Air conditioner and control valve in variable displacement compressor

BACKGROUND OF THE INVENTION

The present invention relates to an air conditioner having a refrigerant circuit. More particularly, the present invention pertains to a displacement control valve used in a variable displacement compressor in a refrigerant circuit.

A typical refrigerant circuit of a vehicle air conditioner includes a condenser, an expansion valve, an evaporator and a compressor. The compressor receives refrigerant gas from the evaporator. The compressor then compresses the gas and discharges the gas to the condenser. The evaporator transfers heat to the refrigerant in the refrigerant circuit from the air in the passenger compartment. The pressure of refrigerant gas at the outlet of the evaporator, in other words, the pressure of refrigerant gas that is drawn into the compressor (suction pressure Ps), represents the thermal load on the refrigerant circuit.

Variable displacement swash plate type compressors are widely used in vehicles. Such compressors include a displacement control valve that operates to maintain the suction pressure Ps at a predetermined target level (target suction pressure). The control valve changes the inclination angle of the swash plate in accordance with the suction pressure Ps for controlling the displacement of the compressor. The control valve includes a valve body and a pressure sensing member such as a bellows or a diaphragm. The pressure sensing member moves the valve body in accordance with the suction pressure Ps, which adjusts the pressure in a crank chamber. The inclination of the swash plate is adjusted, accordingly.

In addition to the above structure, some control valves include an electromagnetic actuator, such as a solenoid, to change the target suction pressure. An electromagnetic actuator urges a pressure sensing member or a valve body in one direction by a force that corresponds to the value of an externally supplied current. The magnitude of the force determines the target suction pressure. Varying the target suction pressure permits the air conditioning to be finely controlled.

Such compressors are usually driven by vehicle engines. Among the auxiliary devices of a vehicle, the compressor consumes the most engine power and is therefore a great load on the engine. When the load on the engine is great, for example, when the vehicle is accelerating or moving uphill, all available engine power needs to be used for moving the vehicle. Under such conditions, to reduce the engine load, the compressor displacement is minimized. This will be referred to as a displacement limiting control procedure. A compressor having a control valve that changes a target suction pressure raises the target suction pressure when executing the displacement limiting control procedure. Then, the compressor displacement is decreased such that the actual suction pressure Ps is increased to approach the target suction pressure.

The graph of

FIG. 17

illustrates the relationship between suction pressure Ps and displacement Vc of a compressor. The relationship is represented by multiple lines in accordance with the thermal load in an evaporator. Thus, if the suction pressure Ps is constant, the compressor displacement Vc increases as the thermal load increases. If a level Ps

1

is set as a target suction pressure, the actual displacement Vc varies in a certain range (ΔVc in

FIG. 17

) due to the thermal load. If a high thermal load is applied to the evaporator during the displacement limiting control procedure, an increase of the target suction pressure does not lower the compressor displacement Vc to a level that sufficiently reduces the engine load.

Thus, the compressor displacement is not always controlled as desired as long as the displacement is controlled based on the suction pressure Ps.

SUMMARY OF THE INVENTION

Accordingly, it is an objective of the present invention to provide an air conditioner and a control valve used in a variable displacement compressor that accurately control the compressor displacement regardless of the thermal load on an evaporator.

To achieve the above objective, the present invention provides an air conditioner including a refrigerant circuit. The refrigerant circuit has a condenser, a decompression device, an evaporator and a variable displacement compressor. The compressor has a discharge pressure zone, the pressure of which is a discharge pressure, and a suction pressure zone, the pressure of which is a suction pressure. The refrigerant circuit further has a high pressure passage extending from the discharge pressure zone to the condenser and a low pressure passage extending from the evaporator to the suction pressure zone. A displacement control mechanism controls the displacement of the compressor based on the pressure difference between the pressure at a first pressure monitoring point located in the refrigerant circuit and the pressure at a second pressure monitoring point located in the refrigerant circuit. The first pressure monitoring point is located in a section of the refrigerant circuit that includes the discharge pressure zone, the condenser and the high pressure passage. The second pressure monitoring point is located in a section of the refrigerant circuit that includes the evaporator, the suction pressure zone and the low pressure passage.

The present invention also provides a control valve for controlling the pressure in a crank chamber of a compressor to change the displacement of the compressor. The compressor has a discharge pressure zone, the pressure of which is a discharge pressure, a suction pressure zone, the pressure of which is a suction pressure, and an internal gas passage that includes the discharge pressure zone, the crank chamber and the suction pressure zone. The control valve comprises a valve housing, a valve body, a pressure receiver and an actuator. The valve body is located in the valve housing to adjust the size of an opening in the internal gas passage. The pressure receiver actuates the valve body in accordance with the pressure difference between the discharge pressure and the suction pressure thereby causing the pressure difference to seek a predetermined target value. The actuator urges the valve body by a force, the magnitude of which corresponds to an external command. The urging force of the actuator represents the target value of the pressure difference.

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 a first embodiment of the present invention;

FIG. 2

is a schematic diagram illustrating a refrigerant circuit including the compressor of

FIG. 1

;

FIG. 3

is a cross-sectional view illustrating a control valve of

FIG. 1

;

FIG. 4

is a schematic cross-sectional view showing part of the control valve shown in

FIG. 3

;

FIG. 5

is a cross-sectional view taken along line

5

-

5

of

FIG. 1

;

FIG. 6

is an enlarged partial cross-sectional view illustrating a check valve of

FIG. 5

;

FIG. 7

is a flowchart showing a main routine for controlling a displacement;

FIG. 8

is a flowchart showing a normal control procedure;

FIG. 9

is a flow chart showing an exceptional control procedure;

FIG.

10

(

a

) is a timing chart showing changes of the duty ratio Dt of a voltage applied to a control valve during the exceptional control procedure;

FIG.

10

(

b

) is a timing chart showing changes of a discharge pressure Pd and a suction pressure Ps during the exceptional control procedure;

FIG.

10

(

c

) is a timing chart showing changes the compressor torque during the exceptional control procedure;

FIG. 11

is a cross-sectional view illustrating a control valve according to a second embodiment of the present invention;

FIG. 12

is a schematic cross-sectional view showing part of the control valve shown in

FIG. 1

;

FIG. 13

is a schematic diagram illustrating a refrigerant circuit according a third embodiment of the present invention;

FIG. 14

is an enlarged partial cross-sectional view illustrating a check valve in the compressor of

FIG. 13

;

FIG.

15

(

a

) is a timing chart showing changes of the duty ratio Dt of a voltage applied to a control valve during the exceptional control procedure;

FIG.

15

(

b

) is a timing chart showing changes of a discharge pressure Pd and a suction pressure Ps during the exceptional control procedure;

FIG.

15

(

c

) is a timing chart showing changes the compressor torque during the exceptional control procedure;

FIG. 16

is a schematic diagram illustrating a refrigerant circuit according a fourth embodiment of the present invention; and

FIG. 17

is a graph showing the relationship between the suction pressure Ps and the displacement Vc of a prior art compressor.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

A first embodiment of the present invention will now be described with reference to

FIGS. 1

to

10

(

c

). As shown in

FIG. 1

, a variable displacement swash plate type compressor used in a vehicle 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 cylinder block

1

and a front housing member

2

.

