EP1637736A2 - Regelventil für einen Verdichter variabler Verdrängung - Google Patents

Regelventil für einen Verdichter variabler Verdrängung Download PDF

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Publication number
EP1637736A2
EP1637736A2 EP05019466A EP05019466A EP1637736A2 EP 1637736 A2 EP1637736 A2 EP 1637736A2 EP 05019466 A EP05019466 A EP 05019466A EP 05019466 A EP05019466 A EP 05019466A EP 1637736 A2 EP1637736 A2 EP 1637736A2
Authority
EP
European Patent Office
Prior art keywords
valve
refrigerant
port
shaft
compressor
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP05019466A
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English (en)
French (fr)
Inventor
Hisatoshi Hirota
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
TGK Co Ltd
Original Assignee
TGK Co Ltd
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by TGK Co Ltd filed Critical TGK Co Ltd
Publication of EP1637736A2 publication Critical patent/EP1637736A2/de
Withdrawn legal-status Critical Current

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/14—Control
    • F04B27/16—Control of pumps with stationary cylinders
    • F04B27/18—Control of pumps with stationary cylinders by varying the relative positions of a swash plate and a cylinder block
    • F04B27/1804—Controlled by crankcase pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/22—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves
    • F04B49/225—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves with throttling valves or valves varying the pump inlet opening or the outlet opening
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/14—Control
    • F04B27/16—Control of pumps with stationary cylinders
    • F04B27/18—Control of pumps with stationary cylinders by varying the relative positions of a swash plate and a cylinder block
    • F04B27/1804—Controlled by crankcase pressure
    • F04B2027/1822—Valve-controlled fluid connection
    • F04B2027/1827—Valve-controlled fluid connection between crankcase and discharge chamber
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/14—Control
    • F04B27/16—Control of pumps with stationary cylinders
    • F04B27/18—Control of pumps with stationary cylinders by varying the relative positions of a swash plate and a cylinder block
    • F04B27/1804—Controlled by crankcase pressure
    • F04B2027/184—Valve controlling parameter
    • F04B2027/1854—External parameters
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/14—Control
    • F04B27/16—Control of pumps with stationary cylinders
    • F04B27/18—Control of pumps with stationary cylinders by varying the relative positions of a swash plate and a cylinder block
    • F04B27/1804—Controlled by crankcase pressure
    • F04B2027/1863—Controlled by crankcase pressure with an auxiliary valve, controlled by
    • F04B2027/1872—Discharge pressure

