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Title:
ENGINE VIBRATION ABSORBER
Document Type and Number:
WIPO Patent Application WO/2019/202542
Kind Code:
A1
Abstract:
This invention relates to a device for absorbing vibration at an engine of a vehicle, transmitted to the engine by wheel inputs, resulting in a reduction of vibration at the engine and the vehicle body. The vibration absorber device comprises an absorber mass fixed to the engine via a biasing means. The mass and biasing means are selected to have a natural frequency which coincides with a natural frequency of a bounce mode of an engine mount system. Consequently, operatively, at least a portion of vibrations transmitted from external sources to the engine, through the engine mounts, is absorbed by the absorber device.

Inventors:
NEL CARL (ZA)
Application Number:
PCT/IB2019/053216
Publication Date:
October 24, 2019
Filing Date:
April 18, 2019
Export Citation:
Click for automatic bibliography generation   Help
Assignee:
UNIV NORTHWEST (ZA)
International Classes:
F16F7/104
Foreign References:
DE3402401A11985-07-25
GB433532A1935-08-09
EP0428949A11991-05-29
EP0103692A21984-03-28
GB433532A1935-08-09
US4641809A1987-02-10
EP0428949A11991-05-29
Attorney, Agent or Firm:
DM KISCH INC. (ZA)
Download PDF:
Claims:
CLAIMS:

1. A vibration absorber device for an engine mount system, which engine mount system comprises engine mounts and an engine, mounted via the engine mounts, to a vehicle, the vibration absorber device comprising an absorber mass operatively fixed to the engine via a biasing means, the mass and biasing means selected to have a natural frequency which coincides with a natural frequency of a bounce mode of the engine mount system, so that, operatively, at least a portion of vibrations transmitted from external sources, through the engine mounts, to the engine, is absorbed by the absorber device.

2. A vibration absorber device according to claim 1 , wherein the natural frequency of the mass and biasing means is selected to coincide with a hop mode natural frequency of a wheel suspension system of the vehicle, and wherein the natural frequency falls within the range of 10 to 13 Hz.

3. A vibration absorber device according to claim 1 , wherein the natural frequency of the mass and biasing means falls within the range of 5 to

10 Hz.

4. A vibration absorber device according to any one of the preceding claims, further comprising a base arrangement, wherein a first end portion of the biasing means is fixed to the base arrangement and wherein a second end portion of the biasing means is fixed to the absorber mass.

5. A vibration absorber device according to claim 4, wherein the base arrangement comprises one of a frame and housing fixed to the engine, so that the base arrangement moves substantially in unison with the engine.

6. A vibration absorber device according to claim 4, wherein the base arrangement is integrally formed with a structural part of the engine. 7. A vibration absorber device according to any one of the preceding claims, further comprising a guide arrangement arranged to limit the movement of a centre of mass of the absorber mass relative to the engine, to oscillation along a one-dimensional axis. 8. A vibration absorber device according to claim 7, wherein the one- dimensional axis is substantially perpendicular relative to a road surface supporting the vehicle in use.

9. A vibration absorber device according to any one of claims 7 and 8, wherein the guide arrangement comprises a linear motion bearing arrangement, comprising a guide in the form of an elongate slide, rail or shaft, fixed to the base arrangement and projecting substantially parallel to the one-dimensional axis, and a bearing arrangement, such as a pad, bushing, slide or roller, fixed to the absorber mass and arranged to receive the guide.

10. A vibration absorber device according to claim 9, wherein the guide arrangement comprises two parallel guides fixed to the base arrangement, and two bearing arrangements fixed to the absorber mass.

11. A vibration absorber device according to any one of claims 7 to 10, in use arranged so that the one-dimensional axis coincides with a position of a centre of mass of the engine.

12. A vibration absorber according to any one of the preceding claims, wherein the biasing means is arranged operatively below the absorber mass, so that the absorber mass exerts a compressive force on the biasing means during static conditions.

13. A vibration absorber device according to any one of the preceding claims, wherein the biasing means comprises an elastic member.

