Variable displacement compressor and air conditioning apparatus

BACKGROUND OF THE INVENTION

The present invention relates to variable displacement compressors varying displacement in a range from minimum to maximum and air conditioning apparatuses incorporating the compressors.

A typical air conditioning apparatus for vehicles has a refrigerant circuit including a condenser, a pressure reducing device (for example, an expansion valve), an evaporator, and a compressor. The compressor recently adopted is often a variable displacement compressor (particularly, a swash plate type variable displacement compressor) that is flexible to meet various air-conditioning requirements. Generally, a prior-art swash plate type variable displacement compressor varies its displacement by maintaining the pressure acting on an evaporator outlet (suction pressure Ps) at a predetermined target value (target suction pressure). That is, the compressor has a displacement control valve that controls the compressor displacement in a feedback manner in accordance with the suction pressure Ps, which serves as a reference indicator, such that the displacement corresponds to the cooling load of the compressor. More specifically, a pressure sensitive member, such as a bellows or a diaphragm, detects the suction pressure Ps. The movement of the pressure sensitive member positions a valve body to adjust the opening size of the control valve. This varies the pressure (crank pressure Pc) in a swash plate chamber (crank chamber) to alter an inclination angle of the swash plate. That is, the piston stroke is varied in accordance with the inclination angle of the swash plate, which is controlled in a range from a minimum inclination angle θmin to a maximum inclination angle θmax. The compressor displacement is thus adjusted as necessary in a range from minimum to a maximum.

However, a detailed operation analysis regarding this swash plate type variable displacement compressor indicates that the compressor is not capable of ensuring a uniform operational efficiency for the entire range in which the displacement is varied. The operational efficiency of the compressor (or an air conditioning apparatus including the compressor) is represented by a coefficient of performance (COP) and is indicated by the following equation: COP=Q/L. In the equation, Q indicates refrigerating performance (heat absorbing performance of the evaporator), and L indicates the power supplied to the compressor (workload of the compressor). As the COP increases, the operational efficiency of the compressor increases.

FIG. 7

is a graph in which refrigerating performance ratio (Q/Q

0

) is plotted along the horizontal axis (X-axis) and power ratio (L/L

0

) is plotted along the vertical axis (Y-axis). Q

0

indicates a maximum refrigerating performance. If the equation Q=Q

0

is satisfied, the refrigerating performance ratio Q/Q

0

is 100%. In the same manner, L

0

indicates a maximum power supplied to the compressor. If the equation L=L

0

is satisfied, the power ratio L/L

0

is 100%. In the graph, a diagonal broken line extends from the origin (0, 0) to a point indicating a maximum performance: (L

0

/L

0

, Q

0

/Q

0

)=(1, 1). Along this diagonal straight line, the following equation is satisfied: Q/Q

0

=L/L

0

. Based on this equation, the following equation is obtained: Q

0

/L

0

=Q/L=COP. In other words, the area located above the diagonal straight line in the graph of

FIG. 7

indicates a decrease in the COP, as compared to the maximum performance COP (COP=Q

0

/L

0

). In contrast, the area located below the diagonal straight line in the graph indicates an increased COP, as compared to the maximum performance COP (COP=Q

0

/L

0

)

As shown in

FIG. 7

, the graph includes three curves. The curves indicate characteristics of the swash plate type variable displacement compressor operated under different conditions regarding the suction pressure Ps and the like. The conditions are varied among the curves. As indicated by the graph, each curve crosses the diagonal straight line at a point P (referred to as the “points of divergence”). In an area of the power ratio located above each point P, as viewed in the graph, corresponding sections of the curves are located below the diagonal line. These sections of the curves thus indicate a relative increase in the COP, as compared to the maximum performance COP. In contrast, in an area of the power ratio located downward with respect to the points P, corresponding sections of the curve are located above the diagonal line. These sections of the curves thus indicate a relative reduction of the COP, as compared to the maximum performance COP. The power L supplied to the compressor increases as the inclination angle of the swash plate, or the compressor displacement, increases. Accordingly, as is clear from the graph of

FIG. 7

, the operational efficiency of the compressor decreases if the power supplied to the compressor is smaller than the value corresponding to the point P, or if the displacement is relatively small. Further, if the power supplied to the compressor is greater than the value corresponding to the point P, or the displacement is relatively large, the operational efficiency of the compressor is improved.

It is assumed that the lower operational efficiency during the relatively small displacement operation is caused by the following: (a) a reduced piston stroke decreases the sealing effect between the outer surface of each piston and the inner wall of the corresponding cylinder bore, thus increasing gas leakage from the cylinder bore to the crank chamber; (b) a greater amount of gas must be supplied to the crank chamber from the discharge chamber to maintain the crank pressure Pc at a relatively high level during lower displacement operation, and the amount of waste gas is increased; and (c) the proportion of mechanical power loss caused by friction for moving movable parts including the swash plate is increased during lower displacement operation.

As described, even though the compressor is capable of controlling of the displacement continuously in the entire range from minimum to maximum, this control is not necessarily advantageous regarding the operational efficiency of the compressor.

SUMMARY OF THE INVENTION

Accordingly, it is an objective of the present invention to provide a variable displacement compressor, the operational efficiency of which is improved by avoiding operation under conditions that reduce operational efficiency, and an air conditioning apparatus employing this variable displacement compressor.

To achieve the above objective, the present invention is a variable displacement compressor that varies the displacement in a variation range including a minimum displacement and a maximum displacement. The compressor includes an acquiring device for acquiring a target value used for controlling the compressor displacement, a switching device, which compares the target value with a predetermined reference value and switches an operational mode in accordance with a result from the comparison such that the displacement corresponding to the target value achieves a coefficient of performance equal to or greater than a predetermined level, and an actuator for varying the displacement in accordance with an instruction from at least the switching device.

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 features of the present invention that are believed to be novel are set forth with particularity in the appended claims. 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 view schematically showing an example of a refrigerant circuit of an air conditioning apparatus;

FIG. 2

is a cross-sectional view showing a swash plate type variable displacement compressor;

FIG. 3

is a cross-sectional view showing a control valve of the compressor of

FIG. 2

;

FIG. 4

is a cross-sectional view schematically explaining an effective pressure receiving area of the control valve of

FIG. 3

;

FIG. 5

is a flowchart showing a main routine of a displacement control procedure;

FIG. 6

is a flowchart showing a normal control routine of the procedure;

FIG. 7

is a graph showing a general variation of a refrigerating performance ratio in relation to a power ratio;

FIG. 8

is a graph corresponding to the graph of

FIG. 7

regarding an embodiment of the present invention;

FIG. 9

is a graph showing variation of a duty ratio of a drive signal in relation to compressor displacement;

FIG. 10

is a graph showing variation of the refrigerating performance ratio in relation to the duty ratio; and

FIG. 11

is a timing chart showing an example of variation of the duty ratio and variation of a passenger compartment temperature.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

An embodiment of an air conditioning apparatus for vehicles according to the present invention will now be described with reference to the attached drawings.

