This invention relates to a device using fluids for controlling mechanical forces such as vibrational forces.

Force-controlling devices are used for example in vehicle suspension systems. Vehicle suspension systems typically employ passive suspension systems, such as a spring in parallel with a damper. A problem with such systems however is that the optimisation of the various performance requirements is limited due to the large masses required in practice to exploit ' inertial forces. The device of US 7316303B addresses this problem by providing a component for building a suspension system with any desired mechanical impedance, but with the overall mass of the suspension system being kept small.

The device disclosed by US 7316303B typically consists of a linear to rotary transducer, connected to a flywheel. Several variations of this device have been proposed, some including for example the use of ball screws or racks and pinions. One disadvantage of these is that there is a considerable number of moving parts.

The present invention therefore seeks to provide a simplified device wherein the number of moving parts is greatly reduced.

According to the present invention, there is provided a device for use in the control of mechanical forces, the device comprising:

first and second terminals for connection, in use, to components in a system for controlling mechanical forces and independently moveable; and

hydraulic means connected between the terminals and containing a liquid, the hydraulic means configured, in use, to produce upon relative movement of the terminals, a liquid flow to generate an inertial force due to the mass of the liquid to control the mechanical forces at the terminals such that they are substantially proportional to the relative acceleration between the terminals. Accordingly, the present invention provides a device wherein the moving liquid acts as storage for kinetic energy. An advantage of the device according to the present invention is its tractability in production.

The hydraulic means may further comprise a housing defining a chamber for containing the liquid, the housing being attached to one of the terminals; and a piston attached to the other terminal and movable within the chamber such that the movement of the piston causes the liquid flow along at least one flow path.

The at least one flow path may be helical. Accordingly, the present invention provides a device wherein a liquid acts both as the linear to rotary transducer as well as storage for rotational kinetic energy.

The at least one flow path may be provided either outside or inside the chamber.

The device may further comprise adjusting means to control the mechanical forces at the terminals such that they are proportional to the relative acceleration between the terminals, the proportionality term being a fixed constant.

The device may further comprise means to restrict the extent of relative movement of the two terminals.

The invention also provides a mass simulator comprising a device according to the above definitions.

The present invention also provides a mechanical damping system, such as a system within a car suspension, employing a device as defined above.

The invention also provides a method for vibration absorption, employing a device as defined above.

Examples of the present invention will now be described with reference to the accompanying drawings, in which:

Figure 1 is a schematic view of an exemplary embodiment of the force- controlling hydraulic device according to the present invention;

Figure 2 is a schematic view of another exemplary embodiment of the force-controlling hydraulic device according to the present invention; Figure 3 shows the pressure drop across the device of Figure 1 as a function of (constant) piston velocity; and

Figure 4 shows the damping force on the piston of the device of Figure 1 as a function of (constant) piston velocity.

We built and tested two prototypes, one provided with a coil external to the cylinder as in Figure 1 and using water as fluid, the other with an internal helical path shaped in the piston itself as shown in Figure 2 and using hydraulic fluid.

Figure 1 illustrates an example of a force-controlling hydraulic device 1 according to the present invention. The device 1 comprises a cylinder 2, a piston 3 movable within the cylinder, and a liquid 4 within the cylinder. The device further comprises a helical tube 5 located outside the cylinder creating a sealed path for the liquid to flow out and back into the cylinder via two orifices (6, 7). Movement of the piston 3 causes liquid 4 to flow through helical tube 5 which generates an inertial force due to the moving mass of the liquid. The cylinder may include one terminal, and the piston may include another terminal. As will be explained below, the inertial force due to a moving mass of liquid caused by relative movement between the terminals controls the mechanical forces at the terminals such that they are substantially proportional to the relative acceleration between the terminals.

The motion of the piston 3 may be restricted by devices such as spring buffers (not shown). Such means may provide a useful safety feature to protect the device if large forces or velocities were generated at the limits of travel of the piston.

The device of Figure 1 is implemented using a through-rod 8. Alternatives using a single rod with a floating piston or a double tube or other similar arrangements are equally feasible. Means to pressurise the fluid (not shown) are envisaged.

Figure 2 illustrates another example of a force-controlling hydraulic device 11 according to the present invention. The device 11 comprises a cylinder 12, a piston 13 movable within the cylinder, and a liquid 14 within the cylinder. The outer surface of the piston 13 has a helical channel, such that, when inserted inside the cylinder, a helical path 15 is formed between the piston and the cylinder. Movement of the piston 13 causes liquid 14 to flow through helical path 15 which generates an inertial force due to the moving mass of the liquid inside the cylinder. In the example of Figure 2 the helical path 15 has a cross-section which is a semi-disc which is convenient for machining. Other cross-sectional shapes may also be employed with advantage to control the damping characteristics of the device. The example of Figure 2 is implemented using a through-rod 15. In the example shown in Figure 1 , the characteristic parameters of device

1 , namely the constant of proportionality with which the applied force at the terminals is related to the relative acceleration between the terminals, can be varied by altering values such as the radii of the piston, cylinder, and helical tube, the length of the cylinder, and liquid density. The effect of such parameters will be detailed below.

Consider the arrangement shown in Figure 1 , where T ₁ is the radius of the piston, r ₂ is the inner radius of the cylinder, r ₃ is the inner radius of the helical tube, r ₄ is the radius of the helix, h is the pitch of the helix, n is the number of turns in the helix, L is the inner length of the cylinder, and p is the liquid density. Further,

A _x = π(r ₂ ² - η ² ) is the cross-sectional area of the cylinder, and

A ₂ = /zr ₃ ² is the cross-sectional area of the tube.

