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

In general , the present invention refers to machines , plants or compression systems , with cascade operation, or with two or more circuits , the heat of the condenser of one circuit is absorbed by the evaporator of the next circuit .

The state of the art is represented by US-A- 5 , 095 , 712 concerning a refrigeration circuit in which the economi zer control is provided together with the variable capacity control , constant cooling is obtained by controlling an economi zer cycle in response to the suction pressure of the compressor, the compressor discharge temperature is controlled by controlling the portion of liquid refrigerant supplied to the interstage line , the refrigeration circuit can be modi fied to include two-stage compressor banks in parallel , the condenser and economi zer are in common .

Furthermore , the state of the art is represented by US-B2- 9 , 816 , 733 relating to a chiller, comprising a condenser, an evaporator, a compressor comprising a first compression stage and a second compress ion stage , a refrigerant duct , the configured refrigerant duct to be in fluid communication with the first compression stage and the second compression stage , an economi zer, wherein the economi zer is configured to form fluid communication with the refrigerant conduit between the first and second compressor stages , the fluid communication is formed through an inj ection port , the inj ection port has an inner surface feature configured to inj ect refrigerant from the economi zer 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 refrigerant running in a direction similar to the flow direction of the refrigerant in the refrigerant duct , and the fluid communication is formed closer to the first compression stage than the second compression stage .

The state of the art does not mention configurations of cascaded thermal cycles compared to the case of a traditional single stage made efficient thanks to a subcooling or desuperheating of the refrigerant fluid leaving the high pressure exchanger of the main circuit and thanks to the internal heat recovery by means of a auxiliary circuit .

Obj ect of the present invention is solving the aforementioned prior art problems by providing a machine with greater ef ficiency in managing 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 compared to a traditional system .

The use of the economi zer allows an improvement in ef ficiency, the greater the higher the compression ratio , the higher the temperature of the fluid and the lower the temperature of the cold source . On the other hand, it does not allow to take advantage of the subcooling induced by the high temperature di f ference since it is already exploited by the economi zer itsel f .

The above and other obj ects and advantages of the invention, as will emerge from the following description, are achieved with a reverse cycle steam compression thermal machine , such as the one described in claim 1 . Preferred embodiments and non-trivial variants of the present invention form the subj ect of the dependent claims .

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

It will be immediately obvious that innumerable variations and modi fications ( 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; is

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 is possible to note that a reverse cycle steam compress ion thermal machine relates to a main circuit 100 connected to an auxiliary circuit 200 , comprising an economi zer 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 7 downstream of and in fluid communication with a second compressor 6 , a second expander 8 downstream of and in fluid communication with the second high pressure exchanger 7 .

The auxiliary circuit 200 is adapted to subcool the main circuit 100 i f the high pressure exchanger 2 is a condenser, or to desuperheat the main circuit i f the high pressure exchanger 2 is a gas chiller .

Advantageously, the common economi zer 3 is configured as a heat exchanger with a first hot branch 1c connected downstream of the first high pressure exchanger 2 , a first cold branch I f connected downstream of the economi zer 3 in common and upstream of the first expander 4 , a second hot branch 2c connected downstream of the economi zer 3 in common and upstream of the second compressor 6 , a second cold branch 2 f connected downstream of the second expander 8 and upstream of the economi zer 3 in common .

The common economi zer 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 subcooling of a saturated liquid that circulates in the first hot branch 1 c in the case in which the high pressure exchanger 2 is a condenser, or, through a more extreme desuperheating of a fluid that circulates in the first hot branch 1c with respect to the initial desuperheating in the case in which the high pressure exchanger 2 is a gas cooler, corresponding to point 3* of the ph diagram of FIG . 2 , and through the entire flow rate of the refrigerant fluid m2 sent to the economi zer 3 in common to receive thermal power from the fluid of the main circuit 100 by means of a complete evaporation o f 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 ph 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 to increase 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 the exchange of thermal power with the environment to al low the cooling of the refrigerant fluid ml , to saturated liquid i f the high pressure exchanger 2 is a condenser, or, with desuperheated gas i f 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 5* of the p-h diagram of FIG . 2 , the entire flow of fluid mixed and 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 means of the second high pressure exchanger 7 , 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 7 works .

The refrigerant fluid m2 reaches a condition of saturated or subcooled liquid i f the second high pressure exchanger 7 is a condenser, or of desuperheated gas i f the second high pressure exchanger 7 is a gas chiller .

The entire flow rate of saturated, subcooled or desuperheated 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 di f ferent 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 suitable for providing the refrigeration output comprises a first 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 (1c) 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 subcooling the main circuit if the high pressure exchanger is a condenser or suitable for desuperheating the main circuit if the high pressure exchanger is a gas chiller.

This thermal machine comprises a second high pressure exchanger (7) downstream of and in fluid communication with a second compressor (6) , the economizer (3) in common, i.e. a common 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 cold branch (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 a single 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 desuperheated steam/gas.

Main circuit operation.

