Engine

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to an engine comprising a connecting rod connected at one end to a piston through a piston pin, a first arm turnably connected at one end to the other end of the connecting rod and at the other end to a crankshaft through a crankpin, a second arm integrally connected at one end to the other end of the first arm, a control rod turnably connected at one end to the other end of the second arm, and a movable eccentric shaft mounted between eccentric positions of rotary shafts to which a power reduced at a reduction ratio of 1/2 is transmitted from the crankshaft, the movable eccentric shaft being connected to the other end of the control rod, the stroke of the piston at an expansion stroke being larger than that at a compression stroke.

2. Description of the Related Art

Such engines are conventionally known, for example, from U.S. Pat. No. 4,517,931 and Japanese Patent Application Laid-open No. 9-228853. In each of these engines, the stroke of the piston at an expansion stroke is larger than that at a compression stroke, whereby a larger expansion work is carried out in the same amount of air-fuel mixture drawn, so that the cycle thermal efficiency is enhanced.

In the conventionally known engine, it is common that the position of a top dead center at each of the intake and exhaust strokes and the position of the top dead center at the compression stroke are different from each other. However, if the position of the top dead center at each of the intake and exhaust strokes is higher in level than the position of the top dead center at the compression stroke, there is a possibility that the interference of each of intake and exhaust valves and a top of the piston with each other occurs. If the position of the top dead center at each of the intake and exhaust strokes is lower in level than the position of the top dead center at the compression stroke to avoid the interference, the top dead center at the compression stroke is further lower and hence, an enhancement in a compression ratio in the engine is not desired and it is difficult to operate the engine at a higher thermal efficiency. On the other hand, if the top dead center at the compression stoke is higher in level than the top dead center at each of the intake and exhaust strokes, there is a possibility that the scavenge provided by the piston is insufficient due to the lower level of the piston at the top dead center at each of the intake and exhaust strokes, and thus, a large amount of burned gas remains within a cylinder, thereby bringing about a reduction in output in a full-load state and the instability of burning in a lower-load state.

SUMMARY OF THE INVENTION

Accordingly, it is an object of the present invention to provide an engine, wherein the stroke of the piston at the expansion stroke is larger than that at the compression stroke and in addition, the top dead center at each of the intake and exhaust stroke and the top dead center at the compression stroke are at the same level, whereby the above-described problems are solved.

To achieve the above object, according to a first feature of the present invention, there is provided an engine comprising a connecting rod connected at one end to a piston through a piston pin, a first arm turnably connected at one end to the other end of said connecting rod and at the other end to a crankshaft through a crankpin, a second arm integrally connected at one end to the other end of said first arm, a control rod turnably connected at one end to the other end of said second arm, and a movable eccentric shaft mounted between eccentric positions of rotary shafts to which a power reduced at a reduction ratio 1/2 is transmitted from said crankshaft, said movable eccentric shaft being connected to the other end of said control rod, the stroke of said piston at an expansion stroke being larger than that at a compression stroke, wherein when various dimensions are represented as described below in an x-y plane constituted by an x-axis extending through an axis of said crankshaft along a cylinder axis and a y-axis extending through the axis of said crankshaft in a direction perpendicular to the x-axis: a length of said connecting rod is represented by L

4

; a length of said first arm is represented by L

2

; a length of said second arm is represented by L

1

; a length of said control rod is represented by L

3

; a length from the axis of said crankshaft to axes of said rotary shafts in a direction of the y-axis is represented by L

5

; a length from the axis of said crankshaft to the axes of said rotary shafts in a direction of the x-axis is represented by L

6

; an angle formed by said connecting rod with respective to the cylinder axis is represented by φ

4

; an angle formed by said first and second arm is represented by a; an angle formed by said second arm with the y-axis within the x-y plane is represented by φ

1

; an angle formed by said control rod with the y-axis is represented by φ

3

; an angle formed by a straight line connecting the axis of said crankshaft and said crankpin with the x-axis is represented by θ; an angle formed by a straight line connecting the axes of said rotary shafts and the axis of said movable eccentric shaft with the x-axis is represented by θp; a value of the angle θp is represented by γ when the angle θ is “0”; a length between the axis of said crankshaft and said crankpin is represented by R; a length of the straight line connecting the axes of said rotary shafts and the axis of said movable eccentric shaft is represented by Rp; a rotational angular speed of said crankshaft is represented by ω; and a ratio of the rotational speed of said movable eccentric shaft to the rotational speed of said crankshaft is represented by η and the rotational direction thereof is represented by η=+0.5 or η=−0.5, the following equation is established:

L

4

·sin φ

4

·

d

φ

4

/

dt+L

2

·cos (α+φ

1

)·

d

φ

1

/

dt−R

·ω·sin θ=0

Wherein

φ

4

=arcsin {L

2

·cos (α+φ

1

)+R·sin θ−δ}/L

4

dφ

4

/dt=ω·[−L

2

·sin (α+φ

1

)·{R·cos (θ−φ

3

)−η·Rp·cos (θp−φ

3

)}/{L

1

·sin (φ

1

+φ

3

)}+R·cos θ)]/(L

4

·cos φ

4

)

φ

1

=arcsin [(L

3

2

−L

1

2

−C

2

−D

2

)/{2·L

1

·(C

2

+D

2

)}]−arctan(C/D)

φ

3

=arcsin {(R·cos θ−L

6

−Rp·cos θp+L

1

·sin φ

1

)/L

3

}

C=L

5

+Rp·sin θp−R·sin θ

D=L

6

+Rp·cos θp−R·cos θ

θp=η·θ+γ

dφ

1

/dt=ω·{R·cos (θ−φ

3

)−ηRp·cos (θp−φ

3

)}/{L

1

·sin (φ

1

+φ

3

)}

and crank angles θ at a top dead center at each of the intake and exhaust strokes and at the top dead center at the compression stroke are determined from said equation, and the length L

1

of said second arm; the length L

2

of said first arm; the length L

3

of said control rod; the length L

4

of said connecting rod; the length L

5

from the axis of said crankshaft to the axes of said rotary shafts in the direction of the y-axis; the length L

6

from the axis of said crankshaft to the axes of said rotary shafts in the direction of the x-axis; the amount δ of offsetting of the cylinder axis from the axis of said crankshaft in the direction of the y-axis; the angle α formed by said first and second arms; the length R between the axis of said crankshaft and said crankpin; the length Rp of the straight line connecting the axes of said rotary shafts and the axis of said movable eccentric shaft and the angle θp when the angle θ is “0”, are determined so that the top dead center at each of the intake and exhaust strokes and the top dead center at the compression stroke are congruous with each other, according to the following equation:

X=L

4

·cos φ

4

+

L

2

·sin (α+φ

1

)+

R

·cos θ

which represents a level X of the piston pin at both said crank angles θ.