14. A vibration absorber device according to claim 13, wherein the elastic member is selected from the group comprising of a metallic spring, a composite spring, a cylinder and piston arrangement defining a closed volume containing a fluid, and a polymeric member.

15. A vibration absorber device according to claim 14, wherein the polymeric member comprises an elastomeric bush or mount.

16. A vibration absorber device according to any one of the preceding claims, wherein a tuning mass is fixed to the absorber mass to tune the natural frequency of the absorber device in a single degree of freedom to coincide with the natural bounce mode frequency of the engine mount system.

17. A vibration absorber device for an engine mount system, which engine mount system comprises engine mounts and an engine, mounted via the engine mounts, to a vehicle, the vibration absorber device comprising an absorber mass operatively fixed to the engine via a biasing means, the mass and biasing means selected to have a natural frequency within the range of 10 Hz to 13 Hz, wherein, operatively, at least a portion of vibrations transmitted from external sources, through the engine mounts, to the engine, is absorbed by the absorber device.

18. A vibration absorber device substantially as herein described, with reference to the accompanying drawings.

Description:
ENGINE VIBRATION ABSORBER

INTRODUCTION AND BACKGROUND

This invention relates to a vibration absorber for a vehicle. More specifically, the invention relates to a device for absorbing vibrations at an engine of a road-going vehicle, which vibrations emanate from wheel inputs, and which device results in a reduction of vibration of the engine and vehicle body. A typical road-going automotive vehicle comprises an engine typically mounted via engine mounts to a chassis. The vehicle furthermore comprises a vehicle body structure which is normally also mounted to the chassis. In use, the vehicle experiences dynamic fluctuating amplitude excitations, in the form of vibrations. A clear distinction needs to be drawn between vibrations emanating from internal sources, and vibrations emanating from external sources. The vibrations emanating from internal sources are caused by mass imbalances of reciprocating and rotating parts within the engine, and by cylinder gas pressure differences which are directly related to engine speed that cause fluctuating dynamic movements of the engine, Naturally, the frequency of the vibrations emanating from the running engine, varies quite drastically with engine rotational speed, which typically ranges between 700 and 6000 revolutions per minute. To inhibit these vibrations from being transmitted from the engine to the chassis or vehicle body, the engine mounts are typically manufactured from elastomers or rubber. The rubber mounts are therefore provided for the purpose of vibration isolation. It will be appreciated that rotating machines mounted to static structures, such as machines used in industrial plants, also require isolation in the form of rubber mounts.

Also known, and as described in GB 433,532, are vibration absorbers incorporating a spring mass and dashpot arrangement which may be fixed to the engine. Such arrangements are tuned to absorb higher frequency vibrations emanating from the aforementioned internal sources which are directly related to engine speed. These known passive vibration absorber systems are generally ineffective, in that they are incapable of absorbing vibrations over the broad frequency spectrum of the internal sources (at different rotational engine speeds). Their aim is therefore to absorb vibrations at the engine (therefore the source of the vibrations), to minimise the transmitting of vibrations through the engine mounts to the chassis or body. More sophisticated systems utilise active vibration absorbers, wherein a spring and mass arrangement as described above is manipulated by an electronic control system, based on sensor feedback. The control system manipulates an actuator for adjusting a variable parameter of the absorber system. These systems account for frequency changes brought about by varying engine rotational speeds. The electronic components of these systems are very expensive and intricate, severely limiting the practicality and feasibility thereof. The aforementioned systems are therefore specifically provided to absorb or reduce vibrations emanating from internal sources, namely the running of the engine which are thus related to engine speed.

As stated, the vehicle, including the engine, also experiences vibrations emanating from external sources. For the purpose of this disclosure, “external sources” include dynamic fluctuating forces transmitted to the vehicle and the engine, through the suspension of the vehicle, mainly from uneven road surfaces, but also from mass unbalances of the wheels (typically caused by uneven tyre wear and dynamic tyre deformation). Furthermore, depending on the diameter of the wheels of a specific vehicle, a specific driving speed will result in a wheel rotational speed that coincides with a“hop mode” natural frequency of the vehicle’s suspension system, resulting in a resonance condition that will be transmitted to the vehicle as an external source. The external forces are transmitted to the engine, via the wheel suspension system and the rubber engine mounts, since only a portion of the vibration caused by the external sources is reduced by the wheel suspension arrangement of the vehicle.

