EP2709862A2 - Vorderes aufhängungssystem - Google Patents

Vorderes aufhängungssystem

Info

Publication number
EP2709862A2
EP2709862A2 EP12726822.5A EP12726822A EP2709862A2 EP 2709862 A2 EP2709862 A2 EP 2709862A2 EP 12726822 A EP12726822 A EP 12726822A EP 2709862 A2 EP2709862 A2 EP 2709862A2
Authority
EP
European Patent Office
Prior art keywords
suspension
vehicle
wheel
resilience
steering
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP12726822.5A
Other languages
English (en)
French (fr)
Inventor
Nicholas Richard Shotter
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
THE FOUR WHEELED MOTORCYCLE COMPANY LIMITED
Original Assignee
Nicholas Richard Shotter
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Nicholas Richard Shotter filed Critical Nicholas Richard Shotter
Publication of EP2709862A2 publication Critical patent/EP2709862A2/de
Withdrawn legal-status Critical Current

Links

Classifications

    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G21/00Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces
    • B60G21/02Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces permanently interconnected
    • B60G21/04Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces permanently interconnected mechanically
    • B60G21/05Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces permanently interconnected mechanically between wheels on the same axle but on different sides of the vehicle, i.e. the left and right wheel suspensions being interconnected
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G21/00Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces
    • B60G21/007Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces means for adjusting the wheel inclination
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G21/00Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces
    • B60G21/02Interconnection systems for two or more resiliently-suspended wheels, e.g. for stabilising a vehicle body with respect to acceleration, deceleration or centrifugal forces permanently interconnected
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G3/00Resilient suspensions for a single wheel
    • B60G3/01Resilient suspensions for a single wheel the wheel being mounted for sliding movement, e.g. in or on a vertical guide
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B62LAND VEHICLES FOR TRAVELLING OTHERWISE THAN ON RAILS
    • B62DMOTOR VEHICLES; TRAILERS
    • B62D9/00Steering deflectable wheels not otherwise provided for
    • B62D9/02Steering deflectable wheels not otherwise provided for combined with means for inwardly inclining vehicle body on bends
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B62LAND VEHICLES FOR TRAVELLING OTHERWISE THAN ON RAILS
    • B62KCYCLES; CYCLE FRAMES; CYCLE STEERING DEVICES; RIDER-OPERATED TERMINAL CONTROLS SPECIALLY ADAPTED FOR CYCLES; CYCLE AXLE SUSPENSIONS; CYCLE SIDE-CARS, FORECARS, OR THE LIKE
    • B62K25/00Axle suspensions
    • B62K25/04Axle suspensions for mounting axles resiliently on cycle frame or fork
    • B62K25/06Axle suspensions for mounting axles resiliently on cycle frame or fork with telescopic fork, e.g. including auxiliary rocking arms
    • B62K25/08Axle suspensions for mounting axles resiliently on cycle frame or fork with telescopic fork, e.g. including auxiliary rocking arms for front wheel
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B62LAND VEHICLES FOR TRAVELLING OTHERWISE THAN ON RAILS
    • B62KCYCLES; CYCLE FRAMES; CYCLE STEERING DEVICES; RIDER-OPERATED TERMINAL CONTROLS SPECIALLY ADAPTED FOR CYCLES; CYCLE AXLE SUSPENSIONS; CYCLE SIDE-CARS, FORECARS, OR THE LIKE
    • B62K5/00Cycles with handlebars, equipped with three or more main road wheels
    • B62K5/10Cycles with handlebars, equipped with three or more main road wheels with means for inwardly inclining the vehicle body on bends
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2200/00Indexing codes relating to suspension types
    • B60G2200/10Independent suspensions
    • B60G2200/13Independent suspensions with longitudinal arms only
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2204/00Indexing codes related to suspensions per se or to auxiliary parts
    • B60G2204/10Mounting of suspension elements
    • B60G2204/30In-wheel mountings
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2204/00Indexing codes related to suspensions per se or to auxiliary parts
    • B60G2204/80Interactive suspensions; arrangement affecting more than one suspension unit
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2204/00Indexing codes related to suspensions per se or to auxiliary parts
    • B60G2204/80Interactive suspensions; arrangement affecting more than one suspension unit
    • B60G2204/82Interactive suspensions; arrangement affecting more than one suspension unit left and right unit on same axle
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2206/00Indexing codes related to the manufacturing of suspensions: constructional features, the materials used, procedures or tools
    • B60G2206/01Constructional features of suspension elements, e.g. arms, dampers, springs
    • B60G2206/50Constructional features of wheel supports or knuckles, e.g. steering knuckles, spindle attachments
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2300/00Indexing codes relating to the type of vehicle
    • B60G2300/12Cycles; Motorcycles
    • B60G2300/122Trikes
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2300/00Indexing codes relating to the type of vehicle
    • B60G2300/12Cycles; Motorcycles
    • B60G2300/124Quads
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B60VEHICLES IN GENERAL
    • B60GVEHICLE SUSPENSION ARRANGEMENTS
    • B60G2300/00Indexing codes relating to the type of vehicle
    • B60G2300/45Rolling frame vehicles
    • BPERFORMING OPERATIONS; TRANSPORTING
    • B62LAND VEHICLES FOR TRAVELLING OTHERWISE THAN ON RAILS
    • B62KCYCLES; CYCLE FRAMES; CYCLE STEERING DEVICES; RIDER-OPERATED TERMINAL CONTROLS SPECIALLY ADAPTED FOR CYCLES; CYCLE AXLE SUSPENSIONS; CYCLE SIDE-CARS, FORECARS, OR THE LIKE
    • B62K5/00Cycles with handlebars, equipped with three or more main road wheels
    • B62K2005/001Suspension details for cycles with three or more main road wheels

