Patent Application
20170307259
The application is based on a Japanese Patent Application No. 2014-217455 filed on Oct. 24, 2014, the contents of which are incorporated herein by reference in its entirety.
The present disclosure relates to an ejector refrigeration cycle device that includes an ejector serving as a refrigerant decompression portion.
Conventionally, an ejector refrigeration cycle device is known to be a vapor compression refrigeration cycle device including an ejector as a refrigerant decompression portion.
This kind of ejector refrigeration cycle device can increase the pressure of a suction refrigerant by the pressurizing effect of the ejector, as compared with a normal refrigeration cycle device in which a refrigerant evaporator pressure in an evaporator becomes substantially equal to a pressure of a suction refrigerant drawn into a compressor. Thus, the ejector refrigeration cycle device can reduce the power consumption in the compressor to thereby improve a coefficient of performance (COP) of the cycle.
Furthermore, Patent Document 1 discloses an ejector refrigeration cycle device that includes an ejector with a swirl space for causing swirling flow in a supercooled liquid-phase refrigerant flowing into a nozzle portion (nozzle passage).
In the ejector disclosed in Patent Document 1, a supercooled liquid-phase refrigerant is swirled in the swirl space to thereby decompress and boil the refrigerant on a swirl center side, so that the refrigerant is converted into a two-phase separated state that contains a larger amount of the gas-phase refrigerant on the center side rather than in an outer region of the swirl space. By allowing such a refrigerant in the two-phase separated state to flow into the nozzle passage, boiling of the refrigerant is promoted in the nozzle passage, thereby improving the energy conversion efficiency when converting the pressure energy of the refrigerant into kinetic energy in the nozzle passage.
Meanwhile, based on the studies by the inventors, it is found that the ejector refrigeration cycle device described in Patent Document 1 might cause noise from the ejector at start-up of the compressor. It should be noted that the term “start-up of the compressor” as used herein includes a period of time immediately after start-up of the compressor, at least from when the compressor does not exhibit its refrigerant discharge capacity till when the compressor exhibits a desired target refrigerant discharge capacity.
Thus, the inventors have examined the causes of noise and found that such noise is caused by the fact that, for example, when activating the ejector refrigeration cycle device at a high outside air temperature, the gas-liquid two-phase refrigerant if not cooled sufficiently might flow out of the radiator at start-up of the compressor and enter the ejector.
This is because the ejector disclosed in Patent Document 1 sets a passage cross-sectional area of a refrigerant inflow passage for guiding the refrigerant from the outside of the ejector into the swirl space to a relatively small value to appropriately swirl the supercooled liquid-phase refrigerant flowing into the swirl space.
Thus, if a gas-liquid two-phase refrigerant flows into the refrigerant inflow passage, the gas-liquid two-phase refrigerant circulating through the refrigerant inflow passage flows at a high velocity, compared with when a supercooled liquid-phase refrigerant with a higher density flows thereinto, and causes friction noise when passing through the refrigerant inflow passage. Further, if the friction noise resonates with the gas-phase refrigerant eccentrically located in a columnar shape on the center side of the swirl space, a so-called air column resonance might make significant noise.
The present disclosure has been made in view of the foregoing points, and it is an object of the present disclosure to provide an ejector refrigeration cycle device that includes an ejector having a swirling-flow generating portion to reduce noise generated from an ejector at start-up of the compressor.
An ejector refrigeration cycle device according to a first aspect of the present disclosure includes: a compressor that compresses and discharges a refrigerant; a radiator that dissipates heat from the refrigerant discharged from the compressor; a swirling-flow generating portion that generates a swirling flow in the refrigerant flowing out of the radiator; an ejector including a body portion, the body portion being provided with a nozzle portion that decompresses the refrigerant flowing out of the swirling-flow generating portion, a refrigerant suction port that draws a refrigerant by a suction effect of the injection refrigerant injected from the nozzle portion at a high velocity, and a pressurizing portion that mixes the injection refrigerant with the suction refrigerant drawn from the refrigerant suction port to pressurize the mixed refrigerant; an evaporator that evaporates the refrigerant, and allows the evaporated refrigerant to flow out to the refrigerant suction port; and a discharge-capacity control unit that controls a refrigerant discharge capacity of the compressor.
The swirling-flow generating portion is configured to have a part forming a swirl space in a rotator shape, and a part forming a refrigerant inflow passage through which the refrigerant flows along a peripheral sidewall of the swirl space and flows into the swirl space. Furthermore, the discharge-capacity control unit increases the refrigerant discharge capacity of the compressor in such a manner that an increase amount in the refrigerant discharge capacity of the compressor per predetermined time period is lower than a reference capacity increase amount at start-up of the compressor.
Thus, the discharge-capacity control unit increases the refrigerant discharge capacity in such a manner that an increase amount in the refrigerant discharge capacity of the compressor per predetermined time period is lower than a predetermined reference capacity increase amount at start-up of the compressor. Therefore, even if the gas-liquid two-phase refrigerant flows into the refrigerant inflow passage, the flow velocity of the gas-liquid two-phase refrigerant is prevented from becoming high, and thereby it can reduce the friction noise that would be caused when the gas-liquid two-phase refrigerant circulates through the refrigerant inflow passage.
As a result, the ejector refrigeration cycle device including the ejector with the swirling-flow generating portion can suppress the generation of noise from the ejector at start-up of the compressor. For example, the reference capacity increase amount may adopt the maximum capacity increase amount per predetermined time period enabled by the compressor, that is, the maximum capacity increase amount per predetermined time period that is determined by a capacity inherent to the compressor.
According to a second aspect of the present disclosure, an ejector refrigeration cycle device includes: a compressor that compresses and discharges a refrigerant; a radiator that dissipates heat from the refrigerant discharged from the compressor; a swirling-flow generating portion that generates a swirling flow in the refrigerant flowing out of the radiator; an ejector including a body portion, the body portion being provided with a nozzle portion that decompresses the refrigerant flowing out of the swirling-flow generating portion, a refrigerant suction port that draws a refrigerant by a suction effect of the injection refrigerant injected from the nozzle portion at a high velocity, and a pressurizing portion that mixes the injection refrigerant with the suction refrigerant drawn from the refrigerant suction port to pressurize the mixed refrigerant; an evaporator that evaporates the refrigerant, and allows the evaporated refrigerant to flow out to the refrigerant suction port; and an inflow rate adjustment portion that adjusts an inflow rate of the refrigerant flowing into the swirling-flow generating portion
The swirling-flow generating portion is configured to have a part forming a swirl space in a rotator shape, and a part forming a refrigerant inflow passage through which the refrigerant flows along a peripheral sidewall of the swirl space and flows into the swirl space. Furthermore, the inflow rate adjustment portion increases the refrigerant inflow rate in such a manner that an increase amount in the refrigerant inflow rate per predetermined time period is lower than a reference flow-rate increase amount at start-up of the compressor.