The suspension wheel hop frequency normally coincides with the natural frequency of the bounce mode of a typical automotive rubber engine mount system, resulting in an undesired resonance condition. As a result of this resonance condition, excessive vibration, as amplitude amplification, is transmitted from the external sources to the engine, in turn resulting in undesired large vibration transmitted from the engine to the vehicle body.

The engine vibration responses caused by external sources are typically much larger than the engine vibration responses emanating from internal sources. The engine vibration responses emanating from external sources cause material fatigue and wear of the body and engine, unwanted noise levels, and discomfort to passengers of the vehicle.

In a bid to inhibit the transmitting of vibrations from external sources to the engine, US patent 4,641 ,809 proposes an engine mount arrangement, to substitute standard rubber engine mounts, which comprises opposing end portions, for fixing between the chassis and engine respectively, and having inwardly facing oblique surfaces between which a wedge-shaped core section is sandwiched. Small masses are fixed to the core section as part of the mount. During compression of the two end portions, the core section is said to rotate in oscillating fashion to dampen vibrations emanating from the external sources.

European patent application 428 949 A1 , on the other hand proposes an engine mount arrangement, to substitute standard rubber engine mounts, which comprises separate internal masses interconnected by elastomeric bodies.

The use of the abovementioned special non-standard rubber engine mounts to isolate an engine from vibrations emanating from external sources, as mentioned above, has proven not to be viable, due to a number of factors. The relative movement between oblique surfaces and oscillating core portions cause unwanted friction and very large material fluctuating stresses inside these special types of mounts, which negatively impacts on the enormous risk of dangerous or catastrophic fatigue failure of these engine mounts.

Relative movement of mass components inside the engine mount is undesirable, due to the size of the static load (the engine’s weight) carried by the mounts. Spatial constraints within engine compartments of vehicles make it difficult to retrofit such engine mounts to existing vehicles. Also, the spatial constraints inhibit relative movement, such as the oscillating movement of a core section, or the relative vibration of internal masses.

Tuning of a spring and mass system to a desired frequency involves selecting a specific mass and a spring having a suitable stiffness. The size of the mass impacts the amplitude of the vibration at the specific frequency. At the desired frequency, a larger mass would have a smaller amplitude of vibration, while a smaller mass would have a larger amplitude of vibration. In practice, in order for an absorber system to be effective, the size of the mass is determined by the mass of the engine, and the maximum allowable dynamic movements experienced by the stiffness element attached to the absorber mass. Due to spatial constraints, the permissible and practical size of additional masses affixed to the engine mounts, and the allowable amplitude of vibration of these masses, are severely limited, negating the effectiveness of these systems. Due in part to the aforementioned limitations, no viable engine mount systems with practically feasible small masses inside or attached to the engine mount have been developed to date. OBJECT OF THE INVENTION

Accordingly, it is an object of the present invention to provide an engine vibration absorber for a vehicle with which the applicant believes the aforementioned problems may at least be alleviated or which may provide a useful alternative for known systems.

SUMMARY OF THE INVENTION According to a first aspect of the invention there is provided a vibration absorber device for an engine mount system, which engine mount system comprises engine mounts and an engine, mounted via the engine mounts, to a vehicle, the vibration absorber device comprising an absorber mass operatively fixed to the engine via a biasing means, the mass and biasing means selected to have a natural frequency which coincides with a natural frequency of a bounce mode of the engine mount system, so that, operatively, at least a portion of vibrations transmitted from external sources, through the engine mounts, to the engine, is absorbed by the absorber device.

The natural frequency of the mass and biasing means may be selected to coincide with a hop mode natural frequency of a wheel suspension system of the vehicle. The natural frequency may therefore fall within the range of 10 to 13 Hz.

