US20090123300A1

US20090123300A1 - System and method for controlling a variable displacement compressor

Info

Publication number: US20090123300A1
Application number: US11/794,998
Authority: US
Inventors: Jeffery Ross Uibel
Original assignee: Alumina Micro LLC
Current assignee: MQ ACQUISITION CORP; Dunan Microstaq Inc
Priority date: 2005-01-14
Filing date: 2006-01-11
Publication date: 2009-05-14
Also published as: EP1836399A1; CN101107446A; KR20070092328A; JP2008527244A; CN100591916C; WO2006076386A1

Abstract

A system and method for controlling the displacement of a variable displacement compressor by feeding back crankcase pressure as part of a control scheme is disclosed.

Description

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a national stage of International Application No. PCT/US2006/000880, which claims priority to U.S. Provisional Patent Application Ser. No. 60/644,097, filed Jan. 14, 2005. The disclosures of which are both incorporated herein by reference.

STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT

This invention was made with United States Government support under cooperative agreement number 70NANB2H3003 awarded by the National Institute of Standards and Technology (NIST). The United States Government has certain rights in the invention.

BACKGROUND OF THE INVENTION

1. Field of the Invention
This invention relates in general to control of variable displacement compressors.
2. Description of Related Art
U.S. Pat. No. 4,428,718 entitled “Variable Displacement Compressor Control Valve Arrangement” to Skinner (Skinner '718), the disclosures of which are hereby incorporated herein by reference, describes a variable displacement compressor, a conventional pneumatic control valve, general function of the variable displacement compressor, and interaction of the control valve with the compressor.
Referring now to the drawings, FIG. 1 shows a variable displacement refrigerant compressor 210 as described by Skinner '718. The variable displacement refrigerant compressor 210 is of the variable angle wobble plate type connected in an automotive air conditioning system having a normal condenser 212, an orifice tube 214, an evaporator 216 and an accumulator 218 arranged in that order between the compressor's discharge and suction sides. The compressor 210 comprises a cylinder block 220 having a head 222 and a crankcase 224 sealingly clamped to opposite ends thereof. A drive shaft 226 is supported centrally in the compressor 210 at the cylinder block 220 and the crankcase 224 by bearings. The drive shaft 226 extends through the crankcase 224 for connection to an automotive engine (not shown) by an electromagnetic clutch 236 which is mounted on the crankcase 224 and is driven from the engine by a belt 238 engaging a pulley 240 on the clutch 236.
The cylinder block 220 has five axial cylinders 242 through it (only one being shown), which are equally spaced about and away from the axis of the drive shaft 226. The cylinders 242 extend parallel to the drive shaft 226 and a piston 244 is mounted for reciprocal sliding movement in each of the cylinders 242. A separate piston rod 248 connects the backside of each piston 244 to a non-rotary, ring-shaped, wobble plate 250.
The non-rotary wobble plate 250 is mounted at its inner diameter 264 on a journal 266 of a rotary drive plate 268. The drive plate 268 is pivotally connected at its journal 266 by a pair of pivot pins (not shown) to a sleeve 276 which is slidably mounted on the drive shaft 226, to permit angulation of the drive plate 268 and the wobble plate 250 relative to the drive shaft 226. The drive shaft 226 is drivingly connected to the drive plate 268. The wobble plate 250 while being angularable with the rotary drive plate 268 is prevented from rotating therewith by a guide pin 270.
The angle of the wobble plate 250 is varied with respect to the axis of the drive shaft 226 between the solid line large angle position shown in FIG. 1, which is full-stroke, to the zero angle phantom-line position shown, which is zero stroke, to thereby vary the stroke of the pistons 244 and thus the displacement or capacity of the compressor 210 between these extremes. There is provided a split ring return spring 272 which is mounted in a groove on the drive shaft 226 and has one end that is engaged by the sleeve 276 during movement to the zero wobble angle position and is thereby conditioned to initiate return movement.
The working ends of the cylinders 242 are covered by a valve plate assembly 280, which is comprised of a suction valve disk and a discharge valve disk, clamped to the cylinder block 220 between the latter and the head 222. The head 222 is provided with a suction area 282, which is connected through an external port 284 to receive a gaseous refrigerant from the accumulator 218 downstream of the evaporator 216. The suction area 282 is open to an intake port 286 in the valve plate assembly 280 at the working end of each of the cylinders 242 where the refrigerant is admitted to the respective cylinders on their suction stroke each through a reed valve formed integral with the suction valve disk at these locations. Then on the compression stroke, a discharge port 288 open to the working end of each cylinder 242 allows the compressed refrigerant to be discharged into a discharge area 290 in the head 222 by a discharge reed valve which is formed integral with the discharge valve disk. The compressor's discharge area 290 is connected to deliver the compressed gaseous refrigerant to the condenser 212 from whence it is delivered through the orifice tube 214 back to the evaporator 216 to complete the refrigerant circuit as shown in FIG. 1.
The wobble plate angle and thus the compressor displacement can be controlled by controlling the refrigerant gas pressure in the sealed interior 278 of the crankcase 224 behind the pistons 244 relative to the suction pressure. In this type of control, the angle of the wobble plate 250 is determined by a force balance on the pistons 244. When the crankcase-suction pressure differential exceeds a set amount (the suction pressure control set-point) the net force on the pistons 244 results in a sufficiently large turning moment about the wobble plate pivot pins (not shown) that the wobble plate angle is reduced (i.e., moved toward the angle shown in phantom in FIG. 1) thereby reducing the compressor capacity by reducing the length of stroke of the pistons 244.
An important element of the variable displacement compressor 210 is a pneumatic control valve 300 inserted into the head portion 222 of the compressor 210. The control valve 300 senses an air conditioning load by sensing the pressure state of the refrigerant gas returning to the compressor 210 (the suction pressure). The control valve 300 is operably connected to the crankcase chamber 278. There are channels in the cylinder block 220 and the head 222 of the compressor 210 for gas flow between the control valve 300 and suction area 282, discharge area 290 and crankcase chamber 278 of the compressor 210. The control valve 300 controls the displacement of the piston 244 within the compressor 210 by controlling the pressure of gas in the crankcase chamber 278 that acts on the backside of the pistons 244 and the wobble plate 250.
The control valve 300 inserts into a stepped, blind control valve cavity 298 formed in the compressor head 222. The blind end of the control valve cavity 298 communicates directly with discharge area 290 through port 292. Control valve cavity ports 294 and 295 communicate with the crankcase chamber 278. A control valve cavity port 296 communicates with the suction area 282. The control valve 300 is sealed into the control valve cavity 298 so that particular features of the control valve 300 align with the ports 292, 294, 295 and 296.
U.S. Pat. No. 6,769,667 entitled “Control Valve for a Variable Displacement Compressor” to Kume et al. (Kume '667), the disclosures of which are hereby incorporated herein by reference, describes a control valve that uses electro/pneumatic control with a solenoid and a suction-pressure referenced bellows.
U.S. Pat. No. 6,390,782 entitled “Control Valve for a Variable Displacement Compressor” to Booth et al. (Booth '782), the disclosures of which are hereby incorporated herein by reference, describes a control valve that uses electro/pneumatic control with a solenoid and a suction-pressure referenced bellows.
A variable set point control valve (variable control valve) 10 is represented in the diagram of FIG. 2 according to the prior art disclosed in Booth '782. In FIG. 2, the variable control valve 10 is depicted in cross-sectional view and has a shape and feature placements appropriate to fit the control valve cavity 298 of the Skinner '718 variable displacement compressor described previously (see FIG. 1). The variable control valve 10 is coupled to a compressor 100, which compresses a gas. The variable control valve 10 controls the amount of gas and the degree to which it is pressurized in compressor 100. In the embodiment shown in FIG. 2, the gas compressed in compressor 100 is a refrigerant such as is used in an air conditioning unit, for instance, an air conditioning unit found in an automobile.
The variable control valve 10 comprises a compressor displacement control portion 30 and a variable set point control portion 80. The compressor displacement control portion 30 controls the flow of the gas from compressor 100 in and out of the variable control valve 10 while the variable set point control portion 80 controls the operation of the compressor displacement control portion 30. A valve body 12 of the variable control valve 10 is formed with many variable control valve functional elements, which will be described later. In the embodiment illustrated in FIG. 2, the valve body 12 is substantially cylindrical in shape as may be inferred from the cross-sectional view shown. O-ring retaining grooves 14 are indicated on the exterior of the valve body 12 in three locations. When the variable control valve 10 is inserted into a control valve cavity of a compressor 100 (see for example, FIG. 1), it is assembled with o-ring seals that allow different pressure sources to be communicated to different portions and ports of the variable control valve 10.
The compressor displacement control 30 comprises a suction pressure chamber 32 formed in the lower end 16 of the valve body 12 which is in gas communication with the suction area 120 of the compressor 100 through a variable control valve suction port 34 formed in the valve body 12 and a suction pressure path 112. A refrigerant circuit line 111 feeds low pressure gas into a compression chamber 114 of the compressor 100 via the suction area 120 and a compressor valve plate 126. The refrigerant circuit line 111 is a line returning low pressure refrigerant gas from the accumulator 144 of an air conditioning system (not shown).
