• 专利附录
  • 专利说明
  • 权利要求
  • 类似技术
  • 同族专利
  • 引用文献
  • 法律状态
  • 优先权
    • 专利摘要:
    Refrigeration device for extracting heat from at least one member (15) by exchanging heat with a working fluid circulating in the working circuit (10) comprising in series: a compression mechanism (13) of the fluid a mechanism for cooling the preferably isobaric or substantially isobaric fluid, an expansion mechanism (8) for the fluid and a fluid heating mechanism (9, 6), in which the compression mechanism (13, 4) is of the centrifugal compression type and consists of two compression stages (13, 4) arranged in series in the circuit (10), the device comprising two respective electric drive motors (2, 7) of the two compression stages (13, 4), the an expansion mechanism consisting of a turbine (8) coupled to the motor (2) of one of the compression stages (13, 4), the turbine (8) of the expansion mechanism being coupled to the engine (2) of compression first stage drive
    公开号:FR3072160A1
    申请号:FR1701041
    申请日:2017-10-09
    公开日:2019-04-12
    发明作者:Fabien Durand
    申请人:Air Liquide SA；LAir Liquide SA pour lEtude et lExploitation des Procedes Georges Claude；
    IPC主号:
    专利说明:

    The invention relates to a low temperature refrigeration device and method.
    The invention relates more particularly to a low temperature refrigeration device between -100 ° C and -273 ° C comprising a working circuit containing a working fluid, the device being intended to extract heat from at least one member by heat exchange with the working fluid circulating in the working circuit, the working circuit comprising in series: a compression mechanism of the preferably isentropic or substantially isentropic fluid, a cooling mechanism of the preferably isobaric or substantially isobaric fluid, a fluid expansion mechanism preferably isentropic or substantially isentropic and a fluid heating mechanism preferably isobaric or substantially isobaric, in which the compression mechanism is of the centrifugal compression type and consists of two compression stages respectively a first stage of compression and second compression stage dis placed in series in the circuit, the device comprising two respective electric drive motors of the two compression stages, the expansion mechanism consisting of a turbine coupled to the motor of one of the compression stages.
    The invention relates in particular to so-called “Turbo Brayton” or “Turbo Brayton coolers” cycle refrigerators.
    The invention relates in particular to cryogenic refrigerators, that is to say reaching temperatures less than or equal to -100 ° C or 173K for example in particular between -100 ° C and -273 ° C).
    Document JP3928230B2 describes a Turbo-Brayton type refrigerator using a high speed motor, a turbine and a compressor located respectively at the two ends of its drive shaft.
    To improve the energy efficiency of a refrigerator, one solution consists in using one or more centrifugal compressors with good yields. A centrifugal compressor achieves good performance if its specific speed is equal to or close to the optimal value. The optimal value having been determined experimentally by a person skilled in the art by combining the performance measurements of a multitude of centrifugal compressors having different specific speeds. This is typically 0.75 when calculated with the system of units defined below.
    If this specific speed is higher or lower than the optimal value, the yield is lower. The specific speed ws of a centrifugal compressor is defined by the formula as follows: ws = wQ ° - 5 / Ahs ° · 75
    In which w is the speed of rotation of the compressor in radians per second, Q the volume flow at the inlet of the compressor in m 3 / s and Ahs the increase in enthalpy through the compression stage (in J / kg) considering compression as isentropic.
    A known device is illustrated in Figure 1. A single motor 2 drives a compressor 13 and a turbine 8. The inventors have found that this type of device does not allow the compressor to operate at a good specific speed. Indeed, the low volume flow inherent in this architecture results in a low specific speed compared to the optimal value.
    Another known solution illustrated in FIG. 2 consists in using a second motor 7 with a centrifugal compressor at one of its ends and in placing this machine upstream from the compressor 13 already present.
    Since there is an optimal overall Ahs enthalpy variation with respect to the refrigerator, this is not changed from the state of the art. This new architecture makes it possible to distribute the overall enthalpy Ahs variation over the two compression stages 4, 13 and consequently to decrease the enthalpy Ahs variation of a compression stage and to increase the specific speed of the two stages of compression 4, 13 and to approach or reach the optimal specific speed.
    The inventors have however found that, in practice, this improvement only benefits the first compression stage 4. In fact if the first compression stage 4 operates at the optimum specific speed, the second compression stage 13 will operate at a specific speed typically two times lower than the optimal specific speed. This has an impact on the performance of this stage (typically minus ten efficiency points) and therefore has a strong impact on the overall performance of the refrigerator.
