the volumes of the N elementary compressors, but with a (variable) heat transfer surface only slightly larger than that of an elementary compressor. Due to the way these plates are arranged, the liquid piston acts simultaneously in each of 1070 the N elementary compressors, leading to the formation of N elementary pistons, the piston speed in each of them being N times less than the speed of the unique piston, and the heat energy corresponding to this power is distributed over a contact area of N times greater. Therefore, compared to the non-thermal sponge compressor, we can achieve an isothermal compression of the same amount of gas, with a hydraulic motor speed (which supplies the liquid agent) of about N times higher 1075 (the same power, distributed over a time interval of N times shorter).

Starting from this constructive scheme, a multitude of similar configurations can be made. Fig.17 shows a cross section of a cylindrical densifier (in many applications a rectangular section is more advantageous), and Fig.19 shows a horizontal flat section of it, at the level of an elementary compressor. The horizontal plates 7.3 and 7.4 separate the compressor from the constant pressure 1080 tank 7g and the liquid piston 7I, respectively. In this type of densifier, the liquid piston consists of a fixed volume of liquid agent (the same type of liquid as the one in the tank 7g), equal to the free volume of the compressor (the volume of gas in the cylinder, immediately after the suction phase). In the first phase, the densifier absorbs a volume of gas through the valve 7a (located in the upper elementary compressor), when the piston 7.5 moves from the TDC to the BDC. At the same time, an 1085 identical volume of liquid, located in the densifier, is transferred into the tank 7L. The compression phase follows, in which, after closing the suction valve, the liquid from the tank 7L enters the densifier cylinder and, according to the law of the communicating vessels, is distributed into the N elementary compressors. The instantaneous gas pressure in these mini compressors is almost the same (the difference is given by the height of the liquid column between the compared compressors. The 1090 compressed gas is exhausted at the level of each elementary compressor, through the windows 7.6a, practiced in the partition wall 7.6. This partition wall, together with the side wall 7.1 and the two vertical intermediate walls 7.6b (Fig.19), borders a ring sector 7s, which communicates freely with the tank 7g, being constantly flooded by the liquid agent with the pressure p _f in the tank. Opening windows 7.6a is done by moving a movable cap (piston 7.7) that is running tight (by means of seals) on wall 7.6 and is 1095 controlled by a differential pressure switch 7p, when the piston 7.5 is in position T and the pressure p _f of the liquid in the densifier is equal to the pressure in the tank 7g. At this time, the gas pressure in each elementary compressor is equal to the pressure p _f , to which is added the pressure given by the liquid column between the measuring point and the elevation of the respective elementary compressor. The displacement of piston 7.7 causes the entire amount of compressed gas in the 1100 densifier to be replaced by liquid agent in the tank 7g (in this way, the entire volume bordered by horizontal plates 7.3 and 7.4 is occupied by the liquid) and causes the level of the liquid in this tank to lower. If piston 7.5 continues to move to the TDC, the amount of fluid between the level T and the TDC level (equal to the total volume of compressed gas during a stroke of the piston) is exhausted through the pipes 7r to another device with pressure p _f (for example, a tank, or a hydraulic generator).

1105 In another configuration, the tank 7g and the piston 7.7 may be missing, the wall 7.3 becomes the outer wall and the windows 7.6a are replaced, each of them with a check valve. The tank 7s is replaced by a simple pipe in the wall of which valves are mounted to each elementary compressor, the lower end being opened through a slot to the lower compressor, and the upper end is opened to the upper gas layer, located in the upper elementary compressor. In this case, the check valves open 1110 successively, the respective mini compressor being immediately flooded by the liquid in the main column, the thickness of the air layer increasing accordingly.

The liquid and, indirectly, the plates in the densifier can be cooled by keeping the liquid in the tank 7g at temperature T _amb or/and by the recirculation, continuous or intermittent, of the liquid agent from the tank 7L. An increase in heat transfer surfaces can be achieved if the gas inlet pipe in the 1115 densifier is supplied by a foam generator. It can also introduce, at the right time, foam or compressed gas, directly into the liquid of each elementary compressor, using thin pipes. An increase in heat transfer surfaces can be achieved if the gas inlet pipe in the densifier is supplied by a foam generator. It can also introduce, at the right time, foam or compressed gas, directly into the liquid of each elementary compressor, using thin pipes. In some configurations, this compressed gas can even come 1120 from the tank 7g. Additional cooling is obtained if different types of metal inserts, strips, metal nets, etc. are inserted into each elementary densifier at its top, or if various types of vertical metal elements are mounted on its ceiling. Due to the natural convective currents, as the gas is compressed it tends to thermally stratify, so that in areas where the instantaneous temperature can reach higher values, the absorption surface of the thermal sponge will be increased.

1125 The densifier in Figure 18 is built on the same principle of overlapping a large number of elementary liquid piston compressors, made by interspersing their upper and lower walls, 7.3s and 7.3i, respectively. Compared to the previous configuration, two types of liquid piston mini-densifiers appear in this type of densifier: a mini-densifier 7c between the upper and lower walls, with higher height, cooled by liquid spray and a mini-densifier 7d between the lower and upper walls, with a lower 1130 height, without spraying. The liquid piston is inserted into the elementary compressors 7d directly, through the windows 7.3f. A vertical skirt 7.3g is inserted to separate a layer of gas into each densifier. The introduction of the liquid piston into the elementary compressors 7c is done by a distributor of agent 7.11, from which the liquid agent is sprayed into the compressor, intensifying its role as a gas coolant. Similar to the previous configuration, a horizontal wall 7.3 separates the densifier area from 1135 the tank 7g, which is in direct communication with a tank 7s located, this time, in the center of the densifier, having a cylindrical shape and being separated from the densifier by the cylindrical wall 7.6, in which the windows 7.6a are executed, on each basic compressor. To each of these windows corresponds a similar window, located at the same level, in the cylindrical wall 7.7, located inside cylinder 7.6, so that these windows overlap in the "open" position and allow gas and liquid to pass 1140 from one compartment to the other. The closing of these windows in the "closed" position is done by turning with an appropriate angle, or by moving vertically the cylindrical wall 7.7, in such a way that the seals mounted on the outer surface of cylinder 7.6 around the windows 7.6a block all gas passageways. The same transfer system is also applied in the last phase of the cycle, in case of fluid exhaust from the cylinder and gas suction with the pressure p,. A sector 7.8 in the side wall, with the 1145 height equal to that of the densifier, has a series of windows 7.8a practiced at each mini-densifier.

During the compression operation, these windows are closed by a piston 7.8b of the appropriate size and shape, fitted with suitable seals and a horizontal displacement system. Through these windows the liquid agent is removed from the elementary compressors (after the compressed gas is removed and the valves 7.6a are closed) and the working gas is introduced at the initial pressure.

1150 In some configurations, when the densifier is carried out by means of parallel plates very close to each other, or by means of small alveoli inserts, or very small mesh woven nets, in the case of liquids whose viscosity exceeds a certain limit, it is effective to implement some devices to accelerate the discharge of liquid from the densifier after the compressed gas exhaust phase. In Fig.20, we represented on a large scale details of a possible configuration for such a device, as well as for a 1155 variant of gas inlet valve and a variant of diffuser. The accelerator is made by making the upper plates

7.3s of each elementary compressor out of three distinct components: an outer peripheral ring 7.3sa, with the outer diameter equal to the inner diameter of cylinder 7.8 to which it is rigidly attached, an inner peripheral ring 7.3sb, with internal diameter equal to the outer diameter of cylinder 7.6 on which it is rigidly fixed, and a movable flat ring 7.3s, with outer diameter greater than the inner diameter of 1160 the peripheral ring 7.3sa and inner diameter smaller than the outer diameter of the inner peripheral ring 7.3sb. The movable ring-sections shall all be fixed on one or more rods 7.9 in a position below the corresponding plate 7.3s in such a way that, through the seals mounted on the edges of the upper surface, air and liquid are not allowed to flow to the lower compressor when the rods 7.9 are in the “closed” position. Moving the entire rod-plate system to a lower position causes wide access paths to 1165 open, in which the friction between the liquid and the plates is greatly diminished. An increase in the efficiency of fluid circulation is achieved when these movements are made at high speeds, with sudden starts and stops, so that the forces generated by the surface tension of the liquid are overcome.

Here is the sequence of the phases in the densifier: the liquid agent is inserted through a gate 1170 7a into the tank 7L, from which it is distributed naturally between the lower plates 7.3i of an elementary compressor 7c and the upper plates 7.3s of the following elementary compressor. The mini-densifiers 7d are thus formed, which also become liquid dispensers for mini-densifiers 7c. The gas layer between these plates is compressed and pushed through the holes 7.10 of the bottom plates inside the compressor 7c located above them. In the configuration in Figure 20, a simple plug 7.1 Od with the 1175 appropriate seal (whose collar steps on springs 7.1 Or) completely obscures the entrance path to the respective mini-compressor 7c. The springs 7.1 Or are pretensioned, so that the valves can be opened at a set pressure. Various devices for regulating the flow of gas passing through these holes can also be implemented. If both air inlet and gas inlet valves are remotely controlled, priority may be given to either of the two types of densifiers.

1180 During operation, regardless of the flow of liquid agent introduced (or the speed imposed to the solid piston), the fluid pressure at the periphery of the master cylinder is equal to that of the gas and liquid in the densifiers 7d, and that of the gas and liquid in the densifiers 7d is slightly lower, this causing the valves 7.10 and 7.11 to open, depending on their adjustment. If the flow of liquid entering the sprinklers is lower than that introduced through the windows 7a, the gas pressure increases, which 1185 also causes the gas inlet from the densifiers 7d to open into the densifiers 7c, causing the gas pressure in these densifiers to increase. When passing through the liquid layer, this gas undergoes additional cooling.

In the configuration shown in Fig.19 and in the additional details of Fig.20, the inlet valve is a sealing plug directed by a pre-loaded spring, and the sprinkler can be a simple disc-shaped cap, with a 1190 horizontal spray holes 7.11 on the side. This valve opens when the difference between the dispenser fluid pressure and the gas pressure in the corresponding elementary compressor is greater than a preset value. This difference remains almost constant throughout the compression period, but may have slight differences from one elementary minicompressor to another.

After a few such cycles of alternating the compression process between the two types of mini- 1195 densifiers, all the mini-densifiers of type 7d transfer the compressed gas to the mini-densifiers of type 7c , and when the pressure reaches the final one, the compressed gas is evacuated to the tank 7g. Using the same idea (for their use in cases where large single-step compression ratios are desired), many configurations of liquid piston densifiers can be made, with one or more solid-piston densifiers as the main subassembly. Any of the densifiers described above, or made on the same 1200 constructive principles, may be used. In these configurations, the functionality of the system and its adaptation to different particular applications depend on the initial number and volume of densifiers, as well as the amount of mechanical energy available. For example, for any of the solid piston densifiers in Fig.5-12, the exhaust of compressed gas by replacing it with a liquid that has the gas pressure is an efficient process for removing compressed gas bags and for extracting and discharging excess heat 1205 from the solid thermal sponge. The process can be implemented by fitting valves in the piston that allow controlled fluid intake and proper extension of the master cylinder. In this way, the solid piston densifier becomes a liquid piston densifier. For example, Fig.21 shows the densifier in Fig.12, in which vertical perforations 5. 11 o are made in the components of the solid thermal sponge (horizontal metal plates 5.11), in such a way that when the plates overlap under compression, these holes also overlap 1210 and form continuous channels. In order to increase the initial volume of gas admitted in the cylinder and to increase the absorption surfaces of the thermal sponge, as well as to decrease the total mass of the densifier, a series of horizontal channels 5.11c are executed between these vertical holes. As the liquid piston only enters into operation at the final stage of compression, when a slower speed is required, the friction forces between the liquid and the solid sponge are lower. An identical solution is 1215 shown in Fig.22, for the solid piston densifier in Fig.7, by making the holes 5.11o and the channels 5.11c. In both cases, in piston 5.2 the valves for fluid intake 5.2s are fitted and the possibility of liquid agent passing through the windows 7a is realized. In Fig.22A, the liquid circulation channels are highlighted, after positioning the solid piston in its upper position.

A considerable increase in the flow of gas circulated under conditions of an isothermal 1220 compression is obtained if, after a corresponding increase in the height of the densifier, an additional thermal sponge 5gs is inserted in the upper part of the densifier, made of metal foam, metallic fabrics, other metal inserts with a large absorption surface (in Fig.21 it is made of woven blankets made of metallic wire, superimposed, without separation intervals, mounted on a horizontal support system, of bars, rods, perforated plates, etc.), which absorbs a large amount of heat energy in all phases of 1225 compression. The liquid piston penetrates into the holes and grooves executed in the horizontal plates, as well as in the alveoli of the additional sponge, only in the final phase of compression, when the forward speed of the liquid piston is quite low and the liquid-sponge friction forces are reduced.

Composite systems may also be made by simultaneously or successively accumulating the effects of a liquid piston and a set of solid pistons. Such configurations are described in Fig.23, 1230 Fig.23a and Fig.23b and support any of the solid piston densifiers described above, if their volume variations can be fully controlled. The densifiers 7g in Fig.23 are equipped with a thermal sponge composed of elastic plates 7.3, mounted on a support 7.3s. They are placed in hermetic bags 7.14, made of elastic materials, or other tear-resistant materials, with high heat transfer coefficient, but slightly deformable, even at low pressures. These bags are attached to a metal plate 7.2s, fitted with 1235 holes 7.2o, connecting the gas in these densifiers to the gas layer 7gs, above the plate 7.2s, layer which contains another thermal sponge, for example, a dense cross-linked metal network. The volume of each bag 7g shall be reduced as the volume of the thermal sponge it contains is reduced and may reach a minimum value when the deformation of the sponge is naturally blocked (when all the inner plates overlap), or by externally controlled devices. The so-created enclosures communicate with the 1240 top gas layer 7gs in the compressor cylinder, but can be separated from it by the valves 7.2ps. The metal plate 7.2s thus separates all the gaseous regions 7s containing thermal sponges with large absorbency area, from the rest of the enclosure 7i. In some configurations, various pipes can be inserted into the enclosure that create additional communication paths, such as pipe 7.16, with rectangular horizontal section, which causes the pressure equalization between layers 7i and 7s to be 1245 done by means of an intermediate liquid piston. The apparatus is equipped with a series of valves for the circulation of gas and liquid flows: valves 7.2a for the initial gas inlet, valves 7.2e for the exhaust of compressed gas, valves 7a for the insertion of the liquid piston into the enclosure, for the circulation of coolant (same as the piston fluid) and to drain the liquid, simultaneously with the gas entering the lower enclosure 7i, the valves 7.2i controlling the liquid piston inlet and outlet in the densifiers 7g, the 1250 valves 7.2ps that can block communication between the densifiers 7g and the layer 7gs, the valves 7.2pm and 7.16pi which controls the communication between the lower and upper enclosure of the cylinder. The play of these valves allows us to choose between a multitude of variants in the functional stages.

In a first configuration, we chose the variant in which the liquid piston successively penetrates 1255 all the densifier enclosures. In the first phase, with the valves 7.2ps and 7a open, with all other valves closed, with an initial pressure p ₁ in the lower enclosure 7i and with an initial pressure p ₂ in the upper enclosure 7gs and in the enclosures 7g, the liquid piston intake takes place in the enclosure 7i. In order to achieve high performance of the installation, it is recommended that the initial gas pressures be as high as possible. At an external pressure p ₁ and an internal pressure p ₂ , the volume of inflatable 1260 bags has a maximum value (a value as close as possible to the total volume of the lower enclosure 7i is recommended), and the elastic elements of the sponges 7g are tensioned, as they take up some of the mechanical energy needed to introduce the gas. In this way, the hydraulic pressure (and power) required to penetrate the liquid piston is high, and the isothermal start speed is much lower than in the case of gases at atmospheric pressure. The liquid compression agent, with an initial pressure p will 1265 gradually replace some of the gas here by compressing it. At the same time, the pressure of the liquid agent is also exerted on the walls of the compressors 7g reducing their volume and increasing the pressure of the gas inside.