The compressor 100 further comprises a piston 116, a crankcase chamber 118, and a discharge area 124. In simple terms, the operation of the compressor 100 is as follows. The refrigerant gas in the compression chamber 114 is compressed by the stroke of the piston 116 as the piston 116 moves towards the compressor valve plate 126. The compressor valve plate 126 admits high pressure gas to the discharge area 124. The refrigerant circuit line 111 is connected to the discharge area 124. The greater the displacement (stroke) 128 along the compression chamber 114 of the piston 116, the greater the pressure and the flow volume of the refrigerant gas as it passes through the compressor valve plate 126. The refrigerant gas then passes from refrigerant circuit line 111 to a condenser 140 where it condenses to a liquid in the condenser coils. The liquid then flows to an evaporator 142, where the liquid expands at an orifice within the evaporator 142, and evaporates. The air passing over the coils gives off heat energy that provides the energy for the state change from liquid to gas. The cooled air is then blown into the passenger cabin of the automobile, or into whatever chamber the air conditioning system is required to cool. After expanding, the refrigerant gas is in a low pressure state and is returned to the compressor 100 through the refrigerant circuit line 111.
The compressor 100 is a variable compressor, meaning that the stroke of piston 116 varies dependent upon the required air conditioning system load. For instance, if a user requires additional cooling of the air passing over the evaporator coils, the flow volume of the refrigerant discharged into the refrigerant circuit line 111 is increased. The stroke 128 of the piston 116 is increased to increase the flow volume.
A pressure is applied within the crankcase chamber 118 to the back of the piston 116. The greater the pressure within the crankcase chamber 118, relative to the suction pressure, the shorter the return stroke 128 of the piston 116 after compression due to the high pressure force exerted against the piston 116 on the return (away from the valve plate 126). Conversely, the lower the pressure within the crankcase chamber 118, relative to the suction pressure, the greater the return stroke of the piston 116 after compression due to the low pressure force exerted against the piston 116. By varying the pressure within the crankcase chamber 118, thus varying the displacement 128 of the piston 116 and ultimately the pressure of the discharge through refrigerant circuit line 111, the temperature of the air from the evaporator is controlled.
The compressor displacement control portion 30 has a middle chamber 40 formed as a bore centered in the valve body 12 leading from the suction pressure chamber 32. A first middle port 42 is formed in the valve body 12 and communicates with a middle chamber 40. The first middle port 42 is in gas communication with the crankcase chamber 118 through a first crankcase pressure path 130. The variable control valve 10 further comprises a pressure sensitive member, a diaphragm 36, exposed to the suction pressure chamber 32. A suction pressure valve, comprising a suction valve closing member, suction valve ball 38, and a suction valve seat 37 formed in the valve body 12, is provided to open and close a gas communication path between the suction pressure chamber 32 and the middle chamber 40.
The suction valve ball 38 is urged against the suction valve seat 37 by a rigid member 41, which is in floating contact with the diaphragm 36. A bias spring 44, retained in the middle chamber 40, urges the suction valve ball 38 off the suction valve seat 37, that is, urges the suction valve portion to open. It is also seen that the bias spring 44 opposes a movement of the diaphragm 36 towards the suction valve seat 37 and so acts as an equivalent pressure, a spring bias pressure, adding to the action of the suction pressure on the pressure receiving area of diaphragm 36. The variable control valve suction pressure valve opens and closes a gas communication path between the suction area 120 and the crankcase chamber 118 of the compressor 100.
A discharge pressure valve portion of the variable control valve 10 is comprised of a discharge valve member, a discharge valve ball 50, and a discharge valve seat 52 formed in the valve body 12. The discharge valve ball 50 is positioned in a discharge pressure chamber 60 formed in an upper end 18 of the valve body 12. A valve insert 64 has a stepped through bore 62 that positions the discharge valve ball 50 in alignment with the discharge valve seat 52. A ball centering spring 58 may be used to further condition the nominal position of the discharge valve ball 50. A particle filter cap 74 sealably covers the end of valve body 12, completing the discharge pressure chamber 60. When the variable control valve 10 is inserted into the compressor 100, the upper end 18 of the valve body 12 is sealed in a blind end of a control valve cavity such as the cavity 298 illustrated in FIG. 1. A discharge pressure path 110 from the discharge area 124 of the compressor is communicated to the blind end of the control valve cavity 298. Discharge pressure gas is thereby communicated to the variable control valve discharge pressure chamber 60 through the filter 74.
The variable control valve 10 has a central stepped bore 70 formed through the valve body 12. The central bore 70 has a large diameter bore portion at the upper end adjacent the discharge chamber 60 where at the discharge valve seat 52 is formed. The central bore 70 and the middle chamber 40 are aligned with each other. A second middle port 56 is formed in the valve body 12 and communicates with the large bore portion of the central bore 70. The second middle port 56 is in gas communication with the crankcase chamber 118 through the second crankcase pressure path 132. When the discharge valve ball 50 is moved off the discharge valve seat 52, discharge pressure gas can flow through the bore 70 to the second middle port 56 and then to the crankcase chamber 118 via the second crankcase pressure path 132.
A valve rod 54, inserted in the central bore 70, partially links the actions of the suction valve portion and the discharge valve portions of the variable control valve 10. The valve rod 54 has a diameter slightly smaller than the small bore portion of the central bore 70. The valve rod 54 freely slides in the central bore 70 yet substantially blocks gas communication between the middle chamber 40 and the discharge chamber 60. The length of the valve rod 54 is chosen so that it simultaneously touches the seated discharge valve ball 50 and the suction valve ball 38 in a fully open (fully unseated) position. This arrangement links the suction and discharge valve portions in a partial open-close relationship. As the suction valve ball 38 moves in a valve-closing direction, the valve rod 54 pushes the discharge ball 50 in a valve-opening direction. As the discharge valve ball 50 moves in a valve closing direction, the valve rod 54 pushes the suction ball 38 in a valve-opening direction.
In the embodiment of FIG. 2, the valve rod 54 is not attached to either of the valve closing balls 38, 50. The valve rod 54 operates to open either the discharge or the suction valve portions of the variable control valve but not to close either. The forces which act to close the discharge valve portion are the pressure of the discharge gas on an effective pressure receiving area of the discharge valve ball 50 and a small spring force imparted by a ball centering spring 58. The force that acts to close the suction pressure valve portion derives from a movement of the pressure sensitive diaphragm 36 via the rigid member 41. Other embodiments of the prior art of U.S. Pat. No. 6,390,782 in which both valve closing members are attached to a coupling means such as the valve rod 54 will be apparent to those skilled in the control valve art. If both valve members are rigidly linked, then a full open-close relationship will exist.
Reference is made specifically now to the variable set point control portion 80 of the variable control valve 10. The variable set point control 80 comprises a closed reference chamber 90 bounded by the variable control valve diaphragm 36, walls 91 formed at the lower end 16 of the valve body 12 when the suction pressure chamber 32 was formed, and a valve end cap 20. The diaphragm 36 is positioned and sealed against an interior step 93 in the suction pressure chamber 32 by a reference valve carrier 81. The diaphragm 36 has a first side 43 with a suction pressure receiving area exposed to suction pressure in the suction-pressure chamber 32 and a second side 39 with a reference pressure receiving area exposed to the reference pressure in the reference chamber 90. The diaphragm 36 is arranged to seal the reference chamber 90 from direct gas communication with the suction pressure chamber 32, the discharge pressure chamber 60, the middle chamber 40, or the central bore 70.
Two pressure bleed passageways, a discharge bleed passageway 68 and a suction bleed passageway 72 are provided in the valve body 12 and align with two holes in the diaphragm 36 that is sealed against a valve body interior step 93. A valve insert 64 has a valve insert bleed hole 69 provided to communicate the discharge chamber 60 with the discharge bleed passageway 68. The bleed passageways, the valve insert bleed hole, and the corresponding diaphragm holes, provide a source of suction pressure gas and discharge pressure gas to the reference chamber 90. The feature depicted of supplying the discharge pressure gas to the reference chamber 90 from the variable control valve discharge pressure chamber 60 is important because this design uses a filter 74 to protect the components and passages in reference chamber 90 from foreign material.
The variable control valve components contained in the reference chamber 90 are illustrated more clearly in FIG. 3. The reference chamber valve means are further illustrated at higher detail level in FIG. 4. The same elements in FIGS. 2-4 are labeled with the same numbers.
Referring now to FIGS. 2-4, the reference valve carrier 81 is formed as a thick-walled cylinder with outside walls that sealably fit against the interior of walls 91 formed at the lower end 16 of the valve body 12. The upper end of reference valve carrier 81 seals against the diaphragm 36. Two small blind chambers, a suction bleed chamber 96 and a discharge bleed chamber 98 are formed in the reference valve carrier 81 from the upper end that is sealed against the diaphragm 36. The open end of the suction bleed chamber 96 aligns with the suction bleed passageway 72 and the open end of the discharge bleed chamber 98 aligns with the discharge bleed passageway 68. Reference chamber valve means are generally indicated as a reference inlet valve 88 and a reference outlet valve 86.