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 ph diagram and subsequently cooled by means of the first high pressure exchanger (2) corresponding to point 2* of the diagram ph. 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 therefore introduced into the high pressure exchanger (2) , which, exchanging thermal power with the environment, allows the fluid to cool, which therefore leads to a saturated liquid condition in the event that the high pressure exchanger (2) is a condenser or a desuperheated 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 economizer (3) in common allows the subcooling of the saturated liquid that circulates in the first hot branch (1c) of said economizer (3) in common in the case in which the high pressure exchanger (2) is a condenser, or a more extreme desuperheating of the fluid circulating in a first hot branch (1c) of said economizer (3) in common (with respect to the initial desuperheating) if the high pressure exchanger (2) is a gas cooler, corresponding to point 3* of the ph 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 desuperheated) 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 ph 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. Auxiliary circuit operation.

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 ph diagram and subsequently cooled by means of the second high pressure exchanger (7) which corresponds to point 7* of the diagram ph.

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 second high pressure exchanger (7) works. The vapor compressed by the second compressor (6) is then introduced into the second high pressure exchanger (7) , which by exchanging thermal power with the environment, allows the cooling of the fluid (m2) , which then leads to a liquid condition saturated or of subcooled liquid in the case in which said second high pressure exchanger (7) is a condenser or with desuperheated gas in the case in which said second high pressure exchanger (7) is a gas cooler. Subsequently the fluid (saturated, subcooled or desuperheated) 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 ph diagram.

From here, the cycle is repeated again as just explained, ie the entire flow of fluid is mixed and compressed completely at the pressure of the second high pressure exchanger (7) to start a new cycle.

In the art there is not, and has never been suggested, a device for the activities/ functions described above equal or similar to the one object of the present invention. The useful effect of the configuration compared to the case of the traditional single stage is therefore further increased, both thanks to the subcooling or desuperheating of the refrigerant fluid leaving the high pressure exchanger of the main circuit and thanks to the internal heat recovery through an auxiliary circuit. In this way, in fact, it is possible to obtain greater efficiency due to an effective management of the distribution of fractionated loads, which uses at most two compression stages in two separate circuits, thus making it less expensive and less complex to manage than a plant, traditional.

The use of the economizer allows an improvement in efficiency, the greater the higher the compression ratio, the higher the temperature of the fluid and the lower the temperature of the cold source. On the other hand, it does not allow to take advantage of the subcooling induced by the high temperature difference since it is already exploited by the economizer itself.

Each expander can assume the configuration of a lamination valve, or alternatively it can assume the configuration of a turbine. In this case, said turbine can be physically/mechanically connected to at least one alternator (i.e. it assumes the configuration of a turbo-alternator) , or alternatively it can be physically/mechanically connected to at least one compressor (i.e. it assumes the configuration of a turbocharger) , suitable for to supply/supply electrical and/or mechanical energy respectively, ensuring the possibility of autonomous adaptation to different load conditions without the aid of external controls .

Thus , the turbine can be either the turbine of a turbocharger ( including an open compressor ) , or the turbine of a turbocharger .

This turbocharger or turbo-alternator can be useful for powering both an internal compressor of the system (which can be added in series or in parallel ) or in any case any internal electronic device , and an engine external to the system ( for example a combustion engine of a car or a vehicle in general ( in particular trains ) in which the air conditioning or heating system is installed) .

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

The theoretical preliminary results obtained showed how the proposed solution potentially of fers both greater energy ef ficiency and an increase in cooling capacity compared to the values that characteri ze the real system with which the data were compared . Once the temperature of the high pressure exchanger ( for example a condenser ) has been set , both the variation in energy ef ficiency and cooling capacity increase as the evaporation temperature decreases and therefore as the gross lift of operating temperatures increases .

Basically, the worse the operating conditions , the greater the savings potential that can be obtained .

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 refrigeration efficiency 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 di f ferent working fluids , showed how the proposed solution allows to obtain an improvement in the Coef ficient of Performance . ( COP ) and the cooling capacity of the plant .

As the gross li ft of the thermodynamic cycle increases , the potential for energy savings achievable with the proposed configuration increases .

Similarly to what has been done in the previous practice , said turbine can be physically/mechanically connected to at least one alternator ( i . e . it assumes the configuration of a turbo-alternator ) , or alternatively it can be physically/mechanically connected to at least one compressor ( i . e . it assumes the configuration of a turbocharger ) .

The intermediate pressure levels at which the expanders work have not been set in advance . Rather, it was decided to veri fy what their optimal values should be , which, together with an appropriate calibration of the fractions , maximi ze 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 an inventive/original 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 ef fect coef ficient of an operator thermodynamic plant , optimi zing the performances at the minimum cost compared to known patent documents of the same sector .

For each application, it will therefore be necessary to appropriately si ze 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 optimi zation 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 proj ect based on theoretical analysis , show how the use of the proposed solution is more appropriate in air conditioning systems or refrigerators of small , medium and large si ze ; at present there are no theoretical limits to a possible use in heat pump operation .