As shown in

FIG. 1

, the air conditioning apparatus has a refrigerant circuit (refrigerating circuit) including a swash plate type variable displacement compressor CM and an external refrigerant circuit

30

. The external refrigerant circuit

30

has, for example, a condenser

31

, an expansion valve

32

, which is a pressure reducing device, an evaporator

33

, a refrigerant passage

35

, and a refrigerant passage

36

. The passage

35

connects an outlet of the evaporator

33

to a suction chamber

21

of the compressor CM, and the passage

36

connects a discharge chamber

22

of the compressor CM to an inlet of the condenser

31

. Refrigerant gas is supplied to the suction chamber

21

from the evaporator

33

through the passage

35

. The compressor CM draws the refrigerant gas from the suction chamber

21

and compresses the gas. The compressed gas is sent to the discharge chamber

22

. The high-pressure gas in the discharge chamber

22

is then supplied to the condenser

31

through the passage

36

. The expansion valve

32

internally controls its opening size in a feedback manner in accordance with the temperature and pressure of refrigerant gas, which are detected by a sensor

34

located in the vicinity of the outlet of the evaporator

33

. The amount of the refrigerant gas supplied from the condenser

31

to the evaporator

33

thus corresponds to cooling load of the compressor CM. In this manner, the amount of the refrigerant flowing in the external refrigerant circuit

30

is directly adjusted.

General Structure of Compressor

As shown in

FIG. 2

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

1

, a front housing member

2

, and a rear housing member

4

. The front housing member

2

is secured to a front end of the cylinder block

1

, which is the left end in FIG.

2

. The rear housing member

4

is connected to a rear end of the cylinder block

1

with a valve plate

3

provided between the rear housing member

4

and the cylinder block

1

. The cylinder block

1

, the front housing member

2

, the valve plate

3

, and the rear housing member

4

form a housing of the compressor CM. A crank chamber

5

is formed in the housing. A drive shaft

6

extends through the crank chamber

5

and is rotationally supported by the housing.

A lug plate

11

is secured to the drive shaft

6

and rotates integrally with the drive shaft

6

. The drive shaft

6

and the lug plate

11

, which are integrally connected to each other, are urged toward the front housing member

2

by a spring

7

and positioned in thrust direction. The drive shaft

6

has a front end connected to an external drive source, which is an engine E of a vehicle in this embodiment, through a power transmitting mechanism PT. In this embodiment, the power transmitting mechanism PT is a clutchless mechanism that transmits power constantly (for example, a combination of a belt and a pulley). A cam plate, which is a swash plate

12

in this embodiment, is accommodated in the crank chamber

5

. The swash plate

12

is operationally connected to the lug plate

11

and the drive shaft

6

by means of a hinge mechanism

13

. The hinge mechanism

13

includes a pair of support arms

14

(only one is shown in

FIG. 2

) and a pair of guide pins

15

(only one is shown in FIG.

2

).

Each support arm

14

projects from a rear side of the lug plate

11

, and each guide pin

15

projects from a front side of the swash plate

12

. The support arms

14

cooperate with the associated guide pins

15

. The drive shaft

6

extends through a through hole formed in the swash plate

12

and contacts with the swash plate

12

by way of the through hole. Accordingly, the swash plate

12

rotates integrally with the lug plate

11

and the drive shaft

6

through the engagement by hinge mechanism

13

and the contact in the through hole. Further, the swash plate

12

inclines with respect to the drive shaft

6

while sliding axially along the drive shaft

6

. An inclination angle reducing spring

16

is provided around the drive shaft

6

and extends between the lug plate

11

and the swash plate

12

. The spring

16

urges the swash plate

12

toward the cylinder block

1

for decreasing the inclination angle of the swash plate

12

. A return spring

17

is provided around the drive shaft

6

and extends between the swash plate

12

and a restriction ring

18

secured to the drive shaft

6

. When the swash plate

12

is inclined by a maximum inclination angle (as indicated by the broken line in FIG.

2

), the spring

17

does not affect the swash plate

12

. However, if the inclination angle of the swash plate

12

decreases (as indicated by the solid line in FIG.

2

), the return spring

17

is compressed between the swash plate

12

and the restriction ring

18

. The spring

17

thus urges the swash plate

12

away from the cylinder block

1

.

A plurality of cylinder bores

1

a

(only one is shown in

FIG. 2

) are formed in the cylinder block

1

. Each cylinder bore

1

a

accommodates a single-headed piston

20

, and the piston

20

moves in the cylinder bore

1

a

. A front end of each piston

20

is connected to the outer periphery of the swash plate

12

through a pair of shoes

19

. The shoes

19

connect the piston

20

to the swash plate

12

. Thus, when the swash plate

12

rotates integrally with the drive shaft

6

, the rotation of the swash plate

12

is converted to linear movement of each piston

20

. The stroke of the piston

20

corresponds to the inclination angle θ of the swash plate

12

. A suction chamber

21

and a discharge chamber

22

are formed by the valve plate

3

and the rear housing member

4

. The suction chamber

21

is encompassed by the discharge chamber

22

. The valve plate

3

includes suction ports

23

, suction valves

24

selectively opening and closing the associated suction ports

23

, discharge ports

25

, and discharge valves

26

selectively opening and closing the associated discharge ports

25

.

Each cylinder bore

1

a

corresponds to one suction port

23

and the associated suction valve

24

as well as one discharge port

25

and the associated discharge valve

26

. When each piston

20

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

21

(a zone in which the suction pressure Ps acts), which is introduced from the outlet of the evaporator

33

, is drawn to the cylinder bore

1

a

through the suction port

23

opened by the associated suction valve

24

. The refrigerant gas in the cylinder bore

1

a

is then compressed to a predetermined pressure when the piston

20

moves from its top dead center to its bottom dead center. The compressed gas is discharged from the cylinder bore

1

a

to the discharge chamber

22

(a zone in which the discharge pressure Pd acts) through the discharge port

25

opened by the associated discharge valve

26

. More specifically, when the drive shaft

6

is rotated by the power from the engine E, the swash plate

12

is rotated as inclined by an angle θ. The angle θ is defined as an angle formed between a hypothetical plane extending perpendicular to the axis of the drive shaft

6

and the swash plate

12

. When the swash plate

12

is rotated, each piston

20

is moved by a stroke corresponding to the inclination angle θ of the swash plate

12

. The pistons

20

repeatedly perform the above operation, which is drawing refrigerant gas to the cylinder bores

1

a

, compression of the gas, and discharge of the gas from the cylinder bores

1

a.