Alternatively, the natural frequency of the mass and biasing means may fall within the range of 5 to 10 Hz.

The absorber device may further comprise a base arrangement. A first end portion of the biasing means may be fixed to the base arrangement while a second end portion of the biasing means may be fixed to the absorber mass.

The base arrangement may comprise one of a frame and housing that may be fixed to the engine. In use, the base arrangement may move substantially in unison or in sympathy with the engine. Alternatively, the base arrangement may be integrally formed with a structural part of the engine.

The system may further comprise a guide arrangement which may be arranged to limit the movement of a centre of mass of the absorber mass relative to the engine, to oscillation along a one-dimensional axis.

The one-dimensional axis may be substantially perpendicular relative to a road surface supporting the vehicle in use. The guide arrangement may comprise a linear motion bearing arrangement. The linear motion bearing arrangement may comprise a guide in the form of an elongate slide, rail or shaft which may be fixed to the base arrangement. The guide may and project substantially parallel to the one- dimensional axis. The linear bearing arrangement may furthermore comprise a bearing arrangement, such as a pad, bushing, slide or roller, which may be fixed to the absorber mass and which may be arranged to receive the guide in use.

The guide arrangement may comprise two parallel guides fixed to the base arrangement, and two bearing arrangements fixed to the absorber mass.

In use, the one-dimensional axis may be arranged to coincide with a position of a centre of mass of the engine.

The biasing means may be arranged operatively below the absorber mass, so that the absorber mass exerts a compressive force on the biasing means during static conditions, in other words, when the vehicle is stationary.

The biasing means may typically comprise an elastic member. The elastic member may be selected from the group comprising of a metallic spring, a composite spring, a cylinder and piston arrangement defining a closed volume containing a fluid, and a polymeric member such as an elastomeric bush or mount. The bush or mount may exhibit visco-elastic behaviour.

A tuning mass may be fixed to the absorber mass to tune the natural frequency of the absorber system to coincide with the natural bounce mode frequency of the engine mount system of a specific vehicle.

The absorber mass and base arrangement may be manufactured from a metal, such as steel or lead.

The ratio of the mass of the engine to the mass of the absorber may be between 13 and 22 to 1 , preferably, between 16 and 18 to 1 and most preferably, may be around 17 to 1. The magnitude of the absorber mass may be adjusted by adding relatively small tuning masses to the absorber mass, thereby adjusting the natural frequency of the absorber system, so that this magnitude of the natural frequency matches the bounce mode natural frequency of the engine mount system .

According to a second aspect of the invention there is provided a vibration absorber device for an engine mount system, which engine mount system comprises engine mounts and an engine, mounted via the engine mounts, to a vehicle, the vibration absorber device comprising an absorber mass operatively fixed to the engine via a biasing means, the mass and biasing means selected to have a natural frequency within the range of 10 Hz to 13 Hz, wherein, operatively, at least a portion of vibrations transmitted from external sources, through the engine mounts, to the engine, is absorbed by the absorber device.

According to a third aspect of the invention there is provided an absorber system tuned to absorb vibrations of an engine of a vehicle at a predetermined frequency in the range of 10 to 13 Hz.

The absorber system may comprise an absorber mass fixed to the engine via a biasing means. The magnitude of the absorber mass may be adjusted so that a natural frequency of the system matches a bounce mode natural frequency of an engine mount system of the vehicle.

BRIEF DESCRIPTION OF THE ACCOMPANYING DIAGRAMS

The invention will now further be described, by way of example only, with reference to the accompanying diagrams wherein: figure 1 is a diagrammatic view of an engine vibration absorber device according to the invention; figure 2 is a side view of an example embodiment of the engine vibration absorber device of figure 1 , in use; figure 3 is front view of the engine vibration absorber of figure 2; and figure 4 is an exploded perspective view of the engine vibration absorber of figure 2. DESCRIPTION OF PREFERRED EMBODIMENTS OF THE INVENTION

An engine 12 is typically mounted via engine mounts 14 to a vehicle 16. The engine 12 and engine mounts 14 together constitute an engine mount system 17. A vibration absorber device for the engine mount system 17 is generally designated by the reference numeral 10 in the figures. The vehicle 16 (shown in figures 2 and 3) may typically be a road-going, automotive vehicle.