Definitions

  • the invention relates to a control valve according to the preamble of claim 1, particularly for a variable displacement compressor in a refrigeration cycle of an automotive air conditioner.
  • Variable displacement compressors capable of varying the compression capacity of refrigerant are employed to obtain an adequate cooling capacity without being constrained by the speed of the vehicle engine driving the compressor.
  • a wobble plate on a shaft driven by the engine is coupled to compression pistons.
  • the stroke of the pistons is varied to vary the discharge amount.
  • the control valve is disposed either between the discharge chamber and the crankcase or between the crankcase and the suction chamber and maintains the differential pressure across a valve such that the flow rate between the discharge chamber and the crankcase at a predetermined value.
  • the differential pressure can be set to the predetermined value by externally changing a value of control current supplied to a solenoid of the control valve.
  • a control valve known from JP 2001-107854 A controls the discharge flow rate to become constant.
  • Two spaced apart pressure sensors in a refrigerant passage toward the suction chamber detect a differential pressure to indirectly measure the drawn in flow rate.
  • the control valve then controls a constant flow rate between the discharge chamber and the crankcase to thereby control the control discharge flow rate.
  • the known system requires expensive pressure sensors and a control device for detecting the differential pressure across the refrigerant circulation passage and for controlling the control valve. This leads to increased costs of the automotive air conditioner.
  • the control valve is configured such that the first valve indirectly measures the discharge flow rate of refrigerant, and such that the second valve is controlled based on a value of the discharge flow rate to thereby control the pressure in the crankcase.
  • the first valve has a structure that opens depending on the flow rate between the discharge chamber and the outlet port, and hence closes when the compressor shifts to the minimum capacity operation to minimize the flow rate, discharge chamber and the outlet port of the compressor has changed e.g. immediately after a transition to the minimum capacity operation, to hold the pressure at the outlet port at the pressure value assume before the transition to the minimum capacity operation has taken place. This allows to abolish a check vale at the outlet port of the compressor for this purpose, and thereby reduces the cost of the compressor.
  • the first valve is also configured to have a larger pressure-receiving area for receiving the discharge pressure on the discharge chamber side than the second valve. This allows to construct a highly responsive variable displacement compressor that is operable when the rotational compressor speed has rapidly changed, to promptly react in a direction of suppressing a change of the displacement.
  • the control valve 10 of Fig. 1 (first embodiment)comprises a first valve 11 operating in dependence on the flow rate discharged from the compressor, a second valve 12 for controlling pressure Pc in the crankcase, a third valve 13 for controlling the amount of leakage of refrigerant, and a solenoid 14 for externally setting the flow rate discharged from the compressor.
  • the first valve 11 is formed in a first body 15 disposed at an upper end location.
  • the first body 15 has a port 16 communicating with the discharge chamber (discharge pressure Pdh).
  • a port 17 communicates with an outlet port of the compressor (discharge pressure Pdl).
  • a refrigerant passage 18 connects these ports 16 and 17.
  • a valve seat 19 is formed in the refrigerant passage 18.
  • a valve element 20 is movably disposed on the side of the valve seat 19 toward the port 17. The valve element 20 is urged by a weak spring 21 in the direction of closing the refrigerant passage 18.
  • the first valve 11 is constructed as a check valve that opens when the discharge pressure Pdh at the port 16 is higher than the discharge pressure Pdl at the port 17 by more than the urging force of the spring 21, and that otherwise closes.
  • the second valve 12 is formed in a second body 22 to which the first body 15 is secured by press-fitting.
  • the second body 22 defines a refrigerant introducing space 22a into which the discharge pressure Pdh is introduced via a refrigerant passage 23 in the first body 15.
  • a strainer 24 covers an inlet port side of the refrigerant passage 23.
  • the second body 22 has a port 25 communicating with the crankcase to discharge refrigerant at a controlled pressure Pc to the crankcase.
  • a valve hole In the centre of the upper part of the second body 22, there is formed a valve hole between the refrigerant introducing space 22a and an internal space 25a communicating with the port 25.
  • a valve element 27 is movably guided by the second body 22 in relation to a valve seat 26 formed at a lower end of the valve hole. The valve element 27 is urged by a spring 28 in a direction away from the valve seat 26.
  • the second valve 12 controls the flow rate at the discharge pressure Pdh to supply pressure Pc to the crankcase
  • a shaft 29 movably extends through an axial through hole 15a.
  • the shaft 29 also extends through the valve hole of the second valve 12.
  • An upper end of the shaft 29 is loosely fitted into the valve element 20.
  • a lower end of the shaft 29 is loosely fitted into the valve element 27 of the second valve 12.
