Improving the power of thermodynamic machines

Abstract

The invention relates to a device comprising a first thermodynamic vapour-compression machine having a refrigerant and a second thermodynamic flow machine having a heat-transfer fluid. The first machine comprises a compressor (5), a condenser (2), an expansion valve (6) and an evaporator (7) which together define, with the refrigerant, for the first machine, a thermal power to be exchanged with the heat-transfer fluid. The second machine comprises a fluid circulator (1) configured to feed the heat-transfer fluid into the second machine in thermal contact with the refrigerant. The fluid circulator (1) is dimensioned so as to be able to deliver the heat-transfer fluid at a maximum flow rate greater than one hundred and twenty percent of the nominal flow rate. The first machine comprises a first pipe with a first cross-sectional widening (3) between the compressor (5) and the condenser (2) and a second cross-sectional widening (4) between the condenser (2) and the expansion valve (6).

Description

Improving the power of thermodynamic machines

[0001] This application concerns the field of thermodynamic machines and more particularly the improvement of the heating or cooling power of vapor compression machines. Background

[0002] The prior art knows thermodynamic vapor compression machines, usable on the one hand as a heat pump for heating to take heat from a cold source and transfer it to a hot sink (and therefore the utility is on the sink side) or usable, on the other hand, as a refrigeration machine to take heat from a cold source and transfer it to a hot sink (and therefore the utility is on the source side).

[0003] A vapor-compression thermodynamic machine commonly comprises a compressor, a condenser, an expansion valve, and an evaporator in series in a closed fluid circuit containing a refrigerant, or "primary fluid." Such a machine exchanges heat via its heat exchangers (evaporator and condenser) with a refrigerant circuit (the cold source) and a heat transfer circuit (the hot sink), respectively. Depending on the application, the refrigerant or heat transfer fluid, called the "secondary fluid," can be either in a gaseous state (air or other gases) or in a liquid state (water or other liquids).

[0004] In all cases, the phenomenon of transfer of heat emitted to the hot well or absorbed from the cold source is ensured externally to the thermodynamic machine by a circulation of fluid which is another fluidic circuit which transports a secondary fluid in thermal contact with one of the exchangers of the machine, by means of a fluidic circulator which allows the flow of the secondary fluid.

[0005] When the secondary fluid is in the liquid phase, the fluid circulator is typically a pump and when the fluid is in the gas phase, the fluid circulator is typically a fan.

[0006] The constituent elements of a machine, namely the compressor, the condenser, the expansion valve and the evaporator, together with the refrigerant used, define a useful power or nominal thermal power for the thermodynamic machine. This nominal thermal power is the quantity of heat transmitted to the secondary fluid per unit of time, expressed in kilowatts (kW), under test conditions according to the standards in force or the rules of the art (nominal operating rate). In particular, the EN 14511-2 standard defines the test conditions for determining the characteristics and performance of air conditioners, liquid chillers and heat pumps with an electric motor-driven compressor for space heating and cooling. Generally, the nominal thermal power is part of the technical specifications reported by the manufacturers of thermodynamic machines.

[0007] A nominal flow rate must be ensured by the circulator. It is deduced in a known manner from the nominal thermal power of the thermodynamic machine and a predefined variation (or difference) in temperature or enthalpy of the secondary fluid at the exchanger terminals. More precisely, the nominal flow rate of the secondary fluid is such that the product of the nominal flow rate by the predefined enthalpy variation is equal to the nominal thermal power. The variation in temperature or enthalpy is predefined in the sense that it is set by the standards in force or the rules of the art. In particular, for a heat pump (HP) with water as the secondary fluid, the EN 14511-2 standard sets the variation in water temperature between the inlet and outlet of the exchanger for different types of HP. Thus, the variation in water temperature is set at 5°C for low-temperature HPs, 8°C for medium-temperature HPs and 10°C for high-temperature HPs.Since the enthalpy variation is equal to the product of the specific heat of water (cp = 4.18 kJ / kg.K) by the temperature variation, we can deduce the predefined enthalpy variation imposed by this standard. For a low-temperature heat pump, the temperature variation of 5°C corresponds to an enthalpy variation equal to 5x4.18 = 20.9 kJ / kg. The nominal flow rate is equal to the quotient of the nominal power (at nominal test conditions, defined by the standard) by this enthalpy difference: nominal flow rate = nominal power / 20.9. Generally, the nominal flow rate is part of the technical specifications reported by the manufacturers of thermodynamic machines.