The vibration absorber device 10 comprises an absorber mass 18 operatively fixed to the engine 12 via a biasing means 20. The absorber mass 18 and biasing means 20 are selected to have a natural frequency which coincides with a natural frequency of a bounce mode of the engine mount system 17, so that, operatively, at least a portion of vibrations transmitted from external sources to the engine 12, through the mounts, is absorbed by the absorber device 10.

In a first and preferable system setup, the natural frequency of the absorber mass 18 and biasing means 20 is selected to coincide with a hop mode natural frequency of a wheel suspension system 22 of the vehicle. Thus, the natural frequency of the absorber mass 18 and the biasing means is selected to fall within the range of 10 to 13 Hz, which range coincides with both the hop mode natural frequency of the wheel suspension system 22 and the natural frequency of the bounce mode of the engine mount system

17.

Since the hop mode natural frequency of the wheel suspension system 22, and the natural bounce mode frequency of a typical automotive engine mount system 17, both fall within the range of 10 to 13 Hz, an undesired resonance condition is caused in use. The wheel hop frequency is mainly excited from external sources, in the form of dynamic fluctuating forces, mainly emanating from uneven road surfaces, but also from mass imbalances of the wheels, typically caused by uneven tyre wear and dynamic tyre deformation. Since the natural frequency of the absorber mass 18 and biasing means 20 falls within this same range in the first system setup, at least a portion of the vibrations transmitted to the engine 12 from these external sources, is effectively reduced by the absorber device 10. The engine vibration absorber device 10 is tuned in a single degree of freedom system, to ensure that the dynamic stiffness coefficient, provided by the biasing means 20 for the vibration absorber 10 provides a similar magnitude of natural frequency as the bounce mode of a typical engine rubber mount system 17, typically within the range of 10 to13 Hz. This is done so that vibration, specifically emanating from external sources, in other words from vibration transmitted through a wheel suspension system 22 to the engine 12 via the engine mounts 14 falling within this frequency range is effectively absorbed by the engine vibration absorber device 10.

In the absence of the absorber device 10, the engine 12 mounted via rubber engine mounts 14 to the vehicle 16, and having a significant mass, therefore may experience amplitude amplification caused by the aforementioned undesired resonance condition. This amplitude amplification results from the forced vibration transmitted from the road surface and wheels 24 to the engine 12, and may in turn, cause fatigue damage to a body 26 and/or frame of the vehicle 16. It was also found that the natural frequency of a human spinal cord is in the same range (10 to 13 Hz). By reducing the amplitudes of the vibration stemming from the external sources, the vibration absorber 10 furthermore improves occupant comfort and noise levels within the vehicle. The applicant has found that the amplitude of the vibration caused by the aforementioned external sources are much larger than the amplitude of vibration caused by internal sources (mass imbalances of reciprocating and rotating parts within the engine 12 and also cylinder gas pressure differences). This relatively larger engine vibration with amplitude amplification is then transmitted through the aforementioned undesired resonance condition from the engine to the vehicle body 26.

The vibration emanating from the external sources is thus effectively reduced by the engine vibration absorber 10. The vibration absorber device 10 is therefore especially effective in reducing material fatigue and passenger discomfort during normal use of road-going vehicles.

It is also foreseen that, by reducing the vibration experienced by the engine, and therefore the forces transmitted to the engine from the external forces, fatigue wear of the engine itself may also advantageously be reduced.

