Control device of variable displacement compressor

BACKGROUND OF THE INVENTION

The present invention relates to a controller of a variable displacement compressor.

The refrigeration circuit of a typical vehicle air conditioner includes a compressor, such as a variable displacement swash plate type compressor. A typical variable displacement swash plate type compressor includes a displacement control mechanism for maintaining the pressure at the outlet of an evaporator, which will be referred to as the suction pressure Ps, at a target value, which will be referred to as target suction pressure. The displacement control mechanism feedback controls the displacement of the compressor, or the inclination angle of the swash plate, by referring to the suction pressure Ps such that the displacement corresponds to the cooling load. A typical displacement mechanism includes a displacement control valve, which is called an internally controlled valve. The internally controlled valve detects the absolute value of the suction pressure Ps by means of a pressure sensitive member such as a bellows or a diaphragm. The internally controlled valve moves a valve body by the displacement of the pressure sensing member to adjust the valve opening size. Accordingly, the pressure in a swash plate chamber (a crank chamber), or the crank pressure Pc is changed, which changes the inclination of the swash plate. However, an internally controlled valve that has a simple structure and a single target suction pressure cannot respond to the changes in air conditioning demands. Therefore, control valves having a target suction pressure that can be changed by external electrical control are becoming standard.

A typical electrically controlled control valve is a combination of an internally controlled valve and an actuator such as an electromagnetic solenoid, which applies an electrically controlled force. Mechanical spring force, which acts on the pressure sensing member is externally controlled to change the target suction pressure. The target suction pressure is changed by controlling a current to the electromagnetic solenoid in an analog or a digital manner. The supplied current is controlled by a controller having a microcomputer that is designed for air conditioning. Specifically, the controller executes a proportional and integral (PI) control procedure or a proportional, integral and differential (PID) control procedure based on temperature information from a temperature sensor located near the evaporator or in a passenger compartment for continuously controlling the current. As a result, the compressor theoretically maintains an ideal displacement, or a displacement that corresponds to the magnitude of the cooling load.

However, to execute a PI control procedure or a PID control procedure for continuously and finely controlling the target suction pressure, the controller, which includes a microcomputer, must continuously receive temperature information from the temperature sensor and compute the current supplied to a control valve. Thus, the controller must have a high-performance microcomputer to bear a high computation load. Even if the controller has a high-performance microcomputer, the controller receives temperature data relatively frequently (at an extremely short cycle). Thus, the controller cannot be used for other purposes, which increases the ratio of cost of the controller in the total cost of the compressor.

In a displacement control procedure in which the absolute value of the suction pressure Ps is used as a reference, changing of the target suction pressure by electrical control does not always quickly change the actual suction pressure to the target suction pressure. This is because whether the actual suction pressure quickly seeks a target suction pressure when the target suction pressure is changed greatly depends on the absolute magnitude of the cooling load. Therefore, even if the target suction pressure is finely and continuously controlled by controlling the current to the control valve, changes in the compressor displacement are likely to be too slow or too sudden.

SUMMARY OF THE INVENTION

Accordingly, it is an objective of the present invention to provide a control device of a variable displacement compressor that has a simple structure and improves the controllability and response of displacement control.

To achieve the foregoing and other objectives and in accordance with the purpose of the present invention, a controller for a variable displacement compressor, which is used for air conditioning a compartment, is provided. The compressor includes a suction pressure zone, a discharge pressure zone, and a control chamber, which is connected to the suction pressure zone and to the discharge pressure zone. The pressure in the control chamber is adjusted for controlling the displacement of the compressor. The controller includes a refrigerant circuit, a control valve, a detection circuit and a pressure difference changer. The refrigerant circuit is connected to the compressor. Two pressure monitoring points are located in the refrigerant circuit. The control valve controls the pressure in the control chamber. The control valve operates based on the actual pressure difference between the pressure monitoring points such that a target value of the pressure difference between the pressure monitoring points, which is externally determined, is maintained. The detection circuit includes a temperature sensor for monitoring a temperature that represents the temperature of the compartment. The detection circuit produces a first detection signal when the sensed temperature exceeds a threshold value and a second detection signal when the sensed temperature falls below the threshold value. The pressure difference changer gradually increases the target value of the pressure difference when the first signal is received from the detection circuit and gradually decreases the target value of the pressure difference when the second signal is received from the detection circuit.

Other aspects and advantages of the invention will become apparent from the following description, taken in conjunction with the accompanying drawings, illustrating by way of example the principles of the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention, together with objects and advantages thereof, may best be understood by reference to the following description of the presently preferred embodiments together with the accompanying drawings in which:

FIG. 1

is a cross-sectional view illustrating a variable displacement swash plate type compressor according to one embodiment of the present invention;

FIG. 2

is a schematic diagram illustrating a refrigeration circuit according to the embodiment of

FIG. 1

;

FIG. 3

is a cross-sectional view illustrating the control valve in the compressor of

FIG. 1

;

FIG. 4

is a schematic cross-sectional view showing an effective pressure receiving area of the control valve shown in

FIG. 3

;

FIG. 5

is a block diagram showing a control system of the embodiment shown in

FIG. 1

;

FIG. 6

is a graph showing the relationship between a detection circuit signal and a monitored temperature;

FIG. 7

is a flowchart showing an irregular interruption routine (1);

FIG. 8

is a flowchart showing an irregular interruption routine (2);

FIG. 9

is a flowchart showing a regular interruption routine (A);

FIG. 10

is a flowchart showing a regular interruption routine (B);

FIG. 11

is a timing chart showing the relationship between a duty ratio Dt and a detection circuit signal (a rising signal);

FIG. 12

is a timing chart showing the relationship between a duty ratio Dt and a detection circuit signal (a falling signal); and

FIG. 13

is a timing chart showing the relationship between a duty ratio Dt and detection circuit signals (rising signals and falling signals).

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

One embodiment according to the present invention will now be described with reference to

FIGS. 1

to

13

.

