In a double-tube vibration damper the working chamber adjacent the piston rod guide member is connected to the compensation chamber surrounding the cylinder through a by-pass valve. The by-pass valve is mounted in the cylinder or in a structural member stationary with respect to the cylinder. The by-pass valve has a cross-sectional area variable in dependency on a pressure signal or an electric signal. The damping characteristics of the vibration damper can be varied e.g. in dependency on the load charged onto the vehicle of which the vibration damper is a part.

SUMMARY OF THE INVENTION 
The present invention is directed to a double-tube vibration damper 
particularly for use in a combination shock absorber spring unit 
comprising 
a cylinder member having an axis and two ends and defining a cavity 
therein; 
a bottom valve unit adjacent a first end of said cylinder member; 
a piston rod guide unit adjacent a second end of said cylinder member; 
a piston rod member extending through said piston rod guide unit; 
a piston unit connected to said piston rod member within said cavity, said 
piston unit dividing said cavity into a first working chamber adjacent 
said first end and a second working chamber adjacent said second end, said 
piston unit further comprising a piston valve unit interconnecting said 
first and second working chambers; 
a container surrounding said cylinder and defining a compensation chamber 
in the annular space between said cylinder and said container, said 
compensation chamber being connected to said first working chamber through 
said bottom valve unit; 
a body of liquid in said cavity; 
a body of liquid and a body of gas within said compensation chamber. 
Such double-tube vibration dampers are well known in the art. 
From German `Offenlegungsschrift` No. 2,911,768 a vibration damper is 
known, in which the working chambers on both sides of the piston are 
interconnected by a by-pass passage. In this by-pass passage there is 
provided a by-pass valve. This by-pass valve comprises a valve member 
which is actuated by an electric signal through an electromagnetic 
actuating device. The by-pass valve is located in the piston rod. 
Moreover, vibration dampers are known which are conbined with gas springs 
and in which the increasing pressure of the gas spring resulting from an 
increased load is used to vary the damping action of the vibration damper. 
Also in this known system the by-pass valve is located in a hollow piston 
rod. The valve is actuated by an air-pressure signal through a membrane or 
a separating piston. The control of the damping characteristic is 
established by either controlling the cross-sectional area of the by-pass 
valve or by controlling the prestress of the valve springs of the piston 
valve unit. 
The known constructions have disadvantages. One disadvantage is the 
expensive construction in view of connection lines, membranes, complicated 
valves and so on. Moreover, in the known constructions valve units must be 
used which are different from the valve system commonly used and easily 
available in the market. Moreover, the piston rods are generally weakened 
by the necessary recesses accommodating the by-pass valve and associated 
parts. This is particularly objectionable when transverse forces are to be 
transmitted through the piston rod. Such transverse forces occur if wheel 
guiding forces are to be transmitted through shock absorbers. Finally, the 
dimensions of the vibration damper are generally increased if a by-pass 
valve is provided within the piston rod. 
It is an object of this invention to provide a vibration damper which 
avoids at least part of the above-mentioned disadvantages. A particular 
object of the invention is to provide a vibration damper with a variable 
damping force which is of simplified construction, less expensive than 
known constructions and comprises a by-pass valve unit easily to be 
assembled. A further object is to provide a vibration damper which can 
easily be housed at the location in which vibration dampers are usually 
mounted. 
In view of at least one of these objects a double-tube vibration damper as 
defined above comprises a by-pass passage interconnecting the cavity and 
the compensation chamber. A by-pass valve unit within said by-pass passage 
is located in the cylinder or in a structural member stationary with 
respect to said cylinder in operation. The cross-sectional area of the 
by-pass valve unit is controllable by a pressure signal or an electric 
signal. 
The various features of novelty which characterize the invention are 
pointed out with particularity in the claims annexed to and forming a part 
of this disclosure. For a better understanding of the invention, its 
operating advantages and specific objects attained by its use, reference 
should be had to the accompanying drawings and descriptive matter in which 
there is illustrated and described a preferred embodiment of the 
invention.