  • An upper part of the shaft 29 has an outer diameter larger than an inner diameter of the through hole 15a of the first body 15, such that a tapered stepped portion 29a is formed at a boundary between the upper part and a lower part.
  • the second body 22 has a hole in the centre of a lower part.
  • the open rim of a bottomed sleeve 30 is tightly connected to the hole.
  • the bottomed sleeve 30 contains a fixed core 31 and a plunger 32 of the solenoid.
  • the core 31 as well is fixed to the hole of the first body 15 and to the bottomed sleeve 30 by press-fitting.
  • the plunger 32 is axially slidable in the bottomed sleeve 30, and is fixed to one end of a shaft 33 which axially extends with clearance through the core 31.
  • the plunger 32 is urged toward the core 31 by a spring 34 to bring the other shaft end into contact with a lower end face of the valve element 27.
  • a coil 35 surrounds the bottomed sleeve 30.
  • a harness 36 leads from the coil 35to the outside of the solenoid 14.
  • the inside of the bottomed sleeve 30 communicates with an internal space communicating with the port 25 via a pressure equalizing hole 37 in the second body 22.
  • Fig. 1 shows a state of the control valve 10 immediately after stoppage of the operation of the automotive air conditioner. This corresponds to the case where after the automotive air conditioner has been in operation, the solenoid 14 is de-energized. The solenoid 14 ceases to create a force attracting the plunger 32 toward the core 31, and hence the second valve 12 is fully opened by the spring 28 acting against the spring 34 in valve-opening direction. Refrigerant at the discharge pressure Pdh is supplied from the port 25 to the crankcase via the strainer 24, the refrigerant passage 23, and the second valve 12. The compressor is shifted to the minimum capacity operation.
  • the valve element 20 of the first valve 11 When the compressor is operating with a predetermined capacity, the valve element 20 of the first valve 11 has been moved away from the valve seat 19.
  • the shaft 29 is pushed downward to close the third valve 13.
  • the solenoid 14 When the solenoid 14 is de-energized to allow the valve element 20 of the first valve 11 to move upward and the valve element 27 of the second valve 12 to move downward (Fig. 1).
  • the shaft 29 receives the high discharge pressure Pdl on the top end and the discharge pressure Pdh lower than the discharge pressure Pdl on the bottom end, which holds the third valve 13 closed. This prevents that the high discharge pressure Pdl may leak via the radial clearance to the upstream side of the second valve 12 the pressure at which has become lower than the discharge pressure Pdl.
  • the pressure at the compressor outlet port can maintain the discharge pressure Pdl assumed before the stoppage of the operation of the automotive air conditioner. This is advantageous for the efficiency of the compressor since it is not necessary to compress refrigerant to the discharge pressure Pdl when the automotive air conditioner later resumes operation. Further, normally, a separate check valve would have to be provided at the outlet port of the compressor for this purpose, but in the present embodiment, the first valve 11 already fulfils the check valve function, such that a separate check valve is not needed, to reduce the cost of the compressor.
  • the compressor starts operation with maximum capacity.
  • the discharge pressure Pdh becomes sufficiently higher than the discharge pressure Pdl (Fig. 3)
  • the differential pressure therebetween causes the valve element 20 to move away from the valve seat 19.
  • the first valve 11 opens and refrigerant at discharge pressure Pdh at port 16 changes into refrigerant at the discharge pressure Pdl, which flows from port 17 to the compressor outlet port.
  • refrigerant flows through the first valve 11 at a flow rate corresponding to a value obtained by multiplying the passage area formed by opening the first valve 11 by the differential pressure across the first valve 11.
  • the valve element 20 of the first valve 11 is moved in valve-opening direction into abutment with the upper end of the shaft 29 which has been lifted upward, as viewed in FIG. 3, by the valve element 27 of the second valve 12. This causes the first valve 11 and the second valve 12 to operate interlocked with each other via the shaft 29.
  • the second valve 12 operates by detecting the differential pressure between the discharge pressure Pdh and the discharge pressure Pdl acting on the first valve 11, and the differential pressure between the discharge pressure Pdh and the pressure Pc.
  • the second valve 12 slightly opens to a position where the differential pressures across the first valve 11 and the second valve 12, the loads of the springs 21 and 28, the urging force of the solenoid 14 dependent on the current value are balanced, whereby the controlled pressure Pc is supplied to the crankcase to place the compressor in the state in which the capacity or displacement is controlled.
  • the second valve 12 controls the pressure in the crankcase such that the differential pressure across the passage having the passage area produced by flow of refrigerant from the discharge chamber through the first valve 11 maintains a differential pressure set by the solenoid 14, to thereby control the compressor discharge flow rate to a constant flow rate. More specifically, e.g. when the engine speed increases to increase the discharge pressure Pdh, the valve element 20 of the first valve 11 urges the valve element 27 of the second valve 12 in opening direction, by the increased amount of the differential pressure. This increases the pressure Pc in the crankcase, and hence the compressor operates in the direction of reducing the displacement, whereby the discharge flow rate is controlled to a predetermined flow rate.