When the gas pressure in the enclosure 7i equals that of the 7s, the inflatable bags return to the rest form, the one with the untensioned elastic elements, the mechanical energy previously 1270 accumulated in the elastic elements of the thermal sponge diminishing the mechanical energy needed for the liquid piston to achieve this compression. Part of the thermal energy produced by compressing the gas in the enclosure 7i is transferred to the walls 7.14 (inflatable bags) and to the walls of the enclosure 7i, with heat transfer surfaces that decrease with the advance of the liquid piston and with the increase of the gas pressure in all compartments, and part of the heat energy resulting from the 1275 compression of the gas in the enclosures 7s and 7g is given to the thermal sponges 7g and 7gs, respectively. The gas pressure in the enclosures 7s can increase, with a speed depending on Young’s module of the plates 7.3, until the plates that make up the sponge overlap and occupy a minimum volume, the mechanical energy accumulated by the elastic plates being increased. In this position of the piston, if inside each inflatable bag there are liquid elements of the thermal sponges that occupy all 1280 the gaps inside the respective bellow (the excess fluid being also discharged into the layer 7gs), the closure of the valves 7.2ps causes the temporary separation of the lower enclosure 7i (where the gas pressure has reached the valuep ₃ ) from regions 7g and 7gs (where the gas pressure has reached the value p ₄ >p ₃ ). In the next phase, the two pressures p ₄ and p ₃ , are equalized, operation that can be performed in different ways, including:

1285 - opening through the windows 7a of a communication path with a hydraulic motor and opening the valves 7.2pi, causes the forces that held the elastic elements of the densifiers 7g to disappear and their return to the rest volume, with the discharge of the elastic energy stored in the mechanical work produced by the engine, simultaneously with the expansion of the gas from the enclosure 7i, with thermal energy absorption from the walls of the enclosure, up to the pressure p ₃ and 1290 filling with the liquid from the enclosure of the inflatable bags; after equalizing the pressures and opening the valve 7.2i, the liquid piston continues to compress the entire volume of gas, from pressure p ₃ , to pressure p _F , then, after opening the valve 7.2e, fill the entire chamber completely with liquid and exhaust the compressed gas.

- opening of the valve 7.2pm and sudden actuation of the liquid piston (e.g. by suddenly 1295 inserting a solid or liquid piston into the pipe bringing the liquid to the inlet 7a) results in the formation of a gas piston that adiabatically compresses the gas from the enclosure 7gs to the pressure p ₄ , with unchanged preservation of the volume of inflatable bags. The mechanical energy carried out to move this volume of gas shall be converted into thermal energy absorbed entirely by the gas in the enclosure. Differences in temperature and pressure between the gases that mix cause unnegligible 1300 losses of exergie. Further movement of the liquid piston results in increased gas pressure in the enclosure, from pressure p ₃ to pressure p _F , then, after opening the valve 7.2e, full liquid filling of the entire chamber and the compressed gas exhaust. The mechanical energy stored in the elastic elements can be recovered in the exhaust phase of the liquid in the enclosure 7i, by directing it to a hydraulic motor.

1305 - if the line 7.16 is filled with liquid, at the opening of the valves 7.2pm, 7.16i and 7.2pi, it is pushed to the layer 7g with the lower pressure (by the pressure of the gas in the chamber 7i and by the pressure exerted by the displacement of the walls of the inflatable bags 7g due to the elastic mechanical energy accumulated in the elastic elements of the thermal sponges) and compresses it to an intermediate pressure between p ₃ and p ₄ , the same for all the gas in the enclosure; the bags 7g are 1310 filled with liquid and return to the rest volume, and the elastic mechanical energy accumulated in the components of the bags is fully recovered, as mechanical compression energy

The thermal sponge of the densifier in Fig.23a is made of plates, or elastic metal strips 5.14, which have a series of ripples with different radii of curvature. They can be arranged in organized structures, or they can be arranged as shown in Fig.23a, in a more or less random way and placed in 1315 deformable hermetic bags 7.14, made of tear-resistant elastic materials. Flermetic bags may be in the form of mattresses, having width I equal to one side of the enclosure of the apparatus and the length equal to a multiple of the other side L. These mattresses are laid in overlapping layers over the entire height of the apparatus, by bending with 180 degrees after each length L without strangling the free movement of the gas. If necessary, mattresses communicate with each other through rigid tubes. In 1320 another configuration, instead of a single very long bag, an appropriate number of mattresses are used, with a convenient thickness (depending on the type of thermal sponge) and with the surface equal to the horizontal section of the cylinder, mattresses that communicate with each other through more rigid tubes. In another configuration, instead of each mattress, tubes of length I or L are used (both dimensions arranged in alternating layers can be used). All of these tubes (and mattresses) 1325 communicate with each other, through rigid, metallic, or deformable tubes, forming a single enclosure. And in the configuration of Fig.23b, the isothermalizer contains layers of cylindrical bags, or rectangular mattresses, but the thermal sponge is made of elastic metal sheet rolled into more, or less helical, parallel scrolls. In all cases, although the bags can be mounted without support, in terms of the speed of heat transfer from the enclosure, it is preferable to arrange them on some supports.

1330 In the initial state, the bags are inflated at an initial pressure p1, which may be different from the atmospheric pressure. The liquid is inserted and discharged into/out of the enclosure through gates 7a with hydraulic pumps, the ratio of inlet and exhaust flow is variable but always unitary (the liquid also plays a cooling role), or superunitary (the liquid plays the role of piston). The operation of this compressor is similar to that of the compressor in Fig.23, with similar components having the 1335 same notation.

On the compressor shown in Fig.24, the introduction of liquid into the compressor is made by spraying. The upper thermal sponge is a wire mesh 7gs, under which the lower thermal sponge is mounted, in the form of a tubular double-walled pipe 7.16, one end of the pipe being located in the area of the wire mesh, and the other near the base of the cylinder. Both ends of the pipe cross the 1340 walls of the cylinder and can be closed through the valves, both on the liquid input/output path 7a and on the gas input/output path 7.2pi and 7.2ps. This pipe can be very long and, if placed in parallel layers (for example, L or / length cylinders, sitting side by side, can take up a large fraction of the total volume of the cylinder. It is equipped with a large number of nozzles 7.10, or/and simple holes with a higher density at the top. The liquid piston (which is also a coolant and is part of a cooling circuit 1345 equipped with a hydraulic pump and a heat exchanger) is inserted through the valve 7a at one end of the pipe, into the space between the walls of the pipe. Spray nozzles are mounted in the walls of the pipe, both toward the inside of the pipe and toward its exterior, which spray liquid in all areas occupied by the gas, compressing it. The constant temperature of the gas is maintained by adjusting the ratio between the flow of the input and the exhaust fluid. At some point, the entire bottom of the cylinder is 1350 occupied by the liquid, the gas accumulating in the layer 7gs. Compression continues with a low liquid flow until final pressure p _F is reached, when the valve 7.12e opens, the liquid flow increases rapidly and the compressed gas is completely exhausted to the tank/consumer. After compression, the liquid is drained by suddenly retracting the lower cylinder head 7.4 and draining the liquid into a tank.

The compressor in Fig.24a is a liquid piston densifier, with no moving solid parts, with a 1355 cylinder 7.1 having a circular section, in which the mass of the thermal sponge is distributed in such a way that the thermal energy from the mechanical energy transformation of the piston is absorbed as evenly as possible. Since, according to the invention, the piston moves at the isothermal speed v _iz , the temperature difference DT is maintained throughout the compression period. In this apparatus, until the desired compression ratio is reached, the liquid agent moves a longer route than in the case of 1360 liquid piston compressors of the state of the art. The cylinder cover 7.2 is fitted with an exhaust valve 7.12r (gas exhaust is made by replacing it with liquid agent) and the lower wall 7.4 with an inlet valve 7.12a and the valves 7.4e for exhaust the fluid from the cylinder, at the end of compression. The thermal sponge is made of concentric vertical cylinders 7.3v (their cross section may not be necessarily circular), arranged at greater distances in the central part of the densifier, but increasingly 1365 closer to its periphery. Furthermore, the peripheral cylinders are provided with elements to amplify the heat absorption (in the figure, the horizontal fins 7.3f). Another component of the thermal sponge is the wire metal mesh 7gs, located at the top of the cylinder. The pressure in the cylinder is kept constant by practicing holes 7.3o at the top of the vertical plates. Such holes can also be practiced at lower levels to control the routes traveled by the liquid agent, to increase the absorption power of excess thermal 1370 energy and to increase the ascending convective currents.

Another advantage of this configuration is to make it possible that after the compressed gas has been exhausted from the densifier, cylinder 7.1 and its cover 7.2 can be lifted for a short period of time, during which the liquid agent is removed and replaced with the working gas.

A new type of isothermalizer capable of performing energy-efficient isothermal transformations 1375 is the gas piston isothermalizer. whose scheme of principle is shown in Fig .25. In principle, this isothermalizer has a first stage, which is composed by one or more isothermalizers 8.1 with solid or liquid piston, which exhausts into a tank 8.2i (which constitutes the second step), the volume of which is significantly higher than that of the isothermalizer 8.1.

In the configuration of Fig.25, the tank 8.2i is inserted in another parallelepiped tank 8.2, 1380 partially filled with a coolant 8.2I. The outer tank 8.2 is filled with liquid up to a certain level, above the upper level of the tank 8.2i. A layer of gas is left in the upper part of the tank 8.2, which, through pipe 8.2c, constantly communicates with the gas in the inner tank. Therefore, this gas layer is also compressed, at the same pressure, by the action of the densifier piston 8.1. To cool this gas, a thermal sponge is installed consisting of a wire metal inserts 8gs, which in turn can be cooled continuously, or 1385 periodically. The fluid level in the tank 8.2 is kept constant by the 8.6M pump. In turn, the liquid in the tank 8.2 is cooled by its inclusion in a system that also contains the liquid-gas heat exchanger FIE. The cooled liquid in the heat exchanger is also directed to cool the thermal sponges of the first stage densifiers. Moreover, for rapid cooling, the first-stage isothermalizer can also be inserted into this tank. Through the valve 8.2r, the amount of liquid in this circuit can be supplemented to cool the sponge 1390 component 8gs, to discharge the gas when it reaches the required pressure, or to transform this agent into a liquid piston.

In the case of compression, at large volumes of the tank 8.2i, if on start from the pressure p ₁ with which the gas is sucked into a densifier 8.1 , it would work for a long time in a non-economic regime, with very low compression ratios. Even in the case, frequently encountered in the state of the 1395 art, where the tank 8.2i is initially filled with liquid and the discharge from the densifier is done at a constant pressure p ₂ , with a high compression ratio, the densifier does not operate in an economical regime: the piston speed frequently has large amplitude changes (also taking into account the suction and exhaust phases, deployed at high speeds). A possible solution is that in the first phase, the densifier works with the most economical compression ratio, by discharging in a tank with constant 1400 pressure p, , intermediate between p ₁ and p ₂ , and after the end of the liquid from the tank, the compression continues in the closed tank, with the gradual and slow increase in the compression ratio. In the first phase, the discharge is done under constant pressure, each gas tranche replacing a liquid tranche with the same volume in the tank, the liquid being taken over by a hydraulic motor, to recover the energy consumed during the discharge phase. Another recommended solution is like this first 1405 phase, in which the densifier works with a constant compression ratio, discharging into a tank with constant pressure p„, to have a limited duration, after which the operation of discharging a volume of compressed gas from the condenser 8.1 to be accompanied by the removal of a lower volume of liquid from the tank 8.2, which results in a gradual increase in the gas compression ratio in the tank 8.2. These operations can be performed over time so that the pressure p ₂ is reached with the discharge of 1410 the last slice of liquid.

For an isothermal compression, the average temperature of the gas in each compression stage must be constant, equal to the setpoint T _iz . The most efficient solution for maintaining the gas in the tank 8.2i at this value is to introduce into the tank a non-deformable thermal sponge, the envelope of which consists of the walls of the tank 8.2i itself, permanently cooled by an adjustable system, 1415 which maintains the average temperature of the gas, equal to the setpoint T _lz .

In the second phase of compression, the gas discharging from the last densifier of the first stage begins each time the gas pressure equals the p _r pressure in the storage tank. During the exhaust, the volume of compressed gas between the solid piston and the tank acts on the gas in the tank as a gas piston, increasing with each cycle the compression ratio. In the tank 8.2i, part of the 1420 thermal energy contained in the compressed gas from that given by the piston during the exhaust (exhaust which does not occur at constant pressure, but is in fact a compression with a small compression ratio) it is taken over by the thermal sponge and is given over to the cooling system, including during the next compression phase of the solid piston compressor, when its exhaust valve is closed and the tank 8.2i becomes a simple heat exchanger. Due to the imposition of an isothermal 1425 speed v _iz1 for the densifier 8.1 for the entire compression duration t _iz1 and an isothermal speed v _iz2 , for the same piston for the entire compression duration t _iz2 of the gas exhaust (equivalent to an isothermal compression of the gas in the tank), the temperature of the gas in the tank can be maintained at a value T _z , close to the ambient temperature. During the following phases executed by the piston, the suction and the compression (with a slightly higher compression ratio), the exhaust valve is closed and 1430 the gas in the tank continues to cool, suffering also some pressure reduction.

If, during a cycle, the pressure in the first stage of the densifier starts from an initial value p, (usually atmospheric pressure) and increases to the instantaneous pressure p _r in the tank, and the pressure in the tank increases slightly in the intervals in which, after opening the exhaust valve, the compressor piston 8.1 exhausts compressed gas into this tank. In the case of closed tanks, this 1435 exhaust operation is in fact a process of compressing the gas contained in compressor and tank, to the volume of the tank. If the difference between the two volumes is large, the compression ratio is very small, and the work done by the piston is approximately equal to that required for constant pressure exhaust.

In another configuration, when the densifier 8.1 terminates its exhaust operation, another 1440 similar isothermal compressor can reach the exhaust phase, the gas in this compressor having also the temperature of T _iz and a pressure equal to the new pressure in the tank. The exhaust phase for this compressor requires a higher mechanical energy consumption, therefore the thermal energy that must be absorbed by the tank sponge increases, and the time required for the operation is a little longer. At the end of this phase, another isothermal densifier, after his compression is completed, is 1445 ready for a new exhaust operation, at a slightly higher pressure and with a slightly longer duration. This densifier is followed by as many densifiers as necessary for the first densifier to complete the isothermal compression cycle, with a compression ratio higher than in the first cycle, and be ready for the next exhaust. The sequence of these phases is repeated until the desired pressure is reached in the tank 8.2, then the gas in the tank is transferred through the valve 8.2s into a storage tank 8.7, by 1450 replacing it with pressurized fluid.

The compressors in the first stage must provide gas at its outlet with a temperature equal to T _iz and a variable pressure, always equal to the gas pressure in the second stage, a requirement which can be met by different compressor combinations, if their valves are operated by an automatic system, by commands released according to various parameters permanently measured. For 1455 example, in Fig.25, if the pressure p _r in the tank 8.2i is continuously measured with a piezoelectric transducer, the control system can determine on the basis of the gas law, what should be the instantaneous pressure p _r (variable from one cycle to another) at the inlet of the adiabatic compressor C1 , for the outlet temperature to reach T _lz and gives a command to the solenoid valve at the outlet of the isothermal compressor C2 to open exactly when the gas pressure in the compressor reaches this 1460 value. These state variables can be provided also by a quasi-isothermal compressor, which absorbs gas at temperature T _amb and pressure of p _amb and exhaust it in the tank R, with temperature 7 _r and pressure p _r . The compression ratios of the two adiabatic compressors do not change in this configuration, which is useful, especially for high values of T _iz . For T _iz values close to T _amb , the configuration in Fig.26 is useful. Flere, the first step ensures, through a quasi-adiabatic compressor 1465 (for example, a low-pressure blower), a significant flow rate of gas and ensures in the tank R the same isothermal temperature T _iz , at pressure p _r . One or more isothermal compressors take gas from this tank and compress it isothermically at this temperature, their exhaust valve being opened automatically when the pressure in the compressor reaches the value p _r .