A recess is formed in the center of the cylinder block

1

. A spring

7

and a rear thrust bearing

9

B are located in the recess. The spring

7

urges the drive shaft

6

forward (to the left as viewed in

FIG. 1

) through the thrust bearing

9

B. A lug plate

11

is secured to the drive shaft

6

in the crank chamber

5

. A front thrust bearing

9

A is located between the lug plate

11

and the inner wall of the front housing member

2

.

The front end of the drive shaft

6

is connected to an external drive source, which is an engine E in this embodiment, through a power transmission mechanism PT. The power transmission mechanism PT includes a belt and a pulley. The mechanism PT may be a clutch mechanism, such as an electromagnetic clutch, which is electrically controlled from the outside. In this embodiment, the mechanism PT has no clutch mechanism. Thus, when the engine E is running, the compressor is driven continuously.

A drive plate, which is a swash plate

12

in this embodiment, is accommodated in the crank chamber

5

. The swash plate

12

has a hole formed in the center. The drive shaft

6

extends through the hole in the swash plate

12

. The swash plate

12

is coupled to the lug plate

11

by a hinge mechanism

13

. 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

has a guide hole and projects from the rear side of the lug plate

11

. Each guide pin

15

projects from the swash plate

12

. The guide hole of each support arm

14

receives the corresponding guide pin

15

. The hinge mechanism

13

permits the swash plate

12

to rotate integrally with the lug plate

11

and drive shaft

6

. The hinge mechanism

13

also permits the swash plate

12

to slide along the drive shaft

6

and to tilt with respect to a plane perpendicular to the axis of the drive shaft

6

. The swash plate

12

has a counterweight

12

a

, which is angularly spaced by 180 degrees 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

. A stopper ring

18

is fixed on the drive shaft

6

behind the swash plate

12

. A spring

17

is fitted about the drive shaft

6

between the stopper ring

18

and the swash plate

12

. When the swash plate

12

is at the maximum inclination angle position shown by the broken line in

FIG. 1

, the spring

17

does not apply force to the swash plate

12

. However, as the swash plate

12

is moved toward the minimum inclination angle position shown by the solid line in

FIG. 1

, the force of the spring

17

increases. The spring

17

is not fully contracted when the swash plate

12

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

Several cylinder bores

1

a

(only one shown) are formed about the axis of the drive shaft

6

in the cylinder block

1

. A single headed piston

20

is accommodated in each cylinder bore

1

a

. Each piston

20

and the corresponding cylinder bore

1

a

define a compression chamber. Each piston

20

is coupled to the swash plate

12

by a pair of shoes

19

. The swash plate

12

coverts rotation of the drive shaft

6

into reciprocation of each piston

20

.

A suction chamber

21

and a discharge chamber

22

are defined between the valve plate assembly

3

and the rear housing member

4

. The suction chamber

21

forms a suction pressure zone, the pressure of which is a suction pressure Ps. The discharge chamber

22

forms a discharge pressure zone, the pressure of which is a discharge pressure Pd. The valve plate assembly

3

has suction ports

23

, suction valve flaps

24

, discharge ports

25

and discharge valve flaps

26

. Each set of the suction port

23

, the suction valve flap

24

, the discharge port

25

and the discharge valve flap

26

corresponds to one of the cylinder bores

1

a

. 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

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

via the corresponding discharge port

25

and discharge valve

26

.

The inclination angle 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, which is based on pressures in the compressor. 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). In this embodiment, the crank pressure Pc is adjusted by a crank pressure control mechanism, which will be discussed below. Accordingly, the inclination angle of the swash plate

12

is adjusted to an angle between the maximum inclination and the minimum inclination. The inclination angle of the swash plate

12

defines the stroke of each piston

20

and the displacement of the compressor.

The contact between the counterweight

12

a

and a stopper

11

a

of the lug plate

11

prevents further inclination of the swash plate

12

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

16

and

17

.

The crank pressure control mechanism is located in the compressor to regulate the crank pressure Pc. As shown in

FIGS. 1 and 2

, the mechanism includes a bleed passage

27

, a supply passage

28

and a control valve

200

. The bleed passage

27

connects the crank chamber

5

with the suction chamber

21

to conduct refrigerant gas from the crank chamber

5

to the suction chamber

21

. The supply passage

28

connects the discharge chamber

22

with the crank chamber

5

to conduct refrigerant gas from the discharge chamber

22

to the crank chamber

5

. The control valve

200

is located in the supply passage

28

. The control valve

200

adjusts the flow rate of refrigerant gas supplied from the discharge chamber

22

to the crank chamber

5

through the supply passage

28

to control the crank pressure Pc. The bleed passage

27

and the supply passage

28

form an internal gas passage for circulating refrigerant gas in the compressor.

As shown in

FIGS. 1 and 2

, the refrigerant circuit of the vehicle air conditioner includes the compressor and an external circuit

30

, which is connected to the compressor. The external circuit

30

includes a condenser

31

, a decompression device and an evaporator

33

. The decompression device, which is a temperature-type expansion valve

32

, adjusts the flow rate of refrigerant supplied to the evaporator

33

based on the temperature or the pressure detected by a heat sensitive tube

34

, which is located downstream of the evaporator

33

. The temperature or the pressure at the downstream of the evaporator

33

represents the thermal load on the evaporator

32

. The external circuit

30

includes a low pressure pipe

35

, which extends from the evaporator

33

to the suction chamber

21

of the compressor, and a high pressure pipe

36

, which extends from the discharge chamber

22

of the compressor to the condenser

31

.

The difference between the discharge pressure Pd and the suction pressure Ps corresponds to the flow rate of refrigerant in the refrigerant circuit. That is, the pressure difference increases as the flow rate increases. In this embodiment, a first pressure monitoring point P

1

is located in the discharge chamber

22

, which is the most upstream section of the high pressure pipe

36

. A second pressure monitoring point P

2

is located in the suction chamber

21

, which is the most downstream section of the low pressure pipe

35

. In other words, the first pressure monitoring point P

1

is defined in the discharge pressure zone, which is a high pressure zone in the compressor, and the second pressure monitoring pint P

2

is defined in the suction pressure zone, which is the low pressure zone in the compressor. Detecting the difference (Pd−Ps) between the refrigerant gas pressure at the first monitoring point P

1

(the discharge pressure Pd) and the refrigerant gas pressure at the second monitoring point P

2

(the suction pressure Ps) permits the flow rate of refrigerant in the refrigerant circuit, or the compressor displacement, to be indirectly detected. The control valve

200

uses the pressure difference (Pd−Ps) as a parameter for controlling the compressor displacement.

The first pressure monitoring point P

1

need not be located in the discharge chamber

22

but may be at any location where the pressure is the discharge pressure Pd. That is, the first monitoring point P

1

may be in the discharge chamber

22

, in the condenser

31

or in the high pressure pipe

36

. Similarly, the second pressure monitoring point P

2

need not be located in the suction chamber

21

but may be at any location where the pressure is the suction pressure Ps. That is, the second monitoring point P

2

may be in the suction chamber

21

, in the evaporator

33

or in the low pressure pipe

35

.