[0008] The fluid circulator is sized to be able to provide such a nominal flow rate of secondary fluid for a characteristic pressure drop of the secondary circuit. Generally, the selected circulator operates within a certain flow rate range and can ensure a maximum flow rate greater than the nominal flow rate selected. However, in practice, this maximum flow rate does not exceed 120% of the nominal flow rate.

[0009] Thus in the prior art: - the flow rate of the liquid secondary fluid in a circulation is ensured by a pump sized for a corresponding nominal flow rate, for a temperature (or enthalpy) difference fixed by the standards or rules of the art at the terminals of the exchanger, at the nominal thermal power of the machine; - the flow rate of the gaseous secondary fluid is ensured by a fan or a system of fans sized for a corresponding nominal flow rate, for a temperature (or enthalpy) difference fixed by the standards or rules of the art at the terminals of the exchanger, at the nominal thermal power of the machine. The fluidic circulator is therefore sized for a nominal thermal power of the thermodynamic machine.

[0010] In conventional devices, the increase in the flow rate of the secondary fluid in contact with the exchangers of the thermodynamic machine, at best maintains the power constant and at worst results in a loss of power. There is therefore nothing to be expected from oversizing the pump of a circuit of the liquid secondary fluid (water or other) of a thermodynamic machine or from oversizing the fan of a circuit of the gaseous secondary fluid (air or other) of a thermodynamic machine in terms of energy efficiency or improvement of the thermal power obtained from a thermodynamic machine. Furthermore, since such oversizing of the fluidic circulator represents an additional cost, greater bulk and overconsumption, it is carefully avoided in the prior art.

[0011] In some prior art, unconventional vapor compression thermodynamic machines, pipe bulges or "inner section bulges of a pipe" or "section bulges" are provided in the fluid circuit containing the primary refrigerant, which improve the thermal efficiency of a vapor compression machine. Examples of such bulges are described, for example, in patent document EP 3071901 B1. These bulges typically comprise, for a given inner section pipe before and after the bulge, an inlet chamber in which the section increases and an outlet chamber in which the section decreases.

[0012] These bulges are typically bulges in which the inlet and outlet chambers are fluidically connected by at least one tube on the fluid circuit of the machine. A bulge that includes more than one tube operates a division of pipe (and refrigerant flow) into sub-pipes, conveniently called tubes, in parallel on the fluid circuit between the inlet and outlet chambers. This type of bulge is conveniently called a "tube bulge". It is understood that a tube bulge does not involve circulation of a secondary fluid unlike a multi-tube heat exchanger.

[0013] The aforementioned bulges are usually arranged in a gaseous or liquid medium, in a so-called high-pressure part of a thermodynamic machine, namely between a compressor and a condenser (arrangement of the bulge in a gaseous medium) or between a condenser and an expander (arrangement of the bulge in a liquid medium). These bulges improve the power and the COP of a vapor compression thermodynamic machine, when they are inserted in such a machine.

[0014] There is a constant need for improvement of thermodynamic machines in terms of energy efficiency or thermal power. General presentation

[0015] In this context, the invention relates to a device comprising a first thermodynamic machine for vapor compression of a refrigerant fluid and a second thermodynamic machine for circulation of a heat transfer fluid, in which the first machine comprises a compressor, a condenser, an expansion valve and an evaporator, defining, together and with the refrigerant fluid, for the first machine, a nominal thermal power, in which the second machine comprises a fluidic circulator configured to deliver the heat transfer fluid in the second machine in thermal contact with the refrigerant fluid, in which the nominal thermal power, exchanged with the second machine, is defined by the product of a nominal flow rate by a predefined enthalpy variation of the heat transfer fluid, in which the fluidic circulator is dimensioned to be able to deliver the heat transfer fluid at a maximum flow rate greater than one hundred and twenty percent of the nominal flow rate,wherein the first machine comprises a first conduit, extending in width according to an internal section and extending in length between the compressor and the expander via the condenser, wherein the first conduit comprises a first section bulge between the compressor and the condenser, wherein the first conduit comprises a second section bulge between the condenser and the expander, wherein the first section bulge comprises a first inlet chamber in which the section increases, a first outlet chamber in which the section decreases and first tubes fluidly connecting the first inlet chamber to the first outlet chamber, wherein the second section bulge comprises a second inlet chamber in which the section increases, a second outlet chamber in which the section decreases and at least one second tube fluidly connecting the second inlet chamber to the second outlet chamber.,

[0016] A "vapor compression thermodynamic machine" is a system operating using a refrigerant fluid which undergoes cyclical transformations (compression then condensation then expansion then evaporation) during which there is an exchange of energy with two thermodynamic machines circulating heat transfer fluid.