As shown in

FIG. 1

, a variable displacement swash plate type compressor includes a cylinder block

1

, a front housing member

2

, which is secured to the front end face of the cylinder block

1

, and a rear housing member

4

, which is secured to the rear end face of the cylinder block

1

. A valve plate assembly

3

is located between the cylinder block

1

and the rear housing member

4

. The cylinder block

1

, the front housing member

2

, the valve plate assembly

3

and the rear housing member

4

are secured to one another by bolts

10

(only one is shown) to form the compressor housing. In

FIG. 1

, the left end of the compressor is defined as the front end, and the right end of the compressor is defined as the rear end. A crank chamber

5

is defined between the cylinder block

1

and the front housing member

2

. A drive shaft

6

extends through the crank chamber

5

and is supported through radial bearings

8

A,

8

B by the housing. A recess is formed in the center of the cylinder block

1

. A coil spring

7

and a rear thrust bearing

9

B are located in the recess. A lug plate

11

is secured to the drive shaft

6

to rotate integrally with the drive shaft

6

. A front thrust bearing

9

A is located between the lug plate

11

and the inner wall of the front housing member

2

. The drive shaft

6

is supported in the axial direction by the rear bearing

9

B, which is urged forward by the spring

7

, and the front bearing

9

A.

The front end of the drive shaft

6

is connected to an external drive source, which is a vehicle engine E in this embodiment, through a power transmission mechanism PT. In this embodiment, the power transmission mechanism PT is a clutchless mechanism that includes, for example, a belt and a pulley. The power transmission mechanism PT therefore constantly transmits power from the engine E to the compressor when the engine E is running. Alternatively, the mechanism PT may be a clutch mechanism (for example, an electromagnetic clutch) that selectively transmits power when supplied with a current.

As shown in

FIG. 1

, a cam plate, which is a swash plate

12

in this embodiment, is located in the crank chamber

5

. A hole extends through the middle of the swash plate

12

. The drive shaft

6

extends through the hole. The swash plate

12

is connected with the lug plate

11

and the drive shaft

6

through a coupling guide mechanism, which is a hinge mechanism

13

in this embodiment, to rotate integrally with the drive shaft

6

. The hinge mechanism

13

includes two support arms

14

(only one is shown) and two guide pins

15

(only one is shown). Each support arm

14

projects from the rear side of the lug plate

11

. Each guide pin

15

projects from the front side of the swash plate

12

. The support arms

14

and the guide pins

15

cooperate to permit the swash plate

12

to rotate integrally with the lug plate

11

and the drive shaft

6

. Contact between the drive shaft

6

and the wall of the swash plate center hole permits the swash plate

12

to slide along the drive shaft

6

and to tilt with respect to the axis of the drive shaft

6

. The swash plate

12

has a counterweight

12

a

located at the opposite side of the drive shaft

6

from the hinge mechanism

13

.

A spring

16

is located between the lug plate

11

and the swash plate

12

. The spring

16

urges the swash plate

12

toward the cylinder block

1

, or in a direction decreasing the inclination angle of the swash plate

12

. A stopper ring

18

is fixed on the drive shaft

6

behind the swash plate

12

. A return spring

17

is fitted about the drive shaft

6

between the stopper ring

18

and the swash plate

12

. When the inclination angle is great as shown by the broken line in

FIG. 1

, the spring

17

does not apply force to the swash plate

12

. When the inclination angle is small as shown by the solid line in

FIG. 1

, the spring

17

is compressed between the stopper ring

18

and the swash plate

12

and urges the swash plate

12

away from the cylinder block

1

, or in a direction increasing the inclination angle. The normal length of the spring

17

and the location of the stopper ring

18

are determined such that the spring

17

is not fully contracted when the swash plate

12

is inclined by the minimum inclination angle θmin (for example, an angle from one to five degrees).

Several cylinder bores

1

a

(only one shown) are formed in the cylinder block

1

about the drive shaft

6

. The rear end of each cylinder bore

1

a

is blocked by the valve plate assembly

3

. A single headed piston

20

is reciprocally accommodated in each cylinder bore

1

a

. Each piston

20

and the corresponding cylinder bore

1

a

define a compression chamber, the volume of which is changed according to reciprocation of the piston

20

. The front portion of each piston

20

is coupled to the swash plate

12

by a pair of shoes

19

. Therefore, when the swash plate

12

rotate integrally with the drive shaft

6

, rotation of the swash plate

12

reciprocates each piston

20

by a stroke that corresponds to the angle θ.

A suction chamber

21

, which is included in a suction pressure zone, and discharge chamber

22

, which is included in a discharge pressure zone, are defined between the valve plate assembly

3

and the inner wall of the rear housing member

4

. The suction chamber

21

is located approximately in the center of the rear housing member

4

, and the discharge chamber

22

surrounds the suction chamber

21

. The valve plate assembly

3

includes a suction valve flap plate, a port plate, discharge valve flap plate and a retainer plate. The valve plate assembly

3

has suction ports

23

and discharge ports

25

, which correspond to each cylinder bore

1

a

. The valve plate assembly

3

also has suction valve flaps

24

, each of which corresponds to one of the suction ports

23

, and discharge valve flaps

26

, each of which corresponds to one of the discharge ports

25

. Each cylinder bore

1

a

is connected to the suction chamber

21

through the corresponding suction port

23

and is connected to the discharge chamber

22

through the corresponding discharge port

25

. Refrigerant gas is drawn from the outlet of the evaporator

33

to the suction chamber

21

, where the pressure is a suction pressure Ps. When each piston

20

moves from the top dead center position to the bottom dead center position, refrigerant gas in the suction chamber

21

flows into the corresponding cylinder bore

1

a

via the corresponding suction port

23

and suction valve flap

24

. When each piston

20

moves from the bottom dead center position to the top dead center position, refrigerant gas in the corresponding cylinder bore

1

a

is compressed to a predetermined pressure and is discharged to the discharge chamber

22

, where the pressure is a discharge pressure Pd, via the corresponding discharge port

25

and discharge valve

26

. The highly pressurized refrigerant in the discharge chamber

22

flows to the condenser

31

.

Power from the engine E is transmitted to and rotates the drive shaft

6

. Accordingly, the swash plate

12

, which is inclined by an angle θ, is rotated. The angle θ is defined by the swash plate

12

and an imaginary plane that is perpendicular to the drive shaft

6

. Rotation of the swash plate

12

reciprocates each piston

20

by a stroke that corresponds to the angle θ. As a result, suction, compression and discharge of refrigerant gas are repeated in the cylinder bores

1

a.