    This can be demonstrated in the following numerical example (where it is considered that the mechanical power and the speed of rotation of the two motors 2, 7 are identical).
    In this example, the mechanical power P2 of the second compression stage 13 is equal to 150% of the mechanical power P1 of the first compression stage 4 due to the presence of the turbine 8 which helps the second motor 2 typically up to 50 % of engine power. Given that the mechanical power of a centrifugal compressor P is equal to the product of the mass flow m by the enthalpy increase Ah (P = m.Ah) and that the mass flow of the two compression stages is identical, then l increase in enthalpy of the second compression stage Ah2 is equal to 150% of the increase in the enthalpy of the first compression stage: Ah2 = 150% Ah1.
    Ah being the increase in actual enthalpy (measured) through the compression stage (in J / kg) (that is to say, the compression is not necessarily isentropic).
    If, moreover, it is considered that the yields of the two compression stages are identical then Ahs2 = 150% Ahs1.
    The volume flow Q2 of the second compression stage 13 is equal to 56% of the volume flow Q1 of the first compression stage since the compression rate is typically 1.8 at the level of the first compression stage 4: Q2 = 56% Q1
    The specific speed ws1 of the first compression stage 4 is equal to w.Q1 0 · 5 / Ahs1 0 · 75
    The specific speed ws2 of the second compression stage 13 is therefore equal to ws2 = w.Q2 ° - 5 / Ahs2 0 · 75 = w. (56% Q1) ° - 5 / (150% Ahs1) 0 · 75 =
    55% .w.Q1 ° - 5 / Ahs1 0 · 75 = 55% .ws1.
    Assuming that the specific speed ws1 of the first compression stage is equal to the optimal specific speed then the ^ second stage operates at 55% of the optimal specific speed.
    This does not optimize the performance of the system.
    An object of the present invention is to overcome all or part of the drawbacks of the prior art noted above.
    To this end, the device according to the invention, moreover conforms to the generic definition given in the preamble above, is essentially characterized in that the turbine of the expansion mechanism is coupled to the drive motor of the first stage compression.
    That is, the device can be a reverse Brayton cycle refrigerator using a two-stage centrifugal compressor that is connected in series and two preferably electric motors to drive the compressors. The low pressure stage (first compressor) and the expansion turbine are mounted on the rotor of a single motor (first motor) and in which the high pressure stage is mounted on the rotor of the second motor.
    Furthermore, embodiments of the invention may include one or more of the following characteristics:
    - the electric drive motor of the first compression stage comprises an output shaft one of whose ends carries and drives in rotation by direct coupling the first compression stage and whose other end carries and is driven in rotation by coupling direct by the turbine,
    - the two motors are identical or similar,
    the cooling mechanism comprises an intermediate cooling exchanger situated between the first compression stage and the second compression stage, for cooling the fluid leaving the first compression stage before it enters the second compression stage,
    - the motors are high-speed motors, that is to say motors whose product of power P in kW by speed N in revolutions per minute squared (PN 2 ) is between 5.10 10 and 5.10 12
    - the rotation speed of the two motors is identical,
    - the mechanical power of the two motors is identical,
    - the drive motor of the second compression stage also mechanically drives an additional circulator or compressor configured to circulate a coolant from the motor or the motors,
    the two compression stages each consist of a centrifugal compressor having a specific optimum speed determined maximizing the energy efficiency of the compressor and in that the device is configured to maintain the specific speed of the compressors between 70% and 130% and preferably between 80% and 120% of the optimal specific speed and even more preferably between 90% and 110% of the optimal specific speed,
    - the two compression stages consist of centrifugal compressors each having a specific optimum speed determined maximizing the energy efficiency of the compressor, each compressor having a determined volume flow and a determined mechanical power and in that the ratio between the volume flow of the first compressor and the volume flow rate of the second compressor is between 1.1 and 2.5 and preferably equal to 1.8 and in that the ratio between the mechanical power driving the first compressor and the mechanical power driving the second compressor is between 1 , 1 and 2.5 preferably equal to 1.5 and in that the ratio of the rotational speeds of the two motors is between 0.5 and 1.5 and preferably equal to one,
    the device comprises an electronic control member of the device and comprising a data storage and processing member, the electronic control member being configured to control in particular one at least of the motors,
    - the working circuit is preferably closed,
    - the two motors are of the electric type,
    - the motors have the same electromagnetic stators and / or the same electromagnetic rotors and / or the same bearings and / or the same cooling systems,
    the cooling mechanism comprises at least one cooling exchanger situated between the second compression stage and the turbine, for cooling the fluid leaving the second compression stage before it enters the turbine,
    - at least one of the cooling exchangers is a counter-current exchanger also thermally exchanging with the working fluid after it leaves the turbine and / or after heat exchange with the member to be cooled
    - the second compression stage drive motor comprises an output shaft which carries and drives in rotation by direct coupling the second compression stage,
    the expansion turbine or turbines are of the centripetal expansion type,
    the engine output shafts are mounted on magnetic type or dynamic gas type bearings, said bearings being used to support the compressors and turbine respectively,
    - the heating mechanism includes a common heat exchanger in which the working fluid travels against the current depending on whether it is cooled or heated,
    the working circuit comprises a reservoir forming a buffer capacity for storing the working fluid,
    the working fluid is in the gas phase and constitutes a pure gas or a mixture of pure gases among: helium, neon, nitrogen, oxygen, argon, carbon monoxide, methane, or any other suitable fluid,
    - the working fluid is subjected in the thermodynamic working cycle circuit (temperature T, entropy S) of the reverse Ericsson type.