  • valve element 20 urges the valve element 27 of the second valve 12 in the direction of further closing the same. This reduces the pressure Pc in the crankcase and hence the compressor operates in the direction of increasing the displacement thereof, whereby the discharge flow rate is controlled to the predetermined flow rate.
  • the discharge pressure Pdh varies sensitively in response to a change in the compressor speed. Therefore, when the engine speed rapidly increases to rapidly increase the compressor speed as well, the discharge pressure Pdh also rapidly increases.
  • the control valve 10 since the pressure receiving area of the valve element 20 of the first valve 11 is set to be larger than that of the valve element 27 of the second valve 12, the force of the valve element 20 of the first valve 11 urging the valve element 27 of the second valve 12 in the direction of further opening the same by a change in the differential pressure across the valve element 20 is instantaneously increased, and as shown in Fig. 5, the second valve 12 operates instantaneously more intensely than during normal opening operation, whereby the compressor is promptly controlled in the direction of reducing the displacement.
  • valve element 20 of the first valve 11 also operates instantaneously more intensely in valve-closing direction, and hence the second valve 12 also operates instantaneously more intensely in valve-closing direction, whereby the compressor is promptly controlled in the direction of increasing the displacement.
  • the control valve 10 is capable of promptly restoring the compressor to a predetermined displacement.
  • the pressure which the second valve 12 senses and to which it responds is changed. That is, while the second valve 12 of the first embodiment senses the differential pressure between the discharge pressure Pdh on the discharge side and the pressure Pc in the crankcase, the second valve 12 of the control valve 10a senses the differential pressure between the discharge pressure Pdh on the discharge side and the suction pressure Ps in the suction chamber.
  • the second body 22 has a port 38 communicating with the suction chamber (suction pressure Ps).
  • the second body 22 movably guides a shaft 39 which is integral with the valve element 27 of the second valve 12.
  • the outer diameter of the shaft 39 is approximately equal to an effective discharge pressure Pdh receiving diameter of the second valve 12.
  • the valve element 27 receives the discharge pressure Pdh on the discharge side.
  • the lower end of the shaft 39 receives the suction pressure Ps to sense the differential pressure (Pdh-Ps).
  • the outer diameter of the shaft 39 is only required to be equal to the inner diameter of the valve hole of the second valve 12.
  • the outer diameter of the shaft 39 is made larger than the inner diameter of the valve hole of the second valve 12 to such an extent that there is no substantial influence on the operation of the second valve 12.
  • a spring 28 at the lower end of the shaft 39 urges the valve element 27 of the second valve 12 in valve-opening direction.
  • the bottomed sleeve 30 communicates with the port 38 to receive the suction pressure Ps.
  • the control valve 10a operates similar to the control valve 10 of Fig. 1..
  • the control valve 10b of Figs 7-11 differs from the first and second embodiments in the construction of the first valve 11. While in the control valves 10 and 10a the first valve 11 provides a passage area dependent on the flow rate of refrigerant, the first valve 11 of the control valve 10b of Figs 7-11 does not vary the passage area according to the flow rate within a normal control region and after the first valve 11 once is open.
  • the first valve 11 valve element 20 is movable in the refrigerant passage 18 of the first body 15, and has (Fig 8) a plurality of guides 40 integrally formed at the outer periphery for guiding the valve element 20 in axial direction in the refrigerant passage 18, and has refrigerant passages 41 between the valve element 20 and the inner wall of the refrigerant passage 18, the passage area of which does not vary even when the flow rate is varied and when the valve lift varies.
  • a hollow cylindrical valve seat-forming member 42 is disposed upstream of and opposed to the valve element 20.
  • the valve seat-forming member 42 is press-fitted in the port 16 (discharge pressure Pdh).
  • the shaft 29 forming the valve element of the third valve 13 is axially movable in the valve element 20 of the first valve 11.
  • a spring 21 between the valve element 20 and the shaft 29 urges the valve element 20 and the shaft 29 in respective directions of moving apart to thereby maintain the closed states of the first valve 11 and the third valve 13 when the solenoid 14 is de-energized.
  • the second valve element 27 is integral with the shaft 33 of the solenoid 14.
  • the spring 28 urging the valve element 27 in valve-opening direction is interposed between the core 31 and the plunger 32.
  • the shaft 33 terminates by a transmission shaft 43 extending through the valve hole of the second valve 12.