A control valve

200

shown in

FIG. 3

is actuated by the pressure difference (Pd−Ps), which acts on the control valve

200

. The control valve

200

includes an inlet valve mechanism

50

and an electromagnetic actuator, which is a solenoid

100

in this embodiment. The inlet valve mechanism

50

adjusts the opening size of the supply passage

28

. The solenoid

100

applies a force that corresponds to the value of a supplied current to the inlet valve mechanism

50

through a rod

40

, which has a circular cross section. The rod

40

includes a divider

41

, a coupler

42

and a guide

44

. A part of the guide

44

that is located adjacent to the coupler

42

functions as a valve body

43

. As shown in

FIG. 4

, the cross-sectional area SB of the divider

41

is greater than the cross-sectional area of the coupler

42

. The cross-sectional area SD of the guide

44

and the valve body

43

is greater than the cross-sectional area SB of the divider

41

.

As shown in

FIG. 3

the control valve

200

has a valve housing

45

. The housing

45

includes an upper housing member

45

b

and a lower housing member

45

c

. The upper housing member

45

b

defines the shape of the inlet valve mechanism

50

. The lower housing member

45

c

defines the shape of the solenoid

100

. A plug

45

a

is fitted to an upper opening of the upper housing member

45

b

to close the opening. A valve chamber

46

and a guide hole

49

are formed in the upper housing member

45

b

. A pressure sensing chamber

48

is defined by the upper housing member

45

b

and the plug

45

a

. The upper housing member

45

b

has a wall that separates the pressure sensing chamber

48

from the valve chamber

46

. The guide hole

49

extends through the wall. Part of the guide hole

49

that opens to the valve chamber

46

functions as a valve hole

47

.

The rod

40

extends through the valve chamber

46

, the guide hole

49

and the pressure sensing chamber

48

. The rod

40

moves axially to selectively connect and disconnect the valve chamber

43

with the valve hole

47

. The diameter of the guide hole

49

is constant in the axial direction. The cross-sectional area SB of the guide hole

49

is equal to the cross-sectional area SB of the divider

41

of the rod

40

. Therefore, the divider

41

, which is located in the guide hole

49

, separates the pressure sensing chamber

48

from the valve chamber

46

. Hereinafter, the cross-sectional area of the guide hole

49

and the valve hole

47

will be referred to as SB, which also represents the cross-sectional are of the divider

41

.

A radial port

51

is formed in the upper housing member

45

b

and is connected to the valve chamber

46

. The valve chamber

46

is connected to the discharge chamber

22

through the port

51

and an upstream section of the supply passage

28

. A radial port

52

is also formed in the upper housing member

45

b

and is connected with the valve hole

47

. The valve hole

47

is connected to the crank chamber

5

through the port

52

and a downstream section of the supply passage

28

. The ports

51

,

52

, the valve chamber

46

and the valve hole

47

form a part of the supply passage

28

that is in the control valve

200

.

The valve body

43

is located in the valve chamber

46

. The cross-sectional area SB of the valve hole

47

is greater than the cross-sectional area SC of the coupler

42

and is smaller than the cross-sectional area SD of the guide

44

(see FIG.

4

). A step defined between the valve chamber

46

and the valve hole

47

functions as a valve seat

53

to receive the valve body

43

. When the valve body

43

contacts the valve seat

53

, the valve hole

47

is disconnected from the valve chamber

46

. When the valve body

43

is separated from the valve seat

53

as shown in

FIG. 3

, the valve hole

47

is connected to the valve chamber

46

.

A pressure receiver, which is a cup-shaped movable spool

54

in this embodiment, is located in the pressure sensing chamber

48

and moves axially. The spool

54

divides the pressure sensing chamber

48

into a high pressure chamber

55

and a low pressure chamber

56

. The spool

54

does not permit gas to flow between the higher pressure chamber

55

and the low pressure chamber

56

. The cross-sectional area SA of the bottom wall of the spool

54

is greater than the cross-sectional area SB of the divider

41

and the guide hole

49

(see FIG.

4

).

The higher pressure chamber

55

is connected to the discharge chamber

22

, in which the first pressure monitoring point P

1

is located, through a port

55

a

formed in the plug

45

a

and a first pressure introduction passage

37

. The low pressure chamber

56

is connected to the suction chamber

21

, in which the second pressure monitoring point P

2

is located, through a port

56

a

formed in the upper housing member

45

b

and a second pressure introduction passage

38

. Therefore, the higher pressure chamber

55

is exposed the discharge pressure Pd and the low pressure chamber

56

is exposed to the suction pressure Ps. The upper and lower surfaces of the spool

54

receive the discharge pressure Pd and the suction pressure Ps, respectively. The distal end of the rod

40

, which is located in the low pressure chamber

56

, is fixed to the spool

54

. The spool

54

, the high pressure chamber

55

and the low pressure chamber

56

form a pressure difference detection mechanism. A return spring

57

is located in the high pressure chamber

55

. The return spring

57

urges the spool

54

from the high pressure chamber

55

toward the low pressure chamber

56

.

The solenoid

100

includes a cup-shaped cylinder

61

, which is fixed in the lower housing member

45

c

. A stationary iron core

62

is fitted into an upper opening of the cylinder

61

. The stationary core

62

forms part of the inner walls of the valve chamber

46

and defines a plunger chamber

63

in the cylinder

61

. A plunger

64

is located in the plunger chamber

63

. The plunger

64

is moved axially. The stationary core

62

has guide hole

65

through which the guide

44

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

65

and the guide

44

. The space communicates the valve chamber

46

with the plunger chamber

63

. Thus, the plunger chamber

63

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

46

is exposed.

The lower portion of the guide

44

extends into the plunger chamber

63

. The plunger

64

is fixed to the lower portion of the guide

44

. The plunger

64

integrally moves with the rod

40

in the axial direction. A buffer spring

66

is located in the plunger chamber

63

and urges the plunger

64

toward the stationary core

62

.

A coil

67

is located about the stationary core

62

and a plunger

64

. A controller

70

supplies electricity to the coil

67

through a drive circuit

72

. The coil

67

generates an electromagnetic force F between the stationary core

62

and the plunger

64

. The magnitude of the force F corresponds to the value of the supplied electricity. The force F urges the plunger

64

toward the stationary core

62

, which moves the rod

40

. Accordingly, the valve body

43

is moved toward the valve seat

53

.

The force of the buffer spring

66

is weaker than the force of the return spring

57

. Thus, when electricity is not supplied to the coil

67

, the return spring

57

moves the plunger

64

and the rod

40

to an initial position shown in

FIG. 3

, which causes the valve body

43

to maximize the opening size of the valve hole

47

.