The operation according to the configuration of the first feature will be described below with reference to

FIG. 5

diagrammatically showing the arrangements of the piston pin, the connecting rod, the crankshaft, the crankpin, the first arm, the second arm, the control rod and the movable eccentric shaft. When the coordinates (Xpiv and Ypiv) of the movable eccentric shaft are determined, a moving speed (dX/dt) of the piston pin is determined by differentiating the position of the piston pin in the direction of the x-axis determined by {X=L

4

·cos φ

4

+L

2

·sin (α+φ

1

)+R·cos θ}, and an equation provided when dX/d=0 has four solutions in a range of −2π<θ<2π. The four solutions are associated with the motion of a 4-cycle engine, whereby crank angles providing a top dead center at a compression stroke, a top dead center at each of intake and exhaust strokes, a bottom dead center after an expansion stroke and a bottom dead center after the intake stroke are determined and used to determine various positions of the piston pin in the directions of the x-axis and the y-axis. When the position of the piston pin at the top dead center in the direction of the x-axis at compression stroke is represented by Xctdc; the position of the piston pin in the direction of the x-axis at the top dead center at each of the intake and exhaust strokes is represented by Xotdc; the position of the piston pin ion the direction of the x-axis at the bottom dead center after an expansion stroke is represented by Xebdc; and the position of the piston pin in the direction of the x-axis at the bottom dead center after the intake stroke is represented by Xibdc, a stroke Scomp at the compression stroke and a stroke Sexp at the compression stroke are represented by (Scomp=Xctdc−Xibdc) and (Sexp=Xotdc−Xebdc), respectively, and the length L

1

of the second arm, the length L

2

of the first arm, the length L

3

of the control rod, the length L

4

of the connecting rod, the length L

5

from the axis of the crankshaft to the axes of the rotary shafts in the direction of the y-axis; the length L

6

from the axis of the crankshaft to the axes of the rotary shafts in the direction of the x-axis; the amount δ of offsetting of the cylinder axis from the axis of the crankshaft in the direction of the y-axis; the angle α formed by the first and second arms; the length R between the axis of the crankshaft and the crankpin; the length Rp of the straight line connecting the axes of the rotary shafts and the axis of the movable eccentric shaft and the angle θp when the angle θ is “0”, are determined so that Scomp<Sexp is satisfied and Xctdc=Xotdc is satisfied. Thus, the stroke of the piston at the expansion stroke can be set larger than that at the compression stroke and in addition, the top dead center at each of the intake and exhaust strokes and the top dead center at the compression stroke can be set at the same level. As a result, it is possible to prevent the occurrence of the interference of each of an intake valve and an exhaust valve and a top of the piston with each other; to provide an enhancement in compression ratio in the engine to enable the operation at a higher thermal efficiency, and to achieve the sufficient scavenge by the piston and to prevent a reduction in output in a full-load state and the instability of burning in a lower-load state.

According to a second feature of the present invention, in addition to the first feature, a locus of movement of the piston pin is determined to be fallen into a range between the x-axis and one of tangent lines parallel to the x-axis and tangent to a locus described at the expansion stroke by a point of connection between the connecting rod and the first arm, which is closest to the x-axis. With such feature, it is possible to reduce the friction of the piston and suppress a piston slap sound. More specifically, when the piston is at the expansion stroke, a large load is applied to the piston, but if the change in attitude of the piston is increased due to the large load at that time, the friction is increased and the piston slap sound is magnified. However, the above-described determination of the locus of movement of the piston pin ensures that the connecting rod is always inclined to one side at the expansion stroke, notwithstanding that the piston receives the large load at the expansion stroke, whereby the change in attitude of the piston can be suppressed to reduce the friction of the piston and to suppress the generation of the piston slap sound.

According to a third feature of the present invention, in addition to the second feature, the range of the crank angle at the expansion stroke is set larger than that at the intake stroke, and the range of the crank angle at the exhaust stroke is set larger than that at the compression stroke. With such configuration, it is possible to avoid the degradation of inertia vibration due to an increase in acceleration of the piston. More specifically, during lowering of the piston, the stroke at the expansion stroke is larger than that at the intake stroke, and during lifting of the piston, the stroke at the exhaust stroke is larger than that at the compression stroke. In the setting in which the top and bottom dead centers are alternated with each other at the crank angle of 180 degrees, the speed of the piston at each of the expansion and exhaust strokes at which the stroke is larger is higher than that at each of the intake and compression strokes at which the stroke is smaller, and the acceleration of the piston is increased due to such a large difference between the speeds, thereby bringing about the degradation of inertia vibration. However, by setting the range of the crank angle at each of the expansion and exhaust strokes at which the stroke is larger at a value larger than the range of the crank angle at each of the intake and compression strokes at which the stroke is smaller, as described above, the speed of the piston at each of the stokes can be further uniform to suppress the variation in acceleration of the piston at the bottom dead center after the intake and expansion strokes and the variation in acceleration of the piston at the top dead center after the intake and expansion strokes to avoid the degradation of inertia vibration.

According to a fourth feature of the present invention, in addition to the third feature, the ranges of the crank angles at the expansion and exhaust strokes are set at values exceeding 180 degrees, respectively. With such configuration, the speed of the piston at each of the intake, compression, expansion and exhaust strokes can be further uniform to more effectively suppress the variation in acceleration of the piston at the bottom dead center after the intake and expansion strokes and the variation in acceleration of the piston at the top dead center after the intake and expansion strokes, thereby more effectively avoiding the degradation of inertia vibration.

