WO2024156698A1

WO2024156698A1 - Improving the power of thermodynamic machines

Info

Publication number: WO2024156698A1
Application number: PCT/EP2024/051527
Authority: WO
Inventors: Joelle Rizk; Bruno DEBOUX; Martin LAUBERT; Marco Cavallari; Georges Nehme
Original assignee: Dynaes
Priority date: 2023-01-23
Filing date: 2024-01-23
Publication date: 2024-08-02
Also published as: FR3145206A1

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

Improvement in the power of thermodynamic machines

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

Background

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 well (and therefore the utility is found from well side) or usable, on the other hand, in a refrigeration machine to take heat from a cold source and transfer it to a hot well (and therefore the utility is on the source side).

A thermodynamic vapor compression machine commonly includes a compressor, a condenser, an expander 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 coolant circuit (the cold source) and a heat transfer circuit (the hot sink), respectively. Depending on the application, the coolant or heat transfer fluid, called "secondary fluid", can be either in the gaseous state (air or other gases) or in the liquid state (water or other liquids).

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

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.

The constituent elements of a machine, namely the compressor, the condenser, the expander and the evaporator define together and with the refrigerant used, 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 kilowatt (kW), under the test conditions following the standards in force or the rules of the art (nominal operating speed). In particular, standard EN 14511-2 defines the test conditions for determining the characteristics and performances of air conditioners, liquid chillers and heat pumps with compressor driven by electric motor for heating and refrigeration of premises.
Generally, the nominal thermal power is one of the technical specifications reported by the manufacturers of thermodynamic machines.

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 terminals of the exchanger. 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 temperature or enthalpy variation is predefined in the sense that it is fixed by the standards in force or the rules of the art. In particular, for a heat pump (PAC) with water as secondary fluid, standard EN 14511-2 sets the variation in water temperature between the inlet and outlet of the exchanger for different types of PAC. Thus, the variation in water temperature is set at 5°C for low temperature heat pumps, at 8°C for medium temperature heat pumps and at 10°C for high temperature heat pumps. The enthalpy variation being 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 the nominal test conditions, defined by the standard) by this enthalpy difference: nominal flow rate = nominal power / 20.9.
Generally, the nominal flow rate is one of the technical specifications indicated by the manufacturers of thermodynamic machines.

The fluidic circulator is sized to be able to provide such a nominal flow rate of secondary fluid for a characteristic pressure loss 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.

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

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

In certain non-conventional thermodynamic vapor compression machines of the prior art, pipe bulges or “inner section bulges of a pipe” or “section bulges” are provided in the fluid circuit containing the primary refrigerant fluid. ", 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 include, for a pipe of given interior section before and after the bulge, an inlet chamber in which the section increases and an outlet chamber in which the section decreases.

These bulges are typically bulges in which the inlet and outlet chambers are fluidly connected by at least one tube on the fluidic circuit of the machine. A bulge that includes more than one tube divides the pipe (and the refrigerant flow) into sub-pipes, conveniently called tubing, in parallel on the fluid circuit between the inlet and outlet chambers. This type of bulge is conveniently called a “tubed bulge.” It is understood that a tubular bulge does not involve circulation of a secondary fluid unlike a multi-tubular heat exchanger.

The aforementioned bulges are usually arranged in a gaseous medium or in a 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 COP of a thermodynamic vapor compression machine, when inserted into such a machine.

There is a constant need for improvement of thermodynamic machines in terms of energy efficiency or thermal power.

General presentation

In this context, the invention relates to a device comprising a first thermodynamic machine for vapor compression of a refrigerant and a second thermodynamic machine for circulation of a heat transfer fluid, in which the first machine comprises a compressor, a condenser, a expander and an evaporator, defining, together and with the refrigerant, for the first machine, a nominal thermal power, in which the second machine comprises a fluidic circulator configured to deliver the heat transfer fluid into the second machine in thermal contact with the refrigerant , 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 according to a maximum flow rate greater than one hundred and twenty percent of the nominal flow rate, in which the first machine comprises a first pipe, extending in width along an interior section and extending in length between the compressor and the expander via the condenser, in which the first pipe comprises a first section bulge between the compressor and the condenser, in which the first pipe comprises a second section bulge between the condenser and the expander, in which 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, in which 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.