The inclination angle θ is determined according to the equilibrium of various moments including a rotation moment caused by centrifugal force generated by the swash plate

12

, a moment caused by the force of the spring

16

(and the spring

17

), a moment caused by the force of inertia generated by reciprocating movement of each piston

20

, and a gas pressure moment. The gas pressure moment is generated in accordance with the pressure in each cylinder bore

1

a

and the pressure in the crank chamber

5

(crank pressure Pc), which act on opposite sides of the piton

20

. The gas pressure moment thus acts either to increase or decrease the inclination angle θ of the swash plate

12

, in accordance with the crank pressure Pc. In this embodiment, the crank pressure Pc is adjusted by the displacement control valve, which will be described later, thus altering the gas pressure moment. This adjusts the inclination angle θ of the swash plate

12

to a desired value in a range from a minimum inclination angle θmin to a maximum inclination angle θmax. The maximum inclination angle θmax is mechanically determined by a counterweight

12

a

of the swash plate

12

abutting against a restricting portion

11

a

of the lug plate

11

. The minimum inclination angle θmin is determined in accordance with the force of the spring

16

and the force of the return spring

17

acting against the spring

16

when the gas pressure moment is substantially maximized in the direction in which the inclination angle is decreased.

The inclination angle θ of the swash plate

12

is thus controlled in accordance with the crank pressure Pc. A mechanism for controlling the crank pressure Pc is formed by a bleed passage

27

and a supply passage

28

, which both extend in the housing of the compressor, and the control valve CV, which is an actuator. The bleed passage

27

connects the suction chamber

21

to the crank chamber

5

. The supply passage

28

connects the discharge chamber

22

to the crank chamber

5

. The control valve CV is provided in the supply passage

28

. The amount of high-pressure gas supplied to the crank chamber

5

through the supply passage

28

is altered by adjusting the opening size of the control valve CV. The crank pressure Pc is determined in accordance with the amount of gas supplied through the supply passage

28

into the crank chamber

5

and the amount of gas released through the bleed passage

27

from the crank chamber

5

. If the crank pressure Pc is altered, the difference between the pressure in each cylinder bore

1

a

and the crank pressure Pc, which act on opposite sides of the associated piston

20

, is also changed. The inclination angle θ of the swash plate

12

is thus altered to vary the piston stroke, or the compressor displacement.

Control Valve Controlling Compressor Displacement and Refrigerant Flow

Generally, as the compressor displacement increases and the refrigerant flow rate in the refrigerant circuit increases, the pressure loss per unit length of the circuit, or refrigerant passage, is increased. More specifically, as the refrigerant flow rate in the refrigerant circuit increases, the pressure loss (pressure difference) between a pair of pressure monitoring points P

1

, P

2

located along the refrigerant circuit increases. Thus, the compressor displacement is detected indirectly by determining the pressure difference ΔP(t) between the points P

1

and P

2

. In this embodiment, as shown in

FIG. 1

, an upstream pressure monitoring point P

1

is located in the discharge chamber

22

, which is a most upstream section of the passage

36

. Further, a downstream pressure monitoring point P

2

is located in the passage

36

at a position spaced from the point P

1

at a predetermined distance. The gas pressure PdH detected at the point P

1

(the discharge pressure Pd) is introduced to the control valve CV via a first passage

37

. The gas pressure PdL detected at the point P

2

is introduced to the control valve CV via a second passage

38

. The control valve CV mechanically detects the pressure difference ΔP(t) between the points P

1

and P

2

(ΔP(t)=PdH−PdL). The control valve CV adjusts its opening size in accordance with the detected pressure difference ΔP(t), thus executing a feedback control procedure for the compressor displacement.

As shown in

FIG. 3

, the control valve CV includes an inlet valve portion located in an upper section of the valve CV and a solenoid portion

60

located in a lower section of the valve CV. The inlet valve portion adjusts the opening size (restriction size) of the supply passage

28

connecting the discharge chamber

22

to the crank chamber

5

. The solenoid portion

60

is an electromagnetic urging mechanism that urges a movable rod

40

located in the control valve CV in accordance with an external, electric control signal. The movable rod

40

includes a distal portion

41

, which receives the pressure difference ΔP(t), a connecting portion

42

, a valve body

43

, which is located substantially in the middle of the rod

40

, and a guide rod section

44

, which forms a proximal portion of the rod

40

. The valve body

43

forms part of the guide rod section

44

. The cross-sectional area of the distal portion

41

is defined as SB, that of the connecting portion

42

is defined as SC, and that of the guide rod section

44

(including the valve body

43

) is defined as SD. In this case, the following equation is satisfied: SC<SB<SD.

A valve housing

45

of the control valve CV includes a lid

45

a

, an upper body section

45

b

, which substantially forms the contour of the inlet valve portion, and a lower body section

45

c

, which forms the contour of the solenoid portion

60

. A valve chamber

46

and a communication passage

47

are formed in the upper body section

45

b

. A pressure sensitive chamber

48

is formed by the upper body section

45

b

and a lid

45

a

that is secured to an upper portion of the section

45

b

, as viewed in FIG.

3

. The movable rod

40

extends through the valve chamber

46

, the communication passage

47

, and the pressure sensitive chamber

48

and moves in an axial direction (the vertical direction as viewed in FIG.

3

). The valve chamber

46

is connected with the communication passage

47

when the rod

40

is located at a certain position. However, the communication passage

47

is blocked from the pressure sensitive chamber

48

by a partition (forming part of the valve housing

45

) located between the passage

47

and the chamber

48

. In other words, a guide hole

49

is formed in the partition for guiding the rod

40

, and the diameter of the guide hole

49

is equal to the diameter of the distal portion

41

of the rod

40

. Further, the communication passage

47

is formed by the guide hole

49

, and the diameter of the communication passage

47

is equal to the diameter of the distal portion

41

. Thus, the cross-sectional area of the rod

40

, that of the communication passage

47

and that of the guide hole

49

are all SB.

As shown in

FIG. 3

, the valve chamber

46

has a bottom formed by an upper side of a fixed iron core

62

, which will be described later. A port

51

extends radially through a wall section of the valve housing encompassing the valve chamber

46

. The port

51

connects the discharge chamber

22

to the valve chamber

46

through an upstream section of the supply passage

28

. In the same manner, a port

52

extends radially through a wall section of the valve housing encompassing the communication passage

47

. The port

52

connects the communication passage

47

to the crank chamber

5

through a downstream section of the supply passage

28

. Thus, in the control valve CV, the port

51

, the valve chamber

46

, the communication passage

47

, and the port

52

form part of the supply passage

28

connecting the discharge chamber

22

to the crank chamber

5

. The valve chamber

46

accommodates the valve body

43

of the movable rod

40

. The diameter of the communication passage

47

is larger than the diameter of the connecting portion

42

of the rod

40

but smaller than the diameter of the guide rod section

44

. A step between the valve chamber

46

and the communication passage

47

thus forms a valve seat

53

, and the communication passage

47

functions as a valve hole. If the movable rod

40

is moved from the position of

FIG. 3

(lowermost position) to an uppermost position at which the valve body

43

is received by the valve seat

53

, the communication passage

47

is closed. In other words, the valve body

43

of the movable rod

40

functions as an inlet valve body that adjusts the opening size of the supply passage

28

to a desired degree.