According to a fifth feature of the present invention, in addition to any of the first to fourth features, the movable eccentric shaft is mounted on the rotary shafts having the axes disposed at locations spaced within the x-y plane apart from the axis of the crankshaft by the lengths L

5

and L

6

in the directions of the y-axis and the x-axis, respectively, so that it is displaced from the axes of the rotary shafts by a distance corresponding to a radius Rp, and wherein when the length R between the axis of the crankshaft and the crankpin is set at 1.0, the length L

1

of the second arm is set in a range of 1.7 to 4.5; the length L

2

of the first arm is set in a range of 0.6 to 5.2; the length L

3

of the control rod is set in a range of 4.3 to 6.9; the length L

5

between the axis of the crankshaft and the rotary shafts in the direction of the y-axis is set in a ranger of 2.3 to 4.0; the length L

6

between the axis of the crankshaft and the rotary shafts in the direction of the x-axis is set in a range of 0.00 to 3.35; and the radius Rp is set in a range of 0.25 to 1.80, as well as the angle a formed by the first and second arms is set in a range of 105 to 180 degrees. With such configuration, it is possible to provide the configuration of the fourth feature, thereby more effectively avoiding the degradation of inertia vibration.

The above and other objects, features and advantages of the invention will become apparent from the following description of the preferred embodiments taken in conjunction with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIGS. 1

to

7

show a first embodiment of the present invention.

FIG. 1

is a partially cutaway front view of an engine;

FIG. 2

is a vertical sectional view of the engine, which corresponds to a sectional view taken along a line

2

—

2

in

FIG. 3

;

FIG. 3

is a sectional view taken along a line

3

—

3

in

FIG. 2

;

FIG. 4

is a sectional view taken along a line

4

—

4

in

FIG. 3

;

FIG. 5

is an illustration diagrammatically showing the disposition of a link mechanism;

FIG. 6

is a diagram showing operative states of the link mechanism sequentially;

FIG. 7

is a diagram showing a variation in position of a piston pin corresponding to a crank angle;

FIG. 8

is a sectional view of essential portions of an engine according to a second embodiment;

FIG. 9

is an illustration showing the state of a link mechanism at expansion stroke in a third embodiment;

FIG. 10

is an illustration showing the state of the link mechanism at the expansion stroke when a range of crank angle at each of the intake and compression strokes is set larger than that each of the expansion and exhaust strokes;

FIG. 11

is a graph showing the position of a piston provided at each of the strokes by the link mechanism shown in

FIG. 10

;

FIG. 12

is a graph showing a variation in acceleration of the piston provided at each of the strokes by the link mechanism shown in

FIG. 10

;

FIG. 13

is an illustration showing a state of a link mechanism at expansion stroke in a fourth embodiment;

FIG. 14

is a graph showing the position of a piston provided at each of the strokes by the link mechanism shown in

FIG. 13

;

FIG. 15

is a graph showing a variation in acceleration of the piston provided at each of the strokes by the link mechanism shown in

FIG. 13

;

FIG. 16

is an illustration showing a state of a link mechanism at expansion stroke in a fifth embodiment;

FIG. 17

is a graph showing the position of a piston provided at each of the strokes by the link mechanism shown in

FIG. 16

;

FIG. 18

is a graph showing a variation in acceleration of the piston provided at each of the strokes by the link mechanism shown in

FIG. 16

;

FIG. 19

is an illustration showing a state of a link mechanism at expansion and exhaust strokes in a sixth embodiment;

FIG. 20

is a graph showing the position of a piston provided at each of the strokes by the link mechanism shown in

FIG. 19

;

FIG. 21

is a graph showing a variation in acceleration of the piston provided at each of the strokes by the link mechanism shown in

FIG. 19

; and

FIG. 22

is an illustration diagrammatically showing the disposition of a link mechanism for explaining dimensions of various portions.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

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

FIGS. 1

to

7

. Referring first to

FIGS. 1

to

3

, an engine according to the first embodiment is an air-cooled single-cylinder engine used, for example, in a working machine or the like, and includes an engine body

21

which is comprised of a crankcase

22

, a cylinder block

23

protruding in a slightly upward inclined state from one side of the crankcase

22

, and a cylinder head

24

coupled to a head portion of the cylinder block

23

. Large numbers of air-cooling fins

23

a

and

24

a

are provided on outer surfaces of the cylinder block

23

and the cylinder head

24

. A mounting face

22

a

on a lower surface of the crankcase

22

is mounted on an engine bed of each of various working machines

The crankcase

22

comprises a case body

25

formed integrally with the cylinder block

23

by a casting process, and a side cover

26

coupled to an open end of the case body

25

, and a crankshaft

27

are rotatably carried at its opposite ends on the case body

25

and the side cover

26

with ball bearings

28

and

29

and oil seals

30

and

31

interposed therebetween. One end of the crankshaft

27

protrudes as an output shaft portion

27

a

from the side cover

26

, and the other end of the crankshaft

27

protrudes as an auxiliary-mounting shaft portion

27

b

from the case body

25

. Moreover, a flywheel

32

is fixed to the auxiliary-mounting shaft portion

27

b

; a cooling fan

35

for supplying cooling air to various portions of the engine body

21

and a carburetor

34

is secured to an outer surface of the flywheel

32

by a screw member

36

, and a recoil-type engine stator

37

is disposed outside the cooling fan

36

.

A cylinder bore

39

is defined in the cylinder block

23

, and a piston

38

is slidably received in the cylinder bore

39

. A combustion chamber

40

is defined between the cylinder block

23

and the cylinder head

24

, so that a top of the piston is exposed to the combustion chamber

40

.

An intake port

41

and an exhaust port

42

are defined in the cylinder head

24

, and lead to the combustion chamber

40

, and an intake valve

43

for connecting and disconnecting the intake port

41

and the combustion chamber

40

to and from each other and an exhaust valve

44

for connecting and disconnecting the exhaust port

42

and the combustion chamber

40

to and from each other, are openably and closably disposed in the cylinder head

24

. A spark plug

45

is threadedly fitted into the cylinder head

24

with its electrodes facing to the combustion chamber

40

.

The carburetor

34

is connected to an upper portion of the cylinder head

24

, and a downstream end of an intake passage

46

included in the carburetor

34

communicates with the intake port

41

. An intake pipe

47

leading to an upstream end of the intake passage

46

is connected to the carburetor

34

and also connected to an air cleaner (not shown). An exhaust pipe

48

leading to the exhaust port

42

is connected to the upper portion of the cylinder head

24

and also connected to an exhaust muffler

49

. Further, a fuel tank

51

is disposed above the crankcase

22

in such a manner that it is supported on a bracket

50

protruding from the crankcase

22

.