Thus, the inflow rate adjustment portion is adapted to increase the refrigerant inflow rate in such a manner that an increase amount in the refrigerant inflow rate per predetermined time period is lower than the predetermined reference flow-rate increase amount at start-up of the compressor. Therefore, even if the gas-liquid two-phase refrigerant flows into the refrigerant inflow passage, the flow velocity of the gas-liquid two-phase refrigerant is prevented from becoming high, and thereby it can reduce the friction noise that would be caused when the gas-liquid two-phase refrigerant circulates through the refrigerant inflow passage.
As a result, the ejector refrigeration cycle device including the ejector with the swirling-flow generating portion can suppress the generation of noise from the ejector at start-up of the compressor. For example, the reference flow-rate increase amount may adopt the maximum flow-rate increase amount per predetermined time period enabled by the inflow rate adjustment portion.
It should be noted that the term “start-up of the compressor” as used in the present disclosure includes a period of time immediately after start-up of the compressor, at least from when the compressor does not exhibit its refrigerant discharge capacity till when the compressor exhibits a desired target refrigerant discharge capacity. The number of refrigerant inflow passages is not limited to one, but may be plural.
A first embodiment of the present disclosure will be described below with reference to
The ejector refrigeration cycle device 10 forms a subcritical refrigeration cycle in which a high-pressure side refrigerant pressure does not exceed the critical pressure of the refrigerant, using a hydrofluorocarbon (HFC)-based refrigerant (e.g., R134a) as the refrigerant. Obviously, a hydrofluoroolefin (HFO)-based refrigerant (e.g., R1234yf) or the like may also be adopted as the refrigerant. Further, refrigerating machine oil for lubricating a compressor 11 is mixed into the refrigerant, and part of the refrigerating machine oil circulates through the cycle together with the refrigerant.
Among components of the ejector refrigeration device 10, the compressor 11 draws and pressurizes the refrigerant into a high-pressure refrigerant and then discharges the pressurized refrigerant. The compressor 11 is installed in an engine room together with an internal combustion engine (engine) (not shown) for outputting a traveling driving force. The compressor 11 is driven by a rotational driving force output from the engine via a pulley, a belt, etc.
More specifically, in this embodiment, the compressor 11 adopts a variable displacement compressor that can be configured to adjust a refrigerant discharge capacity by changing its discharge displacement. The discharge displacement (refrigerant discharge capacity) of the compressor 11 is controlled by a control current output from a controller 60, to be described later, to a discharge displacement control valve of the compressor 11.
Here, the term “engine room” as used in this embodiment means an exterior space that accommodates the engine and is enclosed by a vehicle body, a firewall 50 to be described later, and the like. The engine room is also called an engine compartment. A discharge port of the compressor 11 is connected to a refrigerant inflow port of a condensing portion 12a of a radiator 12.
The radiator 12 is a heat-dissipation heat exchanger that cools the refrigerant by exchanging heat between a high-pressure refrigerant discharged from the compressor 11 and a vehicle exterior air (outside air) blown by a cooling fan 12d, thereby dissipating heat from the high-pressure refrigerant. The radiator 12 is installed at the front side of the engine room in the vehicle.
More specifically, the radiator 12 in this embodiment is configured as a so-called subcool condenser that includes the condensing portion 12a, a receiver 12b, and a supercooling portion 12c. The condensing portion 12a condenses the refrigerant by exchanging heat between the high-pressure gas-phase refrigerant discharged from the compressor 11 and the outside air blown from the cooling fan 12d, thereby dissipating heat from the high-pressure gas-phase refrigerant. The receiver 12b separates the refrigerant flowing out of the condensing portion 12a into gas and liquid phase refrigerants to store therein an excessive liquid-phase refrigerant. The supercooling portion 12c supercools the liquid-phase refrigerant by exchanging heat between the liquid-phase refrigerant flowing out of the receiver 12b and the outside air blown by the cooling fan 12d.
The cooling fan 12d is an electric blower that has the number of revolutions (blown air volume) controlled by a control voltage output from the controller 60. A refrigerant outflow port of the supercooling portion 12c in the radiator 12 is connected to a refrigerant inflow port 31a of an ejector module 13.
The ejector module 13 functions as a refrigerant decompression portion that decompresses the high-pressure liquid-phase refrigerant in the supercooled state flowing out of the radiator 12, and also as a refrigerant circulation portion (refrigerant transport portion) that draws (transports) the refrigerant flowing out of an evaporator 14, to be described later, by a suction effect of the refrigerant flow injected at a high velocity, thereby circulating the refrigerant.
Furthermore, the ejector module 13 in this embodiment also functions as a gas-liquid separator for separating the decompressed refrigerant into gas and liquid phase refrigerants.
That is, the ejector module 13 in this embodiment is configured as a “gas-liquid separator integrated ejector” or a “gas-liquid separating function-equipped ejector”. In this embodiment, to clarify a difference from an ejector not having a gas-liquid separator (gas-liquid separation space), an integrated (modularized) configuration of the ejector and gas-liquid separator will be hereinafter referred as the “ejector module”.
The ejector module 13 is installed in the engine room, together with the compressor 11 and the radiator 12. Note that the respective up and down arrows in
More specifically, as shown in
Specifically, the refrigerant inflow/outflow ports formed in the body portion 30 include the refrigerant inflow port 31a, a refrigerant suction port 31b, a liquid-phase refrigerant outflow port 31c, and a gas-phase refrigerant outflow port 31d. The refrigerant inflow port 31a allows the refrigerant exiting the radiator 12 to flow thereinto. A refrigerant suction port 31b draws the refrigerant flowing out of the evaporator 14. The liquid-phase refrigerant outflow port 31c allows the liquid-phase refrigerant separated by a gas-liquid separation space 30f formed in the body portion 30 to flow out to the refrigerant inlet side of the evaporator 14. The gas-phase refrigerant outflow port 31d allows the gas-phase refrigerant separated in the gas-liquid separation space 30f to flow out to the suction side of the compressor 11.