A movable wall

54

, or a partition, is provided in the pressure sensitive chamber

48

and moves axially in the chamber

48

. The movable wall

54

axially divides the pressure sensitive chamber

48

into a pair of sections, which are a P

1

pressure chamber (first pressure chamber)

55

and a P

2

pressure chamber (second pressure chamber)

56

. The movable wall

54

moves in accordance with the pressure difference between the P

1

pressure chamber

55

and the P

2

pressure chamber

56

. The cross-sectional area of the movable wall

54

is defined as SA and is larger than the cross-sectional area SB of the communication passage

47

or the guide hole

49

(SB<SA). The P

1

pressure chamber

55

is constantly connected to the discharge chamber

22

and the upstream pressure monitoring point P

1

through the first passage

37

. The P

2

pressure chamber

56

is constantly connected to the downstream pressure monitoring point P

2

through the second passage

38

. That is, the discharge pressure Pd is applied to the P

1

pressure chamber

55

and is referred to as the pressure PdH. The pressure PdL acting on the point P

2

is applied to the P

2

pressure chamber

56

. Accordingly, an upper side of the movable wall

54

is exposed to the pressure PdH, and a lower side of the wall

54

is exposed to the pressure PdL, as viewed in FIG.

3

. The distal portion

41

of the movable rod

40

projects into the P

2

pressure chamber

56

. The movable wall

54

is secured to a distal end of the distal portion

41

. A buffer spring

57

is located in the P

2

pressure chamber

56

for urging the movable wall

54

toward the P

1

pressure chamber

55

.

The solenoid portion

60

of the control valve CV includes an accommodating cylinder

61

having a closed end. The fixed iron core

62

is fitted in an upper section of the cylinder

61

to define a solenoid chamber

63

in the cylinder

61

. The solenoid chamber

63

accommodates a movable iron core

64

, which is also referred to as a plunger. The movable core

64

moves axially in the solenoid chamber

63

. A guide hole

65

extends axially in the middle of the fixed core

62

. The guide hole

65

receives the guide rod section

44

of the movable rod

40

, which moves axially in the guide hole

65

. A slight clearance, or a slit

65

a

, is formed between the wall of the guide hole

65

and the guide rod section

44

. A valve chamber

46

is connected to the solenoid chamber

63

through the slit

65

a

. That is, the solenoid chamber

63

is exposed to the discharge pressure Pd, which also acts in the valve chamber

46

. The solenoid chamber

63

receives the proximal portion of the movable rod

40

. A proximal end of the guide rod section

44

extends in the solenoid chamber

63

. This end of the guide rod section

44

is securely fitted in a hole formed in the middle of the movable core

64

through crimping. The movable rod

40

thus moves integrally with the movable core

64

.

A return spring

66

is provided between the fixed core

62

and the movable core

64

. The return spring

66

urges the movable core

64

away from the fixed core

62

, thus pressing the movable core

64

and the movable rod

40

downward, as viewed in FIG.

3

. The force f

2

of the return spring

66

is greater than the force f

1

of the buffer spring

57

. The return spring

66

thus acts to return the movable core

64

and the movable rod

40

to a lowermost position (an initial position when current supply is nullified). A coil

67

is wound around the fixed core

62

and the movable core

64

. The coil

67

is supplied with a drive signal sent from a driver

71

in response to an instruction of a controller

70

. The coil

67

generates electromagnetic force F corresponding to current supply from the driver

71

. The electromagnetic force F draws the movable core

64

toward the fixed core

62

, thus moving the movable rod

40

toward the P

1

pressure chamber

55

. The current supply to the coil

67

may be determined by an analog current control procedure or a duty control procedure, in which a duty ratio Dt of the drive signal is altered as needed. In this embodiment, the duty control procedure is employed. The opening size of the control valve CV increases as the duty ratio Dt of the drive signal decreases. That is, the opening size of the control valve CV decreases as the duty ratio Dt of the drive signal increases.

The opening size of the control valve CV is determined in accordance with the position of the movable rod

40

, which forms the valve body

43

. The operational conditions and characteristics of the control valve CV are made clear by analyzing various forces acting on the movable rod

40

.

As viewed in

FIG. 3

, the upper side of the distal portion

41

of the rod

40

receives a downward force generated in accordance with the pressure difference between the points P

1

, P

2

and diminished by the upward force f

1

of the buffer spring

57

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

54

is SA, and the pressure receiving area of the lower side of the movable wall

54

is SA−SB. A lower side of the distal portion

41

(the pressure receiving area of-which is SB−SC) receives an upward force caused by the crank pressure Pc. Pressures acting on the valve body

43

, the guide rod section

44

, and the movable core

64

will hereafter be analyzed with reference to

FIG. 4

, which schematically shows pressures acting on the movable rod

40

. As shown in

FIG. 4

, an imaginary cylindrical surface extending axially from the wall of the communication passage

47

(as indicated by broken lines) divides the upper side of the valve body

43

into a radially inner section and a radially outer section. The crank pressure Pc acts downward on the inner section (the area of which is SB−SC), and the discharge pressure Pd acts downward on the outer section (the area of which is SD−SB), as viewed in FIG.

4

. Since the pressure acting on the upper side of the movable core

64

is equilibrated with the pressure applied to the lower side of the movable core

64

, the discharge pressure Pd, to which the solenoid chamber

63

is exposed, urges the guide rod section

44

upward at an area corresponding to the cross-sectional area SD of the guide rod section

44

. Further, as shown in

FIG. 3

, the guide rod section

44

of the movable rod

40

(including the valve body

43

) receives the upward electromagnetic force F and the downward force f

2

of the return spring

66

, which acts against the electromagnetic force F.