A driving gear

52

is integrally formed on the crankshaft

27

at a location closer to the side cover

26

of the crankcase

22

, and a driven gear

53

meshed with the driving gear

52

is secured to a camshaft

54

rotatably carried in the crankcase

22

and having an axis parallel to the crankshaft

27

. Thus, a rotating power from the crankshaft

27

is transmitted to the camshaft

4

at a reduction ratio of 1/2 by the driving gear

52

and the driven gear

53

meshed with each other.

The camshaft

54

is provided with an intake cam

55

and an exhaust cam

56

corresponding to the intake valve

43

and the exhaust valve

44

, respectively, and a follower piece

57

operably carried on the cylinder block

23

is in sliding contact with the intake cam

55

. On the other hand, an operating chamber

58

is defined in the cylinder block

23

and the cylinder head

24

, so that an upper portion of the follower piece

57

protrudes from a lower portion of the operating chamber

58

; and a pushrod

59

is disposed in the operating chamber

58

with its lower end abutting against the follower piece

57

. On the other hand, a rocker arm

60

is swingably carried on the cylinder head

24

with its one end abutting against an upper end of the exhaust valve

44

biased in a closing direction by a spring, and an upper end of the pushrod

59

abuts against the other end of the rocker arm

60

. Thus, the pushrod

59

is operated axially in response to the rotation of the intake cam

55

, and the intake valve

43

is opened and closed by the swinging of the rocker arm

60

caused in response to the operation of the pushrod

59

.

A mechanism similar to that between the intake cam

55

and the intake valve

43

is also interposed between the exhaust cam

56

and the exhaust valve

44

, so that the exhaust valve

44

is opened and closed in response to the rotation of the exhaust cam

56

.

Referring also to

FIG. 4

, the piston

38

, the crankshaft

27

and a movable eccentric shaft

61

carried in the crankcase

22

of the engine body

21

for displacement in a plane extending through a cylinder axis C and perpendicular to an axis of the crankshaft

27

, are connected to one another through a link mechanism

62

.

The link mechanism

62

comprises a connecting rod

64

connected at one end to the piston

38

through a piston pin

63

, a first arm

66

turnably connected at one end to the other end of the connecting rod

64

and at the other end to a crankpin

65

of the crankshaft

27

, a second arm

67

integrally connected at one end to the other end of the first arm

66

, and a control rod

69

turnably connected at one end to the other end of the second arm

67

and at other end to the movable eccentric shaft

61

. The first and second arms

66

and

67

are integrally formed as a subsidiary rod

68

.

The subsidiary rod

68

includes a semi-circular first bearing portion

70

provided at its intermediate portion to come into sliding contact with half of a periphery of the crankpin

65

, and a pair of bifurcated portions

71

and

72

provided at its opposite ends, so that the other end of the connecting rod

64

and one end of the control rod

69

are sandwiched therebetween. A semicircular second bearing portion

74

included in the crank cap

73

is in sliding contact with the remaining half of the periphery of the crankpin

65

of the crankshaft

27

, and the crank cap

73

is fastened to the subsidiary rod

68

.

The connecting rod

64

is turnably connected at the other end thereof to one end of the subsidiary rod

68

, i.e., to one end of the first arm

66

through a connecting rod pin

75

, which is press-fitted into the other end of the connecting rod

64

inserted into the bifurcated portion

71

at one end of the subsidiary rod

68

and which is turnably fitted at its opposite ends into the bifurcated portion

71

at the one end of the subsidiary rod

68

.

The control rod

69

is turnably connected at one end to the other end of the subsidiary rod

68

, i.e., to the other end of the second arm

67

through a cylindrical subsidiary rod pin

76

, which is passed relatively turnably through one end of the control rod

69

inserted into the bifurcated portion

72

at the other end of the subsidiary rod

68

, and which is clearance-fitted at its opposite end into the bifurcated portion

72

at the other end of the subsidiary rod

68

. Moreover, a pair of clips

77

,

77

are mounted to the bifurcated portion

72

at the other end of the subsidiary rod

68

to abut against the opposite ends of the subsidiary rod pin

76

for inhibiting the removal of the subsidiary rod pin

76

from the bifurcated portion

72

.

The crank cap

73

is fastened to the bifurcated portions

71

and

72

by disposed pair by pair at opposite sides of the crankshaft

27

, and the connecting rod pin

75

and the subsidiary rod pin

76

are disposed on extensions of axes of the bolts

78

,

78

.

The cylindrical movable eccentric shaft

61

is mounted between eccentric positions of a pair of rotary shafts

81

and

82

coaxially disposed and having axes parallel to the crankshaft

27

. Moreover, the rotary shaft

81

is rotatably carried on a support portion

83

mounted to the side cover

26

of the crankcase

22

, and the rotary shaft

82

is rotatably carried on a support portion

84

mounted to the case body

25

of the crankcase

22

.

A follower sprocket

85

is fixed to the rotary shaft

81

, and driving sprocket

86

is fixed to the crankshaft

27

at a location corresponding to the follower sprocket

85

. An endless chain

87

is reeved around the driving sprocket

86

and the follower sprocket

85

. Thus, a rotational power reduced at a reduction ratio of 1/2 is transmitted from the crankshaft

27

to the rotary shafts

81

and

82

, and the movable eccentric shaft

61

mounted between the rotary shafts

81

and

82

is rotated in one rotation about axes of the rotary shafts every time the crankshaft

27

is rotated in two rotations.

By rotating the movable eccentric shaft

61

in the above manner, it is ensured that the stroke of the piston

38

at an expansion stroke is larger than that at a compression stroke. The dimensional relationship in the link mechanism for this purpose will be described with reference to FIG.

5

.