The absorber mass 18 is fixed to the engine 12 via a base arrangement 28 (only schematically shown in figures 1 to 3). A first end 30 (or end portion) of the biasing means 20 is fixed to the base arrangement 28, while a second end 32 (or end portion) of the biasing means 20 is fixed to the absorber mass 18. In one example embodiment, the base arrangement 28 is in the form of a frame or housing which is fixed to the engine 12 (as shown in figures 1 to 3), so that operatively, the base arrangement 28 moves substantially in unison with the engine 12. In such a case the absorber device 10 can easily be retrofitted to existing vehicles, by simply fixing (such as through the use of bolts) the base arrangement to a structural portion of the engine. In an alternative embodiment, the base arrangement 28 may be integrally formed with a structural part of the engine 12 (not shown). As shown in the figures, the shape of the base arrangement 28 is such that it provides a surface, substantially below the absorber mass, to which the first end 30 of the biasing means is fixed, so that the absorber mass 18 is arranged to rest on the biasing means 20, and so that the absorber mass 18 exerts a compressive force on the biasing means 20 during static conditions, such as when the vehicle 16 is stationary or parked. This arrangement is especially suitable during impact loading transmitted through the wheel suspension system 22 of the vehicle 16, caused by uneven road surfaces such as potholes, railway crossings and the like. The biasing means 20 therefore provides the required stiffness coefficient to support the absorber mass 18 during dynamic excitation. A small amount of internal damping is required during transient conditions, such as when the engine is switched on or off. Too much damping, however would be undesired as it would negatively affect the ability of the vibration absorber 10 to reduce vibration during steady-state conditions and caused by external forces.

The biasing means 20 naturally also provides an internal damping coefficient (albeit small), which is diametrically illustrated by the absorber damper 34 in figure 1.

Therefore, as is shown in figures 1 to 3, the absorber device 10 may be represented by an equivalent mass 18, spring 20 and damper 34.

The positioning of the vibration absorber device 10 relative to the engine 12 is such that a one-dimensional axis 36 extending substantially perpendicularly relative to the road surface and coinciding with a position of a centre of mass 38 of the absorber mass 18, coincides with a position of a centre of mass 40 of the engine 12.

A guide arrangement 42 (such as a linear motion bearing arrangement) (as best illustrated in figure 4), comprising a guide 44 and bearing arrangement 46 (see figures 3 and 4), is provided to guide the movement of the absorber mass 18 and to limit angular movement of the centre of mass 38 of the absorber mass 18 relative to the centre of mass 40 of the engine 12. The guide 44 typically takes the form of an elongate slide, rail or shaft, while the bearing arrangement 46 takes the form of a pad, bushing, slide or roller. The guide arrangement 42 therefore ensures that the absorber mass 18 is restricted to vibration movements parallel to the one-dimensional axis 36, during excitation. The centre of mass 38 of the absorber mass 18 therefore vibrates or oscillates along the one-dimensional axis 36 which is substantially perpendicular to a road surface supporting the vehicle 16. The guide 44 therefore projects from the base arrangement 28 substantially parallel to the one-dimensional axis 36.

The centre of mass 38 of the absorber mass 18 and the centre of mass 40 of the engine 12 therefore remains aligned during vibration of the absorber mass 18, ensuring effectiveness of the vibration absorber device 10 in absorbing engine vibrations stemming from the external sources.

The biasing means 20 may be in the form of an elastic member. The biasing means 20 is preferably in the form of a polymeric member such as an elastomeric rubber mount or bush 48 (as shown in figures 2 to 4). Alternatively (not shown) the biasing means 20 may be in the form of a spring (such as a steel or composite spring) or a cylinder and piston arrangement, defining a volume filled with a fluid.

The engine mount system 17, may also be represented by an equivalent spring coefficient schematically shown as a spring 62 and an equivalent damping coefficient schematically shown as damper 64 in figure 1. The wheel suspension system 22 comprises a conventional rubber tyre 50 and a rim 52 with a shock absorber 54 which is attached to the vehicle 16 as diagrammatically illustrated in figures 1 to 3. The wheel suspension system 22 with equivalent unsprung mass 55 (which includes the mass of the rubber tyre 50, suspension and rim 52) have an equivalent spring coefficient diagrammatically illustrated as a spring 56, and an equivalent damping coefficient diagrammatically illustrated as a damper 58 in figure 1. The wheel suspension system 22, as shown in figures 2 and 3, is a conventional suspension arrangement.

