The present invention relates to a reverse cycle steam compression thermal machine.

In general, the present invention relates to machines, plants or compression systems, operating in a booster configuration with two or more circuits, the heat of the condenser or gas cooler of one circuit is absorbed by the evaporator of the next circuit.

The closest state of the art is represented by patent application WO2017179083 concerning a reverse cycle steam compression thermal machine and a method of operation, the machine comprising a condenser downstream and in fluid communication with a main compressor, a valve downstream of the condenser, an evaporator downstream and in fluid communication with the lamination valve, an electric generator, an electric charge accumulator, a timed switch configured according to the hourly intervals of production and consumption of the electricity market, connected downstream of the electric generator and before the accumulator, at least one turbine connected to at least one alternator in fluid communication between the evaporator and the main compressor, and at least one heat exchanger having a hot branch connected downstream of the condenser and before of the lamination valve and a cold branch connected downstream of a second lamination valve and upstream of the turbine. The state of the art is also represented by US patent 5,095,712 A concerning a refrigeration circuit in which the economizer control is provided together with the variable capacity control, constant cooling is obtained by controlling an economizer cycle in response to the pressure of compressor suction, the compressor discharge temperature is controlled by varying the portion of liquid refrigerant supplied to the inter-stage line, the refrigeration circuit can be modified to include two-stage compressor banks in parallel, the condenser and economizer are in common.

Furthermore, the state of the art is represented by US patent 9,816,733 B2 relating to a chiller, comprising a condenser, an evaporator, a compressor comprising a first compression stage and a second compression stage, a refrigerant duct, the configured refrigerant duct to be in fluid communication with the first compression stage and the second compression stage, an economizer, wherein the economizer is configured to form fluid communication with the refrigerant conduit between the first and second compressor stages, fluid communication is formed through an injection port, the injection port has an inner surface feature configured to inject refrigerant from the economizer in a direction of the coolant flow into the coolant duct, the inner surface feature has a curve smooth configured to direct the flow of the r in a direction similar to the direction of flow of the refrigerant in the refrigerant duct, and the fluid communication is formed closer to the first compression stage than the second compression stage.

Finally, the state of the art is represented by the patent WO 2008/130412 which includes a main refrigerant circuit and a closed booster refrigerant circuit. A heat exchanger provides additional cooling for the refrigerant circulating through the main circuit, and thus improves the performance of the refrigerant system. Object of the present invention is solving the aforementioned prior art problems by providing a machine equipped with a higher efficiency in the management of the distribution of fractionated loads, which uses at most two compression stages in two separate circuits, thus resulting less expensive and less complex to manage than a traditional system.

Taking a cue from patent application WO2017179083, a variant has been chosen with respect to the use of a single circuit with a common condenser or gas cooler, using instead two parallel circuits, in order to effectively adapt the thermal loads. These and other objects are achieved in accordance with the present invention by means of a third high pressure exchanger 7b of the auxiliary circuit 200 as a hot source for carrying out a defrosting cycle 11 of the evaporator 5 of the main circuit 100 if the latter is operating in thermo- hygrometric conditions which favor the formation of frost on the surfaces of said evaporator.

This represents an advantage in terms of energy compared to traditional solutions that involve the use of electrical resistances to achieve the same effect.

The present invention is advantageous with respect to a typical traditional solution which provides for defrosting by reversing the cycle of the refrigeration circuit, since, by using the heat rejected by the auxiliary circuit, the interruption of service caused by the fact that the exchanger serving the user (hot) must temporarily become cold during the transitional defrost period.

The aforementioned and other objects and advantages of the invention, which will emerge from the following description, are achieved with a reverse cycle steam compression thermal machine, such as the one claimed in Claim 1.

Preferred embodiments and non-trivial variants of the present invention form the subject of the dependent claims.

It is understood that all attached claims form an integral part of the present description.

It will be immediately obvious that innumerable variations and modifications (for example relating to shape, dimensions, arrangements and parts with equivalent functionality) can be made to what is described, without departing from the scope of the invention as appears from the attached claims.

The present invention will be better described by some preferred embodiments, provided by way of non-limiting example, with reference to the attached drawings, in which:

FIG. 1 shows an operating diagram of an embodiment of the reverse cycle steam compression thermal machine according to the present invention; and

FIG. 2 and FIG. 3 show a graph in the pressure-enthalpy diagram of the operating phases of an embodiment of the reverse cycle steam compression thermal machine according to the present invention.