Here, when various dimensions are represented as described below in an x-y plane constituted by an x-axis extending through an axis of the crankshaft

27

along a cylinder axis C and a y-axis extending through the axis of the crankshaft

27

in a direction perpendicular to the x-axis: i.e., a length of the connecting rod

64

is represented by L

4

; a length of the first arm

66

is represented by L

2

; a length of the second arm

67

is represented by L

1

; a length of the control rod

69

is represented by L

3

; a length of from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in a direction of the y-axis is represented by L

5

; a length from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in a direction of the x-axis is represented by L

6

; an angle formed by the connecting rod

64

with respective to the cylinder axis C is represented by φ

4

; an angle formed by the first and second arms

66

and

67

with each other is represented by α; an angle formed by the second arm

67

with the y-axis is represented by φ

1

; an angle formed by the control rod

69

with the y-axis is represented by φ

3

; an angle formed by a straight line connecting the axis of the crankshaft

27

and the crankpin

65

with the x-axis is represented by θ; an angle formed by a straight line connecting the axes of the rotary shafts

81

and

82

and the axis of the movable eccentric shaft

61

with the x-axis is represented by θp; a value of the angle θp when the angle θ is “0” is represented by γ; a length between the crankshaft

27

and the crankpin

65

is represented by R; a length of the straight line connecting the axes of the rotary shafts

81

and

82

and the axis of the movable eccentric shaft

61

is represented by Rp; a rotational angular speed of the crankshaft

27

is represented by ω; and a ratio of the rotational speed of the movable eccentric shaft

61

to the rotational speed of the crankshaft

27

is represented by η and the rotational direction thereof is by η=+0.5, a level X of the piston pin

63

is determined according to

X=L

4

·cos φ

4

+

L

2

·sin (α+φ

1

)+

R

·cos θ (1)

wherein

φ

4

=arcsin {L

2

·cos (α+φ

1

)+R·sin θ−δ}/L

4

φ

1

=arcsin [(L

3

2

−L

1

2

−C

2

−D

2

)/{2·L

1

·(C

2

+D

2

)}]−arctan (C/D)

C=L

5

+Rp·sin θp−R·sin θ

D=L

6

+Rp·cos θp−R·cos θ

θp=η·θ+γ

Here, a speed of the piston pin

63

in a direction of the x-axis is determined according to the following equation by differentiating the above-described equation (1):

dX/dt=−L

4

·sin φ

4

·

d

φ

4

/dt+L

2

·cos (α+φ

1

)·

d

φ

1

/

dt −R

·ω·sin θ (2)

Wherein

dφ

4

/dt=ω·[−L

2

·sin (α+φ

1

)·{R·cos (θ−φ

3

)−η·Rp·cos (θp−φ

3

)}

/{L

1

·sin (φ

1

+φ

3

)}+R·cos θ)}]/(L

4

·cos φ

4

)

φ

3

=arcsin {(R·cosθ−L

6

−Rp·cos θp+L

1

·sin φ

1

)/L

3

}

dφ

1

/dt=ω·{R·cos (θ−φ

3

)−η·Rp·cos (θp−φ

3

)}/{L

1

·sin (φ

1

+φ

3

)}

An equation in a case where dX/d=0 in the above-described equation (2) has four solutions when θ is in a range of −2π<θ<2π. The four solutions are associated with the motion of a 4-cycle engine, and crank angles providing a top dead center at the compression stroke, an top dead center at the intake and exhaust strokes, a bottom dead center after the expansion stroke and a bottom dead center after the intake stroke are determined and used to determine various positions of the piston pin. When the position of the piston pin

63

in the direction of the x-axis at the top dead center at the compression stroke is represented by Xctdc; the position of the piston pin

63

in the direction of the x-axis at the top dead center at the intake and exhaust strokes is represented by Xotdc; the position of the piston pin

63

in the direction of the x-axis at the bottom dead center after the expansion stroke is represented by Xebdc; and the position of the piston pin

63

in the direction of the x-axis at the bottom dead center after the intake stroke is represented by Xibdc, the stroke Scomp at the compression stroke and the stroke Sexp at the expansion stoke are represented by (Scomp=Xctdc−Xibdc) and (Sexp=Xotdc−Xebdc), respectively, and the following dimensions are determined, so that Scomp<Sexp is satisfied and Xctdc=Xotdc is satisfied: the length L

1

of the second arm

67

; the length L

2

of the first arm

66

; the length L

3

of the control rod

69

; the length L

4

of the connecting rod

64

; the length L

5

from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in the direction of the y-axis; the length L

6

from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in the direction of the x-axis; the amount δ of offsetting of the cylinder axis C from the axis of the crankshaft

27

in the direction of the y-axis; the angle α formed by the first and second arms

66

and

67

; the length R between the axis of the crankshaft

27

and the crankpin

65

; the length Rp of the straight line connecting the axes of the rotary shafts

81

and

82

and the axis of the movable eccentric shaft

61

and the angle θp when the angle θ is “0”.

Such determinations ensure that the stroke of the piston at the expansion stroke is larger than that at the compression stroke and moreover, the top dead center at the intake and exhaust strokes and the top dead center at the compression stroke can be identical with each other.

More specifically, the link mechanism

62

is operated as shown in

FIG. 6

at the intake, compression, expansion and exhaust strokes in the engine, and the position X of the piston pin

63

in the direction of the x-axis is varied as shown in

FIG. 7

in accordance with such operation of the link mechanism

62

. Namely, the stroke Sint at the intake stroke and the stroke Scomp at the compression stroke are equal to each other (Sint=Scomp), and the stroke Sexp at the expansion stroke and the stroke Sexh at the exhaust stroke are equal to each other (Sexp=Sexh). Moreover, the stroke Sexp (=Sexh) at the expansion stroke is larger than the stroke Scomp (=Sint) at the compression stroke. Thus, a larger expansion work can be conducted with the same amount of a fuel-air mixture drawn, thereby enhancing the cycle thermal efficiency.

Further, the position Xotdc of the piston pin

63

in the direction of the X-axis at the top dead center at the intake and exhaust strokes and the position Xctdc of the piston pin

63

in the direction of the X-axis at the top dead center at the compression stroke are also congruous with each other.