When the control valve is operated, the movable rod

40

is positioned to satisfy the following condition: the total force acting on the movable rod

40

is zero. If the downward direction is defined as a positive direction, the following equation (1) is obtained based on the above condition:

PdH·SA−PdL

(

SA−SB

)−

f

1

−

Pc

(

SB−SC

)+

Pc

(

SB−SC

)+

Pd

(

SD−SB

)−

Pd·SD−F+f

2

=0 (1)

The following equation (2) is obtained from the equation (1):

(

PdH−PdL

)

SA+PdL·SB−Pd·SB=F+f

1

−

f

2

(2)

More specifically, while deriving equation (2) from equation (1), +Pd·SD is canceled by −Pd·SD such that Pd·SB remains in the equation (2). In other words, regarding the discharge pressure Pd, the effective pressure receiving area of the guide rod section

44

corresponds to the cross sectional area SB of the communication passage

47

, regardless of the cross sectional area SD of the guide rod section

44

. Thus, in this specification and the attached drawings, if the same type of pressure acts on opposite sides of a member such as a rod, the term “effective pressure receiving area” is defined as the pressure receiving area of one side of the member that has an uncanceled effect.

In this embodiment, since the pressure monitoring point P

1

is located in the discharge chamber

22

, the following equation is satisfied: Pd=PdH. If Pd of the equation (2) is substituted by PdH, the following equation (3) is obtained:

PdH−PdL

=(

F+f

1

−

f

2

)/(

SA−SB

) 3)

In the right side of the equation (3), f

1

, f

2

, SA, and SB are definite parameters that are determined when designing the control valve, while the electromagnetic force F is varied in accordance with the current supply to the coil

67

. The equation (3) thus indicates the following two points. Firstly, the control valve CV determines a target value for the pressure difference ΔP(t) between the points p

1

and P

2

(PdH−PdL), or a target pressure difference TPD in relation to which the control valve CV adjusts its opening. The target value can be changed by an external duty control procedure for the coil

67

. In other words, the control valve CV is externally controlled to alter the target pressure difference TPD. The target pressure difference TPD is determined by the solenoid portion

60

, the buffer spring

57

, and the return spring

66

, as indicated by (F+f

1

−f

2

) in the equation (3). Secondly, the condition that the movable rod

40

is positioned to satisfy, or the equation (3), does not include pressure parameters (such as Pc and Pd) other than the pressure difference between the points P

1

and P

2

(PdH−PdL). The movable rod

40

is thus positioned regardless of the absolute value of the crank pressure Pc and that of the discharge pressure Pd. That is, pressure parameters other than the pressure difference between the points P

1

and P

2

(PdH−PdL) do not affect movement of the movable rod

40

. The control valve CV is thus smoothly operated only in accordance with the pressure difference ΔP(t) between the points P

1

and P

2

, the electromagnetic force F, the spring force f

1

, and the spring force f

2

.

The opening size of the control valve CV that has the above operational characteristics is determined as follows. If the current supply to the coil

67

is null (Dt=zero), the force of the return spring

66

is stronger than the force of the buffer spring

57

. The spring

66

thus acts to locate the movable rod

40

at the lowermost position shown in FIG.

3

. In this state, the valve body

43

of the movable rod

40

is spaced from the valve seat

53

by a maximum distance. The inlet valve portion is thus completely open. However, if the current supply to the coil

67

is in accordance with a minimum duty ratio, at least the upward electromagnetic force F becomes stronger than the downward force f

2

of the return spring

66

. An upward force (F−f

2

) is thus generated by the solenoid portion

60

and acts against a downward force generated in accordance with the pressure difference (PdH−PdL), which is diminished by the upward force f

1

of the buffer spring

57

. Accordingly, the valve body

43

of the movable rod

40

is positioned with respect to the valve seat

53

to satisfy the equation (3), thus determining the opening size of the control valve CV. This determines the amount of the refrigerant supplied to the crank chamber

5

through the supply passage

28

. The crank pressure Pc is thus adjusted in accordance with the refrigerant flow in the supply passage

28

and that of the bleed passage

27

, which releases gas from the crank chamber

5

. In other words, if the opening size of the control valve CV is adjusted, the crank pressure Pc is adjusted. Further, as long as the electromagnetic force F remains unchanged, the control valve CV functions as a constant flow valve that determines the target pressure difference TPD in accordance with the current electromagnetic force F. However, if the electromagnetic force F is varied in accordance with the external control procedure to alter the target pressure difference TPD, the control valve CV functions as a variable displacement control valve.

Electronic Control System and Its Procedure

As shown in

FIGS. 1 and 3

, the air conditioning apparatus includes the controller

70

that controls the air conditioning apparatus as a whole. The controller

70

is a computer-like control unit having a central processing unit (CPU), a read-only memory (ROM), a random-access memory (RAM), and an input/output interface (I/O interface). The driver

71

is connected to an output terminal of the I/O interface, and an external information acquiring device

72

is connected to an input terminal of the I/O interface. The controller

70

operates to determine the target duty ratio and to switch the operational mode of the compressor. More specifically, the controller

70

computes a tentative duty ratio DtP (corresponding to a “target duty ratio”) and a final duty ratio Dt in accordance with at least various external information supplied by the external information acquiring device

72

. The controller

70

performs an internal computation based on the tentative duty ratio DtP and outputs the final duty ratio Dt to the driver

71

. That is, the controller

70

instructs the driver

71

to send a drive signal with the final duty ratio Dt to the coil

67

. The electromagnetic force F of the solenoid portion

60

is altered in accordance with the duty ratio Dt of the drive signal supplied to the coil

67

. Also, the target pressure difference TPD, according to which the control valve CV internally adjusts its opening size, is varied in accordance with the duty ratio Dt.

The external information acquiring device

72

includes various sensors such as an A/C switch

73

, a temperature sensor

74

, a temperature adjuster

75

, a vehicle speed sensor

76

, an engine speed sensor

77

, and an accelerator position sensor

78

. The A/C switch

73

is an ON/OFF switch manipulated by a driver or passenger to turn on and off the air conditioning apparatus. The temperature sensor

74

detects the passenger compartment temperature Te(t) (or the temperature of the air exiting from the evaporator, which is varied in relation to the passenger compartment temperature). The temperature adjuster

75

sets a desired temperature Te(set) for the passenger compartment (or the air exiting from the evaporator). The vehicle speed sensor

76

detects the vehicle speed, and the engine speed sensor

77

detects the engine speed. The accelerator position sensor

78

detects the opening size of a throttle valve provided in an engine intake manifold. The opening size of the throttle valve reflects the position of the accelerator, which is depressed by the driver.

The controller

70

executes a duty ratio control procedure for the control valve CV, as will hereafter be described with reference to the flowcharts of

FIGS. 5 and 6

.

The flowchart of

FIG. 5

shows a main routine of an air conditioning control program. When the ignition switch (or START switch) of the vehicle is turned on, the controller

70

is powered to initiate computation. In step S

51

(hereinafter referred to simply as “S

51

”, and other steps are referred to in the same manner), the controller

70

executes various initial settings in accordance with an initial program. For example, the tentative duty ratio DtP and the final duty ratio Dt are each set to a tentative value or an initial value. In the subsequent steps including S

52

, the controller

70

monitors the operational state of the vehicle and internally computes a duty ratio.