[0017] The term "thermodynamic machine with circulation of a heat transfer fluid" means a fluid circuit capable of circulating a heat transfer fluid and transporting heat from a condenser of a vapor compression thermodynamic machine or to an evaporator of a vapor compression thermodynamic machine.

[0018] In the context of the invention, the term "heat transfer fluid" or "secondary fluid" refers to any fluid capable of transporting heat in a thermodynamic machine with circulation of a heat transfer fluid. Thus, these terms refer indistinctly to heat transfer fluids and refrigerant fluids.

[0019] Finally, the words "fluidically connecting" or "fluidically connected", applied to two elements, will be understood as "allowing transport of a fluid" between the two elements, without loss of fluid in the transport and without diversion of fluid between the elements, thus establishing fluidic contact between the two elements.

[0020] Furthermore, since the fluid circulator is sized to be able to flow the heat transfer fluid at a maximum flow rate greater than (i.e. greater than or equal to) one hundred and twenty percent (120%) of the nominal flow rate, i.e. at a maximum flow rate greater than or equal to 1.2 times the nominal flow rate, it is said to be "oversized" compared to a fluid circulator sized to be able to flow the heat transfer fluid at a maximum flow rate strictly less than one hundred and twenty percent of the nominal flow rate. Similarly, a heat transfer fluid flow rate is said to be "oversized" when it is greater than or equal to one hundred and twenty percent (120%) of the nominal flow rate.

[0021] In some embodiments, the device has the following features, alone or in combination, when technically possible:- the second section bulge comprises a single second tube;- the second section bulge comprises several second tubes;- the fluidic circulator is sized to be able to flow the heat transfer fluid at a maximum flow rate greater than one hundred and fifty percent of the nominal flow rate;- the fluidic circulator is sized to be able to flow the heat transfer fluid at a maximum flow rate greater than two hundred percent of the nominal flow rate;- the thermal contact is imposed in the condenser and the fluidic circulator is a pump;- the thermal contact is imposed in the evaporator and the fluidic circulator is a fan;- the thermal contact is imposed in the condenser and the fluidic circulator is a fan;- thermal contact is imposed in the evaporator and the fluidic circulator is a pump.;

[0022] The device of the invention can be used in the following manner: - the first and second machines are operated so as to exchange a real thermal power between the first and second machines, and - the heat transfer fluid is delivered by means of the fluidic circulator at an actual flow rate greater than one hundred and twenty percent of the nominal flow rate.

[0023] The actual thermal power exchanged with the second machine is defined by the product of the actual flow rate by a variation in the actual enthalpy of the heat transfer fluid. Thanks to the bulges of the device and the oversized circulator, the actual thermal power is higher than the nominal thermal power.

[0024] The foregoing and other features and advantages will become apparent from the following detailed description. This detailed description refers to the accompanying drawings.

[0025] This is a schematic diagram of an example of an Air / Water heat pump (PAC). This is a schematic diagram of an example of an Air / Air air conditioner (CLIM). Detailed description

[0026] Particular embodiments of the proposed device are described in detail below, with reference to the examples shown in the accompanying drawings. These embodiments illustrate the characteristics and advantages of the invention. It is however recalled that the invention is not limited to these embodiments.

[0027] This is a schematic diagram of an Air / Water heat pump (PAC), in which the heat pump comprises in series on a first pipe 8: a compressor 5, a first tubed bulge 3, a condenser 2, a second bulge 4, an expansion valve 6, and an evaporator 7. In addition, on a second pipe 9 in thermal contact with the condenser 2 is arranged a fluidic circulator of pump type 1 in series on the second pipe 9.