The operation of the first embodiment will be described below. The engine includes the link mechanism which is constituted by the connecting rod

64

connected at one end to the piston

38

through the piston pin

63

, the first arm

66

turnably connected at one end to the other end of the connecting rod

64

and at the other end to the crankshaft

27

through the crankpin

65

, the second arm

66

integrally connected at one end to the other end of the first arm to constitute the subsidiary rod

68

by cooperation of the first arm, and the control rod

69

turnably connected at one end to the other end of the second arm

67

. The movable eccentric shaft

61

for supporting the other end of the control rod

69

is mounted between the eccentric positions of the rotary shafts

81

and

82

to which the power reduced at the reduction ratio of 1/2 is transmitted from the crankshaft

27

, and the stroke of the piston

38

at the expansion stroke is larger than that at the compression stroke. In such engine, the following various dimensions are determined properly: the length L

1

of the second arm; the length L

2

of the first arm

66

; the length L

3

of the control rod

69

; the length L

4

of the connecting rod

64

; the length L

5

from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in the direction of the y-axis; the length L

6

from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in the direction of the x-axis; the amount δ of offsetting of the cylinder axis C from the axis of the crankshaft

27

in the direction of the y-axis; the angle α formed by the first and second arms

66

and

67

; the length R between the axis of the crankshaft

27

and the crankpin

65

; the length Rp of the straight line connecting the axes of the rotary shafts

81

and

82

and the axis of the movable eccentric shaft

61

and the angle θp when the angle θ is “0”, so that the top dead center at the intake and exhaust strokes and the top dead center at the compression stroke are congruous with each other.

Therefore, it is possible to prevent the occurrence of interferences of the intake valve

43

and the exhaust valve

44

and the top of the piston

38

with each other and to provide an enhancement in compression ratio in the engine to achieve the operation of the engine at a higher thermal efficiency. It is also possible to achieve the sufficient scavenge by the piston

38

to prevent a reduction in output in a full-load state and prevent the instability of the combustion in a lower-load state.

The first and seconds arms

66

and

67

constitute the subsidiary rod

68

having the semi-circular first bearing portion

70

placed into sliding contact with the half of the periphery of the crankpin

65

by cooperation with each other. The connecting rod

64

is turnably connected to one end of the subsidiary rod

68

, and the control rod

69

is turnably connected at one end to the other end of the subsidiary rod

68

. The crank cap

73

having the semi-circular bearing portion

74

placed into sliding contact with the remaining half of the periphery of the crankpin

65

is fastened to the pair of semi-circular bifurcated portions

71

and

72

integrally provided on the subsidiary rod

68

in such a manner that the other end of the connecting rod

64

and the one end of the control rod

69

are sandwiched between the semi-circular bifurcated portions

71

and

72

. Thus, it is possible to enhance the rigidity of the subsidiary rod

68

mounted to the crankpin

65

.

In addition, the connecting rod pin

75

press-fitted into the other end of the connecting rod

64

is turnably fitted at its opposite ends into one

71

of the bifurcated portions, and the subsidiary rod pin

76

relatively rotatably passed through one end of the control rod

69

is clearance-fitted at its opposite ends into the other bifurcated portion

72

. Therefore, the portion from the piston

38

to the subsidiary rod

68

and the control rod

69

are assembled separately into the engine, and the subsidiary rod

68

and the control rod

69

can be then connected to each other. In this manner, the assembling operation can be facilitated, while enhancing the assembling accuracy and as a result, an increase in size of the engine can be avoided.

Moreover, since the connecting rod pin

75

and the subsidiary rod

76

are disposed on the extensions of the axes of the bolts

78

for fastening the crank cap

73

to the subsidiary rod

68

, the subsidiary rod

68

and the crank cap

73

can be constructed compactly, whereby the weight of the subsidiary rod

68

and the crank cap

73

can be reduced, and the loss of a power can be also suppressed.

FIG. 8

shows a second embodiment of the present invention, wherein portions or components corresponding to those in the first embodiment are designated by the same reference numerals and symbols.

A driven gear

90

fixed to the rotary shaft

81

is meshed with a driving gear

52

which is provided on the crankshaft

27

, so that it is meshed with the driven fear

53

fixed to the camshaft

54

. Thus, a rotational power reduced at a reduction ratio of 1/2 is transmitted from the crankshaft

27

through the driving gear

52

and the driven gear

90

to the rotary shafts

81

and

82

, and the movable eccentric shaft

61

mounted between the rotary shafts

81

and

82

is rotated about the axes of the rotary shafts

81

and

82

in one rotation every time the crankshaft

27

is rotated in two rotations.

Moreover, the movable eccentric shaft

61

of the second embodiment rotates in the direction opposite to that the movable eccentric shaft

61

of the first embodiment rotates. That is, in the second embodiment, rotational direction of the movable eccentric shaft

61

is represented by η=−0.5 when its rotational speed is η.

Also in the second embodiment, the top dead center at the intake and exhaust strokes and the top dead center at the compression stroke can be made congruous with each other to provide an effect similar to that in the first embodiment by properly determining the length L

1

of the second arm

67

; the length L

2

of the first arm

66

; the length L

3

of the control rod

69

; the length L

4

of the connecting rod

64

; the length L

5

from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in the direction of the y-axis; the length L

6

from the axis of the crankshaft

27

to the axes of the rotary shafts

81

and

82

in the direction of the x-axis; the amount δ of offsetting of the cylinder axis C from the axis of the crankshaft

27

in the direction of the y-axis; the angle a formed by the first and second arms

66

and

67

; the length R between the axis of the crankshaft

27

and the crankpin

65

; the length Rp of the straight line connecting the axes of the rotary shafts

81

and

82

and the axis of the movable eccentric shaft

61

and the angle θp when the angle θ is “0”.

When the piston

38

is at the expansion stroke, a large load is applied to the piston

38

due to the combustion in the combustion chamber

40

, but if the change in attitude of the piston

38

is increased due to the large load at that time, the friction is increased and the piston slap sound is magnified. Therefore, an arrangement designed to prevent such disadvantage from being arisen will be described in a third embodiment.

To suppress the friction and the piston slap sound, a locus of movement of the piston pin

63

is determined to be fallen into a range between the x-axis and one (which is closest to the x-axis) of tangent lines parallel to the x-axis and tangent to a locus described at the expansion and compression strokes by a point of connection between the connecting rod

64

and the first arm

66

, i.e., the center of the connecting rod pin

75

.

More specifically, at the expansion and exhaust strokes, the link mechanism

62

is operated as shown in

FIG. 9

between a state in which the piston

38

is at the top dead center (a state shown by a solid line) and a state in which the piston

38

is at the bottom dead center (a state shown by a dashed line), and the center of the connecting rod pin

75

describes a locus

95

1

shown by a thin solid line at the expansion stroke and describes a locus

95

2

shown in a thin solid line at the next exhaust stroke, so that a locus

95

provides an endless configuration as a whole. The locus of movement of the piston pin

63

is determined to be fallen into a range between the x-axis and one

96

of a pair of tangent lines parallel to the x-axis and tangent to the locus

951

at the expansion stroke, which is closest to the x-axis.