A permanent non-deformable thermal sponge can be placed in the densifier 8.2i, as its piston 1470 is a gas piston. Any solution used at the prior art to reduce the polytropic coefficient can be chosen for its realization, but superior efficiencies are obtained when installing thermal sponges with a heat absorption surface as large as possible, where there is the possibility of easy circulation of gas and liquid through the channels and holes of the sponge, reducing to the maximum mechanical energy losses by friction, losses which would be converted into thermal energy which should be eliminated to 1475 the environment and where there is an efficient coolant well distributed in the enclosure, the flow rate of which is correlated with the thermal capacity of the sponge and can be adjusted as the pressure in the tank increases (i.e. the heat to be eliminated). In the case of this type of densifier, in the absence of a moving piston, the surface of the thermal sponge in direct contact with the gas and can be greatly increased. For example, the tank 8.2 may be the primary of a plate heat exchanger (Fig.27), the 1480 secondary of which is part of a cooling/heating circuit equipped with circulator pumps and another heat exchanger, which gives the absorbed heat to another medium, thus benefiting from a large heat exchange area. Of the same large surface area, but with a higher heat transfer coefficient, has a part the secondary of the exchanger in which a fluid at the saturation limit is inserted, which by evaporation, followed by a condensation in an external condenser eliminates excess heat.

1485 Any of the quasi-isothermal compressors of the state of the art, or any of the isothermalizers described above, may become a second stage of a gas-piston compressor, if preceded by a first compression stage, providing output gas with temperature T _iz and variable pressure, always equal to the gas pressure of the second stage. For example, solid-piston compressors of Fig.4 - Fig.12, as well as liquid-piston compressors of Fig.17 - Fig.18 can be stopped in the position where the heat sponge 1490 reaches its most favorable position, with the most favorable distances between the sponge elements, both in terms of heat absorption speed and fluid circulation from the inside. In this position, by starting the first compression stage, these devices become gas piston isothermalizers, without changing the favorable position of the thermal sponge elements. Also, any of the pistons of the combined isothermalizers units in Fig.21 - Fig.24, can be stopped at the time considered most favorable and 1495 converted into gas piston isothermalizers. Compared to the state-of-the-art compressors, the advantage of these isothermalizers is to keep a large area of heat exchange permanently, even at high pressures.

For the tank 8.2i in Fig.25 we have chosen a parallelepiped configuration, the gas 8.2a in this tank being cooled by a fixed sponge, composed of tubular cylindrical supports 8.8v of some sprinklers 1500 inside which coolant for sprinklers circulates, and of horizontal metal plates 8.8o fixed to these supports. The spray of the sprinkler and that accumulated at the base of the tank in a layer whose level is kept constant by the pump 8.6m forms the itinerant liquid sponge which, together with the liquid in the tank 8.2, is the section inside the tank of a cooling circuit. This inner section also includes the sprinklers 8.8a mounted on the top wall of the tank, and the sprinklers mounted on the vertical 1505 supports 8.8v, each spraying liquid or foam, in the corresponding horizontal plane between the solid sponge plates. In turn, horizontal plates may have perforations to create longer routes for fluid leakage. To achieve this, the surface of the plates gradually decreases, from the lower plates to the upper ones.

The density and complexity of the thermal sponge system differ greatly, from one application 1510 to another, depending on the final compression ratio and the temperature T _iz . For example, for many applications it is sufficient, as with quasi-isothermal compressors of prior art, only by a itinerant liquid sponge distributed by ceiling sprinklers. In contrast, the input and output flow of the liquid agent are equal, the place of the solid or liquid piston that would lead to the decrease of the heat exchange surfaces being taken over by the gas piston, which keeps these surfaces unchanged. The tank may 1515 also be cooled with foam, if a foam generator is mounted on the pipe that introduces the gas into the tank and/or a connection is installed between the compressed gas pipeline at the inlet and the liquid layer mixed with surfactants at the bottom of the tank. On this pipe and on those supplying the sprinklers are fitted the blowers S, necessary to create a local overpressure and to adjust the flow rate of the gas-piston. In this case too, there are advantages over the state-of-the-art compressors: 1520 keeping a large area of heat exchange at all times, superior possibilities for distributing and regenerating the foam.

Another strategy that can be applied for an isothermal compression is to use a single compressor, but to accelerate the piston speed and obtain in the tank 8.2i a temperature T _iz +, higher than T _lz , so that during the time when the compressor piston 8.1 performs the admission and 1525 isothermal compression phases at the temperature of T _iz , the gas in the tank 8.2i is cooled below the temperature of T _iz .

The tank 8.2i in Fig.26 is also parallelepiped, the gas 8.2a in this tank being cooled by a system composed of a deformable metal band 8.2b, (similar to the strips used for transporting small materials), with a width almost equal to the width of the tank, which runs permanently on a roller 1530 system mounted inside, or outside of it. In turn, the entire system is inserted into a parallelepiped tank 8.2 filled with coolant 8.2I. Due to the mechanical energy received from the outside by the drive rollers, the metal band passes a winding path, mostly inside the tank 8.2i (Fig.26A) and another part in the tank 8.2, having only two crossing positions, or a route alternately distributed in both tanks. When passing the strip through the metal walls, are provided sealing gaskets or sealing rollers with a smaller 1535 diameter. The longer this route and the closer the adjacent portions of the band segments are, the higher the heat energy absorbed from the compressed gas at a given speed. For longer strip lengths, the rollers can be arranged in such a way that the metal strip has the cooled portions, outside the tank, as long as possible. In Fig.26A is represented another configuration of the placement of the drive roller: they are mounted side by side, bonded to each other, so that the two sets of rollers are 1540 constituted in the two walls, upper and lower, and the metal strip passes as tightly as possible between two such rollers. Seals are required in this configuration for peripheral rollers only. In this situation, after passing between two rollers, some segments of the metal strip may pass an additional route through the coolant. There is also presented a configuration in which all but two rollers are located inside the tank.

1545 The outer tank 8.2 is filled with fluid to a certain level, above the level at which the tank 8.2i and the cooling strip drive systems are located. At the top of the tank 8.2 remains a layer of gas which, through the pipe 8.2c, constantly communicates with the gas in the inner tank and into which the metal inserts 8gs are mounted. By achieving equal pressure between gas and liquid layers, the tank 8.2i can be made with much thinner walls, regardless of the total compression ratio. Secondly, most 1550 importantly, seals between fixed and moving parts, or between moving parts, are not subject to high pressures, which allows the volume of fluid 8.2I entering the tank 8.2i and the equal volume of gas passing into the top layer of the tank 8.1 to be reduced to low values. Also here, the fluid level in the tank 8.2 is kept constant with the pump 8.6m. This way, liquid infiltration in the second stage helps to improve the quality of cooling. In turn, the liquid in the tank 8.2 is cooled by including it in a system that 1555 also contains the pump 8.6M and the FIE liquid-gas heat exchanger. When the pressure in the second stage of the compressor reaches the desired value, the compressed gas can be exhausted by the valve 8.2s into the constant pressure tank 8.7, by inserting additional liquid into the tank, with the pressure equal to the final pressure of the gas, until the entire amount of gas has been transferred..

1560 Another type of compressor capable of performing energy-efficient isothermal compressions is the gas piston densifier shown in Fig.27. In principle, this densifier also consists of an isothermal compressor 8.1 (here, a solid piston densifier with thermal sponge made of horizontal plates 8.1a, mounted on the carry-supports of the harmonic type 8.1 r), or any combination of compressors and densifiers that exhaust in the manifold 8.3. For efficient control of the flow and temperature of the gas 1565 absorbed by the second step, in Fig.27 we also installed a polytropic compressor 8.4, which can be a screw compressor, followed by a heat exchanger. Over the course of a cycle, the pressure in the densifier starts from an initial value pi (usually atmospheric pressure) and increases to the instantaneous pressure in the tank pr, and the pressure in the tank increases slightly at intervals where, after opening the exhaust valve, the compressor piston exhausts the compressed gas. In the 1570 configuration in Fig.27 we chose the solution of an additional cooling of the compressed gas, by inserting it in the form of bubbles into the discharge line 8.3. Also, the liquid in this pipe, which is constantly recirculated, ensures the evacuation of compressed gas from the densifier.

The gas from the other compressors that exhaust into line 8.3 can be cooled in the same way, or by means of an additional FIE heat exchanger. The conditions ensuring a superior efficiency of 1575 the system are: a large and unobstructed section of the gas inlet and outlet routes, a sensitive, fast opening with small pressure losses of the discharge path, the existence of a liquid fraction of the thermal sponge whose volume is so adjusted (before, or during operation) that each time the piston reaches the TDC, the dead volume of the compressor cylinder is zero. The inlet and exhaust valves 8.1s are opened at the control of the controller. The exhaust valve may open at a fixed manifold gas

1580 pressure (fixed position of the piston T-point), in which case the compressed gas is directed to the constant pressure tank 8.7 (each gas tranche will replace a tranche with the same volume of hydraulic fluid), or at a variable pressure (any pressure above the initial pressure pi ) equal to that in the tank 8.2, in which case the compressed gas is directed to the tank 8.2 where it will still be compressed, and after reaching the desired pressure in this tank, it can be directed to the constant pressure tank 8.8. 1585 The best efficiency of using the energy received by the system is obtained when in the first step an isothermal compression is carried out, up to a pressure p _f seeking to achieve a temperature difference DT as small as possible (when the main compressor of this step is a densifier). There are many applications where there is also a demand for heat (cogeneration systems, energy storage systems with thermal energy storage, etc.). In such situations, as well as where a high speed of 1590 energy conversion is required, the use of adiabatic compressors is justified. In these applications, the first stage of the device may contain an adiabatic compressor and an heat exchanger FIE, with the primary at constant pressure, equal to that of the collector, having the role of reducing the temperature of the gas at the compressor outlet to a temperature as close as possible to T _amb , and the heat energy given to the agent in the heat exchanger secondary be stored in most cases, the recovery of excess 1595 thermal energy in the heat exchanger justifies the use of simpler (and cheaper) compressors, with higher output temperatures but with higher flow rates.

In both types of applications, in this compression stage gas is obtained at pressure p _f1 and temperature 7), close to T _amb . The second compression stage is made as a gas piston densifier with a very low compression ratio on a cycle (the inlet pressure p _f1 and intermediate pressure p _f2 being close 1600 in value), associated with an efficient system for removing excess thermal energy. Until the inlet valve is opened, the volume of the densifier is maintained approximately constant, the gas pressure suffering small oscillations caused by cooling. In addition, the total surface area of solid surfaces that take thermal energy from the gas being compressed does not change. After reaching the pressure p _F , the gas is transferred into the tanks 8.8, by means of a liquid piston, but can also be made solid piston 1605 configurations.

A simple solution for making the gas piston densifier 8.2 is to use a plate heat exchanger, whose seals are sized for pressures greater than p _f2 . Through the secondary 8.2a of the exchanger 8.2, a coolant circulates continuously in closed circuit, at a speed depending on the instantaneous compression ratio. This circuit includes, externally, a fluid/ambient HE heat exchanger. The input of 1610 the primary circuit is coupled to the collector line which in turn is coupled, via exhaust valves, with the compressors in the first stage, or to the HE heat exchanger, while the output of the primary circuit (used for discharging the compressed gas) is usually closed.

A considerable increase in the efficiency of the entire system, by reducing the difference DT, can be achieved by introducing a refrigerant in a state of equilibrium liquid/vapor, in the secondary of 1615 the densifier, and implementing in this system the necessary pipes for this secondary to become a thermal tubes cooling system (gravitational or with capillarity). In another variant, by putting a refrigerant in a near equilibrium liquid/vapor state in the secondary, it can become the evaporator of a Rankine-cycle, or ORC thermal engine. To realize the thermal engine, the heat exchanger HE is replaced by a condenser, which receives the agent through the turbine 8.4 and exhausts the 1620 condensate with the pump 8.6. In this configuration, the speed of all pistons in the compression phase can be increased, because the additional mechanical energy consumed, due to the increase in temperature difference DT, is fully recovered in the thermal engine. It is also easy to implement a combined system of the two configurations, in which the energy recovery system only activates after a temperature limit is exceeded.

1625 The existence of gas tanks under constant pressure 8.7 creates the possibility of a new compression stage. After storing the gas in constant pressure tanks 8.7 and 8.8, the compression may continue with the extraction of compressed gas from these tanks and its compression to a higher pressure value, the additional compressed gas may be stored in the tanks from which it was extracted (via a intermediate tank), or in tank 8.2.

1630 A variant of this densifier is shown in Fig.28. The first stage of this densifier consists of two identical densifiers 8.1 , similar to that of Fig.27, and the second stage, from the tank 8.2a, into which a tank 8.2g is inserted, made in the form of a comb. The combination of the two tanks is a heat exchanger 8.2. Each of the two tanks is a gas piston densifier, the exhaust of which can be directed during the working regime, to a cooling system, containing an ambient-gas heat exchanger and a 1635 blower C for driving the gas or, after reaching the desired pressure, toward constant pressure tanks 8.7. The system also offers the possibility that, for large compression ratios (with the consequence of increasing the heat to be exhaustefd) it will go into forced mode, in which only one of the first-stage densifiers is in operation, and its second stage is cooled by the heat exchanger formed by the tank of the second densifier and the heat exchanger concerned.

1640 In Fig.29 another type of gas piston densifier is shown, consisting of compressors system 8.1 and the tank 8.2. An advantageous solution for cooling the walls of the tank is to install this tank inside a larger tank with liquid, in which the pressure is maintained at all times equal to that of the tank 8.2. This allows the walls of this tank to be thinner, allowing faster heat escape.

Gas supply at temperature T, and variable pressure p _f1 is made by a system of densifiers, 1645 compressors and heat exchangers, with a common collector 8.3, similar to that of Fig.27. Inside the tank is installed a thermal sponge according to the invention, designed according to the characteristics and requirements of the system, adapted to the gas supply system and to the cooling system. In Fig.29, the main part of the sponge is a system 8.8v of bars of various thicknesses and vertical plates of various widths, arranged at sufficiently small distances from each other (to achieve a good capture 1650 of the heat accumulated by the gas in the tank), but large enough to allow a slight leakage of the coolant. For high compression ratios, the thermal sponge can be made of yarns, preferably metallic, thin and deformable (like textile yarns). These bundles of wires are mounted on the upper wall of the tank 8.2. A high density of the wires ensures a very good absorption of thermal energy, and the malleability of the wires ensures an easier drainage of water droplets. This type of sponge can also be 1655 mounted on the other types of gas piston isotherms described above, if accompanied by an efficient cooling system.

The liquid required to cool the gas is distributed between the bottom of the tank 8.2, between the heat exchanger FIE, the sprinkler system 8.8a, the pump body 8.6 and the piping system 8.3. The system allows the installation of any type of sprinkler from the state of the art, their number and 1660 distribution, the flow rate and pressure difference, the dispersion angle, the size of the drops produced and other characteristics, being chosen according to the characteristics of the application. In many applications, a sprinkler system is enough to generate a dense and permanent fog of very small drops, the support of sprinklers taking on the role of solid sponge. It is also advisable to create areas with different temperatures that generate ascending gas currents.