Turning to FIG. 4, the reference inlet valve 88 is comprised of a reference inlet valve closing member 162, a reference inlet through hole 160, and a reference inlet valve seat 164. The reference inlet through hole 160 is formed from an interior surface of the cylindrical reference valve carrier 81 through to the discharge bleed chamber 98. The reference inlet valve seat 164 is formed around the inlet through hole 160 where it emerges from the reference valve carrier 81 that is into the reference chamber 90. The reference inlet valve closing member 162 is attached to an inlet valve push rod 167, which is part of an inlet solenoid actuator 94. When an electrical current signal is applied to inlet solenoid leads 85, an inlet valve push rod 167 is pulled into the center of the solenoid actuator 94, urging reference the inlet valve closing member 162 against the reference inlet valve seat 164, closing off the reference inlet through hole 160. The reference inlet through hole 160 communicates reference chamber 90 with discharge bleed chamber 98. Thus, opening and closing the reference inlet valve 88 by means of electrical signals applied to the inlet solenoid actuator 94 controls the flow of discharge pressure gas to the reference chamber 90.
An inlet solenoid leaf spring 168 is arranged to bias the inlet valve push rod 167 in a retracted position as is illustrated in FIG. 4. This inlet solenoid spring bias configuration means that the reference inlet valve 88 will open the reference chamber 90 to the flow of discharge pressure gas in the absence of an electrical signal to energize the coil of the inlet solenoid actuator 94. The depicted reference inlet valve is said to be normally open. The opposite arrangement of spring biasing the reference inlet valve 88 to a normally closed condition is an alternate configuration of the reference inlet valve means that may also be employed successfully in another embodiment of the prior art of U.S. Pat. No. 6,390,782.
The reference outlet valve 86 is comprised of a reference outlet valve closing member 172, a reference outlet through hole 170, and a reference outlet valve seat 174.
The reference outlet through hole 170 is formed from an interior surface of the cylindrical reference valve carrier 81 through to the suction bleed chamber 96. The reference outlet valve seat 174 is formed around the outlet through hole 170 where it emerges from the reference valve carrier 81 that is into the reference chamber 90. The reference outlet valve closing member 172 is attached to an outlet valve push rod 177, which is part of the outlet solenoid actuator 92. When an electrical current signal is applied to the outlet solenoid leads 87, the outlet valve push rod 177 is pulled into the center of the solenoid actuator 92, pulling the reference outlet valve closing member 172 away from the reference outlet valve seat 174, opening the reference outlet through hole 170. The reference outlet through hole 170 communicates the reference chamber 90 with the suction bleed chamber 96. Thus, opening and closing the reference outlet valve 86 by means of electrical signals applied to the outlet solenoid actuator 92 controls the flow of suction pressure gas to the reference chamber 90.
An outlet solenoid leaf spring 178 is arranged to bias the outlet valve push rod 177 in an extended position as is illustrated in FIG. 4. This outlet solenoid spring bias configuration means that the reference outlet valve 86 will close the reference chamber 90 to the flow of suction pressure gas in the absence of an electrical signal to energize the coil of the outlet solenoid actuator 92. The depicted reference outlet valve 86 is therefore normally closed. The opposite arrangement of spring biasing the reference outlet valve 86 to a normally open condition is an alternate configuration of the reference outlet valve means that may also be employed successfully in another embodiment of the prior art of U.S. Pat. No. 6,390,782
It should also be appreciated that, while solenoid actuators are discussed herein and depicted in FIGS. 2-4, any electrically-driven physical actuator means could be employed to open and close the reference inlet valve 88 and the reference outlet valve 86.
The variable set point control portion 80 further comprises an electronic control unit 82, a pressure sensor 84, an electrical circuit carrier 83, and variable control valve electrical leads 89. The pressure sensor 84 is an optional feature of the embodiment of the prior art of U.S. Pat. No. 6,390,782. It is a transducer device that produces an electrical signal that is related to a gas pressure impinging on its sensitive portion. The pressure sensor 84 is mounted on the electrical circuit carrier 83 so as to respond to the gas pressure within the closed reference chamber 90. It is not necessary for the practice of the prior art of U.S. Pat. No. 6,390,782 that the pressure sensor 84 be mounted directly in the interior of the reference chamber 90. An alternative embodiment could mount the pressure sensor at some other position as long as the pressure sensitive portion of the sensor 84 is brought into gas communication with the reference chamber 90.
The electronic control unit 82 is an optional feature of the embodiment of the prior art of U.S. Pat. No. 6,390,782. The control unit 82 may contain electronic circuitry to control the reference chamber valve means or to receive and process the electrical signals produced by the pressure sensor 84. In an embodiment of this optional feature of the prior art of U.S. Pat. No. 6,390,782, the electrical components of the control unit 82 are co-located with the pressure sensor 84 by means of the electrical circuit carrier 83. Other functions of the optional control unit 82 will be described later.
The variable control valve electrical leads 89 are routed from the electrical circuit carrier 83 through a sealed opening in the valve end cap 20. The number of electrical leads needed by the variable control valve 10 will depend on the functions performed by the optional electronic control unit 82 and the device characteristics of the optional pressure sensor 84. When neither the electrical control unit 82 nor the reference chamber pressure sensor 84 are employed, then variable control valve electrical leads 89 need comprise only those needed to carry electrical signals to activate the reference chamber valve means.
The variable set point control portion 80 controls the operation of the compressor displacement control portion 30. By controlling a pressure within the reference chamber 90, the variable set point control 80 is able to regulate the open/close conditions of the suction pressure valve portion and the discharge pressure valve portion of the variable control valve 10. For instance, if the pressure in the reference chamber 90 exerts a force against the diaphragm 36 which is less than the force exerted by the pressure in the suction pressure chamber 32 and the bias spring 44, the diaphragm 36 will distort into the reference chamber 90, that is in the direction of the reference inlet let actuator 94. This motion moves the suction valve ball 38 from the suction valve seat 37, thus opening the flow of gas from the first crankcase pressure path 130 to the suction pressure chamber 32. At the same time, the discharge pressure valve portion is closed by the pressure of discharge gas forcing the discharge valve ball 50 onto the discharge valve seat 52. By opening the flow through the suction valve portion of the variable control valve 10, gas from the crankcase chamber 118 will flow into the suction pressure chamber 32 and out to the suction area 120 of the compressor 100 via the suction pressure path 112. With the bleeding of gas out of the crankcase chamber 118, less force is exerted on piston 116 giving piston 116 greater displacement. The flow of refrigerant gas flowing into the evaporator of the system is thus increased.
If the pressure in the reference chamber 90 exerts a force against diaphragm 36 which is greater than the force exerted by the pressure in the suction pressure chamber 32 and the bias spring 44, the diaphragm 36 will distort into the suction pressure chamber 32, that is, in the direction of the suction valve seat 37. This action closes the variable control valve suction valve portion and, at the same time, opens the variable control valve discharge valve portion by pushing the discharge valve ball 50 away from the discharge valve seat 52 by means of the valve rod 54. As the discharge valve portion is opened, high pressure gas from the discharge pressure path 110 flows through the discharge pressure chamber 60, the stepped central bore 70, the second middle port 56 and the second crankcase pressure path 132 to the crankcase chamber 118. Pressure will build up in the crankcase chamber 118, thus applying a force against the piston 116. The displacement 128 of the piston 116 is thus restricted and the amount of the refrigerant gas passing into the evaporator of the system is reduced.
The force that the bias spring 44 exerts on the diaphragm 36 is an important design variable for the overall performance of the variable control valve 10. It has been found through experimentation that it is most beneficial if the spring force is adjusted to be equivalent to from 2 to 20 psi of suction pressure, and most preferably, from 4 to 10 psi. This range of spring bias force allows for sufficient operational range of the variable control valve 10 in the condition of very low compressor capacity usage, that is, when the compressor is near full de-stroke operation.
The pressure within the reference chamber 90 is controlled by the opening and closing of the reference outlet valve 86 and the reference inlet valve 88. Each of these are optionally controlled by the pressure sensor 84 and the electronic control unit 82. Specifically, the pressure within the reference chamber 90 is in gas communication with the pressure sensor 84. The pressure sensor 84, interfaced to the electronic control unit 82, measures the pressure of the gas in the reference chamber 90 and communicates that pressure to the electronic control unit 82. The electronic control unit 82 receives control signals and information from a compressor control unit 146. Passenger comfort level settings and other information about environmental conditions and vehicle operation conditions are received by the compressor control unit 146. The compressor control unit 146 uses stored compressor performance algorithms to calculate a necessary amount of gas to be compressed within the compression chamber 114 by the piston 116 to cause a desired condition to occur, namely that the passenger comfort level settings are optimally achieved within the constraints imposed by environmental and vehicle operational factors.
The calculated compressor displacement requirements, the pressure information from pressure sensor 84, and known physical response characteristics of the variable control valve 10 elements are utilized by variable control valve performance algorithms to calculate a necessary pressure within the reference chamber 90 to meet the compressor displacement requirements. This calculated reference pressure, necessary to meet the requirements determined by the compressor control unit, is called a predetermined reference pressure. The variable displacement compressor 100 is thereby controlled by the determining of the predetermined reference pressure and the maintenance of the gas pressure in the reference chamber 90 to this predetermined pressure level.