  • the transmission shaft 43 is movable in a hole formed between the refrigerant passage 23 and the port 17 connected to the compressor outlet port.
  • the first valve 11 and the third valve 13 When the solenoid 14 is de-energized (Fig. 9), the first valve 11 and the third valve 13 have the valve element 20 and the shaft 29 urged by the spring 21 in respective directions of moving away from each other.
  • the valve element 20 is seated on the end face of the valve seat-forming member 42 and the shaft 29 is seated on the opening end of the hole holding the transmission shaft 43, both being fully closed.
  • the spring 28 urges the plunger 32 against the spring 34 away from the core 31.
  • the valve element 27 is moved downward.
  • the second valve 12 is fully open.
  • Refrigerant (pressure Pdh) from the discharge chamber is completely supplied from the port 25 to the crankcase via the refrigerant passage 23, and the second valve 12, so that the compressor is operated with the minimum capacity. If this de-energized state of the solenoid 14 corresponds to a state of the compressor which has stopped operation after firstly having operated with a predetermined capacity, the discharge pressure Pdh at port 16 becomes lower than the discharge pressure Pdl at the port 17 connected to the compressor outlet port so that the differential pressure (Pdh-Pdl) holds the first valve 11 (check valve functions) fully closed.
  • a predetermined energization current is supplied to control the compressor to a predetermined capacity, and that the control valve 10b is in a balanced state (Fig. 11). If the engine speed increases to increase the discharge pressure Pdh, the valve element 20 is lifted by an amount corresponding to the increase in the differential pressure across the first valve 11, thereby urging the valve element 27 of the second valve 12 via the shaft 29 in valve-opening direction. This increases the pressure Pc in the crankcase so that the compressor operates in the direction of reducing the capacity, whereby it is controlled to a predetermined discharge flow rate.
  • the control valve 10c in Fig. 12 (fourth embodiment) is distinguished from the third embodiment, in that a port 44 for introducing refrigerant into the second valve 12 is provided independently of the port 16 for introducing refrigerant into the first valve 11.
  • This port 44 is formed in a side of the second body 22. O-rings are provided on axially opposite sides of the valve element 27, with the port 44 located therebetween.
  • control valve 10c in Fig. 12 is applied to a variable displacement compressor equipped with an oil separator downstream of the discharge chamber whereby refrigerant at another discharge pressure Pdh2 is supplied from the oil separator to the port 44 instead.
  • the control valve 10d of Fig. 13 (fifth embodiment) is distinguished from the fourth embodiment in that the port 44 for introducing refrigerant into the second valve 12 and the port 25 for delivering refrigerant have reversed locations.
  • control valve 10d controls a flow rate determined by the passage area of the refrigerant passages 41 between the valve element 20 and the inner wall of the refrigerant passage 18 and by the differential pressure (Pdh-Pdl) on opposite sides of the refrigerant passage 41.
  • the value of the differential pressure (Pdh-Pdl) is set by the solenoid 14, and will be held at the predetermined value by the first valve 11 and the second valve 12 operating in an interlocked manner to control the pressure Pc in the crankcase. As a result, the flow rate through the first valve 11 to the compressor outlet port is held constant.
  • the control valve 10e in Fig. 14 (sixth embodiment) is distinguished from the fifth embodiment by an improvement of avoiding influences of internal refrigerant leakage on the variable displacement control.
  • a port 45 communicating with the suction chamber is formed between the port 17 receiving the discharge pressure Pdl and the port 25 (controlled crankcase pressure Pc). This lengthens the distance between the port 17 (discharge pressure Pdl) and the port 25 (controlled pressure Pc).
  • the transmission shaft 43 of the shaft 33 as well is lengthened, so that as an additional component a shaft 46 is interposed between the transmission shaft 43 and the third valve 13.
  • One end of the movable part of the solenoid 14 is supported by the transmission shaft 43 alone.
  • the third valve 13 is open, and hence even if the refrigerant at the discharge pressure Pdl leaks via the clearance between the shaft 46 and the first body 15 guiding the shaft 46, the leaked refrigerant flows via the port 45 into the suction chamber, but does not flow to the port 25 connected to the crankcase. Therefore, refrigerant leakage into the crankcase which could directly determine the displacement of the compressor does not occur, and hence the pressure Pc in the crankcase does not vary due to such leakage. This results in an accurate displacement control.
  • refrigerant leakage also may occur between the port 25 (controlled pressure Pc) and the port 45 (suction pressure Ps) through the clearance between the transmission shaft 43 and the guiding bore of the first body 15.
  • this clearance defines a smaller passage area than an orifice provided within the compressor at a location between the crankcase and the chamber pressure, for allowing refrigerant to flow from the crankcase into the suction pressure, and hence it does not adversely affect the displacement control of the compressor. If the passage area of the orifice is pre-set by taking the mentioned clearance into account, influences of refrigerant leakages through the clearance can be substantially eliminated.