The inclination angle θ of the swash plate

12

is determined according to various moments acting on the swash plate

12

. The moments include a rotational moment, which is based on the centrifugal force of the rotating swash plate

12

, a spring force moment, which is based on the force of the springs

16

and

17

, a moment of inertia of the piston reciprocation, and a gas pressure moment. The gas pressure moment is generated by the force of the pressure in the cylinder bores

1

a

and the pressure in the crank chamber

5

(crank pressure Pc). Depending on the crank pressure Pc, the gas pressure moment acts either to increase or decrease the inclination angle θ of the swash plate

12

.

The gas pressure moment is adjusted by changing the crank pressure Pc by a displacement control valve, which will be discussed below. Accordingly, the inclination angle θ of the plate

12

is adjusted to an angle between the minimum inclination θmin and the maximum inclination θmax. Contact between a counterweight

12

a

on the swash plate

12

and a stopper

11

a

of the lug plate

11

prevents further inclination of the swash plate

12

from the maximum inclination θmax. The minimum inclination θmin is determined based primarily on the forces of the springs

16

and

17

when the gas pressure moment is maximized in the direction in which the swash plate inclination angle θ is decreased.

As described above, the crank pressure Pc is related to changes of the inclination angle θ of the swash plate

12

. A mechanism for controlling the crank pressure Pc includes a bleed passage

27

, a supply passage

28

and the control valve. The passages

27

,

28

are formed in the compressor housing. The bleed passage

27

connects the suction chamber

21

with the crank chamber

5

. The supply passage

28

connects the discharge chamber

22

with the crank chamber

5

. The control valve regulates the supply passage

28

. Specifically, the opening of the control valve is adjusted to control the flow rate of highly pressurized gas supplied to the crank chamber

5

through the supply passage

28

. The crank pressure Pc is determined by the ratio of the gas supplied to the crank chamber

5

through the supply passage

28

and the flow rate of refrigerant gas conducted out from the crank chamber

5

through the bleed passage

27

. As the crank pressure Pc varies, the difference between the crank pressure Pc and the pressure in the cylinder bores

1

a

varies, which changes the inclination angle θ of the swash plate

12

. Accordingly, the stroke of each piston

20

, or the compressor displacement, is varied.

A slight clearance (not shown) exists between the inner wall of the each cylinder bore

1

a

and the corresponding piston

20

. Each clearance connects the corresponding cylinder bore

1

a

with the crank chamber

5

. The discharge pressure zone, which is connected to the crank chamber

5

, includes the discharge chamber

22

and one or more of the cylinder bores

1

a

in which the piston

20

is in the compression stroke. When each piston

20

compresses the gas in the associated cylinder bore

1

a

, some of the refrigerant gas in the cylinder bore

1

a

leaks into the crank chamber

5

through the clearance between the cylinder bore

1

a

and the piston

20

. The leaking gas is referred to as blowby gas. The blowby gas increases the pressure of the crank chamber

5

. The discharge pressure zone in this embodiment includes the discharge chamber

22

and the cylinder bores

1

a.

As shown in

FIGS. 1 and 2

, a refrigeration circuit, or a refrigerant circuit, of a vehicle air conditioner includes the variable displacement swash plate type compressor and an external refrigerant circuit

30

. The external refrigerant circuit

30

includes, for example, a condenser

31

, a decompression device and an evaporator

33

. The decompression device is an expansion valve

32

in this embodiment. The opening of the expansion valve

32

is feedback-controlled based on the temperature detected by a heat sensitive tube

34

at the outlet of the evaporator

33

and the evaporation pressure, or the pressure at the evaporator outlet. The expansion valve

32

supplies liquid refrigerant to the evaporator

33

to regulate the flow rate in the external refrigerant circuit

30

. The amount of the supplied refrigerant corresponds to the thermal load. A downstream pipe

35

is located in a downstream portion of the refrigerant circuit

30

to connect the outlet of the evaporator

33

to the suction chamber

21

of the compressor. An upstream pipe

36

is located in an upstream portion of the refrigerant circuit

30

to connect the discharge chamber

22

of the compressor to the inlet of the condenser

31

. The compressor draws refrigerant gas from the downstream portion of the refrigeration circuit

30

and compresses the gas. The compressor then discharges the compressed gas to the discharge chamber

22

, which is connected to the upstream portion of the circuit

30

.

The greater the displacement of the compressor is, the higher the flow rate of refrigerant in the refrigeration circuit is. The greater the flow rate of the refrigerant is, the greater the pressure loss per unit length of the circuit is. That is, the pressure loss between two points in the refrigeration circuit corresponds to the flow rate of refrigerant in the circuit. Detecting the pressure difference ΔP(t) between two points P

1

, P

2

permits the displacement of the compressor to be indirectly detected. In this embodiment, two pressure monitoring points P

1

, P

2

are defined in the upstream pipe

36

. The first pressure monitoring point P

1

is located in the discharge chamber

22

, which is the most upstream section of the upstream pipe

36

. The second pressure monitoring point P

2

is located in the upstream pipe

36

and is spaced from the first point P

1

by a predetermined distance. A part of the control valve is exposed to the pressure PdH, or the discharge pressure Pd, at the first point P

1

by a first pressure introduction passage

37

. Another part of the control valve is exposed to a pressure PdL at the second point P

2

by a second pressure introduction passage

38

. The control valve feedback process uses the pressure difference expressed by ΔP(t)=PdH−PdL to estimate the compressor displacement and to feedback control the displacement.

The displacement control valve shown in

FIG. 3

mechanically detects the pressure difference between the pressure monitoring points P

1

, P

2

and adjusts the valve opening based on the detected pressure difference.

As shown in

FIG. 3

, the control valve includes an inlet valve and a solenoid. The inlet valve is arranged in an upper portion of the valve, while the solenoid is arranged in a lower portion of the valve. The inlet valve adjusts the opening size (throttle amount) of the supply passage

28

, which connects the discharge chamber

22

to the crank chamber

5

. The solenoid is an electromagnetic actuator for urging a rod

40

located in the control valve based on current supplied from an outside source. The solenoid functions as an actuator

100

for changing a target pressure difference.

The rod

40

includes a distal portion

41

, a coupler portion

42

and a proximal guide portion

44

. The guide portion

44

includes a valve body

43

, which is located in the center of the rod

40

. The diameter of the distal portion

41

, the coupler portion

42

and the guide portion

44

are represented by d1, d2 and d3, respectively. The diameters satisfy the inequality d2<d1<d3. The cross-sectional area SB of the distal portion

41

is represented by π(d1/2)

2

. The cross-sectional area SC of the coupler portion

42

is represented by π(d2/2)

2

. The cross-sectional area SD of the guide portion

44

is represented by π(d3/2)

2

.