1665 In the configuration in Fig.29, a sprinkler system mounted on the top of the densifier was chosen, which spreads the liquid droplets horizontally. A significant amount of this liquid remains attached to the solid sponge elements, helping to limit the temperature increase. In addition, in this configuration, the coolant contributes to the compression and cooling of the gas sucked from the first stage, through the bubble densifiers 8.11 , made right on the coolant transport pipes, by injecting them 1670 with gas from the first stage collector. From here, the gas reaches the sprinklers and is exhausted in the densifier, even in the areas with the highest temperatures of the gas. If the gas from the first stage is injected into a beam of pipes with a sufficiently small diameter it will not form bubbles, but will form successive layers of gas, due to the surface tension of the liquid, alternating with layers of liquid, which improves the conditions of heat transfer between the two media. 1675 The chosen configuration also uses other cooling processes from the technical stage, namely, increasing the heat transfer surfaces by introducing, or creating aqueous foam. To achieve this desideratum, in the liquid phase 8.11 of the thermal sponge at the base of the densifier, and in the liquid trays 8.9 mounted inside the thermal sponge, a number of surfactants are added to reduce the surface tension of the liquid, this favors the formation of foam when compressed gas is introduced into 1680 the liquid from the collector of the first stage.

Isothermal transformations can also be achieved with rotating devices, starting from the rotary compressors of the state of the art, applying the procedures described in this invention. These transformations can also be carried out with rotary devices which, at the technical stage, are most often used as liquid pumps or internal combustion engines, if appropriate measures are taken to 1685 ensure the tightness between the enclosures with different pressures.

As mentioned above, in the case of single-enclosure rotary devices (as with the blade compressor in Fig.30), the isothermal speed can be obtained by continuously changing the angular speed of the rotor, in such a way as to maintain at all times the equality between the instantaneous work delivered to the gas by the piston (in this case, the sliding blade in the rotor) and the 1690 instantaneous thermal energy transferred by the gas to its environment. In the case of rotary devices with several closed enclosures, the change in the angular speed of the rotor has different effects in each of them, so that it is preferable to maintain a constant rotor speed and to modify separately for each enclosure other characteristics that allow this equality to be achieved (e.g. coolant inlet and outlet flow rate). The permanent change of angular speed can be abandoned also in applications 1695 where the power of the device is a decisive factor, applications where efficient thermal sponges should be used, with large absorption surfaces and cooling installations of high performace.

The isothermalizer described in Fig.30 is a variant of the blade compressor, compressor described in detail in the patent application R0128041 (A2). It is characterized by the fact that it uses only one blade in the rotor. It consists of a stator (empty cylinder 6.2), inside which it rotates around its 1700 center shaft, the rotor 6.1. In this configuration, the rotor is empty and its diameter is larger than the radius of the stator. In the rotor is mounted a pocket, usually parallelepiped 6.4, without side walls, with the main walls parallel to the plane formed by the rotor diameters, in which the parallelepiped blade 6.3 is inserted, the length of which is equal to the inner length of the housing in which it is inserted, the height is less than the depth of the pocket and the thickness is equal to the inside 1705 thickness of the pocket (the four side surfaces of the blade slip tightly onto the inside surfaces of the housing). The length of this housing is equal to the inner length of the stator (the surfaces of the base of the blade also slip tightly onto the inner surfaces of the stator bases). As a rule, the rotor is tangent to the inner surface of the stator wall. In the configuration in Fig.30, the radius of curvature of the stator wall is modified, over the entire length of the stator, on a sector 6.5, being equal to the radius of the 1710 rotor. In this way, the contact portion between the rotor and the stator is no longer limited to a straight segment, but extends to a curved surface with the desired width. The blade can slide along the entire height of the notch, and when its tip touches the stator wall, it divides it into two chambers, sealed between them. This extreme position of the blade is ensured by the centrifugal force generated by the rotation of the blade, as well as by the pressure of the fluid 6.11 closed between the blade and the 1715 bottom of the notch (a lubricant, which is also coolant of the gas and which is inserted using a pump, through a flexible pipe 6.41 (Fig.33) and circulates between the housing in the rotor, a heat exchanger and a tank) or/and, as in the current state of the art, by elastic springs. The liquid in this housing can penetrate, due to high pressure, through a groove made in the blade (channel 6.3a in Fig. 31), to the contact surface between the blade and the stator, mitigating the effects of friction and providing a 1720 superior seal. The sealing between the two compartments with different gas pressures can also be improved by fitting elastic seals (6.3b in Fig. 31) whenever possible.

The height of the rotor can be equal to the inside height of the stator, in which case the surfaces of the two bases of the rotor slide over the surfaces of the two stator bases. In the configuration in Fig.33, section 1-1 , this height is higher and the sliding movement between the stator 1725 bases and the rotor walls is provided by bearings 6.91 and segments or seals, etc. The rotor of the machine is mechanically coupled with an engine (electric or mechanical), and in the case of a expander, with a generator or other mechanical load. On both sides of the tangent surface, there are two rectangular slots (6.6d and 6.7d in Fig.30), connected to pipes 6.6 and 6.7 respectively, for the siction and for the exhaust of the working fluid. If the machine acts as a densifier, the inlet 6.6d can be 1730 free, and on the discharge line 6.7 a valve 6.7a is mounted, automatic or operated by a coil 6.7b (Fig.30). If the machine acts as a rarifier, the suction is through a valve or drawer and the exhaust is usually free. In the configurations in which the axis of the stator is vertical, the suction and the exhaust of the working gas is made by cut-outs executed in the two circular plates that constitute the bases of the stator: a cut for intake in the 6.6v area of the lower base and a cut for exhaust in the 6.7v area of 1735 the upper base of the stator.

With these constructive components, the described device is a rotating polytropic compressor that can achieve good performance in certain specific applications. Like any polytropic compressor, it can perform isothermal compression operations when its angular velocity is equal to the isothermal angular velocity u>(t) over the entire duration of a rotation, but even for large temperature differences 1740 ^17 ^” this speed is very low. Achieving higher speeds and, consequently, higher compressed air flow rates is possible by reducing the polytropic coefficient of compression. In the densifier shown in Fig.30, this objective is achieved by injecting an abundant liquid thermal agent, which serves as a liquid thermal sponge when the liquid inlet flow rate is equal to the outlet flow rate, and also as a liquid piston when the input flow rate is higher. The injection process can start outside the machine in a 1745 humidification antechamber (AC from Fig.30B), where the volume is constant and there are no moving parts, it is easier to control. In this antechamber, the working gas introduced via a rotary compressor is cooled, its temperature being brought to the working value T _iz . In the antechamber is abundantly sprayed the coolant, process which continues in the stator enclosure, through the nozzles 6.9b, mounted at the end of the sprinklers 6.9a, supplied from the pipe 6.9. The flow rate of each sprinkler 1750 can be changed with the adjustment tools 6.9v mounted at the entrance to the main line 6.9, or on each sprinkler. The rotor cylinder can also be filled with coolant, directed to sprinklers mounted in the rotor wall. Also, part of the fluid 6.11 is driven by the rotor blade and discharged together with the compressed gas onto the exhaust pipe and after separating it into the pipe 6.7c is collected in the tank 6.10, which is part of a cooling circuit together with the heat exchanger FIE and the pump 6.7M. All 1755 actuators of flow-regulating devices, as well as those that determine the angular speed of the rotor, are controlled by a central device DC that receives signals from piezoelectric pressure transducers 6.8 mounted in the work room. The central device shall be programd in such a way as to ensure the equality between the mechanical power given to the gas and the thermal power given by the gas to its environment.

1760 Due to the fact that at high gas pressures and low speed of rotation of the blade, the separation between the two compartments is more difficult to achieve, the compression ratio of the gas obtained with a single densifier being limited. In order to obtain large compression ratios, it is necessary to series several densifiers, which perform a compression in pressure stages and a judicious distribution, within a full rotation, of the total engine torque. Figure 31 shows how multiple

1765 apparatus can be arranged in series, and Fig.30B, how densifiers can be superimposed. In this latter case, the liquid collection tanks are mounted between densifiers, and the rotors drive of all densifiers is done by a single motor, on the axis of which the gear shift GS are mounted, which makes the transition from the angular speed of the motor to the angular speed of that device.

The isothermalizer in Fig.32 cumulates a series of changes that can be made to the

1770 isotermalizer with a blade, changes that can be applied separately, or cumulating several of them, depending on the objective pursued. These changes are:

- changing the shape of stator 6.2 so that a section parallel to the bases is no longer circular, but the new section allows a continuous and watertight slide of the blade, and leads to a favorable change in the isothermal angular speed L oft). The stator shape and the curve direction of the blade,

1775 determine different isothermal speeds. From this point of view, the construction of the densifier may differ from that of the rarifier.

- entering a radius of curvature for the blade 6.3 (here, consisting of two sections, 6.31 and 6.32) and implicitly, for the corresponding notch 6.4 in the rotor, change which also results in a change in the isothermal angular speed co(t).

1780 - reduction of the rotor radius 6.1 , combined with the realization of a telescopic configuration of the blade 6.3, to increase the useful interior space of the stator. As a rule, this change must be made by making an integer ratio between the length of the inner circumference of the stator and the length of the outer circumference of the rotor.

- making in the rotor of some internal cavities. In the configuration shown in the figure, these

1785 cavities are designed to make, through valve 6.65, a passage between the uncompressed gas tank and the low pressure chamber of the compressor, and through valve 6.66, a path between the high pressure chamber and the compressed gas tank

- inserting the stator into a tank 6.10 filled with coolant 6.101 , a fluid that, with the help of a pump, circulates through the cooling system SR. Sprinklers 6.9b are supplied via pipes 6.9a directly

1790 from the tank 6.10.

- changing the kinematics of the apparatus: the inner cylinder 6.1 is held fixed and the outer cylinder 6.2, together with the tank 6.10 and the cooling system SR, mounted on one of the caps, rotate around it. In Fig.33 the operation of a rotary isothermalizer with a rotor blade is exemplified, when the 1795 outer diameter of the rotor 6.1, with circular section, is equal to the inner radius of the stator 6.2, also with circular section. In the configuration shown in the figure, the friction movement between the rotor and the stator, existing on the machine in Fig.30, is replaced by a rolling motion of the rotor on the inside walls of the stator. During this free rolling motion, the rotor moves along a circular path on the flywheel 6.81 (Fig.33, section 1-1 ) around a shaft perpendicular to the flywheel. In turn, thanks to an 1800 adjustable speed motor system, the flywheel rotates around the axis of the stator, at a distance equal to the length difference between the two rays (at the apparatus shown in the figure, where the radius of the stator is twice the radius of the rotor, this distance is equal to the radius of the rotor). By turning the flywheel, the rotational axis of the rotor moves on a circle with the center on the axis of the stator cylinder, the walls of the two cylinders being permanently in contact on a generator. In this way, the 1805 rotor is driven to rotate around its axis. At a complete rotation of the flywheel, the rotor performs exactly two rolls on the inner wall of the stator, passing through two main points on the circumference of the stator (the position in which the blade is entirely inside the rotor), the position of which is the same at each rotation. On either side of these points are mounted in the rotor wall, inlet holes 6.61 and 6.63, and exhaust holes 6.62 and 6.64, respectively. The opening and closing of these holes is 1810 done by means of rotating drawers 6.83, which are driven by the engine 6.8 via the axes 6.82, with a rotation speed equal to the flywheel rotation speed. The position of the drawers in 4 different positions (l-IV) of the rotor is indicated in Fig.33B, by horizontal bars for the “closed” position and by vertical bars for the “open” position. Cooling of the gas during compression is done by the sprinklers 6.9b mounted in the wall of the stator, or by itscovers 6.22. As with the compressor in Fig.30, coolant 1815 circulation can also be made through the inside of the rotor 6.1 , if it is not used for other purposes. At this ratio of 2:1 between the two diameters, the rotor blade cannot be executed in one piece, and a telescopic blade consisting of two sections 6.31 and 6.32 respectively is required. Also, in the configuration shown, where the rotor height is higher than that of the stator, to allow the rotor to move, the stator covers 6.2 (the two bases) must be movable in relation to the walls 6.21 : they rotate through 1820 bearings 6.91 mounted on the stator walls and through bearings 6.92 mounted on the rotor walls.

Fig. 34 shows how a solid sponge can be implemented, composed of almost parallel plates in such an isothermalizer. The plates are cylindrical metal sheets 6.12, each with a notch along a generator, with an opening slightly larger than the width of the blade, with unequal diameters, with values between stator and rotor diameter, mounted between these two cylinders so that, compared to 1825 the assembly in Fig.30, the central axis of the rotor is moved towards the central axis of the stator, in the plane containing them, with a distance equal to the total thickness of all these plates, without leaving gaps for gas leaks. By simply rotating the rotor about its axis, the cylindrical plates of the thermal sponge are engaged in a rotational motion in which the peripheral points of contact with the rotor and those of contact between successive plates move at the same speed, which would lead to 1830 different angular velocities of the plates and at pressures exerted on the rotor blade. If these plates are light enough and elastic enough, they can be driven by the rotor blade in a rotational motion synchronized with that of the rotor. Another way to ensure that all cylindrical plates have the same angular velocity is their successive reciprocal drive. To achieve this mechanism, gear teeth 6.1 m are mounted from place to 1835 place, on the outer surface of the rotor cylinder, for example in the shape of triangular prisms (see also Fig.35). Each such tooth corresponds, in the same plane perpendicular to the central axis, on each plate of the thermal sponge, a hollow 6.12m (obtained, for example, by punching), or a hole, slightly longer than the rotor tooth. These holes are made in such a way that the holes near the dead center overlap over the corresponding rotor tooth (so the distance between the holes increases as the 1840 diameter of the cylindrical plate increases). In this way, as it rotates, the rotor engages with the first plate, this with the next, and so on, equalizing their angular velocities.

In the configuration in the figure, the rotor is empty and serves to convey the gas with the inlet pressure p _a , through the valve 6.6a and of the one at the exhaust pressure p _r , through the valve 6.7a. The coolant 6.4I is conveyed through the rotor blade 6.3, through the pipes 6.9a and through the 1845 sprinklers 6.9b, from where it is injected between the sponge plates, the liquid in the stator being discharged, by means of a pump, through a hole made in the lower cover of the stator. .

A very similar construction has the rolling piston isothermalizer in Fig.36. It is made with the same design as the state-of-the-art rolling piston compressors, to which has been added the thermal sponge, the sponge cooling system and the measuring-regulation-control devices that ensure the 1850 isothermal angular velocity. In this configuration, the vane 6.112 that separates the different pressure zones is operated from the outside, using the spring 6.113 and performs back and forth movements in the cylinder 6.111 , along a fixed axis. Rotor 6.1 performs a rolling motion inside the stator 6.2. Under these conditions, the 6.12 plates of the thermal sponge are not engaged in the rotational movement, no additional mechanisms are needed for synchronization.

1855 The isothermalizer in Fig. 37 has a design similar to vane pumps. It consists of a stator 6.2 inside which, tangent to one of the generators of its inner surface, the rotor 6.1 rotates. In the body of the rotor are made several notches 6.4, equally spaced, in which slide the parallelepiped blades 6.3, which, some springs mounted at the bottom of the notch, keep them in constant contact, along a generator, with the inner surface of the stator. In this way, the internal volume of the stator is divided 1860 into several regions whose volume undergoes successive increases and decreases, depending on the rotation angle of the rotor. For judicious use of the entire available volume, in the case of densifiers, the extended regions communicate with each other through the stator wall in a 6.6a portion open to the environment (if or to the suction line 6.6 (if The exhaust valve 6.7a, actuated by a solenoid valve 6.7b, is located in the region where the volume of gas between two 1865 successive blades reaches the minimum value. In the case of rarefiers, the direction of rotation of the rotor and the role of the valves are reversed.