The term "thermodynamic vapor compression machine" means a system operating using a refrigerant fluid to which cyclical transformations are made (compression then condensation then expansion then evaporation) during which there is an exchange of energy with two thermodynamic machines. with circulation of heat transfer fluid.

The term “thermodynamic machine with circulation of a heat transfer fluid” means a fluid circuit capable of circulating a heat transfer fluid and making it transport heat coming from a condenser of a thermodynamic machine for vapor compression or destination of an evaporator of a thermodynamic vapor compression machine.

In the context of the invention, the term “heat transfer fluid” or “secondary fluid” means any fluid capable of transporting heat in a thermodynamic machine circulating a heat transfer fluid. Thus, these terms refer indiscriminately to heat transfer fluids and refrigeration fluids.

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 fluid contact between the two elements.

Furthermore, the fluidic circulator being dimensioned to be able to deliver the heat transfer fluid at a maximum flow rate greater (i.e. greater than or equal to) one hundred and twenty percent (120%) of the nominal flow rate, that is to say at a maximum flow rate greater than or equal to 1.2 times the nominal flow rate, it is said to be "oversized" in relation to a fluidic circulator sized to be able to deliver the heat transfer fluid at a maximum flow rate strictly less than one hundred and twenty percent of the nominal flow rate. Likewise, a flow rate of heat transfer fluid is said to be “oversized” when it is greater than or equal to one hundred and twenty percent (120%) of the nominal flow rate.

In certain embodiments, the device has the following characteristics, 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 deliver 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 deliver the heat transfer fluid at a maximum flow rate greater than two hundred percent of the nominal flow rate;
- thermal contact is imposed in the condenser and the fluid circulator is a pump;
- thermal contact is imposed in the evaporator and the fluid circulator is a fan;
- 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.

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

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

The aforementioned features and advantages, as well as others, will become apparent upon reading the detailed description which follows. This detailed description refers to the accompanying drawings.

There is a schematic diagram of an example of an Air/Water heat pump (PAC).
There is a schematic diagram of an example of an Air/Air air conditioner (CLIM).

detailed description

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. However, it is recalled that the invention is not limited to these embodiments.

There 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 bulge with tubing 3, a condenser 2, a second bulge 4, a regulator 6, and an evaporator 7. In addition, on a second pipe 9 in thermal contact with the condenser 2, a pump-type fluid circulator 1 is arranged in series on the second pipe 9.

There represents the principle diagram 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 bulge with tubes 3, a condenser 2, a second bulge 4, a expander 6, an evaporator 7. In addition, on a second pipe 9 in thermal contact with the evaporator 7, a fan-type fluid circulator 10 is arranged, in series on the second pipe 9.

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

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

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

In the second pipe 9 in thermal contact with the machine at the level of 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.

It is known to size the fluidic 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.

The amount of energy required for use is determined by the heat output of the heat pump and the operating time according to the formula Or :
: Energy delivered to the hot well (Water)
: Heating power of the heat pump
: Operating time

To meet the heat requirement, all the components of the heat pump (the compressor, the expander, the evaporator and especially the condenser) are sized to ensure the necessary heat output.

According to the laws of thermodynamics, and for incompressible fluids like water, the heat output is expressed by the formula with :
: Heating power of the heat pump
: Flow rate of the heat transfer fluid
: Specific heat mass at constant pressure of the heat transfer fluid
: Difference in temperature of the heat transfer fluid between the inlet and outlet of the heat pump condenser .

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 mass flow will generate a change in the temperature difference in an inversely proportional manner in order to conserve the calorific power constant.

Table 1.1 indicates the theoretical values calculated for an example of an Air/Water heat pump with a power of 12 kW (or nominal power for an outside air temperature of 7°C, degrees Celsius, a water temperature at condenser inlet of 30°C and a water temperature at the condenser outlet of 35°C), for three different water flow rates.
Table 1.1 – Theoretical example for the Air/Water heat pump

Case	Heating power (kW)	Water flow (kg/s)	(K)
1	12	0.5	5.74
2	12	0.57	5
3	12	1	2.87
4	12	1.5	1.91

In this theoretical example, the temperature difference decreases in a manner inversely proportional to the flow rate.