The control valve has a valve housing

45

. The housing

45

includes a cap

45

a

and an upper portion

45

b

and a lower portion

45

c

. The cap

45

a

is fixed to the end of the upper portion

45

b

. The upper portion

45

b

defines the shape of the inlet valve portion. The lower portion

45

c

defines the shape of the solenoid. A valve chamber

46

and a communication passage

47

are formed in the upper portion

45

b

. A pressure sensing chamber

48

is defined between the upper portion

45

b

and the cap

45

a.

The rod

40

extends through the valve chamber

46

, the communication passage

47

an the pressure sensing chamber

48

. The rod

40

moves axially, or in the vertical direction as viewed in the drawing. The valve chamber

46

is connected to the communication passage

47

depending on the position of the rod

40

. The communication passage

47

is disconnected from the pressure sensing chamber

48

by a wall, which is a part of the valve housing

45

. A guide hole

49

is formed in the wall to receive the rod

40

. The diameter of the guide hole

49

is equal to the diameter d1 of the distal portion

41

. The communication passage

47

is axially aligned with the guide hole

49

, and the diameter of the communication passage

47

is equal to the diameter dl of the distal portion

41

. That is, the area of the communication passage

47

and the area of the guide hole

49

are equal to the area SB of the distal portion

41

.

The bottom of the valve chamber

46

is formed by the upper surface of a fixed iron core

62

. A Pd port

51

extends radially from the valve chamber

46

. The valve chamber

46

is connected to the discharge chamber

22

through the Pd port

51

and the upstream section of the supply passage

28

. A Pc port

52

radially extends from the communication passage

47

. The communication passage

47

is connected to the crank chamber

5

through the downstream section of the supply passage

28

and the Pc port

52

. Therefore, the Pd port

51

, the valve chamber

46

, the communication passage

47

and the Pc port

52

are formed in the control valve and form a part of the supply passage

28

, which connects the discharge chamber

22

with the crank chamber

5

.

The valve body

43

of the rod

40

is located in the valve chamber

46

. The diameter d1 of the communication passage

47

is greater than the diameter d2 of the coupler portion

42

and smaller than the diameter d3 of the guide portion

44

. Thus, a step is formed between the valve chamber

46

and the communication passage

47

. The step functions as a valve seat

53

, and the communication passage

47

functions as a valve hole. When the rod

40

is moved from the position of

FIG. 3

, or the lowermost position, to the uppermost position, at which the valve body

43

contacts the valve seat

53

, the communication passage

47

is disconnected from the valve chamber

46

. That is, the valve body

43

is an inlet valve body that controls the opening size of the supply passage

28

.

A movable wall

54

is located in the pressure sensing chamber

48

. The movable wall

54

divides the pressure sensing chamber

48

into a first pressure chamber

55

and a second pressure chamber

56

. The movable wall

54

does not permit fluid to move between the first pressure chamber

55

and the second pressure chamber

56

. The cross-sectional area SA of the movable wall

54

is greater than the cross-sectional area SB of the guide hole

49

(SB<SA).

The first pressure chamber

55

is constantly connected to the discharge chamber

22

, which is the upstream pressure monitoring point P

1

, by a P

1

port

55

a

formed in the cap

45

a

and the first passage

37

.

The second pressure chamber

56

is constantly connected to the second pressure monitoring point P

2

through a P

2

port

56

a

formed in the upper portion

45

b

and the second passage

38

. The first pressure chamber

55

is exposed to the discharge pressure Pd, which is the pressure PdH. The second pressure chamber

56

is exposed to the pressure PdL at the second pressure monitoring point P

2

. The upper side of the movable wall

54

receives the pressure PdH and the lower side receives the pressure PdL. The distal portion

41

of the rod

40

is located in the second pressure chamber

56

. The distal end of the distal portion

41

is coupled to the movable wall

54

. A spring

57

is located in the second pressure chamber

56

. The spring

57

urges the movable wall

54

toward the first pressure chamber

55

.

The solenoid (the actuator

100

for changing the target pressure difference) includes a cup-shaped cylinder

61

, which is fixed in the lower portion

45

c

. A stationary iron core

62

is fitted into an upper opening of the cylinder

61

. The stationary core

62

defines a solenoid chamber

63

in the cylinder

61

. A movable iron core

64

is located in the solenoid chamber

63

. The movable iron core

64

is moved axially. The stationary core

62

has a guide hole

65

through which the guide portion

44

extends. There is a clearance (not shown) between the guide hole

65

and the guide portion

44

. The clearance communicates the valve chamber

46

with the solenoid chamber

63

. Thus, the solenoid chamber

63

is exposed to the discharge pressure Pd, to which the valve chamber

46

is exposed.

The proximal portion of the rod

40

is located in the solenoid chamber

63

. The lower end of the guide portion

44

is fitted into a hole formed in the center of the movable iron core

64

. The movable iron core

64

is crimped to the guide portion

44

. Thus, the movable core

64

moves integrally with the rod

40

. A spring

66

is located between the stationary core

62

and the movable core

64

. The spring

66

urges the movable core

64

and the rod

40

downward such that the movable core

64

moves away from stationary core

62

.

A coil

67

is wound about the stationary core

62

and the movable core

64

. The coil

67

receives drive signals from a drive circuit

72

based on commands from an ECU

70

for the engine E. The coil

67

generates an electromagnetic force F that corresponds to the value of the current from the drive circuit

72

. The electromagnetic force F urges the movable core

64

toward the stationary core

62

, which lifts the rod

40

. The current to the coil

67

may be varied in an analog fashion. Alternatively, the current may be duty controlled, that is, the duty ratio Dt of the current may be controlled. In this case, a greater duty ratio Dt represents a smaller opening size of the control valve and a smaller duty ratio Dt represents a greater opening size of the control valve.

The opening size of the control valve is determined by the position of the rod

40

. The rod

40

has the valve body

43

, which functions as an inlet valve body. Forces acting on several parts of the rod

40

will now be explained to describe the operating conditions and the characteristics of the control valve.