[0028] The diagram represents the principle of an Air / Air air conditioner (CLIM), in which the air conditioner comprises in series on a first pipe 8: a compressor 5, a first tubed bulge 3, a condenser 2, a second bulge 4, an expansion valve 6, an evaporator 7. In addition, on a second pipe 9 in thermal contact with the evaporator 7 is arranged a fan-type fluid circulator 10, in series on the second pipe 9.

[0029] The condenser and the evaporator are commonly referred to together and throughout the application as the "heat exchangers" of the machine (PAC or CLIM).

[0030] In a first embodiment, with reference to the, the invention is implemented in an Air / Water heat pump, or PAC, which draws energy from a cold source (i.e. outside air) to bring it to a hot well (i.e. water). Its useful use is the exploitation of the energy delivered at the level of the hot well.

[0031] Such a vapor compression machine is constituted in a known manner by the first pipe 8 on which are located the compressor 5, the first tubed bulge 3, the condenser 2, the second bulge 4, the expander 6 and the evaporator 7.

[0032] In the second pipe 9 in thermal contact with the machine at one of its heat exchangers, here the condenser 2, there is a fluid circuit comprising a fluid circulator, here a pump 1, configured to circulate a heat transfer fluid, here water, in the second pipe 9.

[0033] It is known to size the fluid circulator according to a nominal thermal power, here calorific, of the heat pump, according to a calculation known from the prior art and detailed below.

[0034] The amount of energy required for use is determined by the heat pump's heating output and the operating time according to the formula Or : : Energy delivered to the hot well (Water) : Heat pump heating power : Operating time

[0035] To meet the heat requirement, all the components of the heat pump (the compressor, the expansion valve, the evaporator and especially the condenser) are sized to ensure the necessary heating power.

[0036] According to the laws of thermodynamics, and for incompressible fluids such as water, the calorific power is expressed by the formula with : : Heat pump heating power : Flow rate of the heat transfer fluid : Specific heat per unit mass at constant pressure of the heat transfer fluid : Temperature difference of the heat transfer fluid between the inlet and the outlet of the heat pump condenser .

[0037] For a given machine, in all embodiments of the present invention, the heat output is constant for given operating conditions (temperatures of the hot well and the cold source). And since the specific heat of an incompressible heat transfer fluid is also constant, then any variation in the mass flow rate will generate a change in temperature difference in an inversely proportional manner in order to conserve the heat output constant.

[0038] Table 1.1 shows the theoretical values ​​calculated for an example of an Air / Water heat pump with a capacity of 12 kW (or nominal capacity for an outside air temperature of 7°C, degrees Celsius, a condenser inlet water temperature of 30°C and a condenser outlet water temperature of 35°C), for three different water flow rates.Table 1.1 – Theoretical example for the Air / Water heat pumpCaseHeating capacity (kW)Water flow rate (kg / s) (K)1120.55.742120.57531212.874121.51.91

[0039] In this theoretical example, the temperature difference does indeed decrease in a manner inversely proportional to the flow rate.

[0040] Furthermore, it is understood that in the example of an Air / Water heat pump with a nominal power of 12 kW and for a temperature variation of the heat transfer fluid at the condenser terminals of 5 K (according to standard EN 14511-2), the nominal water flow rate is 0.57 kg / s. In the prior art, in this example, the fluidic circulator would be sized to obtain such a nominal flow rate.

[0041] For all experimental results in Tables 1.2 and 1.3, the operating conditions are identical: the outside air temperature is 7°C and the water temperature at the condenser outlet is 35°C.

[0042] Table 1.2 shows the results of real tests carried out with an Air / Water heat pump without bulges, sized at 12 kW nominal power and operating with a refrigerant of type R454C (mass composition: 21.5% R32, 78.5% R1234yf), for two different water flow rates.Table 1.2 – Results of real tests on the Air / Water heat pump of 12 kW nominal powerCaseHeating power (kW)Water flow rate (kg / s) (K)111,40,515,35210,70,942,72

[0043] It is noted that the measured heat output does not increase when the water flow rate increases. There is therefore no incentive for the person skilled in the art to increase the flow rate of heat transfer fluid, or secondary fluid, in a fluid circulator beyond the nominal flow rate (i.e. beyond 0.57 kg / s) corresponding to the nominal power of 12 kW.