Alternatively, if the pressure sensor 84 is not employed, the predetermined reference pressure may be selected from a stored set of reference pressure levels that has been pre-calculated based on the known nominal characteristics of the variable control valve 10 or, in addition, customized for the variable control valve by means of a calibration set-up procedure. In the case of this alternate embodiment of the prior art of U.S. Pat. No. 6,390,782, the calculated compressor displacement requirements are used to determine, in look-up table fashion, the predetermined reference pressure that is optimal for achieving the desired compressor displacement control.
Control of the reference outlet valve 86 and the reference inlet valve 88 comes from the electronic control unit 82 through the actuators 92 and 94, respectively. Dependent upon the outputs of the algorithms within the electronic control unit 82, the electronic control unit 82 will open and close the reference outlet valve 86 by actuating the outlet actuator 92 and open and close the reference inlet valve 88 by the inlet actuator 94. For instance, when the pressure within the reference chamber 90 is to be increased, the inlet actuator 94 will retract the reference inlet valve member 162 allowing high pressure gas to flow from the discharge pressure chamber 60 through the valve insert bleed hole 69, the discharge pressure bleed passageway 68 and the discharge bleed chamber 98 into the reference chamber 90. At the same time, the outlet actuator 92 closes the reference outlet valve 86, thus allowing the pressure in the reference chamber 90 to increase. Inversely, to decrease the pressure in the reference chamber 90, the electronic control unit 82 will actuate the outlet actuator 92 to retract the reference outlet valve member 172 to open flow from the reference chamber 90 through the suction bleed chamber 96 to the suction pressure bleed passage 76 to the suction pressure chamber 32, thereby bleeding off pressure. At the same time, the actuator 94 is signaled by the electronic control unit 82 to extend the reference inlet valve member 162 to close off discharge pressure flow into the reference chamber 90.
By controlling the pressure within the reference chamber 90 to the predetermined reference pressure, the electronic control unit 82, through the actuators 170 and 172, controls the deflection of the diaphragm 36, thus controlling the varying of the displacement 128 of the piston 116. For the embodiment depicted in FIGS. 2-4, the reference chamber pressure can be continuously or periodically monitored by means of the pressure sensor 84. This pressure information can be used as a feedback signal by the control unit 82 in a pressure servo control algorithm to maintain the reference chamber 90 at the predetermined reference pressure within chosen error boundaries.
It is anticipated that an important benefit of the variable control valve design disclosed herein is the ability to maintain valve control performance by tightly maintaining the predetermined reference pressure. The disclosed design also enables the system to electronically change the predetermined reference pressure to a different value, thereby changing the suction pressure set-point about which the variable displacement compressor operates. This allows the vehicle to adjust the compressor control in the face of changing environmental factors to achieve a desired balance of passenger comfort and vehicle performance. The benefits of the balance of passenger comfort and vehicle performance are more fully realized the more responsive the pressure in the reference chamber is to the control.
The responsiveness of the reference pressure control system depends in part on the characteristics of the flow of discharge pressure gas through inlet valve 88 and the flow out of the outlet valve 86 to suction pressure. FIGS. 5 and 6 illustrate some important geometrical feature details of the reference inlet valve 88 and the reference outlet valve 86.
Referring first to FIG. 5, the inlet valve closing member 162 is illustrated in a fully closed position holding off the force of discharge pressure gas impinging an effective pressure receiving area, AI, on the inlet valve member 162. Also indicated in FIG. 5 is the diameter, DI, of the reference inlet port 160 leading from the discharge bleed chamber 98. A large value of DI will promote quick response to commands to increase reference chamber pressure by admitting a large flow of discharge pressure. The size of DI needed to achieve a given reference chamber pressure rise time will depend on the reference chamber gas volume. A larger reference inlet port 160 will be required for a larger reference chamber gas volume to achieve the same increase in reference chamber pressure rise time as for a smaller reference chamber gas volume.
However, a large value of DI necessitates a correspondingly large value of AI, the effective inlet valve member pressure receiving area. This, in turn, would mean that the closing force that would be needed from the inlet valve actuator 94 would also be large. A large closing force might require a physically large actuator or require excessive power to maintain the inlet valve in a closed state. Consequently, the choice of the reference inlet port 160 diameter, DI, and the pressure receiving area, AI, involves a balance of competing requirements.
The effective inlet valve member pressure receiving area, AI, is the net, unbalanced, area of the inlet valve closing member that is exposed to the discharge pressure when the inlet valve is fully closed. That is, the area that effectively receives the force of the discharge pressure, AI, may be calculated by measuring the force exerted on the inlet valve closing member by the discharge pressure, and dividing by the discharge pressure. It has been found through experimentation effective inlet valve pressure receiving area, AI, may be beneficially chosen to be less than 30,000 square microns and preferably, less than 7500 square microns when the reference chamber gas volume is approximately 2 milliliters. Under typical automotive air conditioner compressor operating conditions, a reference inlet valve closing force of less than 1 pound will suffice if the effective inlet valve member pressure receiving area, AI, is less than approximately 7500 square microns.
Referring to FIG. 6, the outlet valve closing member 172 is illustrated in a fully open position with gas flowing out of the reference chamber 90 through an effective gas flow area. Many geometrical designs of the reference outlet port 170 may be chosen to have the same result in terms of the gas volume flow for a given pressure differential between the reference chamber 90 and the suction bleed chamber 96. The effective flow area is chosen to balance competing performance characteristics. In order to insure quick response to a command to lower the reference chamber pressure, it is desirable to have a large outlet valve 86 effective flow area On the other hand, to help restrain rapid pressure increases in the reference chamber when opening the inlet valve 88 to discharge pressure, and to bring down reference pressure overshoots that may occur, it is helpful to have a small outlet valve 86 effective flow area
The effective gas flow area of the reference outlet valve 86 may be beneficially chosen as a ratio to the effective flow area of the inlet valve 88. Alternatively, the diameter, DO, of the reference outlet port 170, may be chosen as a ratio of the reference inlet port 160 diameter, DI. It has been determined by experimentation and analysis that the beneficial range of the ratio DO to DI is from 0.5 to 5.0, and, most preferably, from 0.7 to 2.0. The corresponding beneficial ratio of inlet-port to outlet-port cross-sectional areas, the inlet-to-outlet port areal ratio, is 0.25 to 25.0, and, most preferably, 0.5 to 4.0. When the geometries of inlet and outlet gas flow areas are more complex than the circular passageways illustrated in FIGS. 5 and 6, the gas flow cross-sectional areas may be analyzed or experimentally determined and the inlet-to-outlet port areal ratio design guideline followed.
It has been found through experimentation, for example, that when the reference chamber 90 gas volume is approximately 2 milliliters, a reference outlet port 170 diameter DO of 100 microns is an effective choice when the reference inlet port 160 diameter DI is 100 microns, a reference outlet port diameter to reference inlet port diameter ratio of 1.0. With these parameter values, and under typical automotive air conditioner compressor operating conditions, the reference chamber pressure can be controllably changed, or tracked to a predetermined reference pressure, at the rate of 10 psi/second.
For alternative embodiments of the variable control valve 10 without a pressure sensor, the compressor control unit 146 may periodically recalculate the compressor displacement conditions required to maintain performance of the cooling system. Based on the magnitude and time behavior of changes in these calculations, the compressor control unit 146 may send instruction signals to the variable control valve electronic control unit 82 to increase or decrease the reference chamber pressure to re-establish the pre-determined reference pressure level. It will be appreciated by those skilled in the art that this method of affecting servo control of the pressure in the reference chamber to the predetermined level will be less timely than can be implemented using a direct measurement of reference chamber pressure. Nonetheless, this loose-servo method can be effective and appropriate for a low cost embodiment of the prior art of U.S. Pat. No. 6,390,782.
The functions attributed to the variable control valve electronic control unit 82 and the compressor control unit 146 could be performed by other computational resources within the overall system employing the variable control valve 10, the compressor 100 and the cooling equipment. For example, if the overall system is an automobile with a central processor, then all of the control information and calculations needed to select and maintain the predetermined reference pressure could be gathered and performed by the automobile central processor. Signals to and from the pressure sensor 84 could be routed to an input/output port of the central processor and the reference inlet and outlet valve actuation signals could be sent to the variable control valve 10 from another input/output port of the central processor. Alternately, a compressor control unit 146 could perform all the control functions needed to manage the variable control valve 10. And finally, the variable control valve control unit 82 could be provided with circuitry, memory and processor resources necessary to perform the compressor displacement requirement calculation as well as selecting and maintaining the predetermined reference pressure