The upper surface of the distal portion

41

receives a downward force, which is the resultant of the force fl of the spring

57

and the pressures acting on the upper and the lower sides of the movable wall

54

. The pressure receiving area on the upper side of the wall

54

is represented by SA. The pressure receiving area of the lower side of the wall

54

is represented by (SA−SB). The pressure receiving area of the lower end of the distal portion

41

is represented by (SB−SC). The crank pressure Pc applies an upward force to the lower end of the distal portion

41

. Assume downward forces have positive values. The sum ΣF1 of the forces acting on the distal portion

41

is represented by the following equation.

Σ

F

1=

PdH·SA−PdL

(

SA−SB

)−

f

1−

Pc

(

SB−SC

) Equation I

A downward force f2 of the spring

66

and an upward electromagnetic force F act on the guide portion

44

, which includes the valve body portion

43

.

The pressures that act on the exposed surfaces of the valve body

43

, the guide portion

44

and the movable iron core

64

will now be described with reference to FIG.

4

. The pressures are simplified as follows. First, the upper end surface of the valve body

43

is divided into the inside section and the outside section by an imaginary cylinder, which is shown by broken lines in FIG.

4

. The imaginary cylinder corresponds to the wall of the communication passage

47

. The crank pressure Pc acts in a downward direction on the inside section (area: SB−SC). The discharge pressure Pd acts in a downward direction on the outside section (area: SD−SB). Taking the pressure balance between the upper and lower surfaces of the movable iron core

64

into account, the discharge pressure Pd, to which the solenoid chamber

63

is exposed, acts on the area corresponding to the cross-sectional area SD of the guide portion

44

to urge the guide portion

44

upward. If the total force ΣF2 that acts on the valve body

43

and the guide portion

44

, defining the upward direction as the positive direction, are summed, ΣF2 is expressed by the following equation.

Σ

F

2=

F−f

2−

Pc

(

SB−SC

)−

Pd

(

SD−SB

)+

Pd·SD=F−f

2−

Pc

(

SB−SC

)+

Pd·SB

Equation II

In the process of calculating equation II, −Pc·SD was canceled by +Pc·SD, and the term Pc·SB remained. That is, if the net force based on the discharge pressure Pd that acts on the upper and lower surfaces of the guide portion

44

is viewed as a force that acts on the lower surface of the guide portion

44

, the effective pressure receiving area of the guide portion

44

regarding the discharge pressure Pd is equal to the area SB (SB=SD−(SD−SB)). As far as the discharge pressure Pd is concerned, the effective pressure receiving area of the guide portion

44

is equal to the cross-sectional area SB of the communication passage

47

regardless of the cross-sectional area SD of the guide portion

44

. When pressures of the same kind act on both ends of a member such as a rod, the pressure receiving area having an effect that is not canceled is called the effective pressure receiving surface area.

Since the rod

40

is an integrated member formed by connecting the guide portion

44

to the distal portion

41

with coupler portion

42

, its position is determined by the physical balance of ΣF1=ΣF2. In the equation ΣF1=ΣF2, the terms Pc(SB−SC) can be canceled. As a result, the following equation III is obtained.

(

PdH−PdL

)

SA−Pd·SB+PdL·SB=F+f

1−

f

2 Equation III

Since the first pressure monitoring point P

1

is located in the discharge chamber

22

, the pressure Pd is equal to the pressure PdH (Pd=PdH). If Pd is replaced by PdH, equation III is converted into the following equations IV and V.

(

PdH−PdL

)

SA

−(

PdH−PdL

)

SB=F+f

1−

f

2 Equation IV

PdH−PdL=

(

F+f

1−

f

2)/(

SA−SB

) Equation V

In equation V, f1, f2, SA and SB are fixed parameters that are primarily defined in the steps of mechanical design, and the electromagnetic force F is a variable parameter that changes in accordance with the power supplied to the coil

67

.

As apparent from equation V, the pressure difference ΔP(t) (ΔP(t)=PdH−PdL), is determined only by duty controlling the current supplied to the coil

67

. That is, a target value TPD of the pressure difference is adjusted by externally controlling the control valve.

Equation V contains no pressure parameters such as the crank pressure Pc and the discharge pressure Pd, other than the pressure difference expressed by PdH−PdL. Thus, the crank pressure Pc and the discharge pressure Pd do not influence the position of the rod

40

. In other words, pressure parameters other than the pressure difference do not affect the movement of the rod

40

, and the control valve is regulated based only on the pressure difference ΔP(t), the electromagnetic force F and the spring forces f1, f2.

The opening size of the control valve is determined in the following manner. When no current is supplied to the coil

67

, or when the duty ratio Dt is zero percent, the spring

66

positions the rod

40

at the lowest position shown in FIG.

3

. The valve body

43

is spaced from the valve seat

53

by the greatest distance, which fully opens the control valve. When a current of the minimum duty ratio is supplied to the coil

67

, the upward electromagnetic force F is greater than the downward force f2 of the spring

66

. The net upward force (F−f2) generated by the solenoid and the spring

66

acts against the net downward force of the pressure difference (PdH−PdL) and the spring

57

. As a result, the position of the valve body

43

relative to the valve seat

53

is determined such that equation V is satisfied, which determines the opening size of the control valve.

Accordingly, the flow rate of gas to the crank chamber

5

through the supply passage

28

is determined. Then, the crank pressure Pc is adjusted in accordance with the relationship between the flow rate of gas through the supply passage

28

and the flow rate of gas flowing out from the crank chamber

5

through the bleed passage

27

. That is, controlling the opening size of the control valve controls the crank pressure Pc. When the electromagnetic force F is constant, the control valve functions as a constant flow rate valve and is actuated based on the target pressure difference TPD, which corresponds to the electromagnetic force F. However, since electromagnetic force F can be externally changed to adjust the target pressure difference TPD, the control valve can vary the displacement of the compressor.