[0044] On the other hand, the inventors discovered that, for a heat pump according to the invention operating with bulges and an oversized pump, the heating power increased unexpectedly by increasing the flow rate of the secondary heat transfer fluid, here water. Indeed, although the temperature difference decreases when the flow rate increases, it is observed that the temperature difference does not decrease in a manner inversely proportional to the flow rate: the observed temperature decrease is smaller than the expected decrease.

[0045] Table 1.3 shows the results of tests carried out on the Air / Water heat pump according to the invention, sized at a nominal power of 12 kW operating with an R454C refrigerant with an oversized circulator and bulges, for four different water flow rates.Table 1.3 – Results of real tests on the Air / Water heat pump with bulges and oversized circulator.CaseHeating power (kW)Water flow rate (kg / s) (K)Nominal water flow rate (kg / s)111,240,564,830,57211,70,694,060,57313,331,182,650,57414,031,372,360,57

[0046] It should be noted that, in Table 1.3, cases 2, 3 and 4 correspond to an oversized flow rate, significantly higher than the nominal flow rate (0.57 kg / s), while case 1 corresponds to a flow rate fairly close to the nominal flow rate.

[0047] These results illustrate the fact that, unexpectedly, for heat pumps with bulges, a non-proportional drop in the temperature difference across the heat transfer fluid is obtained for an increase in the flow rate and therefore an increase in the heat pump's heating power. The invention takes advantage of this phenomenon by oversizing the fluid circulator so as to obtain a heat transfer fluid flow rate significantly higher than the nominal flow rate and, thus, a higher heating power.

[0048] In variants of this first embodiment, the heat pump is an air / air, water / air or water / water heat pump.

[0049] In a second embodiment, with reference to the, the invention is implemented in an Air / Air air conditioner (CLIM) which draws energy from a cold source (indoor air) to bring it to a hot sink (outdoor air). Its useful use is the exploitation of the energy taken from the cold source.

[0050] The configuration of the heat engine is the same as in the first embodiment. However, in this second embodiment, the second pipe 9 is in thermal contact with the other heat exchanger of the machine, here the evaporator 7 and a fluid circulator, i.e. a fan 10, adapted to circulate a refrigerant fluid, or secondary fluid, i.e. air, in the second pipe 9 is used.

[0051] It is known to size the fluid circulator according to a nominal thermal power, here refrigeration, of the air conditioner, according to the calculation below.

[0052] The amount of energy required for use is determined by the cooling capacity of the refrigerating machine and the operating time according to the formula Or : : Energy recovered from the cold source (air) : Cooling capacity of the air conditioning : Operating time

[0053] To meet the refrigeration requirement, all the components of the refrigeration machine (the compressor, the expansion valve, the condenser and the evaporator) are sized to ensure the necessary refrigerating capacity.

[0054] According to the laws of thermodynamics, and for compressible fluids such as air, the cooling power is expressed by: , with : : Cooling capacity of the air conditioning : Flow rate of the refrigerant fluid : Difference in enthalpy of the refrigerant fluid between the inlet and the outlet of the evaporator for a constant (kJ / kg).

[0055] For a given machine, in all embodiments of the present invention, the cooling power is constant for given operating conditions (temperatures of the hot well and the cold source). Therefore, the variation of the mass flow rate will generate a change in the enthalpy difference in an inversely proportional manner in order to keep the cooling power constant .

[0056] Table 2.1 shows the theoretical values ​​calculated for an example of an Air / Air air conditioner (AIR) with a nominal capacity of 6.9 kW (or nominal capacity for an outside air temperature of 35°C, degrees Celsius, and an air temperature at the evaporator inlet of 27°C), for five different air flow rates.Table 2.1 – Theoretical example for Air / Air Air Conditioning Case Cooling capacity (kW) Air flow rate (kg / s) (kJ / kg)16,90,417,2526,90,513,836,90,611,546,90,88,6356,91,06.9

[0057] In this theoretical example, the enthalpy difference does indeed decrease in a manner inversely proportional to the flow rate.

[0058] The nominal secondary fluid flow rate is defined by the manufacturer based on the application in which the vapor compression machine is used. Typically, for a residential air conditioning application, the nominal flow rate must ensure a temperature difference across the evaporator of 10 to 15 K, Kelvin. Typically, for a commercial refrigeration application, the nominal flow rate must ensure a temperature difference across the evaporator of 5 to 10 K.