As with the single vane isothermalizer, the change in working gas pressure is due to both the change in the internal volume of these regions due to the movement of the vane and the change in this volume due to the spray of coolant using 6.9b sprinklers fed through pipes 6.9a. The solenoid 1870 valve 6.9b is controlled separately for each volume of gas contained between two successive blades. As with the single-blade isothermal, the working gas can also be cooled by replacing the sprinklers with foam generators. A more efficient use of the internal volume of the stator 6.2 is achieved by a simplified construction of the rotor 6.1 , keeping only its central axis, which can be full (Fig.39), or empty (Fig.38), 1875 on which they are mounted pockets 6.4, in which the blades 6.3 slide. Between the surface at the base of these pockets and blades are inserted elastic springs and/or lubricating fluid 6.4I. On the isothermalizer rotor in Fig.38, a cylindrical tank 6.4r is also mounted with a smaller radius than in the case of the device in Fig.37, and on the outer surface of this cylinder, the flat plates (preferably metal)

6.1 p are mounted radially, which forms the solid thermal sponge, and the radial pipes 6.11, at the end 1880 of which the sprinklers 6.9b are mounted. On the isothermal rotor, can be mounted solid thermal sponges which can have also other configurations. In Fig.39, the thermal sponge is made of metal wires 6.3b, which can occupy all the space that in Fig.37 the rotor of the device occupies.

In both cases, the suction of the working gas is made as in the case of the isothermalizer in Fig.37, through a wide opening 6.6a in the stator wall, but for exhaust a 6.7 wide opening is used, with

1885 width, measured on the circumference of the circle, slightly larger than the distance between two successive blades. The exhaust opening continues with a pipe mounted at the highest elevation of the stator, if its axis is horizontal. At the stator end of this pipe, a layer of liquid is permanently maintained, which ensures a complete emptying of the compressed gas, with the final pressure p _f , located in the interpaletary space next to the exhaust port (interpaletary space with minimum volume) together with 1890 the 6.2I liquid injected by the 6.9b sprinklers. In addition, if there is a pressure difference between the gas in the pipe and the gas in the apparatus when the position of the blades forms a communication path, this layer of liquid acts as a liquid piston, avoiding the loss of exergy that would occur in the case of its absence. Further rotation of the rotor leads to the emptying of the liquid from the inter-blades space into the tank 6.10 and the intake of the working gas. If the stator axis is vertical, the inlet pipe is 1895 mounted on its lower base, through a 6.6v opening, and the discharge pipe is mounted on its upper base, through a 6.7v opening, of the shape and positioning indicated in Fig. .38 and Fig.39, respectively.

Quasi-isothermal compressions can be obtained just as easily, starting from state-of-the-art liquid ring compressors, with obtaining, at the same discharge temperature, higher compressed gas 1900 flow rates, if between the rotor blades of this type of compressor efficient thermal sponges are introduced, similar to those described in Fig.38 and Fig.39.

The construction of the isothermalizer in Fig.40 brings together the characteristics of several types of isothermal described above. It is a solid double-acting piston device consisting of a cylinder

5.1 (not necessarily circular in section) and two covers: one upper 5.1s and one lower 5.1 i. Together, 1905 they delimit a closed enclosure, divided into two compartments by the piston 5.2, the sealing between them being made with elastic gaskets, segments, etc. The piston is moved between a bottom dead center BDC and a top dead center TDCn, where n=1 for the lower compartment and n=2 for the upper compartment. Obviously, BDC1ºTDC2 and TDC1ºBDC2. The position in figure 40 corresponds to the situation TDC1ºBDC2. This movement is due to a drive motor, mounted inside one of the 1910 compartments, or outside, on the cylinder wall. The motor shaft operates one or more profiled cams 6.14, the profile being executed in such a way that the displacement of the piston is done with the isothermal speed v1 _iz . Through kinematic connections, these cams lead to the telescoping of the rods that move the piston. In the configuration in the figure, the cams move a horizontal bar 6.15 over a short distance, which in turn, pressing the bolts 6.13, lead to the rotation of all the splints 5.23 and 1915 5.24, splints that form the telescopic rods of the piston. When the point furthest from the camshaft reaches its highest position, the piston reaches TDC1. During its ascent, the piston compresses the springs 6.16, mounted between the cover 5.1s and the horizontal bar 6.15a, which rests on the bolts 6.13a, mounted on the upper telescopic rods. In this way, the downward displacement of the piston is also determined by the profile of the profiled cam 6.14, with the speed v2 _iz , which may be different 1920 from v1 _lz „ due to the constructive differences between the two compartments (these differences can also cause differences in volume, or pressure, which may require the use of different tanks for the storage of compressed gas).

The telescopic rods in the two compartments also serve as carry-supports for the horizontal plates 5.11 of the two thermal sponges and are made according to the model of the carry-supports in 1925 the isothermalizers in Fig.12. These rods also serve as supports for the 6.9b sprinklers and the 6.9 pipes that feed them (these pipes can be placed right inside the splints that make up the telescopic rods. In this way a continuous cooling of the thermal sponges can be achieved. The cooling circuit is composed of the 6.7M pump which feeds the 6.9 ducts passing through the lower cover, then through the movable piston (strips 5.22), from the lower to the upper circuit, from the liquid layers 8.2s and 1930 8.2i, which collect the liquid from sprinklers, from the 5.22c telescopic rod that collects the liquid from these layers, keeping their volume constant, and from the heat exchanger HE.

In addition, in the upper part of the upper compartment, the free volume that forms in the area of the elastic springs when the piston is in TDC1 , an additional 8gs sponge is mounted, made of a metal wire. Suction and exhaust of gas are made through valves 6.6i, 6.6s and 6.7i, 6.7s. The 1935 compressed gas is exhausted, for both compartments, through pipes containing a layer of liquid, with the role of completely filling the gas bags and with the role of liquid piston for equalizing, without loss of exergy, the pressures.

Fig .41 A and Fig.41 B show some of the modifications by which other state-of-the-art devices, the gear pump and the cam pump, respectively, can be transformed into rotary isothermal densifiers. 1940 In these configurations, the 6.9b sprinklers are mounted in the housing (they can also be mounted in the rotor body 6.14, 6.15, respectively 6.16, 6.17) which inject coolant into the space between the gear teeth. This liquid is the liquid piston of the compressors, the flow rate through each sprinkler being controlled by means of 6.9c valves (for devices with larger volumes, these can be adjustable valves with servomotors), which receive commands from a central unit, depending on the pressure in closed 1945 enclosure corresponding to the respective sprinkler, pressures indicated by piezoelectric transducers 6.8. As in the case of the rotary apparatus described above, a layer of liquid 6.2I is maintained on the discharge line 6.7, to equalize, without loss of exergy, the pressure in the last inter-blades chamber with that in the storage tank.

Fig.42 contains some proposals for the implementation of thermal sponges in the 1950 configuration of some types of scroll compressors and some peristaltic compressors of the prior art, in order to bring as close as possible the polytropic coefficient of the transformations that take place in these devices to the unit value. As with the other isothermalizers described above, the proposed objective can be achieved if, in addition to this process, are applied procedures for the angular velocity modification and for the controlling the cooling processes of the gas in the compression phase by 1955 introducing a suitable liquid, in the spray state, by the introduction or generation of foam, or by the introduction of substances in suspension. In addition, in order to obtain an increase in the compressed gas flow rate, the procedures described above can be implemented for the complete discharge of the compressed gas and for the elimination of the dead volume. Since, for the implementation of these processes for scroll and peristaltic compressors, no procedures other than those described above are 1960 proposed, in the representations in Figure 42, they are not described.

Fig.42A shows a cross section through a compressor, with the two spiral volutes 6.18 and 6.19 (here, Archimedean spirals) interspersed. Usually, one of the volutes is fixed, the other performing an eccentric orbital motion, without rotating, but there are also compressors in which, to ensure a safer seal between the compartments with different pressures, the two volutes rotate 1965 simultaneously, in the same direction, but with different centers of rotation. The sealing between the compartments with different pressures is achieved by using 13.6 spiral-shaped gaskets, mounted on grooves made on the ridges of the two spirals.

The thermal sponge is composed of thin elastic metal plates, having the same spiral shape as the main spirals, about the same length, the same height and the same step, but with a smaller 1970 thickness g. They are located between the two main spirals, the distance between them being a multiple of a whole fraction of the distance between the loops of a single volute. For example, if this distance is b , the distance between two successive spirals of the sponge is b IN, where N is the number of plates that make up the sponge. In this way, between every two loops of the spiral considered mobile, are found a fixed spiral and a number of 2N spirals of the thermal sponge. To 1975 prevent the movement of the thermal sponge plates along the spirals, displacement that may be the result of frictional forces that occur at points where the distance between the two main spirals is zero (contact points), the thermal sponge spirals will be longer, exceeding both extremities the respective end of the main spiral, as in Fig.42A, and will be perforated, through the resulting holes being inserted a rod 6.22, fixed to the respective main wing, which allows the perpendicular sliding of these spirals 1980 under the action of elastic forces, but prevents other types of travel. Configurations in which both main spirals rotate in the same direction, in addition to the known advantages of compressors equipped with this type of thermal sponge, offer another advantage: the sliding friction of the contact points is replaced by a rolling friction, which leads to reduce the stresses exerted on the spirals.

If n is the number of plates between the two spirals, plates that have the thickness g, on both 1985 sides of one of the main spirals (the fixed one, if only one is movable) a thickness reduction with depth Ng is performed along the entire length of this spiral over the entire height t of this volute. At the points where the distance between the two main spirals is minimal, this distance is equal to N g and is filled, in its entirety, by the spirals of the thermal sponge. If the two covers 6.18 and 6.19 are arranged in a horizontal position, the sponge coils are supported with their lower edge on the lower cover of the 1990 compressor, their upper edge being in contact with the upper cover. All these contacts must be tight, at least when the distance between the two main spirals is minimal. The thermal sponge is mounted by temporarily fixing the N thin spirals on each of the faces of the fixed spiral, its thickness thus becoming equal to the thickness of the movable spiral, then the introduction of the movable spiral, followed by the release of the sponge spirals, spirals that will distance due to their elasticity. Due to the 1995 tendency to return to the original shape, the spirals of the thermal sponge will be arranged, more or less evenly, inside the compressor. This configuration has two major shortcomings: the difficulty of eliminating gas leaks between areas with different pressures and the relatively uneven distribution of sponge spirals, the distance between the spirals decreasing successively, as the distance between the main coils decreases.

2000 The compressor shown in Fig.42A manages to remove these deficiencies. First of all, although the spirals of the thermal sponge are also made of elastic metal plates, on each spiral 6.21 is fixed a series of identical plates 6.20, with the width equal to the width of the spirals, but with a much smaller length. When not tensioned, these plates have a shape close to the letter S, as shown in Fig.42C. In this figure is represented unfolded, a portion of the set of main spirals, thin spirals and

2005 elastic plates, when the distance between the main spirals is maximum. The representation of the same assembly, when the distance between the main spirals is minimal is shown in Fig.42B.

The thickness reduction of the fixed spiral is performed only on a portion of its side faces, which leads to the formation of two channels 6.23 (visible in cross section 1-1), one on each side, along the entire length of the spiral. The depth of this channel has the value 2 Ng, if the thickness of 2010 the elastic plates is equal to that of the spirals of the thermal sponge. The height of the thermal sponge spirals is equal to the width of this channel, and the spirals are arranged in such a way that when the distance between the two main spirals is minimal, the sponge spirals, together with the spacer plates penetrate these channels and occupy the entire volume and in the parts in which the main spirals are spaced at a distance L, the sponge spirals are spaced from each other, with a fraction 2015 UN, the same for all interspaces. As the number of contact points is greater than 3, the sponge plates always have at least 3 support points in that channel, so that they will never come into direct contact with either the lower or the upper cover.

With these modifications, the only locations where pressure losses may occur are located at the contact points, if the respective section of the channel in the fixed spiral is not completely occupied 2020 by the thickness of the secondary spirals and spacer plates. The sealing method proposed in Fig.42A, section 1-1 , is the mounting, on the bottom and on the edges of the respective channels, of some membranes made of elastic materials, slightly deformable. The volume formed between this membrane and the channel walls is filled with a fluid and is tightly divided, by deformable walls, into regions with a width not greater than necessary to cover the contact surface between the main spirals. 2025 When a point of contact approaches such a region, the spirals entering the channel press the membrane at the bottom of the channel and push the fluid between the membrane and the bottom of the channel into the regions between the side walls and the membrane, obstructing possible gas leaks, as can be seen in “magnifying glass” 2, which shows an enlarged image of the region in question, when the movable spiral 6.18 steps on the fixed spiral 6.19.

2030 Fig.42D shows one of the possibilities to transform a peristaltic compressor into an isothermalizer, by implementing a thermal sponge. A thermal sponge (Fig.42E shows a cross section through its unfolded shape) consisting of metal plates 5.14, elastic, corrugated, similar to those described in Fig.10a and metal plate 5.11 , flat, rigid, occupying a central position, between two sets of corrugated boards. All these plates have the same dimensions when fully tensioned. In order to 2035 prevent uncontrolled movement of these plates, they are fixed by an elastic cord 5.7, or by an elastic spring, located between the lower and the upper plate. The thermal sponge is inserted into a deformable peristaltic tube, the shell of which is elastic. When the tube is pressed between two jaws, the corrugated plates become flat, and the sponge acquires a shape similar to that shown in Fig.42B, having in addition, the peristaltic tube, which tightly tightens these plates (position 3 in Fig.42E). The 2040 same result is obtained, in unfolded form and when the sponge is made similar to that of the screw compressor (Fig.42C).

One of the disadvantages of compressors with peristaltic tubes is the need to periodically replace, due to wear, their coating. One way to remove this inconvenience is to replace these tubes with metal channels (troughs), for which can be used the configuration used to insert the thermal 2045 sponge into the channels made in the fixed spiral of the scroll compressor. A section through such a channel is shown in Fig.42F. The shaft of trough 6.2, with a rectangular cross section, is a section of an arc of a circle, having its center having the center in the center of rotation of the arm 6.27, the arm on which is mounted, by means of the bearing 6.32, the pressing roller 6.26. As in the case of the peristaltic tube, the ends of the trough are curved and have a path, towards the downstream and 2050 upstream device, outside the range of the pivoting arm. Inside the trough is mounted a thermal sponge, similar to those in Fig.42E, or Fig.42F, and its walls are lined with an elastic sealing membrane, similar to membrane 6.23a in Fig.42A, section 1-1.

Above the thermal sponge, in its completely unstressed state, along its entire length, a lamella is mounted with a width equal to the inner width of the trough. Continuous channels are drilled on the 2055 side edges of the lamella, in which a sealing gasket is mounted, transforming the lamella into a real piston. The material from which the lamella is made (metal, or plastics similar to those from which peristaltic tubes are made) must be sufficiently malleable, so that, under the action of the force exerted by the roller, combined with that exerted by the elastic elements of the sponge, to follow a smooth route, with acceptable radii of curvature, but be hard enough to withstand the stresses to which it is 2060 subjected for a long time. The end of the lamella in the inlet area of the pressure roller must be shaped in such a way as to allow gradual entry of the roller and must be fixed in relation to the trough so as to prevent movement along it. Likewise, the other end of the blade must be shaped in such a way as to allow easy and complete evacuation of the compressed gas. If the material of which it is made does not have sufficient longitudinal elasticity, this end may be allowed to slide freely on the bottom of the 2065 trough.