The internal spaces formed in the body portion 30 include a swirl space 30a, a decompression space 30b, a pressurizing space 30e, and the gas-liquid separation space 30f. The swirl space 30a serves to swirl the refrigerant flowing thereinto from the refrigerant inflow port 31a. The decompression space 30b serves to decompress the refrigerant flowing out of the swirl space 30a. The pressurizing space 30e serves to allow the refrigerant exiting the decompression space 30b to flow thereinto. The gas-liquid separation space 30f serves to separate the refrigerant flowing out of the pressurizing space 30e into gas and liquid phases.
Each of the swirl space 30a and the gas-liquid separation space 30f is formed to have a substantially columnar rotator shape. Each of the decompression space 30b and the pressurizing space 30e is formed as a substantially conical trapezoidal rotator shape that gradually enlarges its diameter from the swirl space 30a side toward the gas-liquid separation space 30f side. All the central axes of these spaces are arranged coaxially. Note that the rotator shape is a tridimensional shape formed by rotating a plane figure about one straight line (central axis) located on the same plane.
A suction passage 13b is formed in the body portion 30 so as to guide the refrigerant drawn from the refrigerant suction port 31b toward the downstream side of the refrigerant flow in the decompression space 30b and the upstream side of the refrigerant flow in the pressurizing space 30e.
A refrigerant inflow passage 31e that connects the refrigerant inflow port 31a to the swirl space 30a extends in the tangential direction of an inner wall surface of the swirl space 30a as viewed from the central axis direction of the swirl space 30a. Thus, the refrigerant flowing from the refrigerant inflow passage 31e into the swirl space 30a flows along the peripheral sidewall of the swirl space 30a and then swirls around the central axis of the swirl space 30a.
A centrifugal force acts on the refrigerant swirling within the swirl space 30a, whereby the refrigerant pressure on the central axis side of the swirl space 30 becomes lower than the refrigerant pressure on the peripheral side thereof. Thus, in this embodiment, during the normal operation of the ejector refrigeration cycle device 10, the refrigerant pressure on the central axis side in the swirl space 30a is reduced to a pressure at which the refrigerant becomes a saturated liquid-phase refrigerant, or a pressure at which the refrigerant is decompressed and boiled (causing cavitation).
The adjustment of the refrigerant pressure on the central axis side in the swirl space 30a in this way can be achieved by adjusting the swirl flow velocity of the refrigerant swirling in the swirl space 30a. Furthermore, the adjustment of the swirl flow velocity can be performed, for example, by adjusting the ratio of the passage cross-sectional area of the refrigerant inflow passage 31e to the cross-sectional area of the swirl space 30a in a direction perpendicular to the axis direction.
Thus, in this embodiment, the passage cross-sectional area of the refrigerant inflow passage 31e is formed to be smaller than the cross-sectional area of the swirl space 30a in the direction perpendicular to the axis direction and thereby is set to a relatively small value. Note that the swirl flow velocity in this embodiment means a flow velocity in the swirl direction of the refrigerant located in the vicinity of the most peripheral part of the swirl space 30a.
A passage formation member 35 is formed within the decompression space 30b and the pressurizing space 30e. The passage formation member 35 is formed in a substantially conical shape that expands as toward the outer peripheral side as the passage formation member 35 is spaced apart from the decompression space 30b. The central axis of the passage formation member 35 is arranged coaxially with the central axis of the decompression space 30b and the like.
A refrigerant passage having an annular cross-sectional shape in the direction perpendicular to the axial direction (a doughnut shape obtained by removing a small-diameter circle from a circle arranged coaxially therewith) is formed between the inner peripheral surface of a part forming the decompression space 30b and pressurizing space 30e of the body portion 30 and a conical side surface of the passage formation member 35.
In such a refrigerant passage, a refrigerant passage part between a part forming the decompression space 30b of the body portion 30 and the tip side part of the conical side surface of the passage formation member 35 is formed to have its passage cross-sectional area throttled as toward the downstream side of the refrigerant flow. The refrigerant passage part with this shape configures a nozzle passage 13a serving as a nozzle portion that isentropically decompresses and injects the refrigerant.
More specifically, the nozzle passage 13a in this embodiment is formed to gradually decrease its passage cross-sectional area from the inlet side of the nozzle passage 13a toward the minimum passage area portion thereof, and to gradually enlarge its passage cross-sectional area from the minimum passage area portion toward the outlet side of the nozzle passage 13a. That is, the nozzle passage 13a in this embodiment changes its refrigerant passage cross-sectional area, like a so-called Laval nozzle.
Here, the above-mentioned swirl space 30a is disposed on the upstream side of the refrigerant flow on the upper side of the nozzle passage 13a. Thus, in the swirl space 30a of this embodiment, the supercooled liquid-phase refrigerant flowing into the nozzle passage 13a is allowed to swirl around the axis of the nozzle passage 13a. Therefore, in this embodiment, a part of the body portion 30 forming the swirl space 30a and a part forming the refrigerant inflow passage 31e configure a swirling-flow generating portion. In other words, in this embodiment, the ejector and the swirling-flow generating portion are configured integrally.
On the other hand, another refrigerant passage part between a part forming the pressurizing space 30e of the body portion 30 and a part on the downstream side of the conical side surface of the passage formation member 35 is formed to gradually enlarge its passage cross-sectional area toward the downstream side of the refrigerant flow. The refrigerant passage part with this shape configures a diffuser passage 13c that serves as a diffuser portion (pressurizing portion) pressurizing a mixture of an injection refrigerant injected from the nozzle passage 13a and a suction refrigerant drawn from the refrigerant suction port 31b.
In the body portion 30, an element 37 is disposed as a driving device for displacing the passage formation member 35 to change the passage cross-sectional area of the minimum passage area portion of the nozzle passage 13a.
More specifically, the element 37 includes a diaphragm that is designed to be displaceable depending on the temperature and pressure of the refrigerant circulating through the suction passage 13b (that is, the refrigerant flowing out of the evaporator 14). The displacement of the diaphragm is transferred to the passage formation member 35 via an operation stick 37a, thereby vertically displacing the passage formation member 35.