Electricity applied to the coil

67

may be changed either by changing the value of the voltage. Alternatively, the electricity may be changed by duty control. In this embodiment, the electricity is duty controlled. A smaller duty ratio Dt of the voltage applied to the coil

67

represents a smaller electromagnetic force F. A smaller force F causes the valve body

43

to increase the opening size of the valve hole

47

.

The opening size of the valve hole

47

by the valve body

43

is determined by the axial position of the rod

40

. The axial position of the rod

40

is determined by various forces acting on the rod

40

. The forces will be described with reference to

FIGS. 3 and 4

. Downward forces as viewed in

FIGS. 3 and 4

move the valve body

43

from the valve seat

53

(a valve opening direction). Upward forces as viewed in

FIGS. 3 and 4

move the valve body

43

toward the valve seat

53

(a valve closing direction).

Forces acting on the part of the rod

40

that is above the coupler

42

, that is, the forces acting on the divider

41

, will now be described. As shown in

FIGS. 3 and 4

, the divider

41

receives a downward force f

2

, which is applied by the return spring

57

, through the spool

54

. The spool

54

receives a downward force based on the pressure difference (Pd−Ps) between the discharge pressure Pd in the high pressure chamber

55

and the suction pressure Ps in the low pressure chamber

56

. The downward force based on the pressure difference (Pd−Ps) acts on the divider

41

. The area of the spool

54

that receives the discharge pressure Pd in the high pressure chamber

55

is equal to the cross-sectional area SA of the bottom wall of the spool

54

. The area of the spool

54

that receives the suction pressure Ps in the low pressure chamber

56

is computed by subtracting the cross-sectional area SB of the divider

41

from the cross-sectional area SA. The divider

41

also receives an upward force based on the pressure in the valve hole

47

, or the crank pressure Pc. The area of the divider

41

that receives the pressure in the valve hole

47

is computed by subtracting the cross-sectional area SC of the coupler

42

from the cross-sectional area SB of the divider

41

. If downward forces are represented by positive values, the net force ΣF

1

acting on the divider

41

is represented by an equation I.

ΣF

1

=

Pd·SA−Ps

(

SA−SB

)

−Pc

(

SB−SC

)

+f

2

Equation I

The forces acting on the part of the rod

40

that is below the coupler

42

, that is, the forces acting on the guide

44

, will now be described. The guide

44

receives an upward force f

1

of the buffer spring

66

and the upward electromagnetic force F, which acts on the plunger

64

. As shown in

FIG. 4

, the upper end surface

43

a

of the valve body

43

is divided into an inner section and an outer section by an imaginary cylinder, which is shown by broken lines in FIG.

4

. The imaginary cylinder corresponds to the wall defining the valve hole

47

. The pressure receiving area of the inner section is represented by SB−SC, and the pressure receiving area of the outer section is represented by SD−SB. The inner section receives a downward force based on the pressure in the valve hole

47

, or the crank pressure Pc. The outer section receives a downward force based on the discharge pressure Pd in the valve chamber

46

.

As described above, the plunger chamber

63

is exposed to the discharge pressure Pd of the valve chamber

46

. The upper surface and the lower surface of the plunger

64

have the same pressure receiving area. Therefore, the forces acting on the plunger

64

, which are based on the discharge pressure Pd, are cancelled. The lower end surface

44

a

of the guide

44

receives an upward force based on the discharge pressure Pd. The pressure receiving area of the lower end surface

44

a

is equal to the cross-sectional area SD of the guide

44

. If the upward forces are represented by positive values, the net force ΣF

2

acting on the guide

44

is represented by the following equation II.

\begin{matrix} \begin{matrix} Σ F2 = Pd \cdot SD - Pd (SD - SB) - Pc (SB - SC) + F + f1 \\ = Pd \cdot SB - Pc (SB - SC) + F + f1 \end{matrix} & Equation II \end{matrix}

In the process of simplifying equation II, −Pc·SD is canceled by +Pc·SD, and the term Pc·SB remains. Thus, the resultant of the downward and upward forces acting on the guide

44

based on the discharge pressure Pd is an upward force, and the magnitude of the resultant upward force is determined based only on the cross-sectional area SB of the valve hole

47

. The area of the part of the guide

44

that effectively receives the discharge pressure Pd, in other words, the effective discharge pressure receiving area of the guide

44

, is equal to the cross-sectional area SB of the valve hole

47

regardless of the cross-sectional area SD of the guide

44

.

The axial position of the rod

40

is determined such that the force ΣF

1

in the equation I and the force ΣF

2

in the equation II are equal. When the force ΣF

1

is equal to the force ΣF

2

(ΣF

1

=ΣF

2

), the following equation III is satisfied.

Pd−Ps=

(

F+f

1

−

f

2

)/(

SA−SB

) Equation III

In equation III, the electromagnetic force F is a variable parameter that changes in accordance with the power supplied to the coil

67

. As apparent from equation III, the rod

40

changes the pressure difference (Pd−Ps) according to changes of the electromagnetic force F. In other words, the rod

40

moves according to the pressure difference (Pd−Ps), which acts on the rod

40

, such that the pressure difference (Pd−Ps) seeks a target value TPD, which is determined by the electromagnetic force F.

The pressures that affect the axial position of the rod

40

are only the discharge pressure Pd and the suction pressure Ps. The force based on the crank pressure Pc does not influence the position of the rod

40

. Therefore, the rod

40

is actuated by the pressure difference (Pd−Ps), the electromagnetic force F and the spring forces f

1

, f

2

.

As described above, the downward force f

2

of the return spring

57

is greater than the upward force f

1

of the buffer spring

66

. Thus, when voltage is not applied to the coil

67

, in other words, when the electromagnetic force F is zero, the rod

40

is moved to the initial position shown in

FIG. 3

, which maximizes the opening size of the valve hole

47

by the valve body

43

. When the duty ratio Dt of the voltage applied to the coil

67

is minimum in a predetermined range, the resultant of the upward electromagnetic force F and the upward force f

1

of the buffer spring

66

is greater than the downward force f

2

of the return spring

57

. The resultant of the upward electromagnetic force F and the upward force f

1

of the buffer spring

66

acts against the resultant of the downward force f

2

of the return spring

57

and the downward force based on the pressure difference (Pd−Ps). The rod

40

is actuated for satisfying equation III. As a result, the position of the valve body

43

relative to the valve seat

53

, in other words, the opening size of the valve hole

47

, is determined. The flow rate of refrigerant gas from the discharge chamber

22

to the crank chamber

5

through the supply passage

28

corresponds to the opening size of the valve hole

47

. The crank pressure Pc is controlled accordingly.

When the electromagnetic force F is constant, the control valve

200

operates such that the pressure difference (Pd−Ps) seeks the target value TPD, which corresponds to the electromagnetic force F. When the electromagnetic force F is adjusted based on a command from the controller and the target pressure difference TPD is changed accordingly, the control valve

200

operates such that the pressure difference (Pd−Ps) seeks the new target value TPD.

As shown in

FIGS. 1

,

5

and

6

, the discharge chamber is connected to the high pressure pipe

36

of the external circuit

30

by a discharge passage

90

, which is formed in the rear housing member

4

. A check valve

92

is located in the discharge passage

90

. The check valve

92

and its mounting structure will be described below.