[0059] The air conditioner used for this example is a residential type air conditioner. The temperature difference of 10 to 15 K corresponds to an enthalpy difference of approximately 10 to 15 kJ / kg (cp ~ 1.005 , rounded to 1 ) and therefore at a nominal flow rate of 0.46 to 0.69 kg / s. The technical data sheet of the Air / Air air conditioner used for this example indicates a nominal flow rate of 0.5 kg / s.

[0060] For all experimental results in Tables 2.2 and 2.3, the operating conditions are identical: the outside air temperature is 35° and the air temperature at the evaporator inlet is 27°C.

[0061] Table 2.2 shows the results of actual tests carried out with the aforementioned Air / Air CLIM, without bulge, sized at 6.9 kW nominal power and operating with R410A refrigerant (mass composition: 50%R32, 50%R125), for two different air flow rates.Table 2.2 – Results of actual tests on the Air / Air CLIM with 6.9 kW nominal power Case Cooling capacity (kW) Air flow rate (kg / s) (kJ / kg)16,930,4116,8126,820,5113.32

[0062] It is found that, in accordance with the theory, the measured cooling capacity is constant within the experimental tolerances. There is therefore no incentive for the skilled person to increase the flow rate of refrigerant, or secondary fluid, in a fluid circulator beyond the nominal flow rate (here, 0.5 kg / s according to the technical data sheet of the Air / Air air conditioner used) corresponding to the nominal power of 6.9 kW.

[0063] On the other hand, the inventors discovered that, for an air conditioner according to the invention operating with bulges and an oversized fan, the cooling power increased unexpectedly by increasing the flow rate of the circulating refrigerant, here air. Indeed, although the enthalpy difference decreases when the flow rate increases, it is observed that the enthalpy difference does not decrease in a manner inversely proportional to the flow rate: the observed decrease in enthalpy is smaller than the expected decrease.

[0064] Table 2.3 shows the results of the Air / Air CLIM according to the invention, sized at a nominal power of 6.9 kW operating with an R410A refrigerant with an oversized circulator and bulges, for four different air flow rates.Table 2.3 – Actual test results for the Air / Air CLIM with bulges and oversized fanCaseCooling capacity (kW)Air flow rate (kg / s) (kJ / kg)Nominal air flow (kg / s)17,470,5114,40,528,20,6113,470,539,40,8411,180,5410,40,9910,470,5

[0065] It should be noted that, in Table 2.3, cases 2, 3 and 4 correspond to an oversized flow rate, significantly higher than the nominal flow rate (0.5 kg / s), while case 1 corresponds to a flow rate fairly close to the nominal flow rate.

[0066] Analyzing the results of the refrigeration machine with the invention, a surprisingly non-proportional decrease in the enthalpy difference ΔH on the refrigerant is observed for an increase in flow rate and therefore an increase in refrigerating power. The invention takes advantage of this phenomenon by oversizing the fluid circulator so as to obtain a refrigerant flow rate significantly higher than the nominal flow rate and, thus, a higher refrigerating power.

[0067] In variants of this second embodiment, the air conditioner is an air / water, water / air or water / water refrigeration machine.

[0068] The invention is susceptible to industrial application in the field of thermodynamic machines.

[0069] Throughout this application, the term "oversized" for a circulation or fluidic circulator, will be understood in relation to a dimensioning of a circulation or fluidic circulator for a thermodynamic machine with vapor compression without bulges having a certain nominal power. Typically, the nominal power is reported by the manufacturer of the machine.

[0070] The teaching of the present application also extends in an equivalent manner to a vapor compression machine whose compressor would be undersized compared to its exchangers and to the fluidic circulator or secondary fluid circulator (in which case the circulator would be oversized compared to the thermodynamic vapor compression machine) and which would contain the bulges described previously.