The element 37 displaces the passage formation member 35 in the direction (downward in the vertical direction) that enlarges the passage cross-sectional area of the minimum passage area portion with increasing temperature (degree of superheat) of the refrigerant flowing out of the evaporator 14. On the other hand, the element 37 displaces the passage formation member 35 in the direction (upward in the vertical direction) that reduces the passage cross-sectional area of the minimum passage area portion with decreasing temperature (degree of superheat) of the refrigerant flowing out of the evaporator 14.
In this embodiment, the element 37 displaces the passage formation member 35 depending on the degree of superheat of the refrigerant flowing out of the evaporator 14 in this way. Thus, the passage cross-sectional area of the minimum passage area portion of the nozzle passage 13a is adjusted such that the degree of superheat of the refrigerant on the outlet side of the evaporator 14 approaches a predetermined reference degree of superheat.
The gas-liquid separation space 30f is disposed under the passage formation member 35. The gas-liquid separation space 30f configures a centrifugal gas-liquid separator that swirls the refrigerant flowing out of the diffuser passage 13c around its central axis to thereby separate it into gas and liquid phase refrigerants by a centrifugal effect.
Further, in this embodiment, the internal capacity of the gas-liquid separation space 30f is set to a level that can store only a very small amount of excessive refrigerant or cannot substantially retain excessive refrigerant even though the flow rate of refrigerant circulating through the cycle is varied due to fluctuations in the load on the cycle. In this way, this embodiment enables the downsizing of the entire ejector module 13.
An oil returning passage 31f is formed in a part of the body portion 30 that forms the bottom surface of the gas-liquid separation space 30f. The oil returning passage 31f allows the refrigerating machine oil of the separated liquid-phase refrigerant to return to the gas-phase refrigerant passage for connecting the gas-liquid separation space 30f to the gas-phase refrigerant outflow port 31d. The gas-phase refrigerant outflow port 31d is connected to the suction port of the compressor 11.
On the other hand, in the liquid-phase refrigerant passage that connects the gas-liquid separation space 30f to the liquid-phase refrigerant outflow port 31c, an orifice 31i is disposed as a decompressor that decompresses the refrigerant flowing into the evaporator 14. The liquid-phase refrigerant outflow port 31c is connected to the refrigerant inflow port of the evaporator 14 via an inlet pipe 15a.
The evaporator 14 is a heat-absorption heat exchanger that exchanges heat between the low-pressure refrigerant decompressed by the nozzle passage 13a of the ejector module 13 and the ventilation air to be blown to the vehicle interior from the blower 42, thereby evaporating the low-pressure refrigerant to exhibit the heat absorption effect. The evaporator 14 is disposed in a casing 41 of an interior air-conditioning unit 40 to be described later.
Here, in the vehicle of this embodiment, the firewall 50 is provided as a partition plate that separates the vehicle interior from the engine room in the vehicle exterior. The firewall 50 also has the function of reducing heat, sound, and the like to be transferred from the engine room into the vehicle interior. The firewall can also be called a dash panel.
As shown in
Here, since the ejector module 13 is disposed in the engine room (exterior space) as mentioned above, the inlet pipe 15a and the outlet pipe 15b are disposed to pass through the firewall 50.
More specifically, the firewall 50 is provided with a circular or rectangular through hole 50a that passes through the engine room side and vehicle interior side of the fire wall. The inlet pipe 15a and the outlet pipe 15b are connected to a connector 51, which is a metal member for connection, and integrated with each other. The inlet pipe 15a and the outlet pipe 15b are arranged to pass through a through hole 50a while being integrated together by the connector 51.
At this time, the connector 51 is positioned on the inner peripheral side or in the vicinity of the through hole 50a. A packing 52 made of an elastic member is arranged in a gap between the outer peripheral side of the connector 51 and an opening edge of the through hole 50a. The packing 52 adopted in this embodiment is one formed of an ethylene-propylene-diene copolymer (EPDM) rubber, which is rubber material with excellent heat resistance.
In this way, the packing 52 is arranged to intervene in the gap between the connector 51 and the through hole 50a, thereby preventing water, noise, or the like from leaking from the engine room into the vehicle interior via the gap between the connector 51 and the through hole 50a.
Next, the interior air-conditioning unit 40 will be described. The interior air-conditioning unit 40 is to blow out the ventilation air having its temperature adjusted by the ejector refrigeration cycle device 10, into the vehicle interior. The interior air-conditioning unit 40 is disposed inside a dashboard (instrumental panel) at the foremost portion of the vehicle interior. Further, the interior air-conditioning unit 40 accommodates in the casing 41 forming its outer envelope, a blower 42, the evaporator 14, a heater core 44, an air mix door 46, and the like.
The casing 41 forms an air passage for the ventilation air to be blown into the vehicle interior. The casing 41 is formed of resin (for example, polypropylene) with some elasticity and excellent strength. An inside/outside air switch 43 is disposed on the most upstream side of the ventilation air flow in the casing 41. The inside/outside air switch acts to switch between the inside air (vehicle interior air) and the outside air (vehicle exterior air) to guide the selected air into the casing 41.
The inside/outside air switch 43 continuously adjusts the opening areas of an inside-air introduction port for introducing the inside air into the casing 41 and an outside-air introduction port for introducing the outside air thereinto by means of an inside/outside air switching door, thereby continuously changing the ratio of the volume of the inside air to that of the outside air. The inside/outside air switching door is driven by an electric actuator for the inside/outside air switching door, and the electric actuator has its operation controlled by a control signal output from the controller 60.
The fan (blower) 42 is disposed on the downstream side of the ventilation air flow of the inside/outside air switch 43 so as to blow the air drawn thereinto via the inside/outside air switch 43 toward the vehicle interior. The blower 42 is an electric blower that drives a multi-blade centrifugal fan (sirocco fan) by the electric motor and has the number of revolutions (blown air volume) controlled by a control voltage output from the controller 60.
The evaporator 14 and the heater core 44 are disposed on the downstream side of the ventilation air flow from the blower 42 in this order with respect to the ventilation air flow. In other words, the evaporator 14 is disposed on the upstream side of the ventilation air flow relative to the heater core 44. The heater core 44 is a heating heat exchanger that heats ventilation air by exchanging heat between an engine coolant and the ventilation air passing through the evaporator 14.
A cold-air bypass passage 45 is formed inside the casing 41 to allow the ventilation air passing through the evaporator 14 to flow downstream while bypassing the heater core 44. The air mix door 46 is disposed on the downstream side of the ventilation air flow relative to the evaporator 14 and on the upstream side of the ventilation air relative to the heater core 44.