Although there are also linear peristaltic compressors, the most used are those with circular or spiral tubes. In all these variants, thermal sponges can be mounted inside the tubes. In the configuration of Fig.42D, two peristaltic tubes 6.25 are used which have the shape of segments from a circular ring and are fixed on a metal bed 6.24, so as to allow the penetration into the circular space 2070 between them, a pressing roller 6.28. This roller is mounted on a movable arm 6.27, which is continuously rotated by a drive motor 6.26. A higher power density can be obtained if several pressing rollers are mounted on the same arm, each such roller performing the action of compression on two circular peristaltic tubes, with different radii of curvature, equipped with an internal thermal sponge. Each of the peristaltic tubes is provided, at the end of the support on which it is mounted, with a check 2075 valve (in some applications, the inlet valves can be dispensed with). A single pressing roller acts on each peristaltic tube. When this roller reaches the peristaltic tube, both valves are closed and the tube is filled with gas at the initial pressure. As soon as the roller passes the first end of the tube, the inlet valve opens so that another portion of gas enters the rear portion of the tube at the initial pressure p„ while the gas in the tube is progressively compressed to the final pressure p _f . At this point, the exhaust 2080 valve opens and the roller acquires the role of evacuating the compressed gas to the user. The two valves close simultaneously when the press roller reaches the second end of the peristaltic tube. Configurations can also be made in which, on the metal bed, there are N circular tubes of equal length, shorter, each provided with its own valves. For each such tube a pressing roller is mounted, at the same distance from the center of the device, on separate arms, which make equal angles between 2085 them. The use of peristaltic tubes can be extended to other types of compressors. For example, on the inner wall of the stator of vane compressors, peristaltic tubes with thermal sponge can be mounted, the tube being pressed and the gas being compressed by the rotational movement of the vanes. The final pressure of the gas in the peristaltic tube is usually different from that of the gas inside the compressor and has a different destination, but by careful sizing, the two compressors can become 2090 two stages of compression of the same process.

Utilizarea tuburilor peristaltice poate fi extinsa si la alte tipuri de compresoare. De exemplu, pe peretele interior al statorului compresoarelor cu palete, pot fi montate tuburi peristaltice cu burete termic, tubul fiind presat si gazul fiind comprimat prin miscarea de rotatie a paletelor. Presiunea finala a gazului din tubul peristaltic este, de regula, diferita de cea a gazului din incinta compresorului si are 2095 o destinatie diferita, dar prin dimensionare atenta, cele doua compresoare pot deveni doua trepte de comprimare ale aceluiasi proces.

Another type of compressor in which peristaltic tubes equipped with thermal sponges can be used are scroll compressors, which by this method simplifies the problem of friction and tightness of the compartments. For example, in the screw compressor in Fig.42A, is made the thickness reduction 2100 of fixed blade, on both sides, up to the level 2 Ng+2g _h (where g ₁ is the wall thickness of the peristaltic tube in which a thermal sponge is inserted whose thickness in fully compressed state is 2 Ng), then the two tubes in fully compressed state are fixed on both sides of this spiral, then the movable spiral is inserted, after which the two thermal sponges are released. Under the action of the forces exerted by the elastic elements of the thermal sponges, the space inside the compressor will be occupied, almost 2105 entirely, by the peristaltic tubes. Note that in this configuration, the distance between the two covers can be increased, so that the tips of the two spirals no longer touch the lid of the opposite spiral. The orbital movement of the movable spiral, or the rotation in the same direction of both spirals, has the effect of a peristaltic pressure and leads to the compression of the gas in these tubes.

Rotary compressors are especially useful for high gas flow rates, for low compression ratios. 2110 To obtain higher pressures can be used the pressure step method. They are ideal for supplying pre compressed working gas, at temperature T _iz , for densifiers with high compression ratios. Description of the possibilities of industrial application of the invention

A third object of the invention is to propose new complex installations, made by incorporating 2115 the types of densifiers and rarefiers described in the invention. By using the new installations, due to the increase in the performance of compression and expansion processes, increases the performance of all technologies in which gas and vapor compression and/or expansion have an important share: transport, gas storage and liquefaction, air treatment and conditioning, transformation into useful mechanical work of thermal energy from conventional sources and, in particular, from renewable and 2120 residual sources, storage in tanks with pressurized fluids of energy from these sources, etc.

Regarding the transformation of thermal energy into useful mechanical work, with the help of aas-onlv thermal motors, the use of densifiers and rarefiers described in this invention brings great advantages over the systems used in the prior art. When used, their large heat exchange area offers the possibility of obtaining high speeds of heat transfer between different components of the system, 2125 which allows to obtain high power densities in relation to the volume or weight of installations and offers the possibility to operate between closer temperature limits between hot and cold sources, greatly expanding the range of usable energy sources. Satisfactory yields can be obtained using as a hot source lower intensity solar energy, industrial or household waste energy sources, with temperatures lower than those of the prior art, geothermal sources (if the atmospheric temperature is 2130 negative, the hot source can be the soil at a depth of a few meters, or a groundwater table), etc., and as cold sources, the energy of ambient air, soil, running water, water of lakes and seas, groundwater, etc. Even heat engines or heat pumps can be made, which operate at ideal yields (Carnot) based on the differences in atmospheric temperature between day and night, between atmospheric air and the water of lakes and seas, or even between two nearby locations, different sunny. At a given power, the 2135 volume of these motors and pumps increases as the gap between hot and cold source temperatures decreases.

Due to the large exchange areas of thermal sponges, motors or refrigeration systems operating in Carnot, Stirling and Ericson cycles, direct or reversed, made with these types of isothermalizers allow to achieve high speeds of heat transfer between different system components, 2140 such as and between them and the external environment, which allows to obtain high power densities. A great advantage of motors and heat pumps made with these types of isothermalizers is the possibility to choose and modify, during operation, the two operating temperatures T _iz1 and T _iz2 inside the densifier, respectively the rarefier, by changing the compression ratio of isentropic devices and the simultaneous change of the speed of the drive motors of the isothermalizers, keeping the optimum 2145 temperature difference DT, or adapting it to the change of the temperature of the external environment, or of other operating conditions. This facility offers, for example, the possibility to exhaust, almost completely, the thermal energy from a finished thermal source, as it happens to the energy storage system exemplified in Fig.45.

Isothermalizers are also useful in other engine and refrigeration configurations that have as 2150 their working agent different types of gases, especially atmospheric air, such as internal combustion engines, operating in open circuit (with the elimination of flue gases together with quantities appreciable heat energy), or closed and other systems that operate after a cycle Otto, Diesel, Atkinson, dual, Brayton, Humphrey, Lenoir, etc. In Fig.43 we represented the T-s diagrams of some of these cycles: Brayton (Fof ig.43A), Otto (Fig.43B), Lenoir (Fig.43C). A common feature of these cycles 2155 (and the others, mentioned above) is that the entropy of the working gas varies between the values of s _a and s _m . The introduction into these thermodynamic cycles of an isothermal compression at the minimum temperature T _a (usually equal to atmospheric temperature), to replace those phases of these systems in which the entropy of the working gas varies between the values of s _m and s _a , without modifying the others, offers the opportunity to increase both the power density (moving all gas 2160 developments in the area of high pressures) and their engine efficiency, while reducing, to almost zero, their thermal pollution (now, gases eliminated in open cycles have, at atmospheric pressure, atmospheric temperature), without making any other changes in the construction of the apparatus, except for the change in the expansion ratio in the turbine, or the adiabatic expander which uses the absorbed thermal energy.

2165 In Fig.43D we represented the thermodynamic transformations in T-s diagram of an Otto,

Diesel, or Lenoir engines, but we introduced an additional change and flattened the 3-4 curve of heat absorption (adding to the isochoric or isobaric evolution, an expansion component). In this way, the thermodynamic evolution of the working gas no longer takes place in a set of cylinders (the same evolution in each of them, but with phase shifts that reduce vibrations), but in a succession of devices 2170 (Fig.43E):

- the D _iz densifier, with variable compression ratio, which moves the entire thermodynamic evolution towards the high pressure area and has as a consequence, the considerable reduction of the thermal pollution, with the efficiency increase resulting from this

- the C1 adiabatic compressor, the compression ratio of which can be changed at the operator's 2175 command to increase power

- CC combustion chamber, which requires the use of high-temperature materials provided with a piston, a fuel supply system 15.5 and an ignition system 15.6, with a variable fuel flow rate depending on the power required

- isentropic turbine T which, together with the combustion chamber piston, harnesses the thermal 2180 energy resulting from the combustion of the fuel

The engine can run on a wide range of fuels, liquids, gases, or powdery materials, and can also use fuels with lower calorific value.

Fig.43F shows one of the possible engine configurations. In this configuration, the adiabatic compressor C1 and the combustion chamber CC are placed in the same cylinder (not necessarily with 2185 a circular section), being separated from each other by a drawer 15.7. When the gas in compressor C1 reaches the prescribed pressure and temperature, drawer 15.7 opens. At this point, the piston of the CC combustion chamber located in TDC begins to move, with increasing speed, and the compressor piston slows down, to stop in TDC, at the end of the first portion of the cylinder. Due to these variable speed movements, the volume of gas between the two pistons has some fluctuations, 2190 but reaches the preset value when the compressor piston reaches the end of the stroke and the drawer 15.7 closes. Fuel spraying can start as soon as a quantity of gas has entered the CC, but it will only ignite after the drawer has been closed. At the time of ignition, the CC piston is already high enough speed so that the combustion of the fuel does not occur at a constant volume. Refueling can take place, through sprayers placed in different positions, until the piston reaches the end of the 2195 stroke. Ideally, the fuel dosing can be done in such a way that the expansion of the gas in this chamber is isothermal, at the highest T _iz temperature allowed by the CC, which determines the obtaining of the maximum (ideal) efficiency of the engine.

An important improvement, which can be applied in all the systems proposed in this invention, and which can be implemented with significant results also in the prior art systems, when the cost 2200 calculations allow this, is to reduce the thermal pollution by using a new heating system of thermal insulation of apparatus, devices and tanks used. Fig.44 shows this active thermal insulation process, which can be applied to increase the energy efficiency of all thermal engine systems, heat pumps and thermal energy storage systems. The method involves the arrangement of thermal insulation materials, intended to limit heat loss from the components of these systems, on frames, or other 2205 structures, mounted around the object to be insulated, so as to form successive layers of insulating plates, layers between which flows thermally insulating fluids (preferably liquids). In Fig.44, the solid plates 15.2 are arranged around the insulating object 15.1 in such a way that between the plates are generated buses for the movement of the fluid in the form of a spiral path, having in the center the object 15.1. In addition to this structure, as well as for the insulation of smaller devices, or for the 2210 insulation of flat surfaces, the route is composed of insulated tubes, laid in the same plane, on spiral paths, identical to the layer paths of Fig. 44, and the end 15.4 of the route continues with a derivation to a parallel plane, in which the tubes are laid in the same way, resulting in a succession of layers parallel to the wall to be insulated, the direction of flow being the same in all layers. When the liquid is stationary, each of its microregions has a temperature approximately equal to that of the environment 2215 in which it is located, the insulation is passive (as in the prior art) and on each heat flow tube, successive layers of material form a total thermal resistance constant. When the liquid circulates on the spiral path, from outside to inside, starting with a temperature equal to that of the outside environment, the temperature of each microregion increases from one layer to another. Different portions of the fluid layers, after absorbing thermal energy from the solid layer they passed through, 2220 change their position, reaching areas where the temperature difference between the solid layer and the liquid layer is smaller, which makes the intensity of heat flow to the outside decreases, correlated with the heat flow retained by the heat transfer fluid.

If the thickness and arrangement of the solid and liquid layers is carefully calculated (and experimentally verified), a circulation velocity of the fluid v _ld can be found, at which, on each of the flow 2225 tubes, the temperatures on both sides of the solid layers are close equal, and the heat flow to the outside should be almost zero on as many flow tubes as possible. On the whole, one can get close to the ideal situation in which, of the total thermal energy that would be transferred to the environment when the liquid is stationary, only the thermal energy from the outer solid layer is transferred, whose temperature is even lower by the higher the number of insulation layers, the rest of the heat is taken 2230 up by the liquid, whose temperature rises from the ambient temperature to near the temperature of the insulated object (the difference being the greater the speed of movement of the liquid is greater than its ideal speed v _id ). After its extraction from the system, this liquid can be stored in a tank and, after extracting the thermal energy, reintroduced into the circuit. The extraction of thermal energy from the recovery liquid can be done immediately after its exit from the system, in a liquid-gas heat exchanger. 2235 This thermodynamic transformation of the gas can be an isobaric one, being part of an isothermal- isobaric-adiabatic motor cycle, similar to the streamlined Lenoir cycle, and the mechanical energy obtained can be supplied to an electrical resistance located in the isolated system 15.1 , maintaining its temperature constant. The efficiency of the system increases if a residual heat source is available, whose T _rez temperature is too low to be efficiently operated for other purposes. In this case, the fluid 2240 used in the thermal insulation system as a recuperator passes, before entering the system, through a heat exchanger 15.3, where it absorbs heat energy from this source, its entry into the insulating system being made after a more consistent insulating layer.

The field in which the use of isothermalizers can make significant progress compared to the prior art is the storage of mechanical and thermal energy from renewable energy sources and waste 2245 heat from many industrial processes that still dissipate this energy into the environment, thermally polluting it. In any of the storage systems used in the prior art (D-CAES, A-CAES, l-CAES systems), in the storage phase, a significant fraction of the thermal energy of the working gas, resulting from the transformation of the mechanical energy of the piston during the compression phase is discharged into the environment at its temperature. In the energy recovery phase, the temperature at which the 2250 expansion takes place in the expander is lower than the ambient temperature. Therefore, this storage- recovery cycle occurs with significant loss of exergy, or at very low speeds. In some configurations, these losses are recovered by a significant supply of energy from other sources, usually from fossil fuels. In all these systems, especially those based on quasi-isothermal compression to high gas pressures, the replacement of compressors and expanders, currently used, with isothermal densifiers 2255 and rarefiers, leads to significant increases in energy efficiency and to improving all the parameters of these installations. Moreover, the invention proposes a series of new system configurations, in which the fraction of thermal energy released to the environment, out of the total energy available for storage, can be significantly reduced.

Fig. 45 shows an A-CAES type energy storage system (by adiabatic compression), suitable 2260 for the apparatus described in this invention. As in the prior art systems, the energy available for storage is used for adiabatic compression of the working gas by means of the isentropic compressor C1 (Fig.45B), from the pressure P _a and the temperature T _a , to the pressure P _m and the temperature T _m (curve 1-2 on the T-s diagram, Fig.45A). Unlike prior art systems, the resulting gas is cooled (curve 2- 3 on the T-s diagram, Fig. 45A), with additional mechanical energy consumption, in the HE gas/liquid 2265 heat exchanger. If T _m is high, the gas can be cooled in stages, with the heat transfer fluid change in each stage. The resulting liquid, at a temperature close to T _m , is stored in the tank R2, and the gas at the temperature T _a is stored in the tank R1. When a load peak occurs on the consumption network, the gas in the tank R1 is expanded to the R _iz 1 rarefier (curve 3-1 on the T-s diagram in Fig.45A), the resulting mechanical energy being taken over by the useful task (usually a electricity generator). The 2270 thermal energy, stored in the R2 tank, can be extracted at any time and can receive various uses. Still this energy can be transformed into mechanical energy using the heat engine described in Fig.45, through a process that can be used to extract heat from any finite thermal energy tank. The stored thermal energy is consumed, almost entirely, for the production of mechanical energy, with the help of a heat engine running in a Carnot cycle, consisting of the R _iz 2 rarefier, the T turbine (or an adiabatic 2275 piston expander), the D _iz densifier and the C2 adiabatic compressor. For the complete extraction of this energy, the compression/expansion ratios of the compressor and the turbine, respectively, as well as the duration of the isothermal expansion process in the Ft _iz 2 rarefier must be modified after each cycle. These changes are made by a controller, which receives signals from the pressure transducers inside the rarefier and from the temperature transducers in the rarefier and the storage tank. If the Ft _iz 1 2280 rarefier and the D _iz densifier are placed in the same R3 tank, filled with a heat transfer agent and the R _iz 1 rarefier starts simultaneously with the heat engine, the expansion of the gas in the rarefier is made by absorbing the thermal energy ceded by the D _iz densifier. The working gas pressure in the two isothermalizers is chosen as high as possible (to obtain a high power density), and the expansion and compression ratios can be optimized. With this storage system can be obtained a stored energy 2285 utilization factor close to 100%.