If the locus of movement of the piston pin

63

is determined as described above, the friction of the piston

38

can be reduced, and the piston slap sound can be suppressed. More specifically, when the piston

38

is at the expansion stroke, a large load is applied to the piston

38

, but if the change in attitude of the piston

38

is increased due to the large load at that time, the friction is increased and the piston slap sound is magnified. However, the above-described determination of the locus of movement of the piston pin

63

ensures that the connecting rod

64

is always inclined to one side at the expansion stroke, notwithstanding that the piston

38

receives the large load at the expansion stroke, whereby the change in attitude of the piston

38

can be suppressed. As a result, the friction of the piston

38

can be reduced, and the piston slap sound can be suppressed.

In the engine in which during lowering of the piston

38

, the stroke at the expansion stroke is larger than that at the intake stroke, and during lifting of the piston

38

, the stroke at the exhaust stroke is larger than that at the compression stroke, as described above, if the link mechanism is set so that the top and bottom dead centers of the piston

38

are retracted at every crank angle of 180 degrees, there is a possibility that the reciprocating speed of the piston at the expansion and exhaust strokes at which the stroke is larger is larger than the reciprocating speed of the piston

38

at the intake and compression strokes at which the stroke is smaller, and the change in acceleration of the piston at the top and bottom dead centers is magnified due to such a speed difference, thereby bringing about a degradation of inertial vibration. Thus, in the engine using the above-described link mechanism

62

, the range of the crank angle at each of the intake, compression, expansion and exhaust strokes can be set at a value other than 180 degrees.

For example, when the link mechanism

62

is set so that it is brought into a state shown by a solid line in

FIG. 10

at the top dead center at the expansion stroke and a state shown by a dashed line in

FIG. 10

at the bottom dead center, the range of the crank angle at each of the intake, compression, expansion and exhaust strokes is as shown in FIG.

11

. The range (=179.8 degrees) of the crank angle at the intake stroke is larger than the range (=153.5 degrees) of the crank angle at the expansion stroke, and the range (=197.7 degrees) of the crank angle at the compression stroke is larger than range (=189.1 degrees) of the crank angle at the exhaust stroke, and the acceleration of the piston

38

in this case is varied as shown in FIG.

12

.

In this case, when the stroke of the piston

38

at the expansion and exhaust strokes is 56 mm; the stroke of the piston

38

at the intake and compression strokes is 37 mm; and a ratio of the volume at the expansion stroke to the volume at compression strokes is 1.5, the largest acceleration (the largest acceleration toward the top dead center) is +6440 m/sec

2

immediately before the expansion stroke changes to the exhaust stroke; the smallest acceleration (the largest acceleration toward the bottom dead center) is −4009 m/sec

2

in the middle of the expansion stroke, as shown in

FIG. 12

, and both (the absolute value of the largest acceleration) and (the absolute value of the smallest acceleration) are large.

Namely, if the range of the crank angle at the intake stroke is larger than the range of the crank angle at the expansion stroke, and the range of the crank angle at the compression stroke is larger than the range of the crank angle at the exhaust stroke, the acceleration of the piston

38

is not reduced and hence, it is impossible to prevent the degradation of inertia vibration.

Therefore, in a fourth embodiment of the present invention, the range of the crank angle at the expansion stroke is set larger than the range of the crank angle at the intake stroke, and the range of the crank angle at the exhaust stroke is set larger than the range of the crank angle at the compression stroke.

Namely, when the link mechanism

62

is set so that it is brought into a state shown by a solid line in

FIG. 13

at the top dead center at the expansion stroke, and a state shown by a dashed line in

FIG. 13

at the bottom dead center, the range of the crank angle at each of the intake, compression, expansion and exhaust strokes is as shown in FIG.

14

. The range (=195.1 degrees) of the crank angle at the expansion stroke is larger than range (=189.9 degrees) of the crank angle at the intake stroke, and the range (=169.7 degrees) of the crank angle at the exhaust stroke is larger than range (=165.3 degrees) of the crank angle at the compression stroke, and the acceleration of the piston

38

in this case is varied as shown in FIG.

15

.

In this case, when the stroke of the piston

38

at the expansion and exhaust strokes, the stroke of the piston

38

at the intake and compression strokes and the ratio of the volume at the expansion stroke to the volume at the compression stroke are set at the same values in the embodiment shown in

FIGS. 10

to

12

, the largest acceleration (the largest acceleration toward the top dead center) is +3377 m/sec

2

at the time when the expansion stroke changes to the exhaust stroke; the smallest acceleration (the largest acceleration toward the bottom dead center) is −2909 m/sec

2

immediately before the exhaust stroke changes to the intake stroke, as shown in

FIG. 15

, and both (the absolute value of the largest acceleration) and (the absolute value of the smallest acceleration) can be reduced remarkably than those in the embodiment shown in

FIGS. 10

to

12

.

Namely, by setting the range of the crank angle at the expansion and exhaust strokes at which the stroke is larger at a value larger than the range of the crank angle at the intake and compression strokes at which the stroke is smaller, the speed of the piston

38

at each of the strokes can be uniform, and the variation in acceleration of the piston at the bottom dead center after the intake and expansion strokes and the variation in acceleration of the piston at the top dead center after the compression and exhaust strokes can be suppressed, thereby avoiding the degradation of inertia vibration.

In addition, in a fifth embodiment of the present invention, the link mechanism

62

is set so that it is brought into a state shown by a solid line in

FIG. 16

at the top dead center at the expansion stroke, and a state shown by a dashed line in

FIG. 16

at the bottom dead center. Thus, the range of the crank angle at each of the intake, compression, expansion and exhaust strokes is as shown in FIG.

17

. The range of the crank angle at the expansion stroke (=178.2 degrees) is larger than the range of the crank angle at the intake stroke (=177.7 degrees), and the range of the crank angle at the exhaust stroke (=185.3 degrees) is larger than the range of the crank angle at the compression stroke (=178.8 degrees), and the acceleration of the piston

38

in this case is varied as shown in FIG.

18

.