The air mix door 46 serves as an air-volume-ratio adjustment portion that adjusts the ratio of the volume of the air passing through the heater core 44 to the volume of the air passing through the cold-air bypass passage 45 in the air passing through the evaporator 14. The air mix door 46 is driven by an electric actuator for driving the air-mix door. The electric actuator has its operation controlled by a control signal output from the controller 60.
A mixing space for mixing air passing through the heater core 44 with air passing through the cold-air bypass passage 45 is provided on the downstream side of the air flow of the heater core 44 and on the downstream side of the air flow of the cold-air bypass passage 45. Thus, the air mix door 46 adjusts the air volume ratio, thereby regulating the temperature of the ventilation air (conditioned air) which has been mixed in the mixing space.
Further, on the most downstream side of the ventilation air flow in the casing 41, openings (not shown) are provided for blowing the conditioned air mixed in the mixing space toward the vehicle interior as a space to be air-conditioned. Specifically, the openings include a face opening for blowing the conditioned air toward the upper body of an occupant in the vehicle interior, a foot opening for blowing the conditioned air toward the feet of the occupant, and a defroster opening for blowing the conditioned air toward the inner surface of a windshield of the vehicle.
The face opening, the foot opening, and the defroster opening have their downstream sides of the ventilation air flow connected to a face air outlet, a foot air outlet, and a defroster air outlet (all air outlets not shown) provided in the vehicle compartment, respectively, via ducts forming respective air passages.
A face door for adjusting an opening area of the face opening, a foot door for adjusting an opening area of the foot opening, and a defroster door (all doors not shown) for adjusting an opening area of the defroster opening are disposed on the upstream sides of the ventilation air flow relative to the face opening, the foot opening, and the defroster opening, respectively.
The face door, foot door, and defroster door serve as an air-outlet mode switch for switching an air outlet mode, and are coupled to electric actuators for driving the air-outlet mode doors via a link mechanism and the like and rotated in cooperation with the respective actuators for driving the air-outlet mode doors. Note that each of the electric actuators also has its operation controlled by a control signal output from the controller 60.
Specifically, the air outlet modes include, for example, a face mode, a bi-level mode, a foot mode, and a defroster mode. In the face mode, the face opening is fully opened to blow the ventilation air toward the upper body of the occupant. In the bi-level mode, both the face opening and foot opening are opened to blow the ventilation air toward the upper body and feet of the occupant. In the foot mode, the foot opening is fully opened with the defroster opening opened only by a small opening degree to blow the ventilation air mainly toward the feet of the occupant in the vehicle compartment. In the defroster mode, the defroster opening is fully opened to blow the ventilation air toward the inner surface of the windshield of the vehicle.
Next, the outline of an electric control unit in this embodiment will be described using
A group of sensors for air-conditioning control is connected to the controller 60 and designed to input detection values therefrom to the controller 60. The group of sensors includes an inside-air temperature sensor 61, an outside-air temperature sensor 62, a solar radiation sensor 63, an evaporator temperature sensor 64, a coolant temperature sensor 65, and a high-pressure side pressure sensor 66. The inside-air temperature sensor 61 detects a vehicle interior temperature (inside air temperature) Tr. The outside-air temperature sensor 62 detects an outside air temperature Tam. The solar radiation sensor 63 detects the solar radiation amount As within the vehicle interior. The evaporator temperature sensor 64 detects the blown-air temperature (evaporator temperature) Tefin of the air blown from the evaporator 14. The coolant temperature sensor 65 detects the coolant temperature Tw of the engine coolant flowing into the heater core 44. The high-pressure side pressure sensor 66 detects a pressure (high-pressure side refrigerant pressure) Pd of the high-pressure refrigerant discharged from the compressor 11.
The input side of the controller 60 is connected to an operation panel 70 (not shown) disposed near the dashboard at the front of the vehicle compartment. Operation signals from various operation switches provided on the operation panel 70 are input to the controller 60. Various operation switches provided on the operation panel 70 include an automatic switch for setting an automatic control operation of the vehicle air conditioner 1, a vehicle interior temperature setting switch for setting a preset temperature Tset of the vehicle interior, and an air-volume setting switch for manually setting the volume of air from the blower 42.
The controller 60 in this embodiment incorporates therein control units for controlling the operations of various control target devices connected to its output side. In the controller 60, a structure (hardware and software) adapted to control the operation of each control target device serves as the control unit for the corresponding control target device.
For example, in this embodiment, the structure for controlling the operation of a discharge displacement control valve of the compressor 11 configures a discharge-capacity control unit 60a for controlling a refrigerant discharge capacity of the compressor 11. Obviously, the discharge-capacity control unit may be configured as a separate controller with respect to the controller 60.
Now, the operation of the vehicle air conditioner 1 with the above-mentioned structure in this embodiment will be described based on
In step S1, first, initialization is performed which includes initializing a flag, a timer, and the like in a memory circuit of the controller 60, and initial alignment of various electric actuators described above. Note that in the initialization at step S1, the controller may read out some of flags and calculated values previously stored when the vehicle air conditioner 1 is stopped or when the vehicle system is shut down.
Then, in step S2, detection signals from the sensor group (61 to 67) for air-conditioning control and operation signals from the operation panel 70 are read in. In subsequent step S3, a target air outlet temperature TAO, which is a target temperature of the ventilation air to be blown into the vehicle interior, is calculated based on the detection signal and operation signal read in step S2.
Specifically, the target air outlet temperature TAO is calculated by the following formula F1:
TAO=Kset×Tset−Kr×Tr−Kam×Tam−Ks×As+C (F1)
where Tset is a vehicle interior preset temperature set by the vehicle interior temperature setting switch, Tr is a vehicle interior temperature (inside air temperature) detected by the inside-air temperature sensor 61, Tam is the outside air temperature detected by the outside-air temperature sensor 62, and As is an amount of solar radiation detected by the solar radiation sensor 63. Kset, Kr, Kam, and Ks are control gains, and C is a constant for correction.
In subsequent steps S4 to S8, the control state of each of the control target devices connected to the controller 60 is determined.
In step S4, first, the number of revolutions (blowing capacity) of the blower 42, that is, a blower motor voltage (control voltage) applied to the electric motor of the blower 42 is determined, and the operation proceeds to step S5. Specifically, in step S4, a blower motor voltage is determined with reference to a control map pre-stored in the controller 60 based on the target air outlet temperature TAO determined in step S3.