As shown in

FIGS. 5 and 6

, a valve pipe

97

for defining the discharge passage

90

protrudes from the periphery of the rear housing member

4

. A seat

91

is formed in the middle of the discharge passage

90

. The check valve

92

is press fitted in the seat

91

. A step

91

a

is formed between the seat

91

and the inlet of the discharge passage

90

to determine the position of the check valve

92

.

The check valve

92

includes a cylindrical case

96

. The case

96

includes a valve seat

93

. A valve hole

93

a

is formed in the valve seat

93

. A valve seat

94

and a spring

95

are housed in the case

96

. The spring

95

urges the valve body

94

toward the valve seat

93

. When the case

96

is press fitted into the seat

91

and contacts the step

91

a

, the check valve

92

is located at the appropriate position in the discharge passage

90

. Several through holes

96

a

are formed in the peripheral wall of the case

96

. A plug

96

c

is fitted into an opening of the case

96

that is opposite to the valve hole

93

a

. The plug

96

c

receives the spring

95

and has a pressure introduction hole

96

b

. Thus, the valve body

94

is exposed to the discharge pressure Pd in the discharge chamber

22

through the valve hole

93

a

. The valve body

94

is also exposed to a pressure Pd′ in the high pressure pipe

36

through the pressure introduction hole

96

b

. The valve body

94

selectively opens and closes the valve hole

93

a

in accordance with the difference between the pressures Pd and Pd′.

When the force based on the pressure difference (Pd−Pd′) is greater than the force of the spring

95

, the valve body

94

is separated from the valve seat

93

as shown in FIG.

5

and opens the valve hole

93

a

. Accordingly, refrigerant gas flows from the discharge chamber

22

to the high pressure pipe

36

. When the force based on the pressure difference (Pd−Pd′) is smaller than the force of the spring

95

, the valve body

94

contacts the valve seat

93

as shown in FIG.

6

and closes the valve hole

93

a

. Accordingly, the discharge chamber

22

is disconnected from the high pressure pipe

36

.

As shown in

FIGS. 2 and 3

, the controller

70

is a computer, which includes a CPU, a ROM, a RAM and an input-output interface. Detectors

71

detect various external information necessary for controlling the compressor and send the information to the controller

70

. The controller

70

computes an appropriate duty ratio Dt based on the information and commands the drive circuit

72

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

72

outputs the instructed pulse voltage having the duty ratio Dt to the coil

67

of the control valve

200

. The electromagnetic force F of the solenoid

100

is determined according to the duty ratio Dt.

The detectors

71

may include, for example, an air conditioner switch, a passenger compartment temperature sensor, a temperature adjuster for setting a desired temperature in the passenger compartment and a throttle sensor for detecting the opening size of a throttle valve of the engine E. The detectors

71

may also include a pedal position sensor for detecting the depression degree of the acceleration pedal of the vehicle. The opening size of the throttle valve and the depression degree of the acceleration pedal represent the load on the engine E.

The flowchart of

FIG. 7

shows the main routine for controlling the compressor displacement. When the vehicle ignition switch or the starting switch is turned on, the controller

70

starts processing. The controller

70

performs various initial setting in step S

71

. For example, the controller

70

assigns a predetermined initial value to the duty ratio Dt of the voltage applied to the coil

67

.

In step S

72

, the controller

70

waits until the air conditioner switch is turned on. When the air conditioner switch is turned on, the controller

70

moves to step S

73

. In step S

73

, the controller

70

judges whether the vehicle is in an exceptional driving mode. The exceptional driving mode refers to, for example, a case where the engine E is under high-load conditions such as when driving uphill or when accelerating rapidly. The controller

70

judges whether the vehicle is in the exceptional driving mode according to, for example, external information from the throttle sensor or the pedal position sensor.

If the outcome of step S

73

is negative, the controller

70

judges that the vehicle is in a normal driving mode and moves to step S

74

. The controller

70

then executes a normal control procedure shown in FIG.

8

. If the outcome of step S

73

is positive, the controller

70

executes an exceptional control procedure for temporarily limiting the compressor displacement in step S

75

. The exceptional control procedure differs according to the nature of the exceptional driving mode.

FIG. 9

illustrates an example of the exceptional control procedure that is executed when the vehicle is rapidly accelerated.

The normal control procedure of

FIG. 8

will now be described. In step S

81

, the controller

70

judges whether the temperature Te(t), which is detected by the temperature sensor, is higher than a desired temperature Te(set), which is set by the temperature adjuster. If the outcome of step S

81

is negative, the controller

70

moves to step S

82

. In step S

82

, the controller

70

judges whether the temperature Te(t) is lower than the desired temperature Te(set). If the outcome in step S

82

is also negative, the controller

70

judges that the detected temperature Te(t) is equal to the desired temperature Te(set) and returns to the main routine of

FIG. 7

without changing the current duty ratio Dt.

If the outcome of step S

81

is positive, the controller

70

moves to step S

83

for increasing the cooling performance of the refrigerant circuit. In step S

83

, the controller

70

adds a predetermined value ΔD to the current duty ratio Dt and sets the resultant as a new duty ratio Dt. The controller

70

sends the new duty ratio Dt to the drive circuit

72

. Accordingly, the electromagnetic force F of the solenoid

100

is increased by an amount that corresponds to the value ΔD, which moves the rod

40

in the valve closing direction. As the rod

40

moves, the force f

2

of the return spring

57

is increased. The axial position of the rod

40

is determined such that equation III is satisfied.

As a result, the opening size of the control valve

200

is decreased and the crank pressure Pc is lowered. Thus, the inclination angle of the swash plate

12

and the compressor displacement are increased. An increase of the compressor displacement increases the flow rate of refrigerant in the refrigerant circuit and increases the cooling performance of the evaporator

33

. Accordingly, the temperature Te(t) is lowered to the desired temperature Te(set) and the pressure difference (Pd−Ps) is increased.

If the outcome of S

82

is positive, the controller

70

moves to step S

84

for decreasing the cooling performance of the refrigerant circuit. In step S

84

, the controller

70

subtracts the predetermined value ΔD from the current duty ratio Dt and sets the resultant as a new duty ratio Dt. The controller

70

sends the new duty ratio Dt to the drive circuit

72

. Accordingly, the electromagnetic force F of the solenoid

100

is decreased by an amount that corresponds to the value ΔD, which moves the rod

40

in the valve opening direction. As the rod

40

moves, the force f

2

of the return spring

57

is decreased. The axial position of the rod

40

is determined such that equation III is satisfied.

As a result, the opening size of the control valve

200

is increased and the crank pressure Pc is raised. Thus, the inclination angle of the swash plate

12

and the compressor displacement are decreased. A decrease of the compressor displacement decreases the flow rate of refrigerant in the refrigerant circuit and decreases the heat reduction performance of the evaporator

33

. Accordingly, the temperature Te(t) is raised to the desired temperature Te(set) and the pressure difference (Pd−Ps) is decreased.