With reference to the figures, it can be noted that a reverse cycle steam compression thermal machine relates to a main circuit 100 connected to an auxiliary circuit 200, comprising an economizer 3 in common with the main circuit and the auxiliary circuit.

The main circuit 100 comprises a first high pressure exchanger 2 downstream of and in fluid communication with a first compressor 1, an evaporator 5 upstream of the first compressor 1 and downstream of and in fluid communication with a first expander 4. The auxiliary circuit 200 comprises a second high pressure exchanger 7a downstream of and in fluid communication with a second compressor 6 and in parallel with a third high pressure exchanger 7b located near the evaporator 5, a second expander 8 downstream of and in fluid communication with the second high pressure exchanger 7a. The auxiliary circuit 200 is able to sub-cool the main circuit 100 if the high pressure exchanger 2 is a condenser, or to de-superheat the main circuit if the high pressure exchanger 2 is a gas cooler. Advantageously, the common economizer 3 is configured as a heat exchanger with a first hot branch lc connected downstream of the first high pressure exchanger 2, a first cold branch If connected downstream of the economizer 3 in common and upstream of the first expander 4, a second hot branch 2c connected downstream of the economizer 3 in common and upstream of the second compressor 6, a second cold branch 2f connected downstream of the second expander 8 and upstream of the economizer 3 in common. The common economizer 3 allows an internal heat exchange of a first refrigerant fluid ml of the main circuit 100 with a second refrigerant fluid m2 of the auxiliary circuit 200. The entire flow rate of the first refrigerant fluid ml undergoes a sub-cooling of a saturated liquid which circulates in the first hot branch lc in the event that the high pressure exchanger 2 is a condenser, or, through a more extreme de- superheating of a fluid which circulates in the first hot branch lc with respect to the initial de superheating in the case in which the high pressure exchanger 2 is a gas cooler, corresponding to point 3* of the p-h diagram of FIG. 2, and through the entire flow rate of the refrigerant fluid m2 sent to the economizer 3 in common to receive thermal power from the fluid of the main circuit 100 through a complete evaporation of the biphasic mixture up to the condition of saturated or superheated steam which circulates in the second hot branch 2c, corresponding to point 9* of the p-h diagram of FIG. 3.

The entire flow rate of refrigerant fluid ml compressed by the first compressor 1, corresponding to point 1* of the p-h diagram of FIG. 2, and subsequently cooled by means of the first high pressure exchanger 2, corresponding to point 2* of the p-h diagram of FIG. 2, allows increasing the steam pressure to a saturation temperature level higher than the ambient temperature in which the first high pressure exchanger 2 works.

At the working pressure of evaporator 5, the corresponding saturation temperature is lower than the ambient temperature in which evaporator 5 works.

The vapor compressed by the compressor 1 introduced into the high pressure exchanger 2 allows exchanging of thermal power with the environment to allow cooling the refrigerant fluid ml, to saturated liquid if the high pressure exchanger 2 is a condenser, or, with de-superheated gas if the high pressure exchanger 2 is a gas chiller.

The entire flow rate of refrigerant fluid ml sent to the first expander 4 to decrease the pressure, corresponding to point 4* of the p-h diagram of FIG. 2, and introduced into the evaporator 5 allows the complete evaporation of the biphasic mixture up to saturated or superheated steam, corresponding to point 1* of the p-h diagram of FIG. 2, the entire flow of fluid fully compressed to the pressure of the first high pressure exchanger 2 allowing you to start a new cycle. The entire flow rate of refrigerant fluid m2 compressed by the second compressor 6, corresponding to point 6* of the p-h diagram of FIG. 3, and subsequently cooled by the second high pressure exchanger 7a, corresponds to point 7* of the p-h diagram of FIG. 3, allows to increase the steam pressure to a level corresponding to a saturation temperature higher than the ambient temperature in which the second high pressure exchanger 7a works. The refrigerant fluid m2 reaches a condition of saturated or subcooled liquid if the second high pressure exchanger 7a is a condenser, or of de superheated gas if the second high pressure exchanger 7a is a gas chiller. The entire flow rate of saturated, subcooled or de-superheated refrigerant m2, sent to the second expander 8, corresponding to point 8* of the p-h diagram of FIG. 3. At least the first expander 4 and/or the second expander 8 can assume the configuration of a rolling valve.