    - The invention also relates to a method of cooling a cold source using a refrigeration device in accordance with any one of the characteristics above or below, in which a heat exchange is carried out between the cooled working fluid after leaving the expansion mechanism and the member to be cooled.
    According to other possible particularities:
    the specific speed of the compressors is maintained between 70% and 130% and preferably between 80% and 120% and even more preferably between 90% and 110% of their optimal specific speeds.
    The invention may also relate to any alternative device or process comprising any combination of the above or below characteristics.
    Other particularities and advantages will appear on reading the description below, made with reference to the figures in which:
    FIG. 1 represents a schematic and partial view illustrating the structure and the operation of a refrigeration device according to a first embodiment of the prior art,
    FIG. 2 represents a schematic and partial view illustrating the structure and the operation of a refrigeration device according to a second embodiment of the prior art,
    - Figure 3 shows a schematic and partial view illustrating the structure and operation of a refrigeration device according to a possible embodiment of the invention.
    The low temperature refrigeration device (between -100 ° C and -273 ° C and for example cryogenic) shown in FIG. 3 comprises a closed working circuit 10 containing a working fluid subjected to a thermodynamic cycle during which the fluid reaches a cryogenic temperature. The cooled working fluid is placed in heat exchange with a member or fluid 15 to extract heat from it (for example directly or via a heat exchanger 9).
    The working circuit 10 comprises, arranged in series: a compression mechanism 13, 4 (preferably isentropic or substantially isentropic of the fluid), a cooling mechanism 3, 5, 6 of the fluid (preferably isobaric or substantially isobaric), a fluid expansion mechanism 8 (preferably isentropic or substantially isentropic) and a heating mechanism 9, 6 of the fluid (preferably isobaric or substantially isobaric).
    The compression mechanism is of the centrifugal compression type, i.e. uses (preferably exclusively) compressors of the centrifugal type.
    The compression mechanism is preferably made up of two compression stages 13, 4 (that is to say two compressors) respectively a first compression stage 13 and a second compression stage 4 arranged in series in the circuit 10.
    The device 1 comprises two motors 2, 7, preferably electric, respective drive of the two compression stages 13, 4.
    The expansion mechanism comprises (or is constituted) a turbine 8 (preferably of the centripetal type) driving the motor 2 (coupled to the motor) from one of the compression stages. More precisely, the turbine 8 of the expansion mechanism helps the motor 2 to drive the first compression stage (that is to say the motor 2 for driving the first compressor 13 of the two compressors in series).
    Thus, the device uses two motors 2, 7 and the second motor drives only at one of its ends a second centrifugal compressor 4. This second compressor 4 is located downstream of the first compressor 13 (the downstream refers to the direction of circulation of the working fluid in the circuit 10).
    This new architecture makes it possible to distribute the overall increase in enthalpy Ahs over the two compression stages and consequently makes it possible to reduce the increase in enthalpy Ahs by one stage and to increase the specific speed of the compression stages to approximate the optimum specific speed for each compressor.
    The overall increase in Ahs enthalpy remains unchanged from the state of the art in Figure 2.
    This overall increase in enthalpy Ahs is distributed between the two compression stages 13, 4 which again makes it possible to increase the specific speed of the compression stages and to approach or reach the optimal specific speed.