DETAIL DESCRIPTION OF THE DRAWING 
The double-tube vibration damper as illustrated in FIG. 1 consists of a 
cylinder 1 filled with damping liquid, the interior of which is divided 
into the two working chambers 8 and 9 by a piston 3 connected with the 
piston rod 2 and comprising a piston valve unit 3a. The container 4 
surrounds the cylinder 1, with spacing, the annular space thus formed 
constituting the compensation chamber 10. The cylinder 1 is centered in 
the container 4 by means of a bottom valve unit 5 and the piston rod guide 
unit 6, while the piston rod seal 7 seals off the interior of the 
vibration damper from the exterior. For the connection of the upper 
working chamber 8 with the compensation chamber 10, which is partially 
filled with damping liquid and partially with gas, the opening 12 is 
provided in the cylinder 1, and its cross-section is varied by a 
pressure-dependently acting control ring 11. Here as actuating device for 
the control ring 11 there serves a tubular or Bourdon spring 14 which is 
connected firmly by one end with the cylinder while the other end has an 
eyelet 15 in which there engages a radially directed projection 16 of the 
control ring 11. In dependency upon the pressure prevailing in the 
compensation chamber 10 the Bourdon spring 14 turns the control ring 11 
and thus also the bore 13 of the control ring 11 in relation to the 
opening 12 in the cylinder wall, and thus achieves a variation of the 
throughflow cross-section. The pressure variation in the compensation 
chamber 10 is effected by a pressure source (not shown) which acts through 
the pressure medium connection 20, here represented as a bore, upon the 
compensation chamber 10 and thus also upon the Bourdon spring 14. The 
pressure variation in the compensation chamber 10 effects a variation of 
the radius of the Bourdon spring 14, whereby the control ring 11 with the 
bore 13 is rotated in relation to the opening 12 and thus a 
pressure-dependent variation of damping force occurs, since the by-pass 
valve formed by the control ring 11 and the opening 12 is effective in 
parallel with the valve units 3a, 5 in both the pulling and the pushing 
direction of the piston rod 2. The pressure variation supplied by way of 
the pressure medium connection 20 to the compensation chamber 10 for the 
actuation of the by-pass valve is substantially greater than the pressure 
fluctuations acting upon the compensation chamber 10 during damping 
operation. 
It is to be noted that on downward movement of the piston rod 2 as shown in 
FIG. 1 the bottom valve unit 5 offers a high flow resistance to the 
damping liquid flowing from the working chamber 9 to the compensation 
chamber 10, while the piston valve unit 3a offers a relatively small flow 
resistance to the damping liquid flowing from the working chamber 9 to the 
working chamber 8. So in both working chambers 8 and 9 a high pressure is 
built up due to the increasing volume of the piston rod 2 within the 
cylinder 1. The liquid under high pressure can escape through the bottom 
valve unit 5 and the opening 12, the cross-sectional area of the opening 
12 being variable in response to the pressure within the compensation 
chamber 10. The opening 12 is substantially in parallel with the bottom 
valve unit 5. 
When the piston rod 2 is moved upwards as seen in FIG. 1 the piston valve 
unit 3a offers a relatively high flow resistance to the liquid flowing 
from the working chamber 8 to the working chamber 9. Liquid from the 
compensation chamber 10 can enter into the working chamber 9 through the 
bottom valve unit 5, the bottom valve unit 5 offering only small flow 
resistance to this latter flow of liquid. So the opening 12 is 
substantially parallel to the piston valve unit 3a. It results from the 
above that by decreasing the cross-sectional area of the opening 12 the 
damping effect of the vibration damper is increased both in the pulling 
and in the pushing direction. On the other hand, when the cross-sectional 
area of the opening 12 is increased the damping effect is decreased both 
in the pulling and the pushing direction 
The source of pressure connected to the pressure medium connection 20 
offers a variable pressure in dependency e.g. of the load on the 
respective vehicle. So the damping effect may be increased if the load is 
increased and vice versa. This is desirable in view of preventing the 
suspension system of the vehicle to swing through the full stroke and come 
to abutment. 
The increase of pressure within the compensation chamber 10 resulting from 
the source of pressure connected to the connection 20 may additionally 
increase the damping effect of the vibration damper. Generally, however, 
the pressure applied through the connection 20 is relatively small as 
compared with the pressure occurring in the working chambers 8 and 9 on 
movement of the piston rod. 