In S

52

, the controller

70

monitors the ON/OFF state of the A/C switch

73

. When the A/C switch

73

is turned on, the controller

70

initiates an exceptional state determining routine (S

53

). In S

53

, the controller

70

judges whether the vehicle is operating in an exceptional state, or an exceptional mode, in accordance with the external information. The term “exceptional mode” indicates a state in which the vehicle, for example, is climbing a slope, which applies an increased load to the engine E. The term also indicates a state in which the vehicle is accelerated for, for example, when passing another vehicle (or at least the driver is rapidly accelerating the vehicle). The controller

70

acquires the detected accelerator position from the external information acquiring device

72

and compares the value with a predetermined reference value. In this manner, the controller

70

determines that the vehicle is operating in the increased load state or the accelerated state (the exceptional state).

If the judgement of S

53

is positive, or if the vehicle is operated in the exceptional state, the controller

70

performs an exceptional state control procedure (S

54

). More specifically, the controller

70

maintains the final duty ratio Dt at zero or a minimum duty ratio Dt(min) during a predetermined time period Δt after detecting the exceptinal state. During the time period Δt, in which the final duty ratio Dt is minimized, the control valve CV is fully opened (maximum opening size), regardless of the pressure difference (PdH−PdL) between the points P

1

and P

2

. The crank pressure Pc is thus rapidly increased, and the inclination angle θ is quickly minimized to minimize the compressor displacement. This reduces the load acting on the engine E, and makes additional engine power available for driving the vehicle. Although the cooling performance of the air conditioning apparatus is temporarily lowered during the time period Δt, which is relatively short, passenger' comfort is not significantly sacrificed in most cases.

If any judgement conditions for the exceptional state determining routine are not satisfied, the judgement of S

53

becomes negative. In this case, it is determined that the vehicle is operating in a normal state, or a normal operational mode. The term “normal operational mode” indicates a state in which any judgement conditions for the non-normal state determining routine are not satisfied and it is assumed that the vehicle is operated in a normal state. When the judgement of S

53

is negative, the controller

70

initiates a normal state control routine RF

6

. In many cases, the controller

70

first performs the normal state control routine RF

6

and then resumes S

52

of the main routine of FIG.

5

.

As shown in

FIG. 6

, if the vehicle is operated in the normal operational mode, the controller

70

executes a feedback control procedure for the air conditioning performance, or the compressor displacement, in accordance with the normal state control routine RF

6

. The control valve CV, which includes the movable wall

54

that is exposed to the pressure difference ΔP(t), adjusts its opening size mechanically or internally in accordance with variation in the pressure difference ΔP(t) (PdH−PdL). Thus, while executing the routine RF

6

, the controller

70

corrects the target pressure difference TPD of the control valve CV in relation to the thermal load currently acting on the evaporator

33

. In other words, the controller

70

regressively corrects the tentative duty ratio DtP for the internal computation and determines the final duty ratio Dt, which is sent to the driver

71

, in accordance with the corrected tentative duty ratio DtP.

More specifically, in S

61

, the controller

70

judges whether the temperature Te(t) detected by the temperature sensor

74

exceeds the target temperature Te(set) set by the temperature adjuster

75

. If the judgement of S

61

is negative, the controller

70

judges whether the detected temperature Te(t) is lower than the target temperature Te(set) in S

62

. If the judgement of S

62

is also negative, it is indicated that the detected temperature Te(t) is equal to the target temperature Te(set). In this case, the cooling performance of the compressor need not be corrected, and the tentative duty ratio DtP remains unchanged.

If the judgement of S

61

is positive, it is assumed that the passenger compartment temperature is relatively high and the cooling load acting on the compressor has increased. Thus, the controller

70

increases the tentative duty ratio DtP by a unit amount ΔD in S

63

. When the duty ratio of the drive signal is altered to the increased value (DtP+ΔD), the electromagnetic force F generated by the solenoid portion

60

is increased accordingly, thus increasing the target pressure difference TPD of the control valve CV. In this state, the force resulting from current pressure difference ΔP(t) does not equilibrate the upward urging force and the downward urging force acting on the movable rod

40

. The movable rod

40

is thus moved toward the P

1

pressure chamber

55

such that the downward force f

2

of the return spring

66

matches the increased upward electromagnetic force F. Accordingly, the valve body

43

of the movable rod

40

is repositioned to satisfy the equation (3). This reduces the opening size of the control valve CV (the supply passage

28

) accordingly, thus lowering the crank pressure Pc. As a result, the difference between the crank pressure Pc and the pressure in the cylinder bore

1

a

, which act on opposite sides of the piston

20

, decreases to increase the inclination angle of the swash plate

12

. This increases the compressor displacement, thus increasing the load acting on the engine. With the displacement increased, the cooling performance of the evaporator

33

is improved, which lowers the passenger compartment temperature Te(t). In this state, the pressure difference ΔP(t) between the pressure monitoring points P

1

and P

2

is increased. The opening size of the control valve CV is then reversely mechanically increased in a feedback manner.

If the judgement of S

61

is negative and the judgement of S

62

is positive, it is assumed that the passenger compartment temperature is relatively low and the cooling load acting on the compressor is decreased. Thus, the sag controller

70

reduces the tentative duty ratio DtP by a unit amount ΔD in S

64

. When the duty ratio of the drive signal is altered to the decreased value (DtP−ΔD), the electromagnetic force F generated by the solenoid portion

60

is reduced accordingly, thus decreasing the target pressure difference TPD of the control valve CV. In this state, the force resulting from the current pressure difference ΔP(t) does not equilibrate the upward urging force and the downward urging force acting on the movable rod

40

. The movable rod

40

is thus moved away from the P

1

pressure chamber

55

such that the downward force f

2

of the return spring

66

matches the decreased upward electromagnetic force F. Accordingly, the valve body

43

of the movable rod

40

is repositioned to satisfy the equation (3). This increases the opening size of the control valve CV (the supply passage

28

) accordingly, thus raising the crank pressure Pc. As a result, the difference between the crank pressure Pc and the pressure in the cylinder bore

1

a

, which act on opposite sides of the piston

20

, increases to decrease the inclination angle of the swash plate

12

. This reduces the compressor displacement, thus decreasing the load acting on the engine. When the displacement is decreased, the cooling performance of the evaporator

33

is decreased, which increases the passenger compartment temperature Te(t). In this state, the pressure difference ΔP(t) between the pressure monitoring points P

1

and P

2

is decreased. The opening size of the control valve CV is then reversely mechanically reduced in a feedback manner.