Control System

As shown in

FIGS. 2

,

3

and

5

, the control valve is connected to a pressure difference changer, which is an engine ECU

70

in this embodiment, through the drive circuit

72

. The engine ECU

70

mainly controls the engine E. As shown in

FIG. 5

, the ECU

70

includes a CPU, a ROM, a RAM, a timer and an input-output interface circuit. The ROM stores various control programs (see flowcharts of

FIGS. 7

to

10

) and initial data. The RAM has a working memory area. The timer generates clock pulse signals by either hardware or software. The clock pulse signals are at least used as regular interruption signals for notifying the CPU of the starting time of regular interruption routines. The input-output interface circuit has input and output terminals. An external information detection apparatus

71

is connected to input terminals. The drive circuit

72

is connected to output terminals. The engine ECU

70

computes an appropriate duty ratio Dt based on the information from the apparatus

71

and commands the drive circuit

72

to output a drive signal having the computed duty ratio Dt. The drive circuit

72

outputs the instructed drive signal having the duty ratio Dt to the coil

67

of the control valve. The electromagnetic force F of the solenoid is determined according to the duty ratio Dt. Accordingly, the opening size of the control valve is continuously adjusted, which quickly changes the crank pressure Pc and the stroke of each piston

20

. The piston stroke represents the compressor displacement and the torque.

The external information detection apparatus

71

includes various sensors. The sensors of the detection apparatus

71

may include, for example, an A/C switch

81

, a vehicle speed sensor

82

, an engine speed sensor

83

, a throttle sensor (or an acceleration pedal sensor)

84

and a detection circuit

85

. The A/C switch

81

is an ON/OFF switch of the air conditioner operated by a passenger. The A/C switch

81

provides the engine ECU

70

with information regarding the ON/OFF state of the air conditioner. The vehicle speed sensor

82

and the engine speed sensor

83

provide the engine ECU

70

with information regarding the vehicle speed V and the engine speed NE. The throttle sensor

84

detects the inclination angle, or the opening size, of a throttle valve located in the intake passage of the engine. The throttle opening size represents the degree of depression Ac(t) of the acceleration pedal in the vehicle.

The detection circuit

85

is located in the vicinity of the evaporator

33

(see

FIG. 2

) and provides the engine ECU

70

with information regarding the temperature in the vicinity of the evaporator

33

. The temperature information will be referred to as a detection circuit signal. The temperature in the vicinity of the evaporator

33

corresponds to the temperature of the surface of the evaporator

33

and to the temperature of the passenger compartment. The detection circuit

85

includes a temperature sensor, which is a thermistor

86

in this embodiment, for monitoring the temperature in the vicinity of the evaporator

33

and a signal output circuit

87

for generating and outputting the detection circuit signal based on changes of the resistance of the thermistor

86

.

The signal output circuit

87

compares the monitored temperature with threshold temperatures. When the monitored temperature falls below one of the threshold temperatures or surpasses another, the circuit

87

outputs the detection circuit signal.

FIG. 6

shows the relationship between the monitored temperature and the detection circuit signal. The threshold temperatures are a lower limit temperature T1 (for example, three degrees centigrade) and an upper limit temperature T2 (for example, four degrees centigrade). The monitored temperature rises due to changes in the relationship between the flow rate of the refrigerant in the evaporator and the compartment temperature. When the monitored temperature surpasses the upper limit temperature T2, the signal output circuit

87

outputs an ON signal (a rising signal).

When the monitored temperature falls below the lower limit temperature T1, the signal output circuit

87

outputs an OFF signal (falling signal). Since the determination values differ when the signal is switched from OFF to ON from when the signal is switched from ON to OFF, there is a hysteresis. The threshold temperatures, which are three degrees centigrade and four degrees centigrade in this embodiment, are determined such that air sent to the passenger compartment is sufficiently cooled without forming frost the evaporator. Frost on the evaporator reduces the cooling efficiency.

A controller of the compressor at least includes the engine ECU

70

, the detection circuit

85

and the control valve.

Duty control procedure by the ECU

70

will be described with reference to flowcharts and timing charts (

FIGS. 7

to

13

). The ECU

70

normally controls the engine E by, for example, controlling the fuel supply amount. In addition, the ECU

70

regularly and irregularly performs interruptions for controlling the air conditioner.

FIG. 7

is a flowchart of an irregular interruption routine (1), which is executed for starting and stopping air conditioning. When the A/C switch

81

is turned on or off and a signal representing the switching reaches the engine ECU

70

, the ECU

70

judges that there is an interrupt request. In this case, the ECU

70

stops controlling the engine E and starts the irregular interruption routine (1).

If the A/C switch

81

is switched from OFF to ON in step S

71

, the ECU

70

moves to step S

72

. In step S

72

, the ECU

70

initializes the duty ratio Dt. That is, the ECU

70

sets the duty ratio Dt to an initial value Dt(ini), which is, for example, fifty percent. The opening size of the control valve corresponds to the initial duty ratio Dt(ini). The crank pressure Pc is changed accordingly and the compressor displacement is set to a predetermined initial level.

If the A/C switch

81

is switched from ON to OFF in step S

71

, the ECU

70

moves to step S

73

. In step S

73

, the ECU

70

sets the duty ratio Dt to zero, which maximizes the opening size of the control valve. Accordingly, the crank pressure Pc is quickly increased and the inclination angle θ is minimized. The compressor displacement is thus minimized. After either steps S

72

, S

73

, the ECU

70

terminates the interruption and starts controlling the engine E again.

FIG. 8

is a flowchart of an irregular interruption routine (2), which is executed when the A/C switch is on. When the signal from the detection circuit

85

changes, the engine ECU

70

judges that there is an interruption request. In this case, the ECU

70

stops controlling the engine E and starts the irregular interruption routine (2). If the ECU

70

receives a rising signal in step S

81

, the ECU

70

moves to step S

82

. In step S

82

, the ECU

70

starts regular interruption routine (A), which is shown in FIG.

9

. If the ECU

70

receives a falling signal in step S

81

, the ECU

70

moves to step S

83

. In step S

83

, the ECU

70

starts a regular interruption routine (B), which is shown in FIG.

10

. After executing either steps S

82

and S

83

, the ECU

70

terminates the interruption routine (2) and starts controlling the engine E again.

When the duty ratio Dt is the initial value Dt(ini), the compressor displacement is changed, which lowers the temperature in the vicinity of the evaporator

33

. When the monitored temperature falls below the lower limit temperature T1, the ECU

70

receives a falling signal from the detection circuit

85

and thus starts the routine (B). The ECU

70

regularly repeats the routine (B) until the ECU

70

receives a rising signal and starts the routine (A). The routine (B) is executed in synchronization with clock signals from the timer.