As with prior art storage systems, one of the factors limiting the increase in the amount of stored energy is the working gas temperature at the outlet of the isentropic compressor, the high temperatures requiring expensive materials. The avoidance of high temperatures is made, in the prior art, by a staged compression: in each stage, the working gas is compressed from the temperature T _a 2290 to the temperature T _m , then it is cooled in a heat exchanger to the value of T _a . The resulting gas with temperature T _a is stored in a high pressure tank, and the liquid with temperature T _m will be stored in a large enough tank.

An efficient solution by which high storage pressures can be reached, without the gas exceeding the temperature T _m , is presented in Fig.46 and Fig.47: the gas aspirated from the 2295 atmosphere is compressed adiabatically by the compressor C1 , to reach the temperature T _iz (curve 1- 2 in Fig.46), then is compressed isothermally, with constant T _iz , by the D _iz 1 densifier, to the point with s _iz entropy (curve 2-3 in Fig.46) and again adiabatically, by the compressor C2, for to reach the temperature T _m (curve 3-4 from Fig.46). After cooling in the FIE heat exchanger (curve 4-5 in Fig. 46), the compressed gas, having a temperature close to Ta, is stored in the tank R1 , and the liquid coolant, 2300 having a temperature close to Tm, is stored in the tank R2. Due to the heat absorbed by the thermal sponge and the walls of the densifier, their temperature 7), gradually changes. Therefore, the temperature T _ίz =T _p +DT differs from one compression cycle to another, the difference being equal to this increase in the ambient temperature of the gas. TIT is established still from the design phase, and is a compromise between efficiency and power density and can be modified during the storage phase, 2305 depending on external conditions. After a long standstill of the installation, at start-up, in the first cycle, T,i = T _amb , and if until then the energy source is not exhausted, the process lasts until, in the last cycle, Ti = T _m -DT, where T _m is the maximum temperature admitted. At this time, the temperature of the liquid in the tank R1 is 7 _m -7i7 that of the tank R2. The thermal energy stored in these tanks can be extracted and transformed into mechanical energy by means of two heat engines, equipped with 2310 isothermalizers. In the configuration of Fig.47, the liquid from the tank R2 is moved to the tank R1 , the functionality of the D _iz 1 densifier is reversed becoming the R _iz 2 rarefier, and together with the turbine T, the adiabatic compressor C2 and the D _iz densifier mounted in the tank R3, form a heat engine that harnesses the heat stored in the R2 tank. The mechanical energy stored in the gas pressure in the tank R4 is harnesses with the help of the R _iz 1 rarefier and with the help of the thermal energy released 2315 by the D _iz densifier.

The process described in Fig.48 is a new energy storage process, based just on the high energy efficiency of densifiers and rarefiers with thermal sponge, proposed in this invention. The advantage of the new method compared to those of the prior art is that almost all the thermal energy generated by the action of the piston is stored, with each compression cycle, in the working gas 2320 temperature, in the thermal sponge with its solid and/or liquid components and in thermal tanks, from where it is taken over, almost entirely, in the expansion phase.

We know that a thermodynamic process that takes place in one direction (for example, from cold to hot) and in which it is consumed/produced a certain amount of energy E, will produce/will consume, during the development in inversely direction of the process, an amount of energy the closer 2325 to E, the closer the conditions of the process are to those during the direct process. This is the principle on which the operation of the system described in Fig. 48 is based, a system which, in the energy storage phase, functions as a refrigeration system, storing thermal energy, both positive and negative, and in the energy recovery process, is a heat engine that works between the two heat sources created in the first phase. The result is an extremely efficient and flexible system, useful in a 2330 wide range of storage applications, for a wide range of energy supply processes, with variable energy and dimensional parameters in a very wide range.

The proposed system consists of three distinct operating subsystems, arranged in three stages. Energy storage is performed by a hybrid system that works after an reversed Carnot cycle (consisting of a heat pump and a refrigerator) and consists of the Diz densifier, the T isentropic 2335 expander, the Riz isothermal rarefier and the C isentropic compressor. The mechanical energy to be stored is taken over by the drive system of the D _iz densifier, in order to isothermally compress its working gas, which is initially at an optimum pressure p _M (in other configurations the storage process can start with taking the gas out of the atmosphere, its compression at to pressure p _M , and storage in a tank R with constant pressure). During the compression process, the gas pressure increases to a 2340 predetermined pressure p _f , at a constant temperature during the cycle: T, _z1 =T _ί1 +DT ₁ , where T _/r is the temperature of the thermal sponge, of the walls of the compressor and of the storage agent in the tank R _d , temperature that increases with each cycle. The isothermal rarefier Riz takes over the gas (expanded in the isentropic turbine T) at a temperature T _Iz2 =T _I2 -DT ₂ (T _l2 is the temperature of the walls, of the thermal sponge of the isothermal rarefier Riz and of the storage agent in the tank Rd) and 2345 expand it isothermally. The adiabatic compressor C ensures the transition of the gas from the variable temperature T _iz2 to the variable temperature T _izh by the appropriate modifications of the compression ratio, coordinated by the controller. The four apparatus structure a reversed Carnot cycle, which is the most efficient cycle for two heat sources with temperatures T _I1 +DT ₁ and T _I2 -DT ₂ .

If we take into account the case that the temperature of the thermal sponge and the walls of 2350 the densifier is T _ih in the first cycle of the process the mechanical energy input of the piston materializes in increasing the pressure in the densifier and in a thermal energy addition, which is taken over by the sponge and the walls of the apparatus. In turn, part of the heat taken up by the walls is transferred to the storage agent in the Rd tank. A part of the thermal energy taken over by the thermal sponge has the same destination, if a cooling circuit is installed between the inside and the outside of 2355 the densifier. Therefore, after each compression cycle, most of the gas state quantities and the temperature of the other components involved in the process change, most often with very small values. If we are required to permanently maintain the isothermal nature of the compression and the same DT ₁ (considered the most economical), each cycle will increase the T _lz temperature, inlet and outlet pressures, as well as the temperatures of the thermal sponge, of walls and of the agent in the 2360 tank. Similar phenomena occur in the rarefier, if we keep permanently the isothermal character of the expansion and the same D T ₂ (the most economical temperature difference).

The heat pump will operate in this mode until T _iz1 and/or T _iz2 reach the predetermined limit values, at which point an additional cooling/heating system of the densifier and rarefier is switched to a steady state mode in which both T _iz1 and T _iz2 , as well as the other state quantities do not change.

2365 Starting from the same idea, storage systems can be conceived with configurations in which to process the gas in a Stirling, Ericson, Rankine cycle, or even other cycles, if these cycles can be completed, in the recovery phase, in the sense conversely, with minimal exergy losses. For example, in the case of the Stirling and Ericson cycles, for acceptable exergetic efficiency, an isentropic compressor and an isentropic expander are required to force the gas to perform the (positive, or 2370 negative) temperature jumps D T ₃ and D T ₄ , equal to the temperature difference between the primary and secondary of heat exchangers that recover thermal energy in the isochoric/isobaric processes.

The second stage of the system consists in the two isolated tanks Rd and Rr, in which the Diz densifier and the Riz rarefier are immersed, respectively. Between the walls of the tank and the device in the tank is the material (liquid or solid) that has the role of accumulating thermal energy, positive or 2375 negative, released by the device to the environment. To increase the storage capacity, this material can be, in the initial phase in a solid state (for example, a salt, a paraffin, even a metal, etc.), which melts after the T _iz exceeds the melting temperature. On the other hand, the material in the tank Rr may initially be in the liquid state and solidify during the gas expansion cycles in the rarefier. When they are in the liquid state, these materials can circulate inside the respective isothermal device, contributing to 2380 the efficiency of heat exchanges. In addition to the thermal storage agent initially in the tanks Rd and Rr, quantities of this material are stored in the additional tanks R1 and R3, respectively. When the temperatures T _iz1 and/or T _iz2 have reached the prescribed values, a transport circuit is opened, through which a flow of storage agent, with temperature T is introduced into the tanks Rd and Rr, replacing a similar amount of agent with temperature T _amb , respectively T _h (introducing some temperature 2385 difference between the apparatus and the environment in the tank), quantity which is directed to be stored in the additional tanks R2 and R4 respectively. From this point on, the energy input (positive or negative) of the isothermalizers is consumed to change the temperatures of the new storage agent, from T _amb , to T _z , so that the gas temperatures in the two enclosures remain unchanged.

To limit thermal energy losses (which can be significant in the case of very high or very low 2390 storage temperatures), the thermal insulation of all system components is an active insulation (third stage of the system) of the type described in Fig.44 . In the configuration in the figure, the cooling fluid is a gas, which yields its recovered thermal energy to a liquid agent, to be stored in the tanks R5 and R6, respectively.

The recovery of stored energy is done by reversing the cycles performed during storage.

2395 A great advantage of the system is its flexibility. Note, for example, that the configuration described contains all the components needed for atmospheric gas storage operations, similar to those in prior art CAES systems. The D.iz densifier, together with the compressor C, the expander T and a system of constant pressure tanks make up such a system, with an energy efficiency superior to the classical systems. If the expander T is removed from the circuit, the system supplies compressed 2400 gas at the desired T _iz temperature, the thermal energy of the gas can be stored together with the compressed gas, or it can be extracted in a exchanger and used for various purposes, same as the system described in Fig.45B. Switching from one configuration to another can be done at any time during the storage process. Moreover, in the initial phase, the R.iz rarefier can also change its sense (and role) and participate in the process of storing the compressed gas, at the temperature of T _amb , or 2405 T _iz . Also, the amount of stored thermal energy can be increased (increasing the amount of mechanical energy addressed for consumers), at any time of the process, from any available thermal source (fossil fuels, or biofuels, solar energy, geothermal energy, waste energy). Can also be created configurations in which some of the energy is stored in compressed gas tanks at high pressures, at T _iz temperature, and another part in warehouses where thermal sponges extracted from the densifier or 2410 rarefier are stored.

The storage system in Fig.49 is similar, its main components being the tank Rd, in which the temperature of the gas and of thermal sponge ts1 it contains are kept at the temperature T _iz1 and the tank Rr, in which the temperature of the gas and thermal sponge ts2 which its contents are kept at a T _iz2 temperature as low as possible. For this, they are equipped with an active Rec insulation, which 2415 causes the gas that retains the thermal energy that could be lost through a passive insulation, to transfer this energy to the liquid agent in the tank R1 , respectively R2, or to a heat engine (respectively, a heat pump), which restores, with the help of a small additional energy input, the stationary temperatures in the two tanks. The Diz _p densifier installed in the Rd tank, forms together with the adiabatic compressor C _p , the Tp turbine and the Riz _p expander installed in the Rr tank, a heat 2420 pump. When starting the system, the Diz ₂ densifier, mounted in the tank Rr, draws air from the atmosphere, compresses it isothermally at the temperature T _iz2 and stores it, under constant pressure, in the tank R, at the atmospheric temperature T _atm . The transition from T _atm temperature to T _iz2 temperature and the reverse transition are performed by the isentropic T2 expander and the C2 compressor. Simultaneously with this compression operation, the heat pump also starts. It absorbs the 2425 heat delivered by the densifier to the tank Rr and transfers it to the tank Rd, together with an amount of heat equivalent to the mechanical work performed for this operation. As a result, the fluid in the Rd tank and its thermal sponge changes its temperature with each cycle of the pump, storing the mechanical energy received from a wind turbine, or from another source of mechanical energy. The T _iz1 temperature will rise to the permissible limit, while the T _iz2 temperature will remain unchanged. The 2430 recovery phase of the two forms of stored energy is done by the rarefier with variable isothermal speed Riz1 , together with the adiabatic compressor C1 and the expander D1 , with variable compression ratio, controlled by a regulation system. The temperature in the tank Rd will gradually decrease reaching, with the emptying of the gas in the tank R, a temperature close to T _atm .

A great advantage of isothermalizers starts from the possibility of these devices to store the 2435 absorbed mechanical energy also in the form of thermal energy, an advantage offered both by a very good insulation (in the active system described above) and by inserting the isothermalizer in the liquid tank, to which it yields its surplus heat, thus limiting the storage temperature. The thermal energy thus stored can then be used as such, or it can be transformed, almost entirely, into mechanical energy. The outstanding energy efficiency of these devices and the possibility of obtaining them at a low cost, 2440 offer the possibility to make very simple, small energy storage systems, useful for small applications.

For example, can be made systems consisting of such a densifier (preferably in two pressure stages), operated by a wind turbine, or by the electricity from the public grid, outside the peak load, which sucks in atmospheric air, and after compression, exhausts it back, at constant pressure, into a high-pressure metal tank (to avoid accidents, the tank is buried, or placed in a safety enclosure). Most 2445 often, the need to extract stored energy occurs during periods of peak load and higher atmospheric temperatures than in the storage phase, which is favorable to increased efficiency. The system must also contain hydraulic fluid reservoirs, a pump-hydraulic motor to recover the energy consumed during exhaust, a system for controlling the temperature of the gas in the storage tanks and, possibly, an electric generator to introduce the surplus energy into the electric grid. The stored energy is released 2450 by reversing the operating cycle, whenever needed. For example, for a compression ratio of 1 :350, a storage capacity of 1 m ³ can provide the daily energy requirement for several families and can return power to the grid. For larger consumers (residential buildings, office buildings, etc.) one of the systems described above can be used, systems that can also meet the need for hot water.

Due to the higher speeds at which heat transfers are made, the isothermalizers can be used 2455 successfully in state-of-the-art air conditioning and refrigeration installations. Moreover, with these isothermalizers, new, simpler and more economical configurations can be made, in which no elements of the system outside the enclosure in which it is mounted are required, which can successfully rival even with the refrigeration installations in which they take place phase changes.

For example, the installations proposed in Fig.50 are composed of two loops. In the first loop, 2460 which consists of the D _iz 1 and R _iz 1 isothermalizers, the adiabatic compressor C1 and the adiabatic expander D1 , the air from the enclosure in which they are located is circulated, its component elements constituting a heat engine (Fig.50C), respectively, after reversing the cycle operating, a heat pump (Fig.50D). The choice is made according to the needs: heating, respectively cooling the air in an enclosure. The second loop, composed of the D _iz 2 and R _iz 2 isothermalizers, of the adiabatic 2465 compressor C2 and the adiabatic expander D2, is a heat pump (Fig.50C), respectively a heat engine (Fig.50D), in which a technological gas, or vapors at the saturation limit is circulated. Its role is to provide the second heat source necessary for the operation of the heat engine in the first loop (Fig.50C), or to compensate part of the energy consumption of the heat pump in the first loop (Fig.50B). The working gas in the installations of the second loop must have good heat transfer 2470 properties, and its average pressure must be chosen in the light of economic considerations. The components of the installation operating at close temperatures and exchanging heat with each other are installed in the same tank R1 and R2, respectively, tanks insulated from the outside and filled with a coolant. In each of the tanks, a pump P ensures the transfer of the coolant (respectively the heating liquid) between the two isothermalizers in the tank. All four isothermals are 2475 made according to the recommendations of this invention: they have adjustable actuations, with the speed of the pistons controlled by the controller, so as to ensure the isothermal speeds of movement, are provided with a thermal sponge with very large heat transfer surface and are provided with heat transfer installations based on a heat transfer liquid, and/or a foam generating installation, the role of heat exchanger of this installation being taken over by the isothermalizer of the other loop, located in 2480 the same enclosure. The ideal operation is achieved when the differences DT between the working temperatures of the two isothermalizers located in the same tank are very small, which causes the power differences between them to be very small.