    Thanks to this new architecture, the two compression stages 13, 4 can operate close to or at the optimum specific speed (and not only the first stage as was the case for the prior art).
    This can be illustrated in the following numerical example where it is considered that the mechanical power and the speed of rotation w of the two motors 2, 7 are identical.
    In this example, the mechanical power P1 of the first compression stage is equal to 150% of the mechanical power P2 of the second compression stage due to the presence of the turbine 8 which helps the first motor 2 typically up to 50% of his power.
    Given that the mechanical power P is equal to the product of the mass flow m by the increase in enthalpy Ah (P = m.Ah), and that the mass flow of the two compression stages is identical, then, the increase d enthalpy Ah1 of the first compression stage is equal to 150% of the enthalpy increase of the second compression stage Ah2, that is to say that Ah1 = 150% Ah2:
    . If, moreover, it is considered that the yields of the two compression stages are identical to then Ahs1 = 150% Ahs2.
    The volume flow Q1 of the first compression stage 13 is equal to 180% of the volume flow Q2 of the second compression stage because the compression rate is typically 1.8 at the level of the first compression stage. That is to say that: Q1 = 180% Q2
    The specific speed ws1 of the first compression stage is given by ws1 = w.Q1 ° - 5 / Ahs1 0 · 75 = w. (180% Q2) ° - 5 / (150% Ahs2) 0 · 75 = 99% .w .Q2 ° - 5 / Ahs2 0 · 75 = 99% .ws2.
    Assuming that the specific speed ws2 of the second compression stage is equal to the optimal specific speed then the first stage operates at 99% of the optimal specific speed.
    That is to say that the specific speeds ws1, ws2 of the first and second compressor 13, 4 (which are identical) are equal to 99% to 100% of the optimal specific speed
    Thus, the architecture according to the invention makes it possible to operate the device so that the two compression stages 13, 4 operate at the optimum specific speed.
    In the example given above, the two motors 2, 7 are identical, the speeds w of the two motors 2, 7 are identical and the specific speeds ws of the two compressors 13, 4 are identical and optimal.
    Of course, the two compression stages 13, 4 can be controlled at different speeds to operate close to or at the optimum specific speed also in the case where the mechanical power and / or the rotational speed of the two motors are distinct.
    The energy efficiency of the refrigeration device is thus improved compared to the prior art.
    The refrigeration device 1 illustrated in FIG. 3 mainly consists of a first compression stage 13 (rotary compressor) whose rotor is driven by the first high speed motor 2. By high speed motor is meant a motor whose product of power P in kW by speed N in revolutions per minute squared (PN 2 ) is greater than 5.10 10 (for example between 5.10 10 and 5.10 12 ). This first high speed motor 2 also receives at the other end of its rotary shaft the regulator 8 (preferably centripetal expansion turbine) which helps the motor 2 to drive the first compression stage 13. The device comprises a second stage of compression 4, the rotor of which is driven by the second high speed motor 7.
    The first compression stage 13 compresses the working fluid (a gas or a mixture of gases) from a low pressure (typically a gas at a pressure of 5 bar abs and a temperature of 15 ° C). The first compression stage 13 transfers the compressed gas via a line 12 of the circuit 10 (for example at a pressure of 9 bars abs and a temperature of 77 ° C). On this line 12 called "medium pressure" can preferably be mounted a cooling exchanger 3 ("intercooler") to evacuate all or part of the compression heat (up to typically 15 ° C for example). The cooling exchanger 3 ensures, for example, a direct or indirect heat exchange with a heat transfer fluid.
    That is to say that, downstream of this cooling exchanger 3, the compression of the working gas can be described as isothermal.
    The second compression stage 4 then compresses the working fluid from the medium pressure (typically 9 bar abs and 15 ° C) and transfers it, via a line 11 (typically at a pressure of 13.5 bar abs and a temperature of 56 ° C). This so-called "high pressure" pipe 11 preferably comprises a heat exchanger 5 ("intercooler") for removing all or part of the heat from the second compression (cooling down to typically 15 ° C. for example). The cooling exchanger 5 provides for example a direct or indirect heat exchange with a heat transfer fluid. That is to say that, downstream of this cooling exchanger 5, the compression of the working gas can be described as isothermal.
    The working fluid is then cooled in an exchanger 6 (for example down to typically -145 ° C). This exchanger 6 can be a countercurrent exchanger, ensuring a heat exchange between the relatively hot working gas at the end of compression and the relatively cold working gas after expansion and heat exchange with the member 15 to be cooled.