BRIEF SUMMARY OF THE INVENTION

This invention relates in general to variable displacement compressors and more specifically to systems and methods for controlling the displacement of a variable displacement compressor. In one embodiment of the present invention, displacement of a variable displacement compressor is controlled by utilizing crankcase pressure.
Various objects and advantages of this invention will become apparent to those skilled in the art from the following detailed description, when read in light of the accompanying drawings.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

FIG. 1 is a cross-sectional view of a variable displacement compressor for use in an automobile from the prior art of U.S. Pat. No. 4,428,718.

FIG. 2 is a cross-sectional view of a variable set point control valve according to a preferred embodiment of the prior art of U.S. Pat. No. 6,390,782.

FIG. 3 shows a cross-section of the variable set point control portion of the variable control valve of FIG. 2.

FIG. 4 shows a cross-section of the reference chamber valve means of the variable control valve of FIGS. 2 and 3.

FIGS. 5 and 6 show cross-sections of the valve members and valve seats of the reference chamber valve means of the variable control valve of FIGS. 2-4.

FIG. 7 is a cross-sectional view of a variable set point control valve according to a first embodiment of the present invention.

FIG. 8 is a cross-sectional view of a variable set point control valve according to a second embodiment of the present invention.

FIG. 9 is a flow chart illustrating a method for controlling the displacement of a variable displacement compressor by utilizing crankcase pressure according to the present invention.

FIG. 10 is a schematic illustration of an arrangement for pneumatic control of a variable displacement compressor according to another embodiment of the invention.

FIG. 11 is a schematic diagram of a compressor system according to the present invention.