When the engine ECU

70

stops controlling the engine E and starts the routine (B), the ECU

70

decreases the current duty ratio Dt by an amount ΔD in step S

101

. A decrease in the duty ratio Dt represents a decrease of the target pressure difference TPD and a decrease of the refrigerant flow rate or a decrease in the compressor displacement. Accordingly, the air conditioning is controlled to lessen cooling.

In step S

102

, the ECU

70

judges whether the current duty ratio Dt, which was computed by subtracting the amount ΔD from the previous duty ratio Dt, is smaller than a predetermined lower limit value Dt(min). If the outcome of step S

102

is negative, the current duty ratio Dt is greater than the lower limit value Dt(min). In this case, the ECU

70

moves to step S

103

and commands the drive circuit

72

to change the duty ratio Dt, which slightly weakens the electromagnetic force F. Accordingly, the target pressure difference TPD is slightly lowered.

Then, since balance between the forces on the rod

40

is not achieved with the current pressure difference ΔP(t), the rod

40

is moved downward, which reduces the force applied by the spring

66

. Thus, the reduced downward force f2 of the return spring

66

is countered by the reduced upward electromagnetic force F, and the valve body

43

is positioned such that equation V is satisfied again. As a result, the opening size of the control valve, that is, the opening size of the supply passage

28

, is increased, which increases the crank pressure Pc. Accordingly, the difference between the crank pressure Pc and the pressure of the cylinder bores

1

a

increases, and the inclination angle θ of the swash plate

12

is decreased. Accordingly, the compressor displacement is decreased. When the discharge displacement of the compressor is decreased, the heat reduction performance of the evaporator

33

is also reduced, the passenger compartment temperature, or the monitored temperature, is increased, and the pressure difference between the points P

1

and P

2

is decreased.

If the outcome of step S

102

is positive, the ECU

70

sets the duty ratio Dt to the lower limit value Dt(min) in step S

104

and commands the drive circuit

72

to operate at the lower limit value Dt(min) in step S

103

. The lower limit value Dt(min) may be zero.

As the routine (B) is repeated, the duty ratio Dt, or the target pressure difference TPD, is gradually decreased. The timing chart of

FIG. 12

shows changes of the duty ratio Dt when the routine (B) is repeated. When receiving a falling signal from the detection circuit

85

, the ECU

70

keeps gradually decreasing the duty ratio Dt by the amount ΔD at a time in synchronization with the timer clock until the ECU

70

receives a rising signal. Accordingly, the duty ratio Dt is gradually decreased to the lower limit value Dt(min) (see the graph of Dt from t3 to t4 in FIG.

12

). Then, as long as the ECU

70

does not receive a rising signal from the detection circuit

85

, the duty ratio Dt is maintained at the lower limit value Dt(min) (see t4 and after in the graph of FIG.

12

).

A decrease in the duty ratio Dt decreases the compressor displacement and reduces the heat reduction performance of the evaporator

33

. Accordingly, the compartment temperature, or the monitored temperature, is gradually increased. When the monitored temperature surpasses the upper limit temperature T2, the engine ECU

70

receives a rising signal from the detection circuit

85

. The ECU

70

then repeats the regular interruption routine (A), which is shown in

FIG. 9

, until the ECU

70

receives a falling signal.

When the engine ECU

70

stops controlling the engine E and starts the routine (A), the ECU

70

increases the duty ratio Dt by the amount ΔD in step S

91

. An increase in the duty ratio Dt increases the target pressure difference TPD, which increases the refrigerant flow rate and the compressor displacement. Accordingly, the cooling performance is increased.

In step S

92

, the ECU

70

judges whether the current duty ratio Dt, which was computed by adding the amount ΔD to the previous duty ratio Dt, is greater than a predetermined upper limit value Dt(max). If the outcome of step S

92

is negative, the current duty ratio Dt is smaller than the upper limit value Dt(max). In this case, the ECU

70

moves to step S

93

and commands the drive circuit

72

to change the duty ratio Dt, which slightly strengthen, the electromagnetic force F. Accordingly, the target pressure difference TPD is slightly raised.

Then, since balance between the forces on the rod

40

is not achieved with the current pressure difference ΔP(t), the rod

40

is moved upward, which increases the force applied by the spring

66

. Thus, the increased downward force f2 of the return spring

66

is countered by the increased upward electromagnetic force F, and the valve body

43

is positioned such that equation V is satisfied again. As a result, the opening size of the control valve, that is, the opening size of the supply passage

28

, is decreased, which decreases the crank pressure Pc. Accordingly, the difference between the crank pressure Pc and the pressure of the cylinder bores

1

a

decreases, and the inclination angle θ of the swash plate

12

is increased. Accordingly, the compressor displacement is increased. When the discharge displacement of the compressor is increased, the heat reduction performance of the evaporator

33

is also increased, the passenger compartment temperature, or the monitored temperature, is decreased, and the pressure difference between the points P

1

and P

2

is increased.

If the outcome of step S

92

is positive, the ECU

70

sets the duty ratio Dt to the upper limit value Dt(max) in step S

94

and commands the drive circuit

72

to operate at the upper limit value Dt(max) in step S

93

. As the routine (A) is repeated, the duty ratio Dt, or the target pressure difference TPD, is gradually increased. The timing chart of

FIG. 11

shows changes of the duty ratio Dt when the routine (A) is repeated. When receiving a rising signal from the detection circuit

85

, the ECU

70

gradually increases the duty ratio Dt by the amount ΔD at a time in synchronization with the timer clock until the ECU

70

receives a falling signal. Accordingly, the duty ratio Dt is gradually increased to the upper limit value Dt(min) (see the graph of Dt from t1 to t2 in FIG.

11

). Then, as long as the ECU

70

does not receive a rising signal from the detection circuit

85

, the duty ratio Dt is maintained at the upper limit value Dt(max) (see t2 and after in the graph of FIG.

11

).

An increase in the duty ratio Dt increases the compressor displacement and increases the heat reduction performance of the evaporator

33

. Accordingly, the compartment temperature, or the monitored temperature, is gradually decreased. When the monitored temperature falls below the lower limit temperature T1, the ECU

70

then repeats the regular interruption routine (B), which is shown in

FIG. 10

, until the ECU

70

receives a rising signal.

The engine ECU

70

continues to gradually increase or decrease the duty ratio Dt, or the target suction pressure TPD, until the ECU

70

receives a signal (a detection circuit signal) that indicates the monitored temperature crosses one of the threshold temperatures from the detection circuit

85

. When receiving such a signal, the ECU

70

reverse the changing direction of the target pressure difference TPD. Thus, the target pressure difference TPD (duty ratio Dt) is alternately increased and decreased.