As described, if the detected temperature Te(t) is not equal to the target temperature Te(set), the controller

70

corrects the tentative duty ratio DtP in S

63

and/or S

64

. This gradually optimizes the target pressure difference TPD of the control valve CV. The control valve CV thus internally adjusts its opening size in a feedback manner in accordance with the target pressure difference TPD. In this manner, the detected temperature Te(t) approaches the target temperature Te(set).

Further, in this embodiment, the controller

70

performs a procedure for restricting an upper limit of the tentative duty ratio DtP, after terminating S

62

, S

63

, or S

64

. This prevents the tentative duty ratio DtP from exceeding the maximum value Dt(max) of an acceptable variation range for the final duty ratio Dt. More specifically, the controller

70

judges whether the tentative duty ratio DtP is larger than the maximum duty ratio Dt(max) in S

65

. If the judgement of S

65

is positive, the controller

70

reduces the tentative duty ratio DtP to the maximum duty ratio Dt(max) in S

66

. Accordingly, once the controller

70

terminates S

65

or S

66

, the tentative duty ratio DtP is always equal to or smaller than the maximum duty ratio Dt(max).

Subsequently, the controller

70

judges whether the tentative duty ratio DtP is equal to or larger than a predetermined reference value DJ in S

67

. If the judgement of S

67

is positive, the coefficient of performance COP obtained with the displacement corresponding to this tentative duty ratio DtP is satisfactory. That is, the reference value DJ indirectly indicates a displacement corresponding to a minimum value of a desired coefficient of performance, which is a threshold value of displacement (how to set the value DJ will be described later). Thus, if the judgement of S

67

is positive, the tentative duty ratio DtP is selected as the final duty ratio Dt (see S

68

). In this case, in the subsequent step S

610

, the controller

70

instructs the driver

71

to send a drive signal representing the final duty ratio Dt to the coil

67

. If the judgement of S

67

is negative, or the tentative duty ratio DtP is smaller than the reference value DJ, the final duty ratio Dt is nullified (see S

69

). In the subsequent step S

610

, the controller

70

instructs the driver

71

to send a drive signal having the nullified final duty ratio Dt (Dt=zero) to the coil

67

. In other words, if the tentative duty ratio DtP for the internal computation is smaller than the reference value DJ, the current supply to the coil

67

is substantially nullified.

In accordance with the flowchart shown in

FIG. 6

, particularly S

67

to S

610

, the compressor displacement is varied continuously as long as a relatively high coefficient of performance COP is ensured. However, if the COP is likely to be relatively low, the compressor displacement is minimized, regardless of the tentative duty ratio for the internal computation. More specifically, the compressor operation is switched between a variable displacement operation and a minimum displacement operation based on the comparison between the tentative duty ratio DtP and the reference value DJ. Selection of the reference value DJ will hereafter be described by way of example.

FIG. 8

is a graph like to the graph of

FIG. 7

, but

FIG. 8

includes only one curve representing the operational characteristics of the compressor.

FIG. 9

is a graph showing the relationship between the actual duty ratio (the final duty ratio Dt) of the drive signal, which is sent to the coil

67

, and the compressor displacement Vc. Since the power L required by the compressor increases as the displacement Vc increases, the graph of

FIG. 9

also shows the relationship between the final duty ratio Dt and the power L indirectly. As shown in

FIG. 9

, although not linearly, the final duty ratio Dt is increased as the displacement Vc, or the power L, is increased. Considering this relationship between the duty ratio Dt and the power L, the vertical axis (y-axis) of

FIG. 8

is changed from the power ratio to the duty ratio, thus obtaining the graph of FIG.

10

. More specifically, in

FIG. 10

, the refrigerating performance ratio (Q/Q

0

) is plotted along the horizontal axis and the final duty ratio Dt is plotted along the axis. The graph includes a curve having a single-dotted broken section and a solid section. The broken section is connected to the solid section by a point of inflection P′. As shown in

FIG. 10

, the refrigerating performance ratio corresponding to the point P′ is defined as B. As shown in

FIG. 8

, the point of divergence P corresponds to the refrigerating performance ratio defined as B.

In this embodiment, the final duty ratio (DJ) corresponding to the point of inflection P′ of

FIG. 10

is selected as the reference value DJ, which is used for the judgement of S

67

. More specifically, if the final duty ratio Dt is equal to the value DJ, the corresponding refrigerating performance ratio is B. As shown in

FIG. 8

, the COP corresponding to the refrigerating performance ratio B is the value indicated by the point P. As in the graph of

FIG. 7

, in an area below the point P of

FIG. 8

, or an area in which the displacement is lower than a value corresponding to the point P (indicated by the dotted area in FIG.

8

), the COP is relatively low. Accordingly, in order to ensure a sufficient COP, it is preferred that the compressor displacement Vc is controlled to avoid an intermediate displacement between the minimum displacement corresponding to the nullified duty ratio (Dt=0) and the displacement corresponding to the point P. Instead, the compressor displacement Vc is minimized regardless of the tentative duty ratio DtP, if the value DtP is smaller than the reference value DJ. In other words, the reference value DJ is used to judge whether the tentative duty ratio DtP for the internal computation leads to a relatively low COP.

As indicated by the graph of

FIG. 10

, the refrigerating performance ratio is substantially nullified when the compressor is operated with the minimum displacement corresponding to the nullified duty ratio (Dt=0). However, as indicated by the graph of

FIG. 8

, the power ratio is not nullified even when the refrigerating performance ratio is nullified. In other words, as shown in

FIG. 8

, the point of the curve corresponding to the nullified duty ratio (Dt=0), at which the compressor is operated at the minimum displacement, is located slightly above from the diagonal straight line, thus indicating that the COP is relatively low. However, it is also indicated that this point of the curve is located relatively close to the diagonal line, although included in the dotted area, as compared to the point C, which is relatively spaced from the line. That is, the COP corresponding to the minimum displacement is still relatively close to the value Q

0

/L

0

as compared to the COP corresponding to the point C, or higher than the COP corresponding to the point C. Accordingly, in order to ensure a relatively high COP, or a relatively high efficiency, it is advantageous to minimize the displacement if the operational state corresponds to the area below the point P of FIG.

8

.