As described above, the duty ratio Dt is optimized in steps S

83

and S

84

such that the detected temperature Te(t) seeks the desired temperature Te(set).

The exceptional control procedure of

FIG. 9

will now be described. In step S

91

, the controller

70

stores the current duty ratio Dt as a restoration target value DtR. In step S

92

, the controller

70

stores the current detected temperature Te(t) as an initial temperature Te(INI), or the temperature when the displacement limiting control procedure is started.

In step S

93

, the controller

70

starts a timer. In step S

94

, the controller

70

changes the duty ratio Dt to zero percent and stops applying voltage to the coil

67

. Accordingly, the opening size of the control valve

200

is maximized by the return spring

57

, which increases the crank pressure Pc and minimizes the compressor displacement. As a result, the torque of the compressor is decreased, which reduces the load on the engine E when the vehicle is rapidly accelerated.

In step S

95

, the controller

70

judges whether the elapsed period STM measured by the timer is more than a predetermined period ST. Until the measured period STM surpasses the predetermined period ST, the controller

70

maintains the duty ratio Dt at zero percent. Therefore, the compressor displacement and torque are maintained at the minimum levels until the predetermined period ST elapses. The predetermined period ST starts when the displacement limiting control procedure is started. This permits the vehicle to be smoothly accelerated. Since acceleration is generally temporary, the period ST need not be long.

When the measured period STM surpasses the period ST, the controller

70

moves to step S

96

. In step S

96

, the controller

70

judges whether the current temperature Te(t) is higher than a value computed by adding a value β to the initial temperature Te(INI). If the outcome of step S

96

is negative, the controller

70

judges that the compartment temperature is in an acceptable range and maintains the duty ratio Dt at zero percent. If the outcome of step S

96

is positive, the controller

70

judges that the compartment temperature has increased above the acceptable range due to the displacement limiting control procedure. In this case, the controller

70

moves to step S

97

and restores the cooling performance of the refrigerant circuit.

In step

597

, the controller

70

executes a duty ratio restoration control procedure. In this procedure, the duty ratio Dt is gradually restored to the restoration target value DtR over a certain period. Therefore, the inclination of the swash plate

12

is changed gradually, which prevents the shock of a rapid change. In the chart of step S

97

, the period from time t

3

to time t

4

represents a period from when the duty ratio Dt is set to zero percent in step S

94

to when the outcome of step S

96

is judged to be positive. The duty ratio Dt is restored to the restoration target value DtR from zero percent over the period from the time t

4

to time t

5

. When the duty ratio Dt reaches the restoration target value DtR, the controller

70

moves to the main routine shown in FIG.

7

.

FIGS.

10

(

a

) to

10

(

c

) are timing charts showing changes of the duty ratio Dt, the discharge pressure Pd at the first pressure monitoring point P

1

, the suction pressure Ps at the second pressure monitoring point P

2

and the compressor torque. When the duty ratio Dt is set to zero percent at time t

3

, the opening size of the control valve

200

is maximized. At the same time, the displacement and the torque of the compressor are minimized. Accordingly, the discharge pressure Pd is lowered as shown by solid line

111

in FIG.

10

(

b

). Then, the check valve

92

disconnects the discharge chamber

22

from the high pressure pipe

36

to prevent back flow of highly pressurized gas from the high pressure pipe

36

to the discharge chamber

22

. Therefore, the discharge pressure Pd is quickly lowered. Since the flow rate of gas from the suction chamber

21

to the cylinder bores

1

a

is decreased and gas flows to the crank chamber

5

to the suction chamber

21

through the bleed passage

27

, the suction pressure Ps is increased as shown by solid line

112

in FIG.

10

(

b

). As a result, the difference between the discharge pressure Pd and the suction pressure Ps is quickly decreased from time t

3

to time t

4

, during which the compressor displacement is minimum. The check valve

92

functions as an accelerator that accelerates the reduction of the pressure difference (Pd−Ps).

The broken line

113

in FIG.

10

(

b

) represents changes of the discharge pressure Pd at the first pressure monitoring point P

1

when the check valve

92

is omitted. In this case, the discharge chamber

22

is constantly connected to the high pressure pipe

36

. To lower the discharge pressure Pd at the first monitoring point P

1

, the gas pressure in a large zone that includes the discharge chamber

22

and the high pressure pipe

36

must be lowered. Thus as shown by broken line

113

in FIG.

10

(

b

), the discharge pressure Pd is slowly decreased from time t

3

to time t

4

. Therefore, the difference between the discharge pressure Pd and the suction pressure Ps is not sufficiently lowered. This means that there is an excessive discrepancy between the pressure difference (Pd−Ps) and the compressor displacement.

The control valve

200

shown in

FIG. 3

operates to satisfy equation III for varying the compressor displacement. When the duty ratio Dt is zero percent, the electromagnetic force F of the solenoid

100

is eliminated. At this time the pressure difference (Pd−Ps) between the pressure monitoring points P

1

, P

2

must satisfy equation IV. Equation IV is the same as equation III except that the electromagnetic force F is zero. As the difference between the force f

1

of the buffer spring

66

and the force f

2

of the return spring

57

is decreased, the target value of the pressure difference (Pd−Ps) when the duty ratio Dt is zero percent approaches zero.

Pd=Ps=

(

f

1

−f

2

)/(

SA−SB

) Equation IV

To quickly and accurately control the compressor displacement according to changes of the duty ratio Dt, the actual pressure difference (Pd−Ps), which acts on the valve body

54

, must quickly and accurately respond to the target pressure difference TPD, which is changed by controlling the change of the duty ratio Dt. In the illustrated embodiment, the check valve

92

is located between the discharge chamber

22

and the high pressure pipe

36

. Therefore, as shown by solid line

111

in FIG.

10

(

b

), the discharge pressure Pd at the first monitoring point P

1

is quickly lowered after time t

3

, at which the duty ratio Dt is set to zero percent, and the actual pressure difference (Pd−Ps) quickly seeks a value that satisfies equation IV. Thus, the actual pressure difference (Pd−Ps), which acts on the spool

54

, greatly deviates from the target value TPD, which corresponds to the duty ratio Dt (zero percent), for a relatively short period. The period required for the actual pressure difference (Pd−Ps) to seek the target pressure difference TPD is in a permissible range (for example, from time t

3

to time t

4

).

At time t

4

, a duty ratio restoration control procedure is started. Then, the opening size of the control valve

200

is gradually decreased such that the actual pressure difference (Pd−Ps) increases in accordance with the increase of the duty ratio Dt. As shown by solid line

115

in FIG.

10

(

c

), the compressor displacement substantially accurately changes in accordance with the increase of the duty ratio Dt from time t

4

to time t

5

, at which the duty ratio restoration control procedure is finished.

If the check valve

92

is omitted from the compressor of

FIG. 1

, the discharge pressure Pd at the first monitoring point P

1

will change as shown by the broken line

113

. That is, after time t

3

, at which the duty ratio Dt is set to zero percent, the discharge pressure Pd is slowly decreased and does not quickly seek a value that satisfies equation IV. At time t

4

, at which the duty ratio restoration control procedure is started, the actual pressure difference (Pd−Ps), which acts on the spool

54

, differs greatly from the target pressure difference TPD, which corresponds to duty ratio Dt (zero percent).