Furthermore, we understand 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 flow rate nominal water flow 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.

For all the 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.

Table 1.2 shows the results of real tests carried out with an Air/Water heat pump without bulges, sized at 12 kW of 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 12 kW Air/Water heat pump at nominal power

Case	Heating power (kW)	Water flow (kg/s)	(K)
1	11.4	0.51	5.35
2	10.7	0.94	2.72

It can be seen that the measured heat output does not increase when the water flow increases. There is therefore no incentive for those 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 12kW.

On the other hand, the inventors discovered that, for a heat pump according to the invention operating with bulges and an oversized pump, the heat output increased unexpectedly by increasing the circulated flow rate of the secondary heat transfer fluid, here water. Indeed, although the temperature difference decreases when the flow rate increases, we observe 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.

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 a refrigerant of type R454C with an oversized circulator and bulges, for four water flow rates different.
Table 1.3 – Real test results on the Air/Water heat pump with bulges and oversized circulator.

Case	Heating power (kW)	Water flow (kg/s)	(K)	Rated water flow (kg/s)
1	11.24	0.56	4.83	0.57
2	11.7	0.69	4.06	0.57
3	13.33	1.18	2.65	0.57
4	14.03	1.37	2.36	0.57

Note 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 quite close to the nominal flow rate. .

These results illustrate the fact that, unexpectedly, for heat pumps with bulges, we obtain a non-proportional reduction in the temperature difference on the heat transfer fluid for an increase in flow rate and therefore an increase in the heat output of the heat pump. The invention takes advantage of this phenomenon by oversizing the fluidic circulator so as to obtain a flow rate of heat transfer fluid significantly higher than the nominal flow rate and, thus, a higher heat output.

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

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 well (Outdoor air). Its useful use is the exploitation of the energy taken from the cold source.

The configuration of the thermal machine 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 coolant. , or secondary fluid, i.e. air, in the second pipe 9 is used.

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

The required amount of energy 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 power of the CLIM
: Operating time

To meet the refrigeration need, all the components of the refrigeration machine (the compressor, the expander, the condenser and the evaporator) are sized to ensure the necessary refrigeration power.

According to the laws of thermodynamics, and for compressible fluids such as air, the cooling capacity is expressed by: , with :
: Cooling power of the CLIM
: Coolant flow rate
: Difference in enthalpy of the coolant between the inlet and outlet of the evaporator for a constant (kJ/kg).

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). As a result, the variation in mass flow will generate a change in the enthalpy difference in an inversely proportional manner in order to keep the cooling capacity constant .

Table 2.1 indicates the theoretical values calculated for an example of an Air/Air air conditioner (CLIM) with a nominal power of 6.9 kW (or nominal power for an outside air temperature of 35°C, degrees Celsius, and a air temperature at the evaporator inlet of 27°C), for five different air flow rates.
Table 2.1 – Theoretical example for Air/Air CLIM

Case	Cooling power (kW)	Air flow (kg/s)	(kJ/kg)
1	6.9	0.4	17.25
2	6.9	0.5	13.8
3	6.9	0.6	11.5
4	6.9	0.8	8.63
5	6.9	1.0	6.9

In this theoretical example, the enthalpy difference decreases in a manner inversely proportional to the flow rate.

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.

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 sheet of the Air/Air conditioner used for this example indicates a nominal flow rate of 0.5 kg/s.

For all the 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.

Table 2.2 shows the results of real tests carried out with the aforementioned Air/Air CLIM, without bulge, sized at 6.9 kW of nominal power and operating with a refrigerant of type R410A (mass composition: 50%R32, 50%R125 ), for two different air flow rates.
Table 2.2 – Results of real tests on the Air/Air air conditioning with 6.9 kW nominal power

Case	Cooling power (kW)	Air flow (kg/s)	(kJ/kg)
1	6.93	0.41	16.81
2	6.82	0.51	13.32

It can be seen that, in agreement with theory, the measured cooling power is constant within the limit of experimental tolerances. There is therefore no incentive for those skilled in the art to increase the flow rate of coolant, or secondary fluid, in a fluid circulator beyond the nominal flow rate (here, 0.5 kg/s according to the air conditioner technical sheet Air/Air used) corresponding to the nominal power of 6.9 kW.