FIG. 11

is a timing chart (in which curves are simplified for convenience of understanding) showing the variation of the final duty ratio Dt and the detected temperature Te(t) during the normal control routine of

FIG. 6

performed when the target temperature Te (set) is maintained at a constant level. As shown in

FIG. 11

, a period in which the final duty ratio Dt is zero alternates with a period in which the final duty ratio Dt is equal to or greater than the reference value DJ. During the period in which the final duty ratio Dt is zero, the displacement Vc of the compressor is no longer variably controlled but is minimized. In contrast, during the period in which the final duty ratio Dt is equal to or greater than the reference value DJ, the displacement Vc of the compressor is variably controlled. While the displacement Vc is controlled in accordance with these alternate periods, the detected temperature Te(t) increases when the displacement Vc is maintained at minimum. However, if the variable control of the displacement Vc is resumed, the detected temperature Te(t) starts to decrease with a relatively short delay. However, the detected temperature Te(t) starts to increase again, toward the target temperature Te(set), without decreasing excessively. In this manner, the passenger compartment temperature is steered foward the target temperature Te(set) though is has slight fluctuation and varies in a relatively small range around the target value Te(set).

This embodiment has the following effects.

The tentative duty ratio DtP for the internal computation of the controller

70

, in which regressive computations are repeated, is considered to be a parameter that indirectly indicates the compressor displacement Vc, or the refrigerant flow in the refrigerant circuit. Thus, if the current tentative duty ratio DtP is compared with the reference value DJ, it is judged whether the coefficient of performance (COP) in a corresponding operational state (displacement Vc) is relatively high or low. Based on this judgement, the compressor operation is switched between the minimum displacement operation and the variable displacement operation. That is, the variable control of the displacement is avoided when the COP is likely to decrease below a minimum acceptable level (in this embodiment, Q

0

/L

0

) This improves the operation efficiency of the compressor and that of the air conditioning apparatus.

In this embodiment, the compressor displacement is controlled in a feedback manner by directly controlling the pressure difference ΔP(t) between the points P

1

and P

2

(PdH−PdL). Accordingly, regardless of the thermal load acting on the evaporator

33

, the displacement is decreased quickly and reliably in response to an external control procedure, as needed when the engine is in the exceptional state.

When the vehicle is operated in the normal operational mode, the tentative duty ratio DtP for determining the target pressure difference TPD is automatically adjusted in relation to the detected temperature Te(t) and the target temperature Te(set). Further, the control valve internally adjusts its opening size in accordance with the pressure difference ΔP(t) between the points P

1

and P

2

. This controls the compressor displacement. In other words, the air conditioning apparatus adjusts the compressor displacement to reduce the difference between the detected temperature Te(t) and the target temperature Te(set), to make the passenger compartment comfortable.

The present invention may be modified as follows.

In the illustrated embodiment, the reference value DJ, on which the compressor operation of switching between the minimum displacement operation and the variable displacement operation, is based, is a predetermined value (a fixed value). However, the reference value DJ may be varied during the control procedure. For example, the reference value DJ may be corrected in accordance with external information including the engine speed, the flow rate of air through the evaporator, the atmospheric temperature, and the insolation amount. The judgement of S

67

is performed in accordance with the corrected reference value DJ.

In the illustrated embodiment, the reference value DJ is selected as the final duty ratio Dt for achieving the maximum performance COP(COP=Q

0

/L

0

), as indicated by the point P of

FIG. 8

, which corresponds to the point P′ of FIG.

10

. However, the reference value DJ may be any value corresponding to a final duty ratio Dt that achieves an intermediate compressor displacement Vc that divides a displacement variation range into a large displacement area and a small displacement area. In other words, the reference value DJ may be any value, as long as the COP corresponding to the value DJ is considered to be a minimum acceptable COP. The variable controlling of the displacement is suspended when necessary to avoid a COP lower than the minimum acceptable COP, thus satisfying the objective of the present invention.

In the illustrated embodiment, if the tentative duty ratio DtP is equal to or greater than the reference value DJ, the compressor displacement is varied continuously by altering the target pressure difference TPD of the control valve CV. However, even if the tentative duty ratio DtP is equal to or greater than the reference value DJ, the compressor may be operated by a predetermined fixed displacement corresponding to a predetermined COP, for example, a fixed displacement corresponding to the COP indicated by the point D of FIG.

8

. That is, the duty ratio is fixed to a value corresponding to the point D′ of

FIG. 10

, which corresponds to the point D. In this case, the final duty ratio Dt of the drive signal is switched between two values, which are zero and the value corresponding to the point D′. This still suppresses variable displacement operation in a relatively small displacement area, when COP is relatively low.

In the illustrated embodiment, the two pressure monitoring points P

1

and P

2

are located along the passage

36

connecting the discharge chamber

22

of the compressor to the condenser

31

. Instead, the points P

1

and P

2

may be located along the passage

35

connecting the evaporator

33

to the suction chamber

21

of the compressor. Alternatively, the upstream point P

1

may be located in the discharge chamber

22

or the passage

36

, and the downstream point P

2

may be located in the suction chamber

21

or the passage

35

. Further, the point P

1

may be located in the discharge chamber

22

or the passage

36

, and the point P

2

may be located in the crank chamber

5

. In addition, the point P

1

may be located in the crank chamber

5

, and the point P

2

may be located in the suction chamber

21

or the passage

35

. In any case, the pressure difference ΔP(t) between the points P

1

and P

2

reflects the amount of the refrigerant flowing in the refrigerant circuit, or the compressor displacement.

Although the illustrated embodiment is applied to a so-called clutchless compressor, the present invention may be applied to a variable displacement compressor to which power is transmitted from an engine E through a power transmitting mechanism PT having a clutch such as an electromagnetic clutch. In this case, it is preferred that the controller

70

minimizes the compressor displacement, regardless of the tentative duty ratio DtP, and disconnects the clutch if the tentative duty ratio DtP is smaller than the reference value DJ. Alternatively, it is preferred that the controller

70

disconnects the clutch immediately if the tentative duty ratio DtP is smaller than the reference value DJ, instead of minimizing the compressor displacement. That is, if it is assumed that the COP of the compressor is likely to drop, the power supply to the compressor is stopped by disconnecting the clutch.

The present invention may be applied to a prior-art variable displacement compressor that varies its displacement in accordance with suction pressure.

In this specification, the term “refrigerant circuit” indicates, as shown in

FIG. 1

, the circuit including the condenser

31

, the expansion valve

32

, the evaporator

33

, and the compressor (including the suction chamber

21

, the cylinder bores

1

a

, and the discharge chamber

22

). In this regard, the cylinder bore

1

a

, which performs suction, compression, and discharge of refrigerant gas, forms part of the refrigerant circuit.

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. 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
4864832	Suzuki	Sep 1989	A
5117643	Sakurai et al.	Jun 1992	A
5145326	Kimura et al.	Sep 1992	A
5702235	Hirota et al.	Dec 1997	A
6010312	Suitou et al.	Jan 2000	A
6200105	Kawaguchi et al.	Mar 2001	B1
6260369	Yokomachi et al.	Jul 2001	B1
6356825	Takenaka et al.	Mar 2002	B1

Variable displacement compressor and air conditioning apparatus

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