In this case, when the stroke of the piston

38

at the expansion and exhaust strokes, the stroke of the piston

38

at the intake and compression strokes and the ratio of the volume at the expansion stroke to the volume at the compression stroke are set at the same values in the embodiment shown in

FIGS. 10

to

12

and the fourth embodiment, the largest acceleration (the largest acceleration toward the top dead center) is +3798 m/sec

2

at the time when the expansion stroke changes to the exhaust stroke; the smallest acceleration (the largest acceleration toward the bottom dead center) is −2212 m/sec

2

immediately before the exhaust stroke changes to the intake stroke, as shown in

FIG. 18

, and both (the absolute value of the largest acceleration) and (the absolute value of the smallest acceleration) can be reduced remarkably than those in the embodiment shown in

FIGS. 10

to

12

.

Also according to the fifth embodiment, the degradation of inertia vibration can be prevented as in the fourth embodiment.

In the fourth and fifth embodiments, however, the acceleration of the piston

38

can be reduced, but the largest acceleration (the largest acceleration toward the top dead center) and the smallest acceleration (the largest acceleration toward the bottom dead center) are imbalanced between the fourth and fifth embodiments. More specifically, in the fourth embodiment, (the absolute value of the largest acceleration)/(the absolute value of the smallest acceleration) is 1.16, and in the fifth embodiment it is 1.72. To reliably prevent the degradation of inertia vibration, it is desirable that (the absolute value of the largest acceleration)/(the absolute value of the smallest acceleration) is a value near to “1”.

The reason why (the absolute value of the largest acceleration)/(the absolute value of the smallest acceleration) is larger than “1” in the fourth and fifth embodiment is considered to be that in the fourth embodiment, the range of the crank angle at the expansion stroke is 195.1 degrees exceeding 180 degrees, while the range of the crank angle at the exhaust stroke is 169.7 degrees smaller than 180 degrees, and in the fifth embodiment, the range of the crank angle at the exhaust stroke is 185.3 exceeding 180 degrees, while the range of the crank angle at the expansion stroke is 178.2 degrees smaller than 180 degrees.

Therefore, in a sixth embodiment of the present invention, the range of the crank angle at the expansion stroke is set larger than the range of the crank angle at the intake stroke, and the range of the crank angle at the exhaust stroke is set larger than the range of the crank angle at the compression stroke, and in addition, the ranges of the crank angles at the expansion and exhaust strokes are set at values exceeding 180 degrees, respectively.

Namely, the link mechanism

62

is set so that it is brought into a state, for example, shown by a solid line in

FIG. 19

at the top dead center at the expansion stroke and a state, for example, shown by a dashed line in

FIG. 19

at the bottom dead center. Thus, the range of the crank angle at each of the intake, compression, expansion and exhaust strokes is as shown in FIG.

20

. The range of the crank angle at the expansion stroke (=191.2 degrees) is larger than the range of the crank angle at the intake stroke (=168.2 degrees), and the range of the crank angle at the exhaust stroke (=190.2 degrees) is larger than the range of the crank angle at the compression stroke (=170.4 degrees), and the acceleration of the piston

38

in this case is varied as shown in FIG.

21

.

According to the sixth embodiment, the speed of the piston

38

at each of the strokes can be further uniform, and the variation in acceleration of the piston at the bottom dead center after the intake and expansion strokes and the variation in acceleration of the piston at the top dead center after the compression and exhaust strokes can be suppressed more effectively, thereby avoiding the degradation of inertia vibration more effectively.

Namely, when the stroke of the piston

38

at the expansion and exhaust strokes, the stroke of the piston

38

at the intake and compression strokes and the ratio of the volume at the expansion stroke to the volume at the compression stroke are set at the same values in the embodiment shown in

FIGS. 10

to

12

, the largest acceleration (the largest acceleration toward the top dead center) is +2467 m/sec

2

immediately before the expansion stroke changes to the exhaust stroke; the smallest acceleration (the largest acceleration toward the bottom dead center) is −2471 m/sec

2

immediately before the exhaust stroke changes to the intake stroke, as shown in

FIG. 21

, and (the absolute value of the largest acceleration)/(the absolute value of the smallest acceleration) ≈1.0 can be achieved.

To ensure that the range of the crank angle at the expansion stroke is set larger than the range of the crank angle at the intake stroke, and the range of the crank angle at the exhaust stroke is set larger than the range of the crank angle at the compression stroke, and in addition, the ranges of the crank angles at the expansion and exhaust strokes are set at the values exceeding 180 degrees, respectively, the dimensions of the various portions in the link mechanism

62

are set as described below.

As shown in

FIG. 22

, the support shaft

61

is displaced to describe a circular locus having a radius Rp about a point spaced within the x-y plane apart from the axis of the crankshaft

27

by the lengths L

5

and L

6

in the directions of the y-axis and the x-axis, respectively, and when the length R between the axis of the crankshaft

27

and the crankpin

65

is set at 1.0, the length L

1

of the second arm

67

is set in a range of 1.7 to 4.5; the length L

2

of the first arm

66

is set in a range of 0.6 to 5.2; the length L

3

of the control rod

69

is set in a range of 4.3 to 6.9; the length L

5

is set in a ranger of 2.3 to 4.0; the length L

6

is set in a range of 0.00 to 3.35; and the radius Rp is set in a range of 0.25 to 1.80, as well as the angle a formed by the first and second arms

66

,

67

is set in a range of 105 to 180 degrees.

By determining the dimensions of the various portions in the link mechanism

62

, the degradation of inertia vibration can be avoided more effectively, as described in the sixth embodiment.

Although the embodiments of the present invention have been described in detail, it will be understood that the present invention is not limited to the above-described embodiments, and various modifications in design may be made without departing from the spirit and scope of the invention defined in the claims.

For example, the sprockets

85

,

86

and the chain

87

have been used to turn the support shaft

61

in each of the above-described embodiments, but a cog belt or the like may be used.

Number	Date	Country	Kind
2002-079736	Mar 2002	JP
2003-050641	Feb 2003	JP

Number	Name	Date	Kind
2259176	Pitney et al.	Oct 1941	A
4517931	Nelson	May 1985	A
4917066	Freudenstein et al.	Apr 1990	A
4957069	Mederer	Sep 1990	A

Number	Date	Country
93 13 192	Dec 1993	DE
9-228853	Sep 1997	JP

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 (2)

US Referenced Citations (4)

Foreign Referenced Citations (2)