As the vehicle 16 travels, an uneven road surface generally causes movements of the wheel suspension system 22 perpendicular to the road surface, exciting the natural frequency of the equivalent mass with spring and damper system. Dynamic fluctuating forces acting in a perpendicular direction to the road surface are transmitted to the engine 12 via the engine mounts 14. The absorber mass 18 and biasing means 20 is selected according to the specific vehicle 16, or model of vehicle, so that the absorber mass 18 and biasing means 20 have a natural frequency matching or coinciding with the bounce mode natural frequency of the engine mount system 17. The natural frequency of the absorber device 10 is a function of both the mass of the absorber mass 18 and the stiffness of the biasing means 20. However, in order for the vibration absorber device 10 to be effective, the dynamic movements (in other words the amplitude of the vibration or oscillation) of the absorber mass 18 needs to be limited. Spatial constraints when the mass becomes too large, and excessive amplitude of vibration when the mass becomes too small, may both render the absorber device 10 ineffective. A ratio of the mass of the engine to the mass of the absorber may be limited to between 13 and 22 to 1. It has been found that a ratio of the mass of the engine to the mass of the absorber of around 17 to 1 provides an acceptable trade-off between spatial constraints of the absorber device 10 and the dynamic movement of the absorber device 10. Additional relatively small tuning masses may be fixed to the absorber mass, to tune the natural frequency of the absorber device 10 in a single degree of freedom to coincide with the bounce mode natural frequency of the engine mount system 17. Excitations of the body 26 and/or frame of the vehicle 16 in this range is thus reduced.

It will be understood that the natural frequency of the absorber device 10, as a single degree of freedom system, is tuned to match the bounce mode natural frequency of the engine mount system 17 (which is determined in the absence of the absorber device 10). The absorber device 10 can therefore practically be tuned as aforementioned by removing the absorber device 10 from the vehicle and tuning the system to match a known (or predetermined) bounce mode natural frequency of the engine mount system, after which the absorber device 10 may be reinstalled to the vehicle.

In one example, the vibration absorber device 10 was installed on an engine of a small front-wheel-drive vehicle, having an engine mass of 179.1 kg and a dynamic equivalent engine mount stiffness coefficient of around 1072 kN/m. Here the magnitude of the absorber mass 18 was 10.41 kg and the stiffness coefficient of the biasing means 20 was around 69 kN/m. The ratio of the mass of the engine to the mass of the absorber mass 18 was around 17.2 to 1. Here both the base arrangement 28 and absorber mass 18 were manufactured from mild steel. It was further found for this vehicle, at a driving speed of around 82 km/h, the wheel rotational speed coincided with the wheel hop natural frequency of the suspension system, which resulted in a resonance condition and an additional external source transmitted to the engine mount system. It was found that the use of the vibration absorber device 10 in the above example had a significant effect in reducing engine 12 vibration, resulting in a large reduction of dynamic forces transmitted to the vehicle’s body 26 and/or frame, compared to the vehicle before the engine vibration absorber device 10 had been mounted. The engine vibration absorber device 10 furthermore caused a significant reduction in noise levels and an improvement in occupant comfort within the cabin of the vehicle. To evaluate the effect that the vibration absorber device 10 has under excitation by internal forces, the vibration was measured while the engine was running at different engine rotational speeds, and while the vehicle was parked. The magnitudes of the dynamic forces were approximately the same with and without the vibration absorber 10. This indicated that the vibration absorber device 10 was effective at reducing the significantly larger vibration amplitudes caused by the external sources, without negative effects emanating from internal sources.

The engine vibration absorber device 10 used in the abovementioned example is shown in figure 4. Here, two guide arrangements 42 having standard parallel guides 44 and bearings arrangements 46 are used, while three standard rubber mounts 48 as the biasing means 20 are used. The absorber mass 18 and base arrangement 28 are manufactured from mild steel and treated with a corrosion protective surface treatment.