If there is no abrupt changes of thermal load, the increases and decreases of the duty ratio Dt, the duty ratio Dt changes along line

131

in the timing chart of

FIG. 13

from a macroscopic viewpoint. Changes of the monitored temperature, or increases and decreases of the monitored temperature between the threshold temperatures T1 and T2, the detection circuit

85

alternately outputs rising signals and falling signals. Every time the circuit

85

switches between a rising signal and a falling signal, the duty ratio Dt repeats increases and decreases with a constant amplitude above and below a center value DtMid(t). The center value DtMid(t) may be variable or constant. For example, a dashed line

132

represents the center value DtMid(t). In other words, while the engine ECU

70

changes the detection circuit signal between On and OFF in a binary fashion, which maintains the duty ratio Dt in the vicinity of the center value DtMid(t) in a certain amplitude.

As described above, the duty ratio Dt, or the pressure difference TPD, is quickly adjusted when the thermal load on the evaporator

33

is changed. The flow rate of refrigerant is adjusted accordingly, which maintains the temperature in the vicinity of the evaporator

33

at a temperature suitable for cooling the passenger compartment.

The illustrated embodiment has the following advantages.

The temperature in the vicinity of the evaporator

33

is maintained at a level suitable for cooling by a simple procedure. That is, the ECU

70

simply increases or decreases in response to a rising signal or a falling signal from the detection circuit

85

. In other words, the procedures for optimizing the temperature in the vicinity of the evaporator

33

are sufficiently simple to be performed as interruptions by the ECU

70

, which reduces the calculation load on the ECU

70

. Thus, there is no need for an expensive controller specialized for air conditioning, and the engine ECU

70

, which is mainly used for controlling the engine E, is used for air conditioning.

In the illustrated embodiment, the threshold temperatures, which are compared with the temperature monitored by the detection circuit

85

, include the lower and upper limit temperatures T1, T2. Also, there is a hysteresis in which the temperature at which a rising signal is generated is different from the temperature at which a falling signal is generated. If there is only one threshold temperature, hunting may occur. Compared to a system having a single threshold temperature, the illustrated embodiment stably controls the compressor displacement without applying an excessive load on the compressor. Hunting of the detection circuit

85

refers to a case where the monitored temperature surpasses and falls below a single threshold temperature and the resulting detection circuit signals are excessively generated during a short time.

The suction pressure Ps is greatly influenced by changes in the thermal load on the evaporator

33

. In the illustrated embodiment, the suction pressure Ps is not directly referred to for controlling the opening size of the displacement control valve. Instead, the pressure difference ΔP(t)(ΔP(t)=PdH−PdL) between the two pressure monitoring points P

1

and P

2

is directly controlled for feedback controlling the compressor displacement.

Therefore, the compressor displacement is quickly controlled from the outside without being influenced by the thermal load on the evaporator

33

.

The control valve shown in

FIG. 3

functions as an internally controlled valve. Specifically, as long as the electromagnetic force F is constant, the control valve shown in

FIG. 3

maintains the target pressure difference TPD, which is determined by the forces F, f1, f2 and the areas SA, SB, and automatically controls the compressor displacement to a level that corresponds to the target pressure difference TPD. The electromagnetic force F can be externally changed for changing the target pressure difference TPD. The compressor displacement is changed accordingly.

It should be apparent to those skilled in the art that the present invention may be embodied in many other specific forms without departing from the spirit or scope of the invention. Particularly, it should be understood that the invention may be embodied in the following forms.

The thermistor

86

and the signal output circuit

87

in the detection circuit

85

may be integrated or separated. If the thermistor

86

and the circuit

87

are separated, the thermistor

86

needs to monitor a temperature, which is the temperature of the evaporator

33

in the illustrated embodiment.

The upper limit value Dt(max) and the lower limit value Dt(min) of the duty ratio Dt, which are used in steps S

94

and S

104

, need not be used.

The upper limit temperature T2 and the lower limit temperature T1 may be replaced by a single threshold temperature.

In the illustrated embodiment, the engine ECU

70

functions as the target pressure difference changer. However, the target pressure difference TPD may be changed by a separate controller. Compared to PI control and PID control, in which the target pressure difference is continuously and finely controlled, the control procedure of the illustrated embodiment is simple, which reduces the cost of the controller.

In the illustrated embodiment, the present invention is applied to a reciprocal piston type compressor. However, the present invention may be applied to rotary compressors such as a variable displacement scroll type compressor disclosed in Japanese Unexamined Patent Publication No. 11-324930.

In the illustrated embodiment, the upstream pressure monitoring point P

1

is located in the discharge chamber

22

, and the downstream pressure monitoring point P

2

is located in the upstream pipe

36

. However, the upstream pressure monitoring point P

1

may be located in the downstream pipe

35

and the downstream pressure monitoring point P

2

may be located in the suction chamber

21

. Alternatively, the upstream pressure monitoring point P

1

may be located either in the discharge chamber or the upstream pipe

36

and the downstream pressure monitoring point P

2

may be located either in the suction chamber

21

or the downstream pipe

35

. Also, the upstream pressure monitoring point P

1

may be located either in the discharge chamber

22

and the upstream pipe

36

and the downstream pressure monitoring point P

2

may be located in the crank chamber

5

. Further, the upstream pressure monitoring point P

1

may be located in the crank chamber

5

and the downstream pressure monitoring point P

2

may be located either in the suction chamber

21

or the downstream pipe

35

.

Therefore, the present examples and embodiments are to be considered as illustrative and not restrictive and the invention is not to be limited to the details given herein, but may be modified within the scope and equivalence of the appended claims.

Number	Name	Date	Kind
5191768	Fujil	Mar 1993	A
5884497	Kishita et al.	Mar 1999	A

Number	Date	Country
4068055	Jun 1994	JP
6-341378	Dec 1994	JP
9-228948	Sep 1997	JP

Control device of variable displacement compressor

Information

Patent Number

Date Filed

Date Issued

Inventors

Original Assignees

Examiners

Agents

CPC

US Classifications

Field of Search

US

International Classifications

Abstract

Description

Claims

Priority Claims (1)

US Referenced Citations (2)

Foreign Referenced Citations (3)