In more detail, the blower motor voltage is determined in such a manner as to take the substantially maximum value in an ultralow temperature range (maximum cooling range) and an ultrahigh temperature range (maximum heating range) of the target air outlet temperature TAO. Furthermore, the blower motor voltage is determined in such a manner as to gradually decrease from the substantially maximum value as the target air outlet temperature TAO goes from the ultralow temperature range or ultrahigh temperature range to an intermediate temperature range.
Then, in step S5, a suction port mode, that is, a control signal to be output to the electric actuator for the inside/outside air switching door is determined, and then the operation proceeds to step S6. Specifically, in step S5, the suction port mode is determined with reference to the control map pre-stored in the controller 60 based on the target air outlet temperature TAO.
More specifically, the suction port mode is basically determined to be an outside-air mode for introducing the outside air. When the target air outlet temperature TAO is in the ultralow temperature range and a high cooling performance is desired, the suction port mode is determined to be an inside-air mode for introducing the inside air.
Then, in step S6, an opening degree of the air mix door 46, that is, a control signal to be output to the electric actuator for driving the air mix door is determined, and then the operation proceeds to step S7.
Specifically, in step S6, the opening degree of the air mix door 46 is calculated such that the temperature of ventilation air blown into the vehicle interior approaches the target air outlet temperature TAO, based on the evaporator temperature Tefin detected by the evaporator temperature sensor 64, the coolant temperature Tw detected by the coolant temperature sensor 65, and the target air outlet temperature TAO.
Then, in step S7, an air outlet mode, that is, a control signal to be output to the electric actuator for driving an air-outlet mode door is determined, and then the operation proceeds to step S8. Specifically, in step S8, the air outlet mode is determined with reference to the control map pre-stored in the controller 60 based on the target air outlet temperature TAO.
In more detail, the air outlet mode is switched from the foot mode to the bi-level mode and then the face mode in this order as the target air outlet temperature TAO decreases from a high-temperature range to a low-temperature range.
Next in step S8, the refrigerant discharge capacity of the compressor 11, that is, a control current output to the discharge displacement control valve of the compressor 11 is determined, and the operation proceeds to step S9. The details of step S8 will be described below using the flowchart of
In step S81 of
Next in step S82, the refrigerant discharge capacity of the compressor 11 in the normal control, that is, a control current output to the discharge displacement control valve of the compressor 11 is determined, and the operation proceeds to step S9. Specifically, in step S82, a target evaporator outlet air temperature TEO of the evaporator 14 is determined with reference to the control map pre-stored in the controller 60 based on the target air outlet temperature TAO.
A target refrigerant discharge capacity of the compressor 11 is determined such that an evaporator temperature Tef in approaches the target evaporator outlet air temperature TEO using a feedback control method, based on a deviation between the target evaporator outlet air temperature TEO and the evaporator temperature Tefin detected by the evaporator temperature sensor.
Next in step S83, the refrigerant discharge capacity of the compressor 11 at startup, that is, a control current output to the discharge displacement control valve of the compressor 11 is determined, and the operation proceeds to step S9. Specifically, in step S83, the target refrigerant discharge capacity of the compressor 11 is determined at the startup, in the same way as in step S82. As indicated by a thick solid line in the control characteristic diagram described in step S83 of
In more detail, in step S83, the refrigerant discharge capacity is increased in such a manner that an increase amount (capacity increase degree) in the refrigerant discharge capacity per predetermined time period (predetermined reference time period) is lower than a predetermined reference capacity increase amount (reference capacity increase degree). Furthermore, in this embodiment, the reference capacity increase amount is defined as the maximum capacity increase amount per predetermined time period enabled by the compressor 11. The maximum capacity increase is represented by a slope of a dashed line in the control characteristic diagram described in step S83 of
In other words, in step S83 of this embodiment, it can be expressed that the refrigerant discharge capacity is gradually increased in such a manner that an actual refrigerant discharge capacity of the compressor 11 does not reach the target refrigerant discharge capacity until the predetermined time elapses. Furthermore, it can be expressed that an actual refrigerant discharge capacity of the compressor 11 is gradually increased until it reaches the target refrigerant discharge capacity, over a longer time than the time period during which the compressor 11 exhibits the maximum capacity increase.
Then, in step S9 shown in
That is, in the air-conditioning control program executed by the controller 60, a routine that includes reading a detection signal and an operation signal, determining the control state of each control target device, and outputting a control signal and a control voltage to each control target device in this order is repeated until the stopping of the operation of the vehicle air conditioner 1 is requested. The air-conditioning control program is executed to allow the refrigerant to flow in the ejector refrigeration cycle device 10 as indicated by thick solid arrows shown in
That is, the high-temperature and high-pressure refrigerant discharged from the compressor 11 flows into the condensing portion 12a of the radiator 12. The refrigerant flowing into the condensing portion 12a exchanges heat with the outside air blown by the cooling fan 12d and dissipates heat therefrom to condense itself. The refrigerant condensed in the condensing portion 12a is separated into gas and liquid phases by the receiver 12b. The liquid-phase refrigerant of the gas and liquid phases refrigerants separated by the receiver 12b exchanges heat with the outside air blown from the cooling fan 12d in the supercooling portion 12c, and further dissipates heat therefrom to be converted into the supercooled liquid-phase refrigerant.
The supercooled liquid-phase refrigerant flowing out of the supercooling portion 12c of the radiator 12 is isentropically decompressed by the nozzle passage 13a formed between the inner peripheral surface of the decompression space 30b and the outer peripheral surface of the passage formation member 35 in the ejector module 13, and is then injected therefrom. At this time, the refrigerant passage area of the minimum passage area portion in the decompression space 30b is adjusted such that the degree of superheat of the refrigerant on the outlet side of the evaporator 14 approaches a reference degree of superheat.
The refrigerant flowing out of the evaporator 14 is drawn into the ejector module 13 from the refrigerant suction port 31b by the suction effect of an injection refrigerant injected from the nozzle passage 13a. The injection refrigerant injected from the nozzle passage 13a and the suction refrigerant drawn via the suction passage 13b flow into and are merged in the diffuser passage 13c.
In the diffuser passage 13c, the kinetic energy of the refrigerant is converted into pressure energy thereof by the enlarged refrigerant passage area. Thus, while the injection refrigerant and suction refrigerant are being mixed together, the mixed refrigerant has its pressure increased. The refrigerant flowing out of the diffuser passage 13c is separated by the gas-liquid separation space 30f into gas and liquid phase refrigerants. The liquid-phase refrigerant separated by the gas-liquid separation space 30f is decompressed by an orifice 31i and then flows into the evaporator 14.