[0071] Thus, according to the invention, it is possible to undersize the compressor of a vapor compression thermodynamic machine relative to the exchangers and the fluid circulator of a fluid circulation for said thermodynamic machine, or to oversize the fluid circulator of a fluid circulation for said thermodynamic machine relative to the compressor and the exchangers, in the presence of bulges. For example: - starting from a machine of known nominal power (generally indicated by the manufacturer of the machine), the fluid circulator will have to be oversized to operate at a fluid flow rate greater than the nominal flow rate.This will allow, after introduction of the bulges, to operate with an oversized flow rate and to obtain a thermal power higher than the nominal power, - starting originally from a machine of given nominal thermal power, with unchanged fluid circulator, the compressor will be undersized to reduce the nominal thermal power of the machine. The fluid flow rate imposed by the unchanged fluid circulator will then be higher than the nominal flow rate calculated with the undersized compressor. This will allow, after introduction of the bulges, to recover the original nominal thermal power with the original fluid circulator but with a less powerful and typically less expensive compressor.

[0072] In the same idea, for a device comprising a thermodynamic machine including a compressor and bulges and another thermodynamic machine including a fluidic circulator, after removing the bulges, a well-dimensioned theoretical compressor and fluidic circulator torque is calculated and then it is checked whether, taking into account the specifications of the compressor and the circulator, they are oversized relatively to each other with respect to the relative dimensioning of the theoretical torque. If so, the machine is in accordance with the invention.

[0073] The teaching of this application extends to other refrigerants than those cited (R410A and R454C), in particular to R407C (mass composition: 50% R134a, 23% R32, 25% R125).

[0074] It is understood for the purposes of this application that a bulge, unless otherwise indicated, may comprise a single tube as well as several tubes.

Claims

Device comprising a first thermodynamic machine for vapor compression of a refrigerant fluid and a second thermodynamic machine for circulation of a heat transfer fluid, in which the first machine comprises a compressor (5), a condenser (2), an expansion valve (6) and an evaporator (7), defining, together and with the refrigerant fluid, for the first machine, a nominal thermal power, in which the second machine comprises a fluidic circulator (1, 10) configured to deliver the heat transfer fluid in the second machine in thermal contact with the refrigerant fluid, in which the nominal thermal power, exchanged with the second machine, is defined by the product of a nominal flow rate by a predefined enthalpy variation of the heat transfer fluid, in which the fluidic circulator (1, 10) is dimensioned to be able to deliver the heat transfer fluid at a maximum flow rate greater than one hundred and twenty percent of the nominal flow rate,wherein the first machine comprises a first duct, extending in width according to an internal section and extending in length between the compressor (5) and the expander (6) via the condenser (2), wherein the first duct comprises a first section bulge (3) between the compressor (5) and the condenser (2), wherein the first duct comprises a second section bulge (4) between the condenser (2) and the expander (6), wherein the first section bulge (3) comprises a first inlet chamber in which the section increases, a first outlet chamber in which the section decreases and first tubes fluidically connecting the first inlet chamber to the first outlet chamber, and wherein the second section bulge (4) comprises a second inlet chamber in which the section increases,a second outlet chamber in which the cross-section decreases and at least one second tube fluidly connecting the second inlet chamber to the second outlet chamber., Device according to claim 1 in which the second section bulge (4) comprises a single second tube. Device according to claim 1 wherein the second section bulge (4) comprises several second tubes. Device according to any one of claims 1 to 3 in which the fluid circulator (1, 10) is sized to be able to deliver the heat transfer fluid at a maximum flow rate greater than one hundred and fifty percent of the nominal flow rate. Device according to claim 4 in which the fluidic circulator (1, 10) is sized to be able to flow the heat transfer fluid at a maximum flow rate greater than two hundred percent of the nominal flow rate. Device according to any one of claims 1 to 5 in which the thermal contact is imposed in the condenser (2) and in which the fluid circulator is a pump (1). Device according to any one of claims 1 to 5 in which the thermal contact is imposed in the evaporator (7) and in which the fluid circulator is a fan (10). Device according to any one of claims 1 to 5 in which the thermal contact is imposed in the condenser (2) and in which the fluid circulator is a fan. Device according to any one of claims 1 to 5 in which the thermal contact is imposed in the evaporator (7) and in which the fluidic circulator is a pump. Use of the device according to any one of claims 1 to 9, in which the first and second machines are operated so as to exchange a real thermal power between the first and second machines, in which the heat transfer fluid is delivered by means of the fluidic circulator at a real flow rate greater than one hundred and twenty percent of the nominal flow rate, in which the real thermal power exchanged with the second machine is defined by the product of the real flow rate by a variation in real enthalpy of the heat transfer fluid, and in which the real thermal power is greater than the nominal thermal power.

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