    The working fluid is then admitted into the expansion stage (turbine 8) which expands the working fluid from the high pressure (typically 13.5 bar abs and a temperature of -145 ° C) to a low pressure ( typically 5 bars abs and a temperature of -175 ° C). The expanded working fluid is then transferred via a pipe into a heat exchanger 9 used to extract heat from the fluid, for example to cool an object or fluid 15. In this exchanger 9 the fluid rises in temperature (for example up to at typically -145 ° C).
    The working fluid can then be reheated in the countercurrent heat exchanger 6 mentioned above (for example up to typically 15 ° C.).
    The compression ratios of the two compression stages 13, 4 can be chosen so that the specific speed ws of the two compression stages is as close as possible to the optimum value.
    The compression ratios of the compressors 13, 4 can preferably be selected so that the motors 2 and 7 are identical. That is to say that, for example the stator and / or the rotor and / or the bearings of the motors are identical.
    Thus by identical or similar motor we mean strictly identical or different motors but having similar or similar technical characteristics (work provided, etc.), in particular their maximum torques are equal or substantially equal. For example, the working production mechanisms of the motors are identical or have identical or close to 130% performance.
    This standardization of motors 2, 7 also has advantages in terms of maintenance (reduced number of different parts, reduced production cost thanks to the scale effect).
    Although simple and inexpensive in structure, the invention improves the efficiency of refrigeration devices.
    权利要求:
    Claims (14)
    [1" id="c-fr-0001]
    1. Low temperature refrigeration device between -100 ° C and -273 ° C comprising a working circuit (10) containing a working fluid, the device being intended to extract heat from at least one member (15) by heat exchange with the working fluid circulating in the working circuit (10), the working circuit (10) comprising in series: a compression mechanism (13) of the fluid, preferably isentropic or substantially isentropic, a cooling mechanism preferably isobaric or substantially isobaric fluid, an expansion mechanism (8) of the preferably isentropic or substantially isentropic fluid and a heating mechanism (9, 6) of the preferably isobaric or substantially isobaric fluid, in which the compression mechanism ( 13, 4) is of the centrifugal compression type and consists of two compression stages (13, 4) respectively a first compression stage (13) and second compression stage ession (4) arranged in series in the circuit (10), the device comprising two respective electric motors (2, 7) for driving the two compression stages (13, 4), the expansion mechanism consisting of a turbine (8) coupled to the motor (2) of one of the compression stages (13, 4), characterized in that the turbine (8) of the expansion mechanism is coupled to the motor (2) for driving the first stage of compression.
    [2" id="c-fr-0002]
    2. Device according to claim 1, characterized in that the electric motor (2) for driving the first compression stage (13) comprises an output shaft, one of the ends of which carries and rotates by direct coupling the first compression stage (13) and the other end of which carries and is driven in rotation by direct coupling by the turbine (8).
    [3" id="c-fr-0003]
    3. Device according to claim 1 or 2, characterized in that the two motors (2, 7) are identical or similar.
    [4" id="c-fr-0004]
    4. Device according to any one of claims 1 to 3, characterized in that the cooling mechanism comprises an exchanger (3) of intermediate cooling located between the first compression stage (13) and the second compression stage (4) , to cool the fluid leaving the first compression stage (13) before entering the second compression stage (4).
    [5" id="c-fr-0005]
    5. Device according to any one of claims 1 to 4, characterized in that the motors (2, 7) are high-speed motors, that is to say motors whose product of the power P in kW by the speed N in revolutions per minute squared (PN 2 ) is between 5.1Ο 10 and 5.10 12
    [6" id="c-fr-0006]
    6. Device according to any one of claims 1 to 5, characterized in that the speed of rotation of the two motors (2, 7) is identical.
    [7" id="c-fr-0007]
    7. Device according to any one of claims 1 to 6 characterized in that the mechanical power of the two motors (2, 7) is identical.
    [8" id="c-fr-0008]
    8. Device according to any one of claims 1 to 7, characterized in that the motor (7) for driving the second compression stage (4) also mechanically drives a circulator (14) or additional compressor configured to circulate a engine or engine coolant (2, 7).
    [9" id="c-fr-0009]
    9. Device according to any one of claims 1 to 8, characterized in that the two compression stages (13, 4) each consist of a centrifugal compressor having a specific optimum speed determined maximizing the energy efficiency of the compressor and what the device is configured to maintain the specific speed of the compressors (13, 4) between 70% and 130% and preferably between 80% and 120% of the optimal specific speed and even more preferably between 90% and 110% of the optimal specific speed.