  • the control valve 10f in Fig. 15 is distinguished from the fifth embodiment in that the valve element 27 of the second valve 12 and the transmission shaft 43 are formed separately from the shaft 33. Both are urged by the spring 28 in opening direction of the second valve 12. This makes it possible to form the valve element 27 relatively thin from a robust material bust, and to enhance the freedom of design.
  • the second body 22 and the core 31 of the solenoid 14 are integral, and the core 31 is press-fitted into the bottomed sleeve 30 which has an open end and a flange.
  • the outer periphery of the flange carries a packing 47 made of a material which is impervious to refrigerant penetration.
  • a screw thread 49 for mounting the control valve 10f in the compressor is formed on an outer peripheral portion, close to the flange, of a casing 48 serving as a yoke of the solenoid.
  • the variable displacement compressor in Fig. 16 includes a hermetically sealed crankcase 51 containing a driven rotating shaft 52. One end of the shaft 52 extends through a sealed bearing device to the outside of the crankcase 51. A pulley 53 transmits the drive force from the engine of the automotive vehicle to the shaft 52. A wobble plate 54 is fitted on the shaft 52, such that the inclination angle of the wobble plate 54 can be varied. Cylinders 55 (one of which is shown in Fig. 16) are arranged around the axis of the shaft 52. Each cylinder 55 contains a piston 56 converting the wobbling motion of the wobble plate 54 into reciprocation.
  • the cylinder 55 is connected via a suction relief valve 57 to a suction chamber 59 and via a discharge relief valve 58 to a discharge chamber 60.
  • the control valve 10f is disposed between the discharge chamber 60 and an outlet port 61 and between the discharge chamber 60 and the crankcase 51.
  • An orifice 62 is provided between the crankcase 51 and the suction chamber 59.
  • the compressor comprises a passage, (broken line in Fig. 16) extending from the discharge chamber 60 to the control valve 10f.
  • the control valve 10f is screwed into a mounting hole of the compressor.
  • the outlet port 61 is connected via a gas cooler 63 and an internal heat exchanger 64 by a high-pressure refrigerant conduit line to an expansion valve 65.
  • the expansion valve 65 is connected by a low-pressure refrigerant conduit line via an evaporator 66, an accumulator 67, and again the internal heat exchanger 64, to an inlet port communicating with the suction chamber 59.
  • the refrigeration cycle is a closed circuit.
  • Each piston 56 connected to the outer peripheral part of the wobbling wobble plate 54 reciprocates parallel to the axis of the shaft 52.
  • Refrigerant at suction pressure Ps in the suction chamber 59 is drawn into the cylinder 55 and is compressed, and is discharged at discharge pressure Pdh into the discharge chamber 60.
  • High-pressure refrigerant in the discharge chamber 60 is decompressed to discharge pressure Pdl when passing through the control valve 10f, and is delivered from the outlet port 61 to the gas cooler 64.
  • Part of the high-pressure refrigerant at the discharge pressure Pdh2 is introduced via the control valve 10f into the crankcase 51.
  • the pressure Pc rises whereby the inclination angle of the wobble plate 54 is set such that the bottom dead centre of the piston 56 is brought to a position where the pressure in the cylinder 55 and the pressure Pc in the crankcase 51 are balanced. Thereafter, refrigerant is returned from the crankcase 51 via the orifice 62 to the suction chamber 59.
  • the first valve 11 detects the flow rate between the discharge chamber 60 and the gas cooler 63.
  • the second valve 12 introduces a flow rate into the crankcase 51 dependent on the detected flow rate, thereby providing control such that the flow rate of the refrigerant sent from the discharge chamber 60 to the gas cooler 63 becomes constant. For example, when the engine speed increases, the discharge pressure Pdh rises. This increases the flow rate from the discharge chamber 60 to the gas cooler 63 via the control valve 10f, to increase the differential pressure across the first valve 11. According to an increase in the differential pressure, the second valve 12 opens, and the flow rate at discharge pressure Pdh2 introduced into the crankcase 51 also increases, whereby the pressure Pc in the crankcase 51 increases.
  • the wobble plate 54 inclination is varied until the wobble plate 54 forms a right angle with the shaft 52 to decrease the stroke of the pistons 56 to reduce the discharge flow rate.
  • the control valve 10f increases the flow rate into the crankcase 51 according to the increase in the flow rate of refrigerant, whereby the pressure Pc in the crankcase 51 is increased to reduce the displacement of the compressor. Therefore, the discharged flow rate of the compressor is controlled to be constant.
  • the compressor has the discharge flow rate controlled such that it is increased, whereby the discharge flow rate is controlled to be constant.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Compressors, Vaccum Pumps And Other Relevant Systems (AREA)
  • Control Of Positive-Displacement Pumps (AREA)
EP05019466A 2004-09-16 2005-09-07 Regelventil für einen Verdichter variabler Verdrängung Withdrawn EP1637736A2 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
JP2004269661 2004-09-16
JP2005161179A JP2006112417A (ja) 2004-09-16 2005-06-01 可変容量圧縮機用制御弁