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 coolant, here air. Indeed, although the enthalpy difference decreases when the flow rate increases, we observe that the enthalpy difference does not decrease in a manner inversely proportional to the flow rate: the observed enthalpy decrease is smaller than the expected decrease.

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 a refrigerant of type R410A with an oversized circulator and bulges, for four different air flow rates .
Table 2.3 – Actual test results for Air/Air A/C with bulges and oversized fan

Case	Cooling power (kW)	Air flow (kg/s)	(kJ/kg)	Rated air flow (kg/s)
1	7.47	0.51	14.4	0.5
2	8.2	0.61	13.47	0.5
3	9.4	0.84	11.18	0.5
4	10.4	0.99	10.47	0.5

Note 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 quite close to the nominal flow rate. .

By analyzing the results of the refrigerating machine with the invention, we observe a surprisingly non-proportional drop in the enthalpy difference ΔH on the refrigerating fluid 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 flow rate of coolant fluid significantly higher than the nominal flow rate and, thus, a higher cooling capacity.

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

The invention is capable of industrial application in the field of thermodynamic machines.

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

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

Thus, according to the invention, it is possible to undersize the compressor of a thermodynamic vapor compression machine in relation to the exchangers and to the fluid circulator of a fluid circulation for said thermodynamic machine, or to oversize the fluid circulator of fluid circulation for said thermodynamic machine relative to the compressor and the exchangers, in the presence of bulges. For example :
- starting from a machine with a known nominal power (generally indicated by the machine manufacturer), the fluid circulator must 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 greater than the nominal power,
- starting originally from a machine with a given nominal thermal power, with an unchanged fluidic circulator, the compressor will be undersized to reduce the nominal thermal power of the machine. The fluid flow rate imposed by the unchanged fluidic circulator will then be greater 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 fluidic circulator but with a less powerful and typically less expensive compressor.

Along the same lines, for a device comprising a thermodynamic machine including a compressor and bulges and another thermodynamic machine including a fluidic circulator, after removing the bulges, we calculate a well-sized theoretical compressor and fluidic circulator couple then we check if, taking into account the specifications of the compressor and the circulator, these are relatively oversized relative to each other with respect to the relative sizing of the theoretical torque. If so, the machine conforms to the invention.

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).

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 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 expander (6) and an evaporator (7), defining, together and with the refrigerant, for the first machine, a nominal thermal power,
in which the second machine comprises a fluid circulator (1,10) configured to deliver the heat transfer fluid into the second machine in thermal contact with the refrigerant,
in which the nominal thermal power, exchanged with the second machine, is defined by the product of a nominal flow rate by a predefined variation in enthalpy 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,
in which the first machine comprises a first pipe, extending in width along an interior section and extending in length between the compressor (5) and the expander (6) via the condenser (2),
in which the first pipe comprises a first bulge of section (3) between the compressor (5) and the condenser (2),
in which the first pipe comprises a second bulge of section (4) between the condenser (2) and the expander (6),
in which 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 fluidly connecting the first inlet chamber to the first chamber of exit, and
in which the second section bulge (4) 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 exit chamber.
Device according to claim 1 in which the second sectional bulge (4) comprises a single second tube.
Device according to claim 1 in which the second sectional bulge (4) comprises several second tubes.
Device according to any one of claims 1 to 3 in which 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 fifty percent of the nominal flow rate.
Device according to claim 4 in which 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 two hundred percent of the nominal flow rate.
Device according to any one of claims 1 to 5 in which 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 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 thermal contact is imposed in the condenser (2) and in which the fluidic circulator is a fan.
Device according to any one of claims 1 to 5 in which thermal contact is imposed in the evaporator (7) and in which the fluid circulator is a pump.
Use of the device according to any one of claims 1 to 9, in which the first and the second machine are operated so as to exchange a real thermal power between the first and the second machine, in which the heat transfer fluid is delivered by means of the fluidic circulator according to 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 real enthalpy variation of the heat transfer fluid, and in which the actual thermal power is greater than the nominal thermal power.