DETAILED DESCRIPTION OF THE INVENTION

Referring again to the drawings, there is illustrated in FIG. 7 a system 410 according to a first embodiment of the present invention. FIG. 7 is a view generally similar to FIG. 2, except as will be discussed below, and similar components are labeled with the same numbers.
The system 410 includes a dividing wall 412. The dividing wall 412 separates the reference chamber 90 into two isolated chambers, an upper chamber 490 a and a lower chamber 490 b. The upper chamber 490 a is defined on opposite ends by the diaphragm 36 and the dividing wall 412. The lower chamber 490 b is defined on opposite ends by the dividing wall 412 and the valve end cap 20.
A lower crankcase port 414 is in fluid communication with the crankcase chamber 118 through a third crankcase pressure path 416. Thus, the pressure sensor 84 is operable to measure the actual pressure of the crankcase chamber 118. As used in this application the term “pressure sensor” means any sensor which measures pressure or any other parameter from which pressure can be inferred. The system 410 controls crankcase pressure to control the displacement of the compressor 100, for example by a method that will be discussed below.
FIG. 8 is a view similar to FIG. 7, except showing a system 420 for controlling the displacement of the compressor 100 by utilizing pressure of the crankcase chamber 118, according to a second embodiment of the present invention, and similar components are labeled with the same numbers.
The system 420 includes a sensor 422. The sensor 422 is an electronic pressure sensor, although such is not required. The sensor 422 is located within the crankcase chamber 118 and is operable to measure the actual crankcase pressure. It must be understood, however, that the sensor 422 need not be located entirely inside the main cavity of the crankcase chamber 118. For example, the sensor 422 may be located in a passage, subchamber, or any other suitable location where the sensor 422 may sense the pressure in the crankcase chamber 118. The sensor 422 is located such that the sensor 422 does not interfere with movement of the piston 116 or other moving parts within the compressor 100.
The sensor 422 is electrically connected to the compressor control unit 146 by a sensor lead 424 for electrical communication. In the present embodiment, the sensor 422 measures pressure in the crankcase 118 and transmits a pressure sensor signal to the control unit 146. The control unit 146 then alters a control signal to the variable control valve 10 to effect a change in the position of the variable control valve 10 based upon the pressure sensor signal received. The system 410 is thus responsive to the measured actual crankcase pressure to control the displacement of the compressor 100.
The present embodiments have been described as having electrical control of the displacement of the compressor 100. This electrical control may be achieved, for example, by use of a computer chip, a pressure transducer measuring actual crankcase pressure, and an electrically actuated device. In such a case, electrical control may be achieved by the computer chip setting a target crankcase pressure, and comparing the target crankcase pressure and the actual crankcase pressure, and determining an amount to move the variable control valve 10 based on the difference of the actual crankcase pressure from the target crankcase pressure. Then electrically actuated device or devices, such as a microvalve or a solenoid operated macro-sized valve, such as the reference inlet valve 88 and/or the reference outlet valve 86 are actuated in a manner to cause the desired control valve response (i.e., as described above in the Background Of The Invention) to change the actual crankcase pressure. For electronic control, where a computer chip is used, a program is preferably written that can cause the variable control valve 10 to respond in a fine-tuned method, as will be further described below.
Although the present embodiments have been described as having electrical control, the system may have, and the method may use, any suitable control, such as pneumatic control, electro-pneumatic control, hydraulic control, or any other suitable control.
An arrangement for pneumatic control of an air conditioning system 600 (which is only partially illustrated) is schematically illustrated in FIG. 10. As shown, a diaphragm or bellows 610 controls the position of a three way valve 620 suitable to selectively increase the pressure in the crankcase 118, maintain the pressure in the crankcase 118, or decrease the pressure in the crankcase 118. In the illustrated example, the interior of the bellows is controlled at a reference pressure to generate a target crankcase pressure. The reference pressure may be generated in any suitable manner, including a variable pneumatic pressure regulator. As illustrated, however, the target crankcase pressure is generated by a thermostatic system 630 with a sensor bulb 632 disposed in an air stream of the air conditioning system generated by a fan 634. The bulb 632 is located in the outlet air stream downstream of the evaporator 216, so that when temperature of the air stream rises, pressure in the bulb 632 (and thus inside the bellows 610) will rise, as a fluid medium inside the bulb 632 heats up and expands. Conversely, when temperature of the air stream falls, pressure inside the bulb 632 and thus target crankcase pressure inside the bellows 610 will also fall. Thus, in this instance, the reference pressure is an inverse of a target crankcase pressure, since when an increase in cooling is desired, as would be the case when temperature of the air stream around the bulb 632, the crankcase pressure of the compressor 100 must go down to increase stroke of the pistons and increase cooling. It should also be noted that, like many conventional thermostatic systems, the thermostatic system 630 may be provided with a mechanism (not illustrated) for adjusting the response of the air-conditioning s system 600 to a given temperature of the air stream within which the bulb 632 is disposed, to enable, for example, a user to select a cooler or warmer airflow from the air-conditioning system 600.
The bellows 610 is disposed within a chamber 640 that is in fluid communication with the crankcase 118, and thus contains actual crankcase pressure. The bellows 610 thus expands or contracts in response to the difference in pressure between the reference pressure within the bellows 610 and the actual crankcase pressure outside the bellows 610, within the chamber 640. The moving end of the bellows 610 is mechanically connected to a moving element 650 of the valve 620.
The valve 620 has an inlet chamber 622 in fluid communication with the discharge 652 of the compressor 100, and an outlet chamber 654 in fluid communication with the suction 120 of the compressor 100, and a load chamber 656 in fluid communication with the crankcase 118.
As the bellows 610 expands, when to the reference pressure is greater than the actual crankcase pressure, the three way valve 620 is adjusted rightward, as viewed in FIG. 10, toward an decrease of pressure position, in which the crankcase 118 is connected via the chambers 656 and 654 of the valve 620 to the suction flow path of the compressor 100, and more specifically to the suction 120. Since the pressure at the suction 120 is the lowest pressure in the air conditioning system, the pressure in the crankcase 118 falls when the crankcase 118 is vented thusly to the suction 120.
As the bellows 610 contracts, when the reference pressure is less than the actual crankcase pressure, the three way valve 620 is adjusted leftward, as viewed in FIG. 10, toward an increase of pressure position, in which the crankcase 118 is connected via the chambers 656 and 652 of the valve 620 to the discharge flow path of the compressor 100, and more specifically to the discharge 124. Since the pressure at the discharge 124 is the highest pressure in the air conditioning system, the pressure in the crankcase 118 rises when the crankcase 118 is connected to the discharge 124.
It will be apparent that the reference pressure could be connected to the chamber 640, outside the bellows 610, and the crankcase 118 connected instead to the inside of the bellows 610, in the reverse of what is shown in FIG. 10. In such a case, all that need be done is to connect the discharge 124 to the chamber 654 and the suction 120 to the chamber 652 for this alternative embodiment to work generally as described above.
Additionally, instead of the illustrated thermostatic system 632, it is contemplated that any suitable arrangement may be provided to generate the reference pressure.
An alternative arrangement which might be suitably substituted for the bellows 610 and the valve 620, would be a three-way pressure actuated microvalve. One valve which might be suitably adapted is the pilot-operated valve 10 described in U.S. Pat. No. 6,694,998 (the '998 patent), the disclosures of which are incorporated herein by reference. (Note: In the following discussion, the reference numbers refer to components previously discussed in this disclosure, unless specifically noted as being reference numbers ‘of the '998 patent’). A reference pressure (either from a thermostatic system 632 or from any other suitable arrangement, including an electrically controlled pilot microvalve, such as the microvalve 9 of the '998 patent) is introduced into the control chamber 125 of the '998 patent, and acts against the axial face of the second end 276 (of the '998 patent) of the slider element 240 of the '998 patent. The port 220 of the '998 patent is connected to the discharge 124 of the compressor 100, the port 230 is connected to the suction 120 of the compressor 100, and the port 226 of the '998 patent is connected to the crankcase 118. As described in the '998 patent, the slider element 240 of the '998 patent will operate to maintain the port 226 of the '998 patent at the pressure in the control chamber 125 of the '998 patent, in this case, the reference pressure (the desired target crankcase pressure, in this case).
Furthermore, while the valve 620 is shown as a three-way valve directly actuated by the bellows 610 and directly controlling the connection between the crankcase 118 and the suction 120 and the discharge 124, a pilot valve and pilot operated valve arrangement is also contemplated. The pilot valve (not shown) could be operated by the difference in pressure between the reference pressure and the actual crankcase pressure to direct fluid pressures to the pilot operated valve. The pilot operated valve (not shown) could be selectively positioned by the fluid pressures of the pilot valve to a pressure increase position connecting the discharge 124 to the crankcase 118, a pressure decrease position connecting the suction 120 to the crankcase 118, and a pressure hold position, in which the crankcase 118 is isolated from the suction and discharge flow paths. Such a pilot and pilot-operated valve arrangement could otherwise operate generally similarly to the system 600 shown in FIG. 10. It is also contemplated that the pilot valve and possibly the pilot operated valve could be microvalves. Of course, the suitability of using microvalves to substitute for the valve 620 illustrated in FIG. 10, either as discussed in this paragraph, or in the previous paragraph depend upon the flow requirements of the particular system in which they are to be installed, and the flow capacity of the valves themselves.
In the case of electro-pneumatic control, the reference pressure in a reference chamber (not shown) may set to obtain to a desired crankcase pressure, similarly as described above for pneumatic control. A diaphragm or bellows may effectively measure the difference between the reference pressure and the actual crankcase pressure by the amount of expansion or contraction of the diaphragm or bellows. A sensor (not shown) is provided which measures the movement of the diaphragm or bellows, generating a signal representative of the difference between the reference pressure and the actual crankcase pressure. An electrically actuated device, such as an electrically actuated micro- or macro-sized valve or valves, may then operate to affect a control valve response, based at least in part upon the sensor's signal, to alter actual crankcase pressure. In another embodiment, the micro- or macro-sized valve or valves directly port or vent pressure from the crankcase 118 to change actual crankcase pressure based upon the sensor's signal.