Publications (1)

Publication Number Publication Date
EP1637736A2 true EP1637736A2 (de) 2006-03-22

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ID=35457502

Family Applications (1)

Application Number Title Priority Date Filing Date
EP05019466A Withdrawn EP1637736A2 (de) 2004-09-16 2005-09-07 Regelventil für einen Verdichter variabler Verdrängung

Country Status (3)

Country Link
US (1) US20060053812A1 (de)
EP (1) EP1637736A2 (de)
JP (1) JP2006112417A (de)

Cited By (1)

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GB2543832A (en) * 2015-10-30 2017-05-03 Aspen Pumps Ltd Pump driver

Families Citing this family (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP4695032B2 (ja) * 2006-07-19 2011-06-08 サンデン株式会社 可変容量圧縮機の容量制御弁
JP4833820B2 (ja) * 2006-12-25 2011-12-07 株式会社鷺宮製作所 容量制御弁および容量可変型圧縮機並びに空気調和装置
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KR101864881B1 (ko) * 2016-07-29 2018-06-05 동일기계공업 주식회사 가변 용량 압축기용 제어 밸브 및 이를 갖는 가변 용량 압축기

Family Cites Families (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP3783434B2 (ja) * 1998-04-13 2006-06-07 株式会社豊田自動織機 容量可変型斜板式圧縮機、及び空調用冷房回路
JP3583951B2 (ja) * 1999-06-07 2004-11-04 株式会社豊田自動織機 容量制御弁
US6390782B1 (en) * 2000-03-21 2002-05-21 Alumina Micro Llc Control valve for a variable displacement compressor
JP2002285956A (ja) * 2000-08-07 2002-10-03 Toyota Industries Corp 容量可変型圧縮機の制御弁
JP2002332962A (ja) * 2001-05-10 2002-11-22 Toyota Industries Corp 容量可変型圧縮機の制御弁
JP3943871B2 (ja) * 2001-07-25 2007-07-11 株式会社テージーケー 可変容量圧縮機および可変容量圧縮機用容量制御弁
JP2004251159A (ja) * 2003-02-19 2004-09-09 Sanden Corp 可変容量斜板式圧縮機の制御弁

Cited By (2)

* Cited by examiner, † Cited by third party
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