The refrigerant flowing into the evaporator 14 absorbs heat from the ventilation air blown from the blower 42 to evaporate itself. In this way, the ventilation air is cooled. On the other hand, the gas-phase refrigerant separated by the gas-liquid separation space 30f flows out of the gas-phase refrigerant outflow port 31d, and is then drawn into and compressed again by the compressor 11.
In the interior air-conditioning unit 40, the ventilation air cooled by the evaporator 14 flows into a ventilation passage on the heater core 44 side and the cold-air bypass passage 45 depending on the opening degree of the air mix door 46. The cold air flowing into the ventilation passage on the heater core 44 side is reheated when passing through the heater core 44, and then mixed with a cold air passing through the cold-air bypass passage 45 in a mixing space. The conditioned air having its temperature adjusted in the mixing space is blown into the vehicle interior via the respective air outlets.
As mentioned above, the vehicle air conditioner 1 in this embodiment can perform air-conditioning of the vehicle interior. Further, in the ejector refrigeration cycle device 10 of this embodiment, the refrigerant pressurized by the diffuser passage 13c is drawn into the compressor 11, thus enabling the reduction in the driving power for the compressor 11, thereby improving the coefficient of performance (COP) of the cycle.
Further, in the ejector module 13 of this embodiment, the supercooled liquid-phase refrigerant flows into and swirls in the swirl space 30a, whereby the refrigerant pressure on the swirl center side of the swirl space 30a is reduced to a pressure at which the refrigerant becomes a saturated liquid-phase refrigerant, or a pressure at which the refrigerant is decompressed and boiled (causing cavitation). The gas-liquid two-phase refrigerant in which the majority of gas-phase refrigerant is located on the swirl central side flows into the nozzle passage 13a.
Thus, the boiling of the refrigerant in the nozzle passage 13a can be promoted due to wall boiling caused by the friction between the refrigerant and the wall surface of the nozzle passage 13a as well as interface boiling caused by a boiling nucleus generated by the cavitation of the refrigerant on the swirl central side. As a result, the energy conversion efficiency can be improved when converting the pressure energy of the refrigerant into the velocity energy thereof in the nozzle passage 13a.
For example, when the outside air temperature is relatively high at start-up of the ejector refrigeration cycle device 10, the gas-phase refrigerant sometimes remains in the radiator. Thus, when the ejector refrigeration cycle device 10 is started up at a high outside-air temperature and the like, the gas-liquid two-phase refrigerant not sufficiently cooled might flow out of the radiator 12 at start-up of the compressor 11. The gas-liquid two-phase refrigerant might flow into the refrigerant inflow passage 31e of the ejector module 13.
It should be noted that the term “start-up of the compressor 11” as used in this embodiment includes a period of time immediately after start-up of the compressor 11, at least from when the compressor 11 does not exhibit its refrigerant discharge capacity till when the compressor 11 exhibits the target refrigerant discharge capacity.
In the ejector module 13 of this embodiment, in order to appropriately swirl the supercooled liquid-phase refrigerant in the swirl space 30a, the passage cross-sectional area of the refrigerant inflow passage 31e is set to a relatively small value as mentioned above.
Thus, if the gas-liquid two-phase refrigerant flows into the refrigerant inflow passage 31e, the gas-liquid two-phase refrigerant circulating through the refrigerant inflow passage 31e flows at a high velocity, as compared to when a supercooled liquid-phase refrigerant with a higher density flows in, which might cause friction noise when it circulates through the refrigerant inflow passage 31e. Further, if the friction noise resonates with the gas-phase refrigerant eccentrically located in a columnar shape on the center side of the swirl space, a so-called air column resonance might make significant noise.
In contrast, in the ejector refrigeration cycle device 10 of this embodiment, as described in control step S83, the refrigerant discharge capacity is increased such that the increase amount in the refrigerant discharge capacity per predetermined time period is lower than the reference capacity increase amount at start-up of the compressor 11.
Therefore, even if the gas-liquid two-phase refrigerant flows into the refrigerant inflow passage 31e, the flow velocity of the gas-liquid two-phase refrigerant is prevented from becoming high, which can reduce the friction noise that would be caused when the gas-liquid two-phase refrigerant circulates through the refrigerant inflow passage 31e. As a result, the noise generated from the ejector module 13 can be reduced at start-up of the compressor 11.
This embodiment adopts as the reference capacity increase amount, the maximum capacity increase amount per predetermined time period that is determined by a capacity inherent to the compressor 11. Thus, the noise generated from the ejector module 13 can be surely reduced when the refrigerant discharge capacity is increased naturally by the maximum capacity increase amount at start-up of the compressor 11.
Further, the reference capacity increase amount is set to such a level of the capacity increase that prevents noise generated from the ejector module 13 at start-up of the compressor 11 from being harsh on the user's ear. Thus, noise generated from the ejector module 13 can be effectively reduced.
As shown in the entire configuration diagram of
The flow-rate adjustment valve 16 is an inflow rate adjustment portion that adjusts the flow rate of the inflow refrigerant flowing into the refrigerant inflow passage 31e forming the swirling-flow generating portion. More specifically, the flow-rate adjustment valve 16 includes a valve body capable of changing the refrigerant passage area and an electric actuator for displacing the valve body. Further, the flow-rate adjustment valve 16 has its operation controlled by a control voltage output from the controller 60.
Thus, as shown in the block diagram of
In the vehicle air conditioner 1 of this embodiment, in step S8′ of the flowchart shown in
Further, in step S85, a valve opening degree of the flow-rate adjustment valve 16, that is, a control progress state to be output to the flow-rate adjustment valve 16 is determined, and then the operation proceeds to step S9. In step S85, if not at start-up of the compressor 11, the valve opening degree of the flow-rate adjustment valve 16 is maximized (in a fully open state). In contrast, at start-up of the compressor 11, the valve opening degree of the flow-rate adjustment valve 16 is gradually increased to attain a refrigerant inflow rate indicated by a thick solid line in the control characteristic diagram of
In more detail, in step S85, at start-up of the compressor 11, the refrigerant inflow rate is increased in such a manner that the increase amount (flow-rate increase degree) in the refrigerant inflow rate per predetermined time period (predetermined reference time period) is lower than the predetermined reference flow-rate increase amount (reference flow-rate increase degree). Furthermore, in this embodiment, the reference flow-rate increase amount is defined as the maximum flow-rate increase amount per predetermined time period enabled by the flow-rate adjustment valve 16.