FIG. 9 is a flow chart illustrating a method 510 for controlling displacement of a variable displacement compressor by utilizing crankcase pressure according to the present invention. For example, the method 510 may be implemented in the system 410, as shown in FIG. 7, in the system 420, as shown in FIG. 8, or in the system 600, as shown in FIG. 10.
In a first step 512 according to the method 510 of the present invention, a reference pressure is set. The reference pressure is a pressure related to the target (desired) crankcase pressure that is expected to result in the compressor 100 operating at a capacity resulting in the desired heat transfer. As discussed above, the reference pressure may be the target crankcase pressure (as in the system 410 and 420) or may be a pressure related to the target crankcase pressure in some fashion (i.e., the reference pressure is a function of the target crankcase pressure), such as the reference pressure of the system 600.
In a second step 514, actual pressure in the crankcase 118 is measured. This actual (measured) crankcase pressure may be acquired in any suitable manner, such as by the sensor 84 in the system 410, by the sensor 422 in the system 420, or by direct connection to a diaphragm or bellows such as the bellows 610 in the system 600.
In a third step 516, the reference crankcase pressure and the actual crankcase pressure are compared. For example, the comparison may be made mechanically, such as by action of a differential pressure across a diaphragm or bellows, such as the bellows 610. Alternatively, a calculation to compare the target crankcase pressure and the crankcase pressure measured may be made by the control unit 146.
For example, one method of comparison is to minimize the differences between the target crankcase pressure and the actual crankcase pressure by using an optimization algorithm. Any suitable optimization algorithm may be used. Many optimization algorithms are available, but they generally fall into three categories: derivative, simulated annealing, and genetic. In one embodiment, a simulated annealing algorithm is utilized where key optimization parameters are developed through prior testing. These parameters are dependent upon the particular system configuration. Another method of comparison would be to use a variable step function that depends on the difference between the target crankcase pressure and the actual crankcase pressure. For example, when the difference is relatively high, a relatively large step toward the target would be taken; that is, a relatively large change in position of the control valve 10 would be commanded (in embodiments utilizing the control valve 10). As the actual crankcase pressure approaches the target pressure, so the difference is relatively smaller, a relatively smaller step would be commanded. As will be discussed below, the steps of the method illustrated in FIG. 9 are repeated in an iterative process, so that each time the third step 516 is repeated, a smaller step will be commanded, although such repetition is not required. As the actual crankcase pressure approaches the target crankcase pressure, smaller steps are taken to minimize the tendency to overshoot and oscillate. The amount of the reduction in step size may be based on the magnitude of difference between target crankcase pressure and actual crankcase pressure, or may be time based, that is reduced stepwise at certain intervals. It is contemplated that the magnitude of a step could be anything from zero (no position modification) to a maximum position modification signal; for example, when utilizing Pulse Width Modulated signals to the associated valve(s) to control the porting of pressure to the crankcase 118 and venting of pressure from the crankcase 118, a zero signal would be zero voltage applied for the full application interval (no pulse), and a maximum signal would be a full power pulse applied for the full application interval. Further, the present invention contemplates that when the actual crankcase pressure is relatively close to the target crankcase pressure, there would be no change in the signal.
The illustrated method 510 then concludes with a step 518 where the position of the control valve 10 is modified based upon the comparison of the reference and actual crankcase pressures, thereby changing the position of the control valve 10 to the desired position. For example, in the systems 410 and 420, the control unit 146 will send an appropriate signal to the reference outlet valve 86 and/or the reference inlet valve 88 to change state based, at least in part, upon the comparison of the target pressure and the actual crankcase pressure to reposition the control valve 10, thereby adjusting the actual crankcase pressure toward the target crankcase pressure. As a further example, in the system 600, the valve 620 is repositioned by the differential pressure acting across the bellows 610 to cause crankcase pressure to change toward the target crankcase pressure.
In an alternative embodiment of the present invention, the method illustrated in FIG. 9 is a continuously iterative process, such that the method 510 continues to loop continuously through the steps 512 to 518, while the variable displacement compressor 100 is in operation, as indicated by the dashed line 520.
In testing, a control program, implementing the method illustrated in FIG. 9, was developed using LabVIEW computer development software (available from National Instrument Corporation, of Austin, Tex.), and was loaded on a computer control system. A compact variable compressor manufactured by Delphi Corporation (Troy, Mich.) was connected to traditional automotive air conditioning system components. An electrical pressure sensor was disposed located in a crankcase chamber of the compact variable compressor. The sensor was suitable to monitor crankcase pressure conditions in the crankcase chamber. A Microstaq™ microvalve, manufactured by Microstaq, Inc. (Bellingham, Wash.) was connected to the compact variable compressor for controlling the pressure in the crankcase chamber. Using crankcase feedback, i.e., a pressure measurement from the sensor, the control program instructed the Microstaq™ microvalve to regulate the pressure in the crankcase chamber. By monitoring crankcase pressure conditions, rather than suction pressure conditions, and using this crankcase pressure as a feedback signal to control compressor displacement, among other inputs, compressor control was achieved that was superior to that achieved by the prior art (which used suction pressure as a signal to control compressor displacement).
Shown in FIG. 11 is a schematic diagram of a compressor system 710 according to the present invention. The compressor system 710 includes a compressor 712. The compressor 712 is a variable displacement compressor in which the capacity of the compressor is controlled by the crankcase pressure, such as the variable displacement refrigerant compressor 210 of FIG. 1, the compressor 100 of FIG. 2, or any suitable compressor. The compressor 712 includes a crankcase port, as indicated at 714, in fluid communication with a crankcase, not shown, within the compressor 712. The compressor 712 includes a discharge port, as indicated at 716, and a suction port, as indicated at 718. The compressor system 710 includes an A/C system 720 having input 722 in fluid communication with the discharge port 716 and having an output 724 in fluid communication with the suction port 718. The A/C system 710 may be, for example, the automotive air conditioning system of FIG. 1 having a normal condenser 212, an orifice tube 214, an evaporator 216 and an accumulator 218 arranged in that order between the compressor discharge port 716 and the compressor suction port 718, the air conditioning unit of FIG. 2 having a condenser 140, an evaporator 142, and accumulator 144, or any suitable air conditioning system.
The compressor system 710 includes a control mechanism 726 having a crankcase interface, as indicated at 728, in fluid communication with the crankcase port 714. The control mechanism 726 has a discharge interface 730 in fluid communication with the discharge port 716. The control mechanism 726 has a suction interface 732 in fluid communication with the suction port 718. The control mechanism includes a valve arrangement for providing selective communication between the crankcase interface 728, the discharge interface 730, and the suction interface 732 in order to control crankcase pressure in a manner previously described. The control mechanism 726 may, for example, include an arrangement such as:
a pneumatic control valve similar to the control valve 300 of FIG. 1;
an electronically controlled valve and control unit similar to the variable control valve 10 and the control unit 146 of FIG. 2;
one or more microvalves and/or one or more macro-sized valves, which valves may be actuated by a differential pressure or may be electrically, pneumatically, or electro-pneumatically operated under the control of an electronic control unit (not shown), and which may be direct acting valves or which may be arranged as pilot valves and pilot-operated valves;
a bellows and valve arrangement, or a diaphragm and valve arrangement, such as the arrangement illustrated in FIG. 10; or
any other suitable control arrangement for controlling the pressure in the crankcase of the compressor 712 which operates the valve portion of the control mechanism to control crankcase pressure depending upon the difference between the actual crankcase pressure and a reference pressure related to a target crankcase pressure.
The present invention provides for significant enhancement of compressor output control, as compared to the prior art. This is achieved by reducing the time between system input (change in the crankcase pressure) and feedback (pressure sensor signal) for a given change. In the prior art compressor control method, suction pressure has been used as a feedback reference. The control valve changes the pressure in the crankcase to change the output of the compressor. Due to such factors as the need for a compressor piston to stoke to change suction pressure, refrigerant compressibility, and the volume of the air conditioning system, I have discovered that there is a relatively long time delay between a change in the position of control valve and the resulting suction pressure change. There is also an inherent instability in the prior art compressor system that tends to drive the compressor to a particular state, for example, minimum output as the variable control is opened and the crankcase pressure increases to maximum. These factors cause the prior art compressor to tend to go to extremes, i.e., to maximum output or to minimum output, with small changes in the control valve setting.
In the present invention, where crankcase pressure is monitored as the feedback reference, the effect of a control valve change upon the pressure in the crankcase, and thereby the compressor output, is recognized much sooner, as compared to the prior art, as the feedback reference is not communicated through the compressor mechanism and the air conditioning refrigerant volume. Consequently, the tendency of the compressor to overshoot is reduced as changes in the crankcase pressure quickly recognized and the variable control valve can be adjusted to minimize or eliminate compressor overshoot.
Although the preferred embodiments have been described in relation to a compressor suitable for use in an automotive air conditioning system, it must be understood that the invention may be practiced with any suitable compressor or compressor system where crankcase pressure controls the capacity of the compressor.
The principle and mode of operation of this invention have been explained and illustrated in its preferred embodiment. However, it must be understood that this invention may be practiced otherwise than as specifically explained and illustrated without departing from its spirit or scope.