Fig.50A and Fig.50B show the temperature-entropy T-s diagrams of the two installations (for the two operating modes). When the enclosure is to be heated, the air in the enclosure is sucked in by 2485 the D _iz 1 densifier in Fig.50C, at atmospheric pressure P _a and atmospheric temperature T _a and isothermically compressed at this temperature, a quantity of thermal energy equal to the mechanical compression energy being transferred, via the heat transfer fluid to the R _iz 2 rarefier. At the exit of the densifier (and from the tank R2), the air passes through a sterilization installation, composed of the adiabatic compressor C1 and the adiabatic expander D1 , from which it leaves with a temperature T _m , 2490 higher than the atmospheric temperature 7 _a and is sucked by the rarefier R _iz 1 from the tank R1.

Flere the air expands isothermally to the pressure P _a and is exhausted into a mixing installation, where by mixing it with the air from the room a comfortable temperature is obtained for the air released in the room. The temperature 7 _m can be chosen within fairly wide limits, depending on the desired flow rate for the air at the outlet of the mixing device. An increase in this temperature leads to 2495 an increase in the energy consumption of the mixing plant (due to an increase in the flow of cold air circulated) as well as an increase in the energy consumption of the associated heat pump (which circulated the technological gas), but this increase is compensated to a very large extent, by increasing the power of the heat engine that circulated atmospheric air. Flere, the corresponding heat pump operates in a Carnot cycle, between the temperature limits 7 _m and T _a .

2500 When the enclosure needs to be cooled, the air in the enclosure is sucked in by the Diz1 densifier in Fig.50D at atmospheric pressure Pa and atmospheric temperature Ta. Due to the fact that the temperature of the thermal sponge is also equal to Ta, at the beginning of compression the temperature of the gas in the densifier increases by a value DT, which depends on the speed of the piston during this time. Then, the piston speed becomes equal to the isothermal one, and the air is 2505 isothermally compressed at the temperature Ta + DT.

When the enclosure needs to be cooled, the air in the enclosure is sucked in by the D _iz 1 densifier in Fig.50D at atmospheric pressure P _a and atmospheric temperature 7 _a . Due to the fact that the temperature of the thermal sponge is also equal to 7 _a , at the beginning of compression the temperature of the gas in the densifier increases by a value DT, which depends on the speed of the 2510 piston during this time. Then, the piston speed becomes equal to the isothermal one, and the air is isothermally compressed at the temperature T _a +AT. At the exit of the D _iz 1 densifier (and of the R2 tank), the air passes through a sterilization installation, composed of the adiabatic compressor C1 and the adiabatic expander D1 , from which it leaves with a temperature T _m -AT, lower than the atmospheric temperature T _a . From here, the air is sucked in by the Ft _iz 1 rarefier from the R1 tank and 2515 isothermally expanded to the pressure P _a , then it is discharged into a mixing installation, where by mixing it with indoor air, cold air is obtained at a comfortable temperature. And this time, the temperature T _m can be chosen within quite wide limits, depending on the desired flow rate for the air from the outlet of the mixing device. The corresponding heat engine operates in a Carnot cycle, or an equivalent cycle (Stirling, or Ericsson) between the temperature limits T _a and T _m .

2520 Both installations can work very well in the limit situation in which T _a =T _m , in which case an adiabatic compressor and an adiabatic expander are abandoned, the isothermalizers acquires a more or less polytropic character, and the outer loop works at a difference of temperature 2DT.

The sterilization system proposed in this invention can be implemented in any state of the art air conditioning system, local or centralized. It is a thermodynamic system for decontamination, by 2525 heat treatment, of the air intended for respiration, which we will call calcinator. Thermodynamic sterilization destroys pathogens by incineration. These systems are easy to implement and can be extremely efficient (100% efficiency, for a sufficiently high incineration temperature, applied for a sufficiently long period of time), therefore, they should not be missing from any air conditioning system. By this process of thermodynamic sterilization, the temperature of the air intended for respiration (air in 2530 a certain location, at ambient temperature and atmospheric pressure) is raised to the calcination temperature (at which the target pathogens are rendered harmless). The increase in air temperature is achieved by compressing it quasi-adiabatically to the pressure corresponding to this temperature, followed by a period of maintaining the air at this temperature in a regenerator, and by a quasi- adiabatic expansion "in the mirror" until the ambient temperature. A cooling, or heating (electrically or 2535 thermodynamically) of the air in the regenerator, allows its delivery to a predetermined temperature (which may be different from that of the environment). In this invention, both the compressor and the expander are electrically driven positive displacement devices. The new process can be used in many areas, is flexible, allows a wide range of powers and dimensions, an easy adjustment of working flows, pressures and temperatures. These systems can also be made to a small size, so they can be applied 2540 to personal protective equipment as well as portable systems. Another great advantage of this system is that most of the mechanical energy consumed by the compressor is returned to the system by the expander, minimizing the energy consumption required for sterilization. This is a crucial advantage over any other sterilization system, and the flexibility of the system makes it easy to implement in any ventilation, heating/cooling and air conditioning system for breathing air. It is compatible and can be 2545 coupled with most other state-of-the-art sterilization systems.

The design of the calcinator is made taking into account its precise destination: the class (or classes) of microorganisms to be combated. This determines the minimum temperature T _m up to which the air must be heated to obtain the calcination effect and the minimum duration t _m to maintain it, necessary for the complete destruction of this/these classes of pathogens. This minimum 2550 temperature T _m can be exceeded, and a permissible value T _adm can be reached. There is a time duration t _d <t _m in which the air temperature is higher than the minimum temperature T _m , sufficient for a complete calcination. This allows the adoption of a nonstop calcination strategy: the gradual increase of the air temperature up to the value of T _adm , above the value of T _m , followed by an immediate decrease (without pause), below this value, so that the time when the temperature value exceeds the 2555 value of T _m is higher than td value. The configuration of the computer is chosen according to the chosen strategy.

Obtaining the temperature required for calcination is done by quasi-adiabatic compression of the sucked gas from the environment made with a reciprocating compressor, and the temperature necessary for the air to be breathed, through a process of its expansion (combined, if necessary, with 2560 other additional thermal processes). The calcinator can be realized with any type of compressor and expander that can meet the requirements of the chosen operating variant, therefore its choice is made according to the performance of volume, weight, cost, convenience, etc. of the whole ensemble. In air conditioning systems, the most suitable are positive displacement compressors and expanders.

In the system described in Fig.51 A we chose a nonstop strategy. In order to reduce the losses 2565 generated by the valves with a reduced passage section, the cylinder of a compressor/expander can be equipped (Fig.51 A) with a single inlet-outlet orifice (to achieve the largest possible diameter of the access path in the cylinder). In the chosen configuration, the cylinders of both devices are connected to the body of a 4-way valve 15.11 , which has three main positions:

- the open position, noted a in Fig. 51 A, characterized by the creation of wide paths, both for the 2570 admission of atmospheric air and for the evacuation of sterilized air, carried out by the proper movement of the pistons

- the compression position, denoted b in Fig.51 A, in which the piston of the compressor 15.8 performs a compression movement and that of the expander 15.10 is stationary in the TDC,

- the expansion position, denoted c in Fig.51 A, in which the piston of the expander 15.10 performs the 2575 expansion operation, and that of the compressor is stationary in the TDC.

The 4-way valve (electrically or mechanically controlled) is a spherical valve 15.11 in which ball 15.9 the gas passageways are made, by creating cavities that also serve for the storage of compressed air 15.12. The valve ball is actuated by means of a shaft and a camshaft, for the correct synchronization of the operating stages.

2580 This type of valve can also be used successfully in any application described in this invention, when it is desired to create wide paths for gas and liquids circulation, thus a reduction of exergy losses (consequently, an increase in energy consumption). A very useful application is to make a new type of isothermalizer. In Fig.51 B, the valve used is a 3-way valve: one way for the isothermal compressor/expander, one for the inlet pipe and the other for the compressed gas discharge pipe. The 2585 discharge pipe is connected directly to the compressed gas storage tank (constant pressure tank) at its bottom and it is permanently filled with liquid. In the cavities created in the ball of the valve, non- deformable thermal sponges are mounted, for example from interwoven wire nets, which have a large heat absorption surface. Sprinklers can also be installed in the outer walls of the valve to inject coolant as these cavities pass in front of them. In turn, the isothermal cylinder is equipped with a thermally 2590 deformable sponge and cooling systems, whose inlet flow is always equal to the outlet flow, so that the amount of liquid in the cylinder is constant, equal to the amount needed to eliminate the dead volume, when the piston is in TDC.

In its continuous or sequential motion, the valve goes through three main positions:

- the open position, denoted a in Fig.51 B, position in which is opened, through one of the ball cavities, 2595 a path with a large passage section, for the admission of the gas in the cylinder. The other cavity of the ball is filled with liquid, in direct connection with the tank. In this phase, a heat transfer takes place between the liquid and the thermal sponge of this cavity.

- the compression position, denoted b in Fig.51 B, position in which, to the lower cavity of the valve, the communication with the suction pipe is suppressed, the cylinder piston moves towards TDC, the

2600 cooling installation is started and the compressed gas is directed, entirely , to this valve cavity. In the upper cavity, a quantity of liquid with a volume equal to that of the cavity (and of the compressed gas from the other cavity) is displaced to the inlet pipe. In this position, immediately after the connection between the lower cavity and the suction pipe has closed, the liquid sprinklers that introduce the heat transfer agent into the cavity come into operation.

2605 - the exhaust position, denoted c in Fig.51 B, position in which the cylinder piston is in the cavity which in the previous positions was in the lower position, arrives in front of the exhaust pipe and the compressed air is replaced with liquid from the pipes and it moves towards the compressed air tank, and the cavity that in the previous positions was in the upper position, arrives in front of the suction pipe, the liquid from the cavity being evacuated through this pipe. After evacuation, the amount 2610 of liquid introduced through the sprinklers is separated and introduced into the cooling circuit, the rest of the liquid (in an amount equal to the amount of liquid that left the tank, is stored in another tank, to be used when the gas stored in the first tank it is directed, under the same pressure, to a user (which can be this densifier, transformed into a rarefier).

Unlike other installations that store compressed gas in constant pressure tanks, the 2615 isothermalazer piston described above no longer consumes mechanical energy to transfer the compressed gas into the tank, therefore a hydraulic motor is not required in the installation configuration to recover displacement energy. Another advantage of this system is the possibility of gas compression in stages: two isothermal stages, one in the cylinder, the other in the valve cavity and a polytropic stage, performed by the liquid piston in the exhaust pipe. For the final cooling of the gas, 2620 the inlet to the discharge pipe is made through a wire mesh 15.18, which reduces the diameter of the air bubbles formed by the penetration of the liquid, bubbles that are cooled more strongly in the exhaust pipe and in the tank.

Fig.52 shows another configuration that allows the heating or cooling of the air in an enclosure. Compared to the system shown in Fig.50C and Fig.50D, the main loop works in a Stirling 2625 cycle, more advantageous, at least for the small installations, due to the abandonment of the adiabatic compressor and expander, more difficult to operate and adjust in case small temperature differences between hot and cold source, replacing them with a single recuperator (devices that have recently reached high performance). In addition, all valves are removed in this loop. In the secondary loop, through which the atmospheric air circulates, because the adiabatic compressor and the expander 2630 also have the role of sterilizing the air, the configuration described above is kept. In Fig.52, R _iz 2 and D _iz 2 are isothermalizers made according to the invention: they are equipped with thermal sponge 15.15 and sprinklers 15.16 to ensure optimal heat transfer and an actuation system, which in addition to imposing isothermal speed during compression and expansion, ensures a correct correlation between the movements of the two pistons. In a first phase, in the D _iz 2 2635 densifier there is gas at temperature T _a , equal to that of the liquid in which the D _iz 2 densifier is immersed, the D _iz 2 piston is in BDC, and the R _iz 2 piston in TDC, blocking the gas entry in this cylinder. Displacement of the D _iz 2 piston with the isothermal velocity corresponding to the temperature T _a , leads to the isothermal compression of the gas in the densifier cylinder. At a certain moment of movement, when a predetermined volume is reached, the piston of the R _iz 2 rarefier starts at the same speed.

2640 In the next phase, the piston of the D _iz 2 densifier reaches the BDC, where it stops, closing the respective end of the regenerator 15.17. Through this operation, the gas in the densifier pass into the rarefier, keeping the same volume, after changing the thermal energy with the regenerator and reaching the temperature T _m . In the next phase, the piston of the R _iz 2 rarefier continues its motion to BDC, with the isothermal velocity corresponding to the temperature T _m . A new transfer phase follows, 2645 in which both pistons move from one end to the other end of the respective cylinders and in which, when passing through the recuperator, the gas returns to T _a temperature.

For larger installations, the collecting of the gas that is introduced into the densifier, as well as the distribution of the final product are done through piping networks, similar to the networks in the state-of-the-art installations. In these installations it is possible to work, in both loops, with higher 2650 pressures, the result being the processing of higher flows.

Another area in which the use of isothermalizers can bring an increase in the performance of the installations used is that of aas liquefaction. The T-s diagram in Fig.53 explains the new principle of operation, applicable to most gases and gas mixtures (air, natural gas, etc.), regardless of the pressure P _a and the temperature T _a at start-up. The proposed process is similar to the Siemens 2655 process: after passing through a treatment unit 15.20, the gas is compressed isothermally (curve 1-2 in Fig.53), in a D _iz 1 densifier (Fig.54), up to a pressure P ₂ (for high pressures several stages may be preferable, without the need for intermediate heat exchangers). The pressure P ₂ corresponds to an entropy s ₂ , slightly higher than the entropy of the critical point. Then, the gas is released adiabatically (curve 2-3 in Fig.53), in a turbine T, to a pressure below the vapor saturation curve, close to P _a , (in this 2660 area, the pressure P _a is boiling pressure ) and a temperature below the critical point. For reasons of anti cavitation protection, it is recommended that, when leaving the turbine, the liquid concentration be only a few percent. The gas is exhausted in a condenser in which, by extracting the latent heat (curve 3-4 from Fig.53), the gas is completely liquefied.

A proposed configuration for such an installation is shown in Fig.54. The condenser of the 2665 installation is the secondary 15.21 of a plate heat exchanger, through whose primary 15.22 a heat transfer agent circulates, which at the pressure P _a , in the vicinity of the temperature T ₁ is in liquid state. The heat exchange between the two regions is all the more intense, the larger the surface of the partition walls and the smaller the distance between the plates. The primary liquid is conveyed by a 15.27 pump and introduced into another 15.25 tank in which the D _iz 2 rarefier of a heat pump operating 2670 in Carnot mode (curve 2'-5'-4'-3 'in Fig. .53) is mounted. In turn, this heat pump (wich also consists of an adiabatic C2 compressor and T2 turbine, as well as a D _iz 2 densifier) transfer the extracted heat, as well as the mechanical work consumed, to another heat sink, at temperature T _a , or at a different temperature. To control the liquefaction flow, the installation may be provided with an additional system for cooling the gas of condenser, consisting of the D _iz 3 densifier (which uses as coolant even 2675 the liquefied product, or the heat transfer agent from the tank 15.25) and the expansion valve 15.26. This system extracts the warmer gas, from the upper area of the condenser and after a slight isothermal compression expand it isentropically to the pressure P _a (curve 6-3 in Fig.53).

The great advantage of this liquefaction process is that, in the case of gas regeneration, it can be applied in the opposite direction, by going through the same steps, recovering, 2680 in case of small temperature differences between the two paths, most of the energy consumed during liquefaction.