It will be understood that the invention focusses on the reduction of vibrations of the engine 12 that emanates from external sources, in other words from vibration transmitted though the wheel suspension system 22 to the engine 12 and base arrangement 28. This vibration falls within the 10 to 13 Hz frequency range and is effectively absorbed by the engine vibration absorber device 10. The dynamic stiffness coefficient for the biasing means 20 (in the form of rubber mount 48) is a function of the magnitude of the absorber mass 18 and the magnitude of the natural frequency of the bounce mode of a specific engine mount system 17, which is normally at or near the typical vehicle 16 suspension wheel hop frequency within a range of 10 td 3

Hz. The bounce mode natural frequency of conventional engine mount systems 17 (which is a direct function of the stiffness of the engine mounts 14) falls within the same range as the hop mode natural frequency of the wheel suspension system 22. The engine mounts 14 need to be stiff enough to prevent excessive sag of the engine mounts 14 under the mass of the engine 12 during static conditions, to limit the permissible displacement of the engine mount 14 under dynamic excitation (stemming from external sources) and to be able to withstand torque transmitted from the engine 12 to the drivetrain (not shown) of the vehicle.

However, the applicant has found that engine vibrations stemming from internal sources (such as those caused by mass imbalances of the reciprocating parts and cylinder gas pressure differences which are related to engine speed) could more effectively be isolated if the natural bounce mode frequency of the engine mount system 17 could be reduced to below 10 Hz. However, to date, providing an engine mount system with a bounce mode natural frequency of below 10 Hz has not been viable, as such a system would not be able to withstand vibrations stemming from the external sources as aforementioned (and since to date, no absorber system could effectively absorb vibrations stemming from external sources). However, with the use of absorber device 10, the engine mount system 17 may have a natural bounce mode frequency falling in the range of 5 to 10 Hz, to more effectively isolate vibrations stemming from internal sources. This will typically be achieved by providing softer or less stiff engine mounts 14. In such a case, the absorber device 10 needs to be provided in a second system setup, where the natural frequency of the absorber mass 18 and biasing means 20 is selected also to fall within the range of 5 to 10 Hz. The absorber mass 18 and biasing means 20 will therefore be specifically selected to result in a natural frequency within this range. Similarly, if an engine mount system 17 with a bounce mode natural frequency exceeding 13 Hz is desired, such as in high performance sports vehicles, the use of the vibration absorber device 10 may effectively limit vibrations transmitted from the engine to the frame or body of the vehicle, preventing damage to the frame or body of the vehicle that would normally occur with an engine mount system 17 having such a high bounce mode natural frequency. Again, the vibration absorber system will be selected so that its natural frequency will match the bounce mode natural frequency of the engine mount system 17.

It will therefore be appreciated that, by providing a vibration absorber device 10 having a natural frequency that coincides with a bounce mode natural frequency of the engine mount system 17, vibrations of the engine 12 that would normally either cause damage to the engine mounts 14 or the frame or body 26 of the vehicle 16, or cause noise or occupant discomfort to the passengers, can effectively be absorbed.

It will further be appreciated by those skilled in the art that the invention is not limited to the precise details as described herein and that many variations are possible without departing from the scope and spirit of the invention.

It will be appreciated that the invention absorbs vibrations that have been transmitted to the engine, instead of seeking to isolate the engine from the chassis or body of the vehicle. In this way, no modifications or alterations are required to the standard engine mounts 14, which means that the negative effects relating to functionality of engine mounts adapted to absorb vibrations emanating from external sources, or to isolate the engine therefrom, as aforementioned, are not present. This also means that the vibration absorber device 10 may easily be retrofitted to existing vehicles. It will readily be appreciated that the vibration absorber device 10 is capable of being used with various types of vehicles, including hatchback or sedan vehicles, pick-up trucks, sports utility vehicles and the like.

Furthermore, the location of the engine, although shown as located in the front of the vehicle in the figures, is not determinative of the functionality of the device 10, and the use of the vibration absorber 10 with mid- or rear- engine vehicles would also be feasible

Lastly, it will be appreciated that the use and effectiveness of the vibration absorber device 10 is independent on the rotational speed of the engine 12, when running.