At least the first expander 4 and/or the second expander 8 can assume the configuration of a turbine connected to at least one alternator / compressor, in order to supply/deliver electrical / mechanical energy.

The heat engine works by automatically adapting to different load conditions without the aid of external controls.

The turbine can be that of a turbocharger / turbo-alternator to power a compressor connected in series or in parallel, any electronic device, a combustion engine, an air conditioning or heating system.

Examples

The reverse cycle steam compression thermal machine with a main circuit and an auxiliary circuit connected in parallel to the main circuit, comprises an economizer in common with the main circuit and the auxiliary circuit, the main circuit capable of providing the refrigeration efficiency high pressure exchanger (2) downstream of and in fluid communication with a first compressor (1), an economizer (3) in common, i.e. a common heat exchanger having in this main circuit a first hot branch (lc) connected downstream of the first high pressure exchanger (2) and before the economizer (3) in common and a first cold branch (If) connected downstream of the economizer (3) in common and upstream of a first expander (4), a low pressure exchanger, i.e. an evaporator (5) upstream of the first compressor (1) and downstream of and in fluid communication with the first expander (4), the auxiliary circuit suitable for sub-cooling the main circuit if the high pressure exchanger is a condenser or suitable for de-superheating the main circuit if the high pressure exchanger is a gas chiller. This thermal machine comprises a second high pressure exchanger (7a) downstream of and in fluid communication with a second compressor (6) and in parallel with a third high pressure exchanger (7b) located near the evaporator (5), the economizer (3) in common, i.e. a heat exchanger having in the auxiliary circuit a second hot branch (2c) connected downstream of the economizer (3) in common and upstream of the second compressor (6), and a second branch cold (2f) connected downstream of a second expander (8) and before the economizer

(3) in common.

The economizer (3) in common is, for example, of the plate or beam type commonly used in the refrigeration field, it can be a heat exchanger with current or counter-current flow.

In an alternative configuration, it is possible to use only one high pressure exchanger common to the main and auxiliary circuit. The refrigerant fluid (ml) of the main circuit

(100) or the refrigerant fluid (m2) of the auxiliary circuit (200) can be in one of the following conditions: liquid, saturated liquid or subcooled liquid; steam/gas, superheated steam/gas or de-superheated steam/gas.

The entire flow rate of refrigerant fluid (ml) of the main circuit (100) is compressed by the first compressor (1) corresponding to point 1* of the p-h diagram and subsequently cooled by means of the first high pressure exchanger (2) corresponding to point 2* of the diagram p-h.

The first compressor (1) therefore increases the vapor pressure to a level such that the corresponding saturation temperature is higher than the ambient temperature in which the first high pressure exchanger (2) works. Similarly, the operating pressure of the evaporator (5) must be such that the corresponding saturation temperature is lower than the ambient temperature in which the aforementioned evaporator (5) works. The vapor compressed by the compressor (1) is then introduced into the high pressure exchanger (2), which by exchanging thermal power with the environment, allows the fluid to cool, which then leads to a saturated liquid condition in the event that the high pressure exchanger (2) is a condenser or a de superheated gas if the high pressure exchanger (2) is a gas cooler. Subsequently, the fluid is sent to the common economizer (3) which by delivering thermal power to the auxiliary circuit (200) by means of said common economizer (3) allows the sub cooling of the saturated liquid that circulates in the first hot branch (lc) of said economizer (3) in common in the case in which the high pressure exchanger (2) is a condenser, or a more extreme de superheating of the fluid circulating in a first hot branch (lc) of said economizer (3) in common (with respect to the initial de-superheating) if the high pressure exchanger (2) is a gas cooler, corresponding to point 3* of the p-h diagram. This process is due to the internal heat exchange of the refrigerant fluid (ml) of the main circuit (100) with the refrigerant (m2) of the auxiliary circuit (200) by means of said economizer (3) in common.