Claims

1. A system for controlling a variable displacement compressor comprising:

a control valve for controlling pressure in a crankcase of a variable displacement compressor; and

a control unit that controls said control valve based on a comparison between a target pressure and said pressure in said crankcase.

2. The system of claim 1 wherein the system further comprises:

a pressure sensor disposed in said crankcase, said pressure sensor being in communication with said control unit to send said control unit a pressure sensor signal, said control unit being responsive to said pressure sensor signal to cause a change in said control valve position.

3. The system of claim 2 wherein said pressure sensor is one of an electronic pressure sensor and a mechanical feedback pressure sensor.

4. The system of claim 2 wherein the said pressure sensor is disposed in a passage or subchamber of said crankcase.

5. The system of claim 1 wherein the control valve is a pilot operated macro-sized valve, the position of which is controlled by a microvalve operated by said control unit.

6. The system of claim 1, wherein the control valve opens to raise pressure in said crankcase, and further comprising a second control valve which opens to lower pressure in said crankcase, said control unit controlling both the control valve and the second control valve based at least in part on pressure in said crankcase.

7. The system of claim 5, wherein the control valve and the second control valve are pilot valves controlling the position of a third control valve, said third control valve being selectively positionable thereby to a position in which the third control valve ports high pressure fluid from a discharge path of the compressor to said crankcase to raise pressure in said crankcase, a position in which the third control valve vents said crankcase to a suction flow path of the compressor to lower pressure in said crankcase, and a position in which said crankcase is isolated from said discharge flow path and isolated from said suction flow path to hold pressure in said crankcase constant.

8. The system of claim 1 further comprising a plurality of pilot valves which control the position of the control valve and the second control valve based at least in part on pressure in said crankcase.

9. The system of claim 1, wherein said control unit is one of a bellows and a diaphragm, exposed on one side to a reference pressure and on another side to actual crankcase pressure, and wherein said control valve position is determined by movement of said one of a bellows and a diaphragm in response to a differential in the pressures acting across said one of a bellows and a diaphragm.

10. The system of claim 1 wherein said control unit is one of an electrical control unit, a pneumatic control unit, an electro-pneumatic control unit, and a hydraulic control unit.

11. The system of claim 1 wherein the system further comprises:

a thermostatic system connected to said control unit, said thermostatic system operable to generate a target crankcase pressure, wherein said control unit is responsive to control said control valve based at least in part upon said target crankcase pressure.

12. The method of claim 19 wherein the controlling the position of the control valve includes the steps of:

d1) determining a desired change of position of a control valve for controlling the amount of the displacement of the compressor based on the difference determined between the target crankcase pressure and the actual crankcase pressure; and

d2) changing the position of the control valve to the desired position.

13. The method of claim 12 wherein the desired change of position determined in step d1) is determined by utilizing a simulated annealing algorithm dependant upon predetermined optimization parameters.

14. The method of claim 12 wherein the desired change of position determined in step d1) is based at least in part on the magnitude of the difference determined in step c).

15. The method of claim 12 wherein the determination performed in step d1) minimizes the difference between the target crankcase pressure and actual crankcase pressure by utilizing one of a derivative optimization algorithm, a simulated annealing algorithm, and a genetic optimization algorithm.

16. The method according to claim 20, comprising the steps of:

a) setting a reference pressure;

b) measuring an actual crankcase pressure;

c) comparing the reference pressure and the actual crankcase pressure to determine a difference between the reference pressure and the actual crankcase pressure; and

d) changing the position of a control valve controlling crankcase pressure, with the change in position of the control valve being based on the comparison of the actual crankcase pressure and the reference pressure.

17. The method according to claim 20, wherein the reference pressure is a target crankcase pressure.

18. The method according to claim 20, wherein the reference pressure is related to a target crankcase pressure as an inverse function.

19. A method for controlling a variable displacement compressor having a control valve for controlling the amount of the displacement of the compressor utilizing crankcase pressure, comprising the steps of:

a) setting a target crankcase pressure;

b) measuring crankcase pressure of a variable displacement compressor;

c) comparing the target crankcase pressure and the measured crankcase pressure to determine a difference between the target crankcase pressure and the measured crankcase pressure; and

d) controlling the position of a control valve for controlling the amount of the displacement of the compressor based on the difference determined between the target crankcase pressure and the measured crankcase pressure.

20. A method for controlling a variable displacement compressor of the type in which the amount of the displacement of the compressor is controlled by varying crankcase pressure, comprising the step of varying a signal to control compressor displacement as a function of a comparison between a target crankcase pressure and a measured crankcase pressure.

21. (canceled)