The form of embodiment according to FIG. 2 differs from that according to 
FIG. 1 essentially in that the by-pass valve is formed by an axial bore 18 
situated in the piston rod guide unit 6 and connecting the working chamber 
8 through a further passage 18a with the compensation chamber 10. In a 
radial bore 19 crossing the axial bore 18 a valve pin 17 is guided with is 
in operative connection with the free end of the Bourdon spring 14. The 
Bourdon spring 14 comprises a curved elongate housing 14a defining a 
chamber 14b therein. This chamber 14b is filled with a vacuum or with a 
gas. One end of the elongate housing is fixed to the piston rod guide unit 
6 at 14c. The radius of curvature of the housing 14a is responsive to the 
pressure differential between the chamber 14b and the compensation chamber 
10. So the position of the valve pin 17 is varied in dependency on the 
pressure within the compensation chamber 10, which pressure is again 
dependent on the pressure of the pressure source connected to the 
vibration damper at 20. The position of the valve pin 17 is responsible 
for the effective cross-sectional area of the axial bore 18. 
It is to be noted that the Bourdon spring 14 of FIG. 1 is similar to the 
Bourdon spring 14 of FIG. 2. From FIG. 2a one can see that on variation of 
the pressure differential between the chamber 14b and the compensation 
chamber 10 the left-hand end of the housing 14a is moved in 
circumferential direction about the axis of the piston rod 2 and this 
circumferential movement effects in the embodiment of FIG. 1 the rotation 
of the control ring 11. 
FIG. 1 shows in a diagrammatic way a signal generator 50 connected to the 
opening 13. This signal generator 50 is provided with a plurality of 
selectively operable sensors 51, 52, 53 and can also be actuated by a 
manual actuator 54. 
In the embodiment of FIG. 2 the pressure medium required for the 
pressure-dependent control of the damping force is supplied through the 
pressure medium connection 20 to the compensation chamber 10 and the 
pressure-dependently variable damping force results due to displacement of 
the valve pin 17 in the radial bore 19 as a result of pressure-dependent 
variation of diameter of the Bourdon spring 14. 
In the form of embodiment according to FIG. 3 the metal bellows 21 is 
firmly connected by the lower end with the outer face of the cylinder 1. 
The control ring 11' is also secured to the metal bellows 21, a sealed 
chamber 21a being defined by the bellows 21 the cylinder 1 and the control 
ring 11'. The control ring 11' is axially slidable along the cylinder 1 
and a sealing e1ement 11'a is provided between the cylinder 1 and the 
control ring 11'. The pressure variation in the compensation chamber 10, 
which takes place through the pressure medium connection 20 connected with 
the pressure source, effects an axial movement of the metal bellows 21, 
whereby the control ring 11' is displaced axially on the outer wall of the 
cylinder 1 and thus the throughflow cross-section between the opening 12 
in the wall of the cylinder 1 and the bore 13 situated in the control ring 
11' is varied in pressure dependence. 
In FIG. 4 there is shown a further form of embodiment for the 
pressure-dependent variation of the damping force. In this case on the 
outer wall of the cylinder 1 there is secured a mounting ring 26 which 
serves for the reception of a hollow rubber body formed of a caoutchouc or 
plastic material 23, represented in the left half of the drawing, or for 
the reception of a foam material ring 24 surrounded with an impermeable 
skin, as represented in the right half. The control ring 11" is also 
housed within the mounting ring 26. Since the hollow rubber body 23 or the 
foam material ring 24 can exert a force only in one direction upon the 
control ring 11", the latter is loaded on the other side by the spring 25. 
For the introduction of the control pressure into the working chamber 10 
the pressure medium connection 20 is provided in the container, and an 
increase of pressure effects an axial compression of the hollow rubber 
body 23 or the foam material ring 24 and correspondingly the throughflow 
cross-section of the by-pass valve formed by the bore 13 and the opening 
12 is reduced. Liquid is collected in the annular space between the 
control ring 11" and the mounting ring 26. The liquid passing through the 
openings 12, 13 enters into this volume of liquid so that foaming of the 
damping liquid leaving the opening 13 is largely suppressed. 
The form of embodiment according to FIG. 5 differs from that according to 
FIG. 4 essentially in that the mounting ring 27 is secured to the inner 
wall of the container 4 and the hollow rubber body 23 is provided with a 
pressure medium connection 28 which opens into the interior of the hollow 
rubber body 23. Pressure variation in the interior of the hollow rubber 
body 23 effects the axial movement of the control ring 111'", the upper 
edge of the control ring 111'" cooperating with the opening 12 in the wall 
of the cylinder 1. The spring 25 has the effect that the control ring 11'" 
constantly bears with its extension piece 29 of somewhat larger diameter 
on the end face of the hollow rubber body 23. In order to guarantee a 
satisfactory outflow of the liquid passing the by-pass valve from the 
working chamber 8 into the compensation chamber 10, the control ring 11'" 
is provided with at least one passage opening 30. 