The duty ratio Dt is gradually increased from time t

4

to t

5

. However, the control valve

200

is fully opened after time t

4

such that the actual pressure difference (Pd−Ps) is lowered to the target pressure difference TPD, which corresponds to the current duty ratio Dt. At time t

6

, the actual pressure difference (Pd−Ps) matches the target pressure difference TPD, which corresponds to the current duty ratio Dt. Although the duty ratio Dt is gradually increased during a period from time t

4

to time t

6

, the control valve

200

is kept fully opened. Thus, as shown by the broken line

114

in FIG.

10

(

c

), the compressor displacement is maintained at the minimum value during the period from time t

4

to time t

6

. After time t

6

, the displacement and the torque of the compressor are suddenly increased due to a decrease of the opening size of the control valve

200

, which produces a shock.

In this manner, if the check valve

92

is omitted, the displacement and the torque of the compressor are not gradually increased as shown by solid line

115

in FIG.

10

(

c

) when the duty ratio Dt is changed from zero percent to the restoration target value DtR. The check valve

92

is very effective for changing the compressor displacement in accordance with changes of the duty ratio Dt.

This embodiment has the following advantages.

The control valve

200

does not directly control the suction pressure Ps, which is influenced by the thermal load on the evaporator

33

. The control valve

200

directly controls the pressure difference (Pd−Ps) between the pressures at the pressure monitoring points P

1

, P

2

in the refrigerant circuit for controlling the compressor displacement. Therefore, the compressor displacement is controlled regardless of the thermal load on the evaporator

33

. During the exceptional control procedure, voltage is not applied to the control valve

200

, which quickly minimizes the compressor displacement. Accordingly, during the exceptional control procedure, the displacement is limited and the engine load is decreased. The vehicle therefore runs smoothly.

During the normal control procedure, the duty ratio Dt is controlled based on the detected temperature Te(t) and the target temperature Te(set), and the rod

40

is actuated in accordance with the pressure difference (Pd−Ps). That is, the control valve

200

not only operates based on external commands but also automatically operates in accordance with the pressure difference (Pd−Ps), which acts on the control valve

200

. The control valve

200

therefore effectively controls the compressor displacement such that the actual temperature Te(t) seeks the target temperature Te(set) and stably maintains the target temperature Te(set). Further, the control valve

200

quickly changes the compressor displacement when necessary.

The check valve

92

is located between the discharge chamber

22

and the high pressure pipe

36

. The check valve

92

permits the compressor displacement to accurately respond to changes of the duty ratio Dt. Therefore, the compressor displacement is accurately controlled in a desired pattern by controlling the duty ratio Dt.

When the compressor displacement is minimum, the check valve

92

disconnects the discharge chamber

22

from the high pressure pipe

36

. Therefore, when the compressor displacement is minimum, a gas circuit is formed within the compressor. The gas circuit includes the cylinder bores

1

a

, the discharge chamber

22

, the supply passage

28

, the crank chamber

5

, the bleed passage

27

and the suction chamber

21

. The refrigerant gas contains atomized oil. The oil is circulated in the gas circuit with the circulation of refrigerant gas and lubricates the moving parts of the compressor. Thus, when the air conditioner is not operating, the moving parts in the compressor are lubricated.

A second embodiment of the present invention will now be described with reference to

FIGS. 11 and 12

. The second embodiment is different from the embodiment of

FIGS. 1

to

10

(

c

) in the structure of the control valve

200

and in that the first pressure introduction passage

37

is omitted. In the second embodiment, an upstream section of the supply passage

28

functions as the first pressure introduction passage

37

. Otherwise, the embodiment of

FIGS. 11 and 12

is the same as the embodiment of the

FIGS. 1

to

10

(

c

). Like or the same reference numerals are given to those components that are like or the same as the corresponding components of the embodiment of

FIGS. 1

to

10

(

c

).

As shown in

FIGS. 11 and 12

, the rod

40

includes a guide

44

. The valve body

43

is formed in the distal portion of the guide

44

. The cross-sectional area of the guide

44

and the valve body

43

is represented by SF.

The housing member

45

b

includes an upper port

80

. The upper port

80

is communicated with the valve chamber

46

and faces the valve body

43

. The valve chamber

46

is connected to the discharge chamber

22

by the upper port

80

and an upstream section of the supply passage

28

. The cross-sectional area SG of the upper port

80

is smaller than the cross-sectional area SF of the valve body

43

. A step defined between the valve chamber

46

and the upper port

80

functions as a valve seat

81

. The upper port

80

functions as a valve hole. When the valve body

43

contacts the valve seat

81

, the upper port

80

is disconnected from the valve chamber

46

.

A radial center port

82

is formed in the upper housing member

45

b

and is communicated with the valve chamber

46

. The valve chamber

46

is connected to the crank chamber

5

through the center port

82

and a downstream section of the supply passage

28

. The valve body

43

adjusts the opening size of the supply passage

28

according to the axial position of the rod

40

.

The lower housing member

45

c

defines the shape of the lower portion of the solenoid

100

. A radial lower port

83

is formed in the lower housing member

45

c

. The lower port

83

is connected to the suction chamber

21

through the second pressure introduction passage

38

. The stationary iron core

62

includes an axial slit

84

. The slit

84

defines a passage that connects the lower port

83

to the plunger chamber

63

between the inner wall of the cylinder

61

and the stationary core

62

. The plunger chamber

63

is therefore exposed to the suction pressure Ps.

The end surface

43

a

of the valve body

43

receives the discharge chamber Pd in the upper port

80

and a crank pressure Pc in the valve chamber

46

. The guide

44

and the plunger

64

receive the suction chamber Ps in the plunger chamber

63

. There is no space between the guide

44

and the inner wall of a guide hole

65

formed in the stationary core

62

. Therefore, the valve chamber

46

is disconnected from the plunger chamber

63

. Unlike the control valve

200

in

FIG. 3

, the control valve of

FIGS. 11 and 12

has no spool

54

. The rod

40

functions as a pressure receiver.

A return spring

85

is located in the plunger chamber

63

. The return spring

85

urges the plunger

64

away from the stationary iron core

62

. When electricity is not supplied to the coil

67

, the return spring

85

moves the plunger

64

and the rod

40

to an initial position shown in

FIG. 11

, which causes the valve body

43

to maximize the opening size of the upper port

80

.

Axial forces acting on the rod

40

will now be described with reference to FIG.

12

. The upper end surface

43

a

of the valve body

43

is divided into an inner section and an outer section by an imaginary cylinder, which is shown by broken lines in FIG.

12

. The imaginary cylinder corresponds to the wall defining the upper port

80

. The pressure receiving area of the inner section is represented by SG, and the pressure receiving area of the outer section is represented by SF−SG. The inner section receives a downward force based on the discharge pressure Pd in the upper port

80

. The outer section receives a downward force based on the crank pressure Pc in the valve chamber

46

.