The total mass of liquid in the helical tube is approximately equal to:

pnτ<rϊ 4h ² + {2πr> ) ² =: m _M (1 )

The total mass of liquid in the cylinder is approximately equal to:

pπ(r ² - r{)L =: m _Lyl (2)

If the piston is subject to a linear displacement equal tojc , then a fluid element in the helical tube may expect an angular displacement θ (rads) approximately equal to:

*«(',' - r,') . . ₍₃₎

ri'^h' + Qnr, ) ¹ The moment of inertia of the total liquid mass in the helical tube about the axis of the piston is approximately equal to w _H r ₄ ^J =: J. Now suppose that device 1 has an ideal behaviour with b representing the proportionality constant wherein the generated inertial force between the terminals is proportional to the relative acceleration between the terminals. Then we would expect:

which gives h = ) ² _

(5)

1 + (h/(2πr. )) ² I + (A /(2/TT ₄ )) ² *2

Let m _UΛ - m _hd + m _cyl for the total liquid mass. Exemplary values are tabulated below for two different liquids used in the embodiment shown in Figure 1. In the following examples we assume r ₄ = r ₂ + r _i ,h = 2r ₃ and L = nh . We also take the outside diameter (OD) of the device equal to 2(r ₄ + r ₃ ).

Table 1. A synthetic oil with p = 1200 kg m -3

Table 2. Mercury with p = 13579 kg nϊ -3

As shown in Tables 1 and 2, the modelling and testing work demonstrated that the produced inertance effect (force proportional to acceleration) could be sufficiently large (the proportionality constant b is greater than 50 kg). Such effect would be needed where the device is placed in parallel with a spring and damper.

Furthermore, the modelling and testing demonstrated that the viscosity of the liquid provides a departure from ideal behaviour. A further parasitic element might be provided by the compressibility of the fluid which might be modelled as a spring in series with the two parallel elements.

In US 7316303B, an ideal device is defined (i.e. the force proportional to relative acceleration) and deviations caused by friction, backlash etc. are regarded as parasitics which can be made as small as needed so that the essential function of the device is achieved. In the case of the present invention however, the non-linear damping caused by liquid viscosity is intrinsic, and will cause a deviation from ideal behaviour at large piston velocities.

The non-linear damping intrinsic to the present invention is "progressive", namely the force increases with a relative velocity at a faster rate than linear. Practical dampers in automotive applications are often regressive, namely the force increases with a relative velocity at a slower rate than linear. Even when using ordinary liquids such as hydraulic fluids, the device according to the present invention can be configured to display an ideal behaviour, using adjusting means. For example, shim packs at the orifices 6, 7 could be employed to achieve a more linear damping characteristic, although this would leave a non-negligible parallel damper. This has the potential to create a convenient integrated device with the behaviour of an ideal device according to the present with a linear damper in parallel. In other circumstances it may be advantageous not to correct for the viscosity effect.

The following details the effects of damping. Let u be the mean velocity of fluid in the helical tube, Ap the pressure drop across the piston, // the liquid viscosity, and / the length of the helical tube, where

We will now calculate the pressure drop Δ/? across the main piston required to maintain a flow in the tube of mean velocity w.This will allow the steady force required to maintain a piston relative velocity x to be calculated, and hence a damping coefficient.

Given that A _t x = A ₂ U , the Reynolds Number (Re) for the tube is equal to

(to, . 2.5.,, . -V-j-ii ₍ 7, with transition from laminar to turbulent flow occurring around (Re) = 2 x 10 ³ . Assuming that u \s small enough so that laminar flow holds, and using the Hagen-Poiseuille formula for a straight tube gives:

u = i- ^t * ₍ 8)

The force on the piston required to maintain a steady relative velocity x is equal to ApA ₁ . This suggests a linear damping rate coefficient equal to:

The pressure drop needed to maintain a turbulent flow, according to Darcy's formula is:

where /is a dimensionless friction factor. For a smooth pipe the empirical formula of Blasius is:

/ = 0.079(Re)-" ⁴ . (11) This gives the following expression for the constant force on the piston required to maintain a steady velocity:

F = Ap A ₁

= 0.066 ⁷⁵ T^h _T "' ⁷⁵

Let the fluid be water with p = 100 kg m ^'3 , μ = 10 ^"3 Pa s. Take / = 7 m, T ₁ = 8 mm, r ₂ ⁼ 20 mm, r ₃ = 4 mm, L = 300 mm. This results in a device with:

m _hel = 0.352 kg,

m _cyl = 0.317 kg, and

b = 155 kg.

The transition to turbulent flow occurs at a piston velocity of x - 0.0119 m s ^'1 and at velocities consistent with laminar flow, the damper rate is

c = 77.6 N s m ^"1 .

The pressure drop and linear force in conditions of turbulent flow are shown in Figure 3 and Figure 4, respectively.

If n, r ₂ and r ₃ are all increased by a factor of 2 and / is reduced by a factor of 4 and then m _hd and b are left unchanged, m _t is increased by a factor of

4 and the damping force in turbulent flow is reduced by a factor of 2 ^{1 25} = 2.38.

The helical tube shown in Figure 1 may be replaced in other embodiments of the invention with different shaped tubes. Furthermore, the liquid path may be provided inside the cylinder, with the piston being shaped to provide for example a helical liquid flow path or several concentric helices. Clearances around the piston may also be employed to provide the flow path inside the cylinder. In practice, the best results appear to be achieved by a helical flow path.

For automotive suspension applications, an embodiment of the present invention in parallel with the regular spring and a damper is advantageous if the suspension is relatively stiff. For ordinary passenger cars, the parallel configuration may need to be combined with other series elements.