In the form of embodiment as shown in FIG. 6 the pressure-dependently 
acting actuating device consists of a hollow body 31 of annular form, the 
inner wall of which is formed as a diaphragm 32. The by-pass valve is 
formed by the axial bore 18 situated in the piston rod guide unit 6 and by 
the valve pin 17 having a circumferential groove 17a which slides in a 
bore arranged transversely of the axial bore 18. A supply of pressure 
medium through the pressure medium connection 20 into the compensation 
chamber 10 has the effect that the diaphragm 32 assumes the position 
entered in chain lines. The circumferential groove 17a is displaced with 
respect to the axial bore 18 and thus the valve pin 17 reduces the 
cross-section of the bore 18, so that a rise occurs in the damping forces. 
FIG. 7 shows a form of embodiment in which the hollow rubber body 23 
carries a valve plate 33 which is pressed by the internal pressure in the 
hollow rubber body against the opening 12 situated in the cylinder 1. The 
pressure medium is supplied to the interior of the hollow rubber body 23 
through the pressure medium connection 28 secured in the wall of the 
containers 4. In this manner a pressure-dependent variation of 
cross-section is produced between the opening 12 and the valve plate 33. 
A very simple variation of the damping force in dependency upon the wheel 
loading in obtained, as shown by FIG. 8, in that the vibration damper 
forms one structural unit with a pneumatic spring. The spring bellows 34 
is secured on the one hand on the container 4 and on the other hand on the 
protective tube 40 of the vibration damper, and closes off the spring 
chamber 35 to the exterior. The protective tube 40 is movable with the 
piston rod 2 so that the volume of the spring chamber 35 is varied in 
response to movement of the piston rod 2, the volume being reduced on 
downward movement of the piston rod 2. Through the compressed-air 
connection 36 compressed air is supplied to or let off from the spring 
chamber 35, according to the state of loading of the vehicle. The 
compressed-air controlling takes place in accordance with the height 
position of the vehicle, that is to say on increasing of the loading, in 
order to achieve a midway position of the piston in the vibration damper, 
compressed air is admitted through the compressed-air connection 36 until 
this position is reached. The increase of pressure in the spring chamber 
35 takes effect in the compensation chamber (not shown here) of the 
vibration damper, through the pressure medium connection 20, and in this 
way effects a load-dependent controlling for the damping force in that, as 
in the forms of embodiment as described above, the by-pass valve is 
controlled by the pressure variation. In order that the actuating device 
of the by-pass valve in the vibration damper may not be influenced 
appreciably, if at all, by pressure fluctuations which occur in the spring 
chamber 35 during driving, a delay member is provided between the spring 
chamber 35 and the compensation chamber of the vibration damper, 
consisting in that the bore 20 is formed as a small constricted bore. 
In FIG. 9 there is shown an electromagnetic controlling of the by-pass 
valve. In this case the control ring 11.sup.IV comprising the cavity 37 is 
arranged on the outer wall of the cylinder 1. A coil 38 secured on the 
cylinder 1 is situated in this cavity 37. For magnetic controlling it is 
especially suitable to use what is called a linear magnet which effects a 
linear displacement of the control ring 11.sup.IV according to the voltage 
applied, so that the opening 12 in the cylinder wall has a stepless 
variation of the passage cross-section by the control ring 11.sup.IV. 
Referring once more to the embodiment of FIG. 8, it is also possible to 
connect the connection 20 with a pneumatic vehicle suspension separate 
from the shock absorber. 
It is to be understood that the damping characteristic may also be varied 
voluntarily by the driver. So the driver may from the dash board send a 
pressure or electric signal to the respective vibration damper. 
It is also possible to vary the damping characteristic by varying the 
by-pass cross-sectional area in dependency on the speed of the vehicle. It 
is further possible to vary the cross-sectional area of the by-pass in 
dependency upon the braking force in order to achieve a reduction in 
pitching under braking. 
Further it is possible to vary the cross-sectional area of the by-pass in 
dependency of the spring force occurring in the vehicle springs which are 
not necessarily pneumatic springs. 
It is also possible to vary the damping effect in dependency of a plurality 
of the above-mentioned or other parameters. 
While specific embodiments of the invention have been shown and described 
in detail to illustrate the application of the inventive principle, it 
will be understood that the invention may be embodied otherwise without 
departing from such principle.