    [10" id="c-fr-0010]
    10. Device according to any one of claims 1 to 9, characterized in that the two compression stages (13, 4) consist of centrifugal compressors each having a specific optimum speed determined maximizing the energy efficiency of the compressor, each compressor having a determined volume flow (Q1, Q2) and a determined mechanical power (P1, P2) and in that the ratio (Q1 / Q2) between the volume flow (Q1) of the first compressor (13) and the volume flow (Q2) of the second compressor (4) is between 1.1 and 2.5 and preferably equal to 1.8 and in that ratio (P1 / P2) between the mechanical power (P1) driving the first compressor (13) and the mechanical power (P2) driving the second compressor (4) is between 1.1 and 2.5 preferably equal to 1.5 and in that the ratio (w1 / w2) of the rotational speeds of the two motors is between 0.5 and 1.5 and preferably equal to one.
    [11" id="c-fr-0011]
    11. Device according to claim 10, characterized in that the product: (w1 / w2). (Q1 / Q2) 0 · 5 . (Ahs2 / Ahs1) 075 between:
    - the ratio (w1 / w2) of the rotation speeds of the two motors, the ratio (Q1 / Q2) to the power 0.5 between the volume flow (Q1) of the first compressor (13) and the volume flow (Q2) of the second compressor (4),
    - the ratio (Ahs2 / Ahs1) to the power 0.75 between the increase in enthalpy through the stage of the second compressor (4) considering the compression as being isentropic and the increase in enthalpy through the stage of the first compressor (13) considering the compression as being isentropic, is between 0.70 and 1.30 and preferably between 0.80 and 1.20 and even more preferably between 0.90 and 1.10 .
    [12" id="c-fr-0012]
    12. Device according to any one of claims 1 to 11, characterized in that it comprises an electronic device (16) for controlling the device and comprising a data storage and processing device, the electronic device (16) being configured to control in particular at least one of the motors (13, 4).
    [13" id="c-fr-0013]
    13. A method of refrigerating a cold source (15) using a refrigeration device (1) according to any one of claims 1 to 12, in which a heat exchange is carried out between the working fluid cooled after it leaves of the expansion mechanism (8) and the member (15) to be cooled.
    [14" id="c-fr-0014]
    14. A refrigeration method according to claim 13 characterized in that the specific speed of the compressors (13, 4) is maintained between 70% and 130% and preferably between 80% and 120% and even more preferably between 90% and 110% of their optimal specific speeds.
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    同族专利:
    公开号 | 公开日
    KR20190040123A|2019-04-17|
    WO2019073129A1|2019-04-18|
    AU2018348842A1|2020-05-14|
    EP3695103B1|2021-09-29|
    FR3072160B1|2019-10-04|
    EP3695103A1|2020-08-19|
    DK3695103T3|2021-12-20|
    JP2020537105A|2020-12-17|
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    法律状态:
    2018-10-22| PLFP| Fee payment|Year of fee payment: 2 |
    2019-04-12| PLSC| Publication of the preliminary search report|Effective date: 20190412 |
    2019-10-28| PLFP| Fee payment|Year of fee payment: 3 |
    2020-10-21| PLFP| Fee payment|Year of fee payment: 4 |
    2021-10-21| PLFP| Fee payment|Year of fee payment: 5 |
    优先权:
    申请号 | 申请日 | 专利标题
    FR1701041A|FR3072160B1|2017-10-09|2017-10-09|REFRIGERATION DEVICE AND METHOD|
    FR1701041|2017-10-09|FR1701041A| FR3072160B1|2017-10-09|2017-10-09|REFRIGERATION DEVICE AND METHOD|
    PCT/FR2018/051974| WO2019073129A1|2017-10-09|2018-08-01|Refrigeration device and method|
    EP18758943.7A| EP3695103B1|2017-10-09|2018-08-01|Refrigeration device and method of refrigeration|
    JP2020519342A| JP2020537105A|2017-10-09|2018-08-01|Refrigeration equipment and method|
    CA3079318A| CA3079318A1|2017-10-09|2018-08-01|Refrigeration device and method|
    AU2018348842A| AU2018348842A1|2017-10-09|2018-08-01|Refrigeration device and method|
    US16/755,147| US20200300510A1|2017-10-09|2018-08-01|Refrigeration device and method|
    DK18758943.7T| DK3695103T3|2017-10-09|2018-08-01|DEVICE AND PROCEDURE FOR COOLING|
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