That is, the maximum flow-rate increase amount corresponds to a flow-rate increase amount obtained when the valve opening degree of the flow-rate adjustment valve 16 is maximized at start-up of the compressor 11. The maximum flow-rate increase amount is represented by a slope of a dashed line in the control characteristic diagram described in
In other words, in step S85 of this embodiment, it can be expressed that the valve opening degree (refrigerant inflow rate) is gradually increased in such a manner as to prevent the valve opening degree of the flow-rate adjustment valve 16 from being maximized until the predetermined time elapses at start-up of the compressor 11. Furthermore, it can also be expressed that the refrigerant inflow rate is gradually increased over a longer time than the time period during which the valve opening of the flow-rate adjustment valve 16 is maximized.
The operations of other components are the same as those in the first embodiment. Therefore, the vehicle air conditioner 1 of this embodiment can also perform the air-conditioning of the vehicle interior in the same manner as that in the first embodiment, and thus can obtain the same effects as those in the first embodiment.
Furthermore, in the ejector refrigeration cycle device 10 of this embodiment, as described in control step S85, the refrigerant inflow rate is increased such that the increase amount in the refrigerant inflow rate per predetermined time period is lower than the reference flow-rate increase amount at start-up of the compressor 11.
Therefore, even if the gas-liquid two-phase refrigerant flows into the refrigerant inflow passage 31e, the flow velocity of the gas-liquid two-phase refrigerant is prevented from becoming high, which can reduce the friction noise that would be caused when the gas-liquid two-phase refrigerant circulates through the refrigerant inflow passage 31e. As a result, like the first embodiment, the noise generated from the ejector module 13 can be reduced at start-up of the compressor 11.
This embodiment adopts as the reference flow-rate increase, the maximum flow-rate increase amount per predetermined time period obtained when the valve opening degree of the flow-rate adjustment valve 16 is maximized. Thus, at start-up of the compressor 11, the noise generated from the ejector module 13 can be surely reduced when the valve opening degree of the flow-rate adjustment valve 16 is maximized.
Further, the reference flow-rate increase amount is set to such a level of the flow-rate increase amount that prevents noise generated from the ejector module 13 from being harsh on the user's ear at start-up of the compressor 11, so that noise generated from the ejector module 13 can be effectively reduced.
In this embodiment, as shown in the entire configuration diagram of
Therefore, the vehicle air conditioner 1 of this embodiment can also perform the air-conditioning of the vehicle interior in the same manner as that in the first embodiment, and thus can obtain the same effects as those in the first embodiment. Further, like the second embodiment, the noise generated from the ejector module 13 can be reduced at start-up of the compressor 11.
The present disclosure is not limited to the above-mentioned embodiments, and various modifications and changes can be made to those embodiments in the following way without departing from the scope and spirit of the present disclosure.
The same goes for the refrigerant inflow rate described in the second embodiment. That is, in the same manner as that illustrated in step S83 of the control characteristic diagram in
Alternatively, a flow-rate adjustment mechanism may be adopted which includes a displacement member to be displaced depending on the temperature and pressure of the refrigerant circulating through a predetermined part in the cycle, and a valve body coupled to the displacement member and adapted to change its refrigerant passage area. The flow-rate adjustment mechanism changes its refrigerant passage area through such a mechanical mechanism. Specifically, the flow-rate adjustment mechanism can be employed which is adapted to detect the degree of superheat of the refrigerant on the outlet side of the radiator 12 based on the temperature and pressure of the refrigerant on the outlet side of the radiator 12 and to increase the valve opening degree with decreasing degree of superheat detected.
For example, whether the compressor 11 is at startup or not may be determined using a pressure (high-pressure side refrigerant pressure) Pd of the refrigerant circulating through the refrigerant flow path that leads from the outlet side of the compressor 11 to the refrigerant inflow port 31a side of the ejector module 13. When a revolution indicator for detecting the number of revolutions of the compressor 11 is installed, whether or not the compressor 11 is at startup may be determined based on a detected value of the revolution indicator.
For example, the above-mentioned embodiments employ the variable displacement compressor as the compressor 11, but the compressor 11 is not limited thereto. The compressor 11 for use may be a fixed displacement compressor that is driven by a rotational driving force output from the engine via an electromagnetic clutch, a belt, etc.
The fixed displacement compressor may adjust the refrigerant discharge capacity by changing an operating rate of the compressor through switching between the connection and disconnection of the electromagnetic clutch. The compressor 11 for use may be an electric compressor that adjusts the refrigerant discharge capacity by changing the number of revolutions of the electric motor.
For example, in the above-mentioned embodiments, the radiator 12 employs a sub-cool type heat exchanger by way of example. Alternatively, a standard radiator configured of only the condensing portion 12a may be adopted. Further, a reservoir (receiver) may be employed along with the standard radiator. The reservoir separates the refrigerant dissipating its heat in the radiator, into gas and liquid phase refrigerants, and stores an excessive liquid-phase refrigerant.
The respective components forming the ejector module 13 are not limited to those disclosed in the above-mentioned embodiments. For example, the components of the ejector module 13, including the body portion 30 and the passage formation member 35, are made of metal, but are not limited thereto and may alternatively be formed of resin.
Further, in the ejector module 13 of the above-mentioned embodiments, the orifice 31i is provided by way of example. However, the orifice 31i may be abolished, and a decompression portion may be disposed in the inlet pipe 15a. Such a decompression portion can be an orifice, a capillary tube, or the like.
Further, the above-mentioned embodiments employ the ejector module 13 of the gas-liquid separator integrated ejector by way of example. However, it is obvious that a standard ejector that does not include a gas-liquid separator integrated therewith may be employed as the ejector.
Further, the ejector module 13 may be disposed on the inner peripheral side of the through hole 50a of the firewall 50. In this case, a part of the ejector module 13 is disposed on the engine room side, and the other part is disposed on the vehicle interior side. Thus, a packing exhibiting the same function as in the first embodiment is desirably disposed in a gap between the outer peripheral side of the ejector module 13 and the opening edge of the through hole 50a.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2014-217455 | Oct 2014 | JP | national |
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/JP2015/004094 | 8/18/2015 | WO | 00 |