Subsequently, the fluid (subcooled or de superheated) present in the first cold branch (If) of said economizer (3) in common is sent to the first expander (4) to reduce its pressure to a predetermined level, which corresponds to point 4* of the p-h diagram. The outgoing fluid is finally introduced into the evaporator (5) which, receiving thermal power from the environment, allows the complete evaporation of the biphasic mixture up to saturated or superheated steam, corresponding to point 5* of the p-h diagram. From here the cycle is repeated again as just explained, i.e. the entire flow of fluid is mixed and compressed completely at the pressure of the first high pressure exchanger (2) to start a new cycle.

The entire flow rate of refrigerant fluid (m2) of the auxiliary circuit (200) is compressed by the second compressor (6) which corresponds to point 6* of the p-h diagram and subsequently cooled by the second high pressure exchanger (7a) or, alternatively, cooled by the third high pressure exchanger (7b), which corresponds to point 7* of the p-h diagram.

The second compressor (6) therefore increases the vapor pressure to a level such that the corresponding saturation temperature is higher than the ambient temperature in which the high pressure exchanger used works.

The vapor compressed by the second compressor (6) is then introduced into the second high pressure exchanger (7a) or, alternatively, it is introduced into the third high pressure exchanger (7b) located near the evaporator (5) to carry out the cycle. The defrost cycle (11) of said evaporator (5) in case of need, which by exchanging thermal power with the environment, allows the cooling of the fluid (m2), which therefore leads to a condition of saturated liquid or subcooled liquid in the if the high pressure exchanger used is a condenser or de-superheated gas if the high pressure exchanger used is a gas cooler. Subsequently the fluid (saturated, subcooled or de superheated) is sent to the second expander (8) to reduce its pressure to a predetermined level, corresponding to point 8* of the p-h diagram. Subsequently, the fluid (m2) present in a second cold branch (2f) of said economizer (3) in common receives thermal power from the fluid of the main circuit (100) by means of said economizer (3) in common, allowing the complete evaporation of the mixture biphasic up to the condition of saturated or superheated steam present in a second hot branch (2c), which corresponds to point 9* of the p-h diagram. From here the cycle is repeated again as just explained, i.e. the entire flow of fluid is mixed and compressed completely at the pressure of the second high pressure exchanger (7a) to start a new cycle. In the art there is no, and has never been suggested, a device for the activities/functions described above equal or similar to that of the present invention.

The proposed solution can be used in both subcritical and transcritical conditions.

The theoretical preliminary results obtained showed how the proposed solution potentially offers both greater energy efficiency and an increase in cooling capacity compared to the values that characterize the real system with which the data were compared.

The operation of the system with the proposed configuration is economically advantageous compared to the traditional system and it can be easily deduced that by increasing the cooling capacity of the system, a greater economic benefit is obtained. The first results of this comparison, which did not concern only the evaluation of the useful effect compared to the traditional system but also the possible use of different working fluids, showed how the proposed solution allows to obtain an improvement in the Coefficient of Performance (COP) and the cooling capacity of the plant. The intermediate pressure levels at which the expanders work have not been set in advance. Rather, it was decided to verify what their optimal values should be, which, together with an appropriate calibration of the fractions, maximize the yield obtained.

The starting idea, however, was to think with the same service provided.

With reference to the technical characteristics of the invention, the described procedure defines a technical method which, in relation to the methods of implementation of the elements combined with each other, provide a useful and convenient result for the plant, as they are easily defined the distinctive elements, adequate and necessary to improve the useful effect coefficient of an operator thermodynamic plant, optimizing the performances at the minimum cost compared to the known patent documents of the same sector. For each application, it will therefore be necessary to appropriately size the internal exchangers, select the most suitable expanders according to the field of use and carefully evaluate the auxiliary parts and the necessary sensors.

Another aspect to be investigated will concern the optimization of internal pressure levels and fractionations as the temperatures of the thermal sources vary. This would allow the construction of a sort of

"map" of the thermodynamic machine which would allow an appropriate regulation system to guarantee the best performance when the operating conditions vary. The results obtained so far, inherent to an initial phase of the project based on theoretical analysis, show how the use of the proposed solution is appropriate in air conditioning systems or refrigerators of small, medium and large size; at present there are no theoretical limits to a possible use in heat pump operation.

Some preferred forms of implementation of the invention have been described, but of course they are susceptible to further modifications and variations within the same inventive idea.

In particular, numerous variants and modifications, functionally equivalent to the preceding ones, which fall within the scope of the invention as highlighted in the attached claims, will be immediately apparent to those skilled in the art.