The guide

44

receives a downward force f

3

of the buffer spring

85

and an upward electromagnetic force F, which acts on the plunger

64

. The suction pressure Ps in the plunger chamber

63

urges the guide

44

and the plunger

64

upward. An effective pressure receiving area of the guide

44

and the plunger

64

that receives the suction pressure Ps in the plunger chamber

63

is equal to the cross-sectional area SF of the guide

44

.

The axial position of the rod

40

is determined such that the sum of the forces is zero. When the sum is zero, the following equation V is satisfied. In equation V, downward forces have positive values.

Pd·SG+Pc

(

SF−SG

)

+f

3

−

Ps·SF−F=

0 Equation V

Equation V can be modified to form the following equation VI.

(

Pd−Ps

)

SG+

(

Pc−Ps

)(

SF−SG

)

=F−f

3

Equation VI

In equation VI, the pressure difference (Pc−Ps) is negligible compared to the pressure difference (Pd−Ps). The area (SF−SG) is negligible compared to the area SG. If the pressure difference (Pc−Ps) and the area (SF−SG) are zero, the following equation VII is satisfied.

Pd−Ps≈

(

F−f

3

)/

SG

Equation VII

As apparent from equation VII, the rod

40

changes the pressure difference (Pd−Ps) according to changes of the electromagnetic force F. In other words, the rod

40

moves according to the pressure difference (Pd−Ps), which acts on the rod

40

, such that the pressure difference (Pd−Ps) seeks a target value TPD, which is determined by the electromagnetic force F. The pressures that affect the axial position of the rod

40

are only the discharge pressure Pd and the suction pressure Ps. The force based on the crank pressure Pc does not influence the position of the rod

40

. Therefore, the rod

40

is actuated by the pressure difference (Pd−Ps), the electromagnetic force F and the spring forces f

3

.

Although the control valve

200

of

FIGS. 11 and 12

has no spool

54

, the control valve

200

operates in the same manner as the control valve

200

of FIG.

3

. The control valve

200

of

FIGS. 11 and 12

is therefore simple and compact.

In the control valve

200

of

FIGS. 11 and 12

, the diameter of the upper port

80

may be equal to the diameter of the valve body

43

. In this case, the supply passage

28

is closed when the valve body

43

enters the upper port

80

. The cross-sectional area SG of the upper port

80

is equal to the cross-sectional area SF of the valve body

43

. Thus, the area SG can be replaced by the area SF in equation V, which satisfies the following equation VIII.

Pd·SF+f

3

−

Ps·SF−F=

0 Equation VIII

Equation V can be modified to form the following equation IX.

Pd−Ps=

(

F−f

3

)/SF Equation IX

Therefore, if the diameter of the upper port

80

is equal to the diameter of the valve body

43

, the control valve operates in the same manner as the control valve of

FIGS. 11 and 12

. That is, the rod

40

moves according to the pressure difference (Pd−Ps) such that the pressure difference (Pd−Ps) seeks a target value TPD, which is determined by the electromagnetic force F. The force based on the crank pressure Pc does not influence the position of the rod

40

and the rod

40

is actuated by the pressure difference (Pd−Ps), the electromagnetic force F and the spring forces f

3

.

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.

FIGS.

13

and

15

(

c

) show a third embodiment. A check valve

92

is located between the suction chamber

21

and the low pressure pipe

35

. The suction chamber

21

is connected to the low pressure pipe

35

through a suction passage

190

formed in the rear housing member

4

. A step

191

a

and a seat

191

are formed at the outlet of the suction passage

190

. The check valve

92

is press fitted in the seat

191

. The step

191

a

determines the axial position of the check valve

92

.

The structure of the check valve

92

is the same as that of

FIG. 5. A

valve body

94

receives a pressure Ps′ in the low pressure pipe

35

from a valve hole

93

a

and the suction pressure Ps in the suction chamber

21

through a pressure introduction hole

96

b

. The valve body

94

selectively opens and closes the valve hole

93

a

in accordance with the difference between the pressures Ps′ and Ps.

When the force based on the pressure difference (Ps′−Ps) is greater than the force of a spring

95

, which acts on the valve body

94

, the valve body

94

is separated from a valve seat

93

and opens the valve hole

93

a

as shown in FIG.

14

. This permits refrigerant gas to flow from the low pressure pipe

35

to the suction chamber

21

. When the force based on the pressure difference (Ps′−Ps) is smaller than the force of the spring

95

, the valve body

94

contacts the valve seat

93

and closes the valve hole

93

a

, which disconnects the low pressure pipe

35

from the suction chamber

21

. Accordingly, gas circulation in the refrigerant circuit is stopped. When the compressor displacement is minimum, the check valve

92

is closed.

In the embodiment of

FIGS. 13 and 14

, refrigerant gas discharged from the discharge chamber

22

is not supplied to the high pressure pipe

36

when the check valve

92

is closed. In this state, refrigerant gas circulates within the compressor.

Timing charts of FIGS.

15

(

a

) to

15

(

c

) correspond to those of FIGS.

10

(

a

) to

10

(

c

). When the duty ratio Dt is set to zero percent at time t

3

, the opening size of the control valve

200

is maximized. At the same time, the displacement and the torque of the compressor are minimized. Accordingly, the discharge pressure Pd is lowered as shown by solid line

117

in FIG.

15

(

b

). Also, the flow rate of refrigerant in the refrigerant circuit is decreased and the pressure Ps′ in the low pressure pipe

35

is lowered. Then, the check valve

92

disconnects the low pressure pipe

35

from the suction chamber

21

to prevent back flow of refrigerant gas from the suction chamber

21

to the low pressure pipe

35

. Refrigerant gas constantly flows from the crank chamber

5

to the suction chamber

21

through the bleed passage

27

. Therefore, as shown by solid line

116

in FIG.

15

(

b

), the suction pressure Ps is quickly increased. As a result, the difference between the discharge pressure Pd an the suction pressure Ps is quickly decreased from time t

3

to time t

4

, during which the compressor displacement is minimum.

In the embodiment of

FIGS. 13

to

15

(

c

), the actual pressure difference (Pd−Ps) quickly and accurately responds to changes of the duty ratio Dt. Therefore, the compressor displacement accurately responds to changes of the duty ratio Dt, which permits the compressor displacement to be accurately controlled along a desired pattern by controlling the duty ratio Dt.

FIG. 16

illustrates a fourth embodiment of the present invention. A check valve

92

is located between the discharge chamber

22

and the high pressure pipe

36

. Another check valve

92

is located between the suction chamber

21

and the low pressure pipe

35

.

Instead of the supply passage

28

, the bleed passage

27

may be regulated by the control valve. In this case, the flow rate of refrigerant gas from the crank chamber

5

to the suction chamber

21

is adjusted by the control valve.

The temperature-type expansion valve

32

may be replaced by a fixed restrictor.

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	Date	Country
406180155	Jun 1994	JP
11-294328	Oct 1999	JP

Air conditioner and control valve in 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 (1)

Foreign Referenced Citations (2)