Definitions

  • vehicle 1 is balanced by gravity countering the centrifugal force and any other destabilising forces such as crosswinds via the stabilising actions of the rider in the same manner as a two wheeled motorcycle
  • first transverse axis Parallelogram or Formed between, first transverse axis, swing arm, third Quadrilateral transverse axis, strut or upright, fourth transverse axis, upper swing arm, fifth transverse axis, straight line distance between first and fifth transverse axes.
  • Traction Roadholding i.e. the time the tyre is in contact with the ground.
  • Each front swing arm is supported by a resilience means
  • Each swing arm is supported by a primary resilience means.
  • Each swing arm supports a secondary resilience means.
  • Each swing arm moves as one with the associated wheel's leaning movement.
  • Embodiment 1/1 First embodiment of the first variation of the invention Figures
  • Embodiment 2/1 Second embodiment of the first variation of the invention
  • Embodiment 3/1 Third embodiment of the first variation of the invention figure
  • Embodiment 1/2 First embodiment of the second variation of the invention is where embodiments 1/1 2/1 3/1 provide a primary suspension movement and each swing arm support a secondary suspension means.
  • Embodiment 2/2 Second embodiment of the second variation of the invention is where the primary suspension movement is provided by a lever arm but can also be applied to the first variation of the invention.
  • vehicle 1 This invention relates to the various embodiments of the motorcycle type vehicle described in EPO 1998472.3, EP10003570.8, and EP03253106.3, hereinafter referred to as vehicle 1, which use paired transversely pivoted swing arms, hereinafter also referred to as swing arms.
  • vehicle 1 When cornering, vehicle 1 is balanced by gravity countering the centrifugal force and any other destabilising forces such as cross winds via the stabilising actions of the rider in the same manner as a two wheeled motorcycle, hereinafter referred to as balance.
  • Vehicle 1 uses paired transversely pivoted swing arms which advantageously:
  • Vehicle l's paired transversely pivoted swing arms provide an increasingly wide track between their associated paired wheels as vehicle 1 increasingly leans away from the upright position.
  • a 460mm track as per Figure 1 when vehicle 1 is upright or vertical increases to, for instance, 531mm when vehiclel as per Figure 2 leans by 30 degrees in either direction.
  • Figure 1 depicts the end view of the said paired wheels 1 and 2 in contact with the ground 3 generating a 460mm track when vehiclel is at 0 degrees angle of lean, i.e. vertical or upright.
  • Figure 2 shows Figure 1 but with vehiclel leaning at 30 degrees from the vertical in both directions with wheels 1 and 2 generating a 531mm track, which is 71mm more than in Figure 1.
  • the track increase is also exponential.
  • a 1 degree increase in lean from 0 degrees to 1 degree causes the track to widen by 0.07mm.
  • the same 1 degree increase in lean from, for instance, 22 degrees to 23 degrees causes the track to widen by 3.6mm.
  • the graph in Figure 3 shows the exponential differences from 0 degrees to 45 degrees in 5 degree increments for the Figure 1 example of a 460mm track. Track dimensions that are bigger than the 460mm Figure 1 example generate bigger exponential differences and track dimensions that are smaller than the 460mm Figure 1 example generate smaller exponential differences. But whatever vehicle l's upright track dimension is, the exponential difference is a feature of vehicle l's paired transversely pivoted swing arms.
  • the exponentially increasing track moves the associated tyres through an exponentially increasing slip angle which generates an exponentially increasing amount of friction between the said tyres and the surface they are in contact with, hereinafter referred to as friction.
  • Figure 4 shows the plan view of Figure 1 combined with Figure 3's exponentially increasing track dimensions to show; the resulting exponentially increasing slip angles 4 and 5 that are generated between the 460mm track and 650.54mm track as vehicle 1 leans away from the upright position to a 45 degree incline respectively.
  • Figure 5 shows the 650.54mm exponentially increased track of Figure 4 combined with Figure 3's exponentially decreasing track dimensions to show the resulting exponentially decreasing slip angles 6 and 7 generated between the 650.54mm track and 460mm track as vehicle 1 leans from a 45 degree incline to the upright position respectively.
  • Figure 6 shows Figure 4 but with a faster rate of lean which produces a greater exponentially increasing slip angle 104 and 105 and resulting friction.
  • Figure 7 shows Figure 5 but with a faster rate of lean which produces a greater exponentially decreasing slip angle 106 and 107 and resulting friction.
  • the travelling speed of vehicle 1 also determines the resulting friction.
  • Figures 6 and 7 can also be interpreted as having the same rate of lean as Figures 4 and 5 respectively but with vehicle 1 travelling slower than in Figures 4 and 5. It can be appreciated that if the rate of lean was great enough the resulting friction would resist the rate of lean by providing vehicle 1 with discernable roll resistance.
  • Roll is a term commonly used to describe a motorcycle's leaning action.
  • Discernable roll resistance is a term herein used to convey that the rider is experiencing more leaning resistance than is normal for a two wheeled motorcycle.
  • the prototype's tyre wear favourably compares to that of a two wheeled motor scooter with the same tyres. This would not be the case if the prototype's tyres were generating a greater slip angle than a two wheeled motor scooter with the same tyres.
  • vehicle l 's angle of lean is minimal and therefore avoids the greater slip angles associated with the greater angles of lean of faster speeds.
  • a lowside destabilising event is when gravity destabilises the balance and causes a two wheeled motorcycle to fall down on its side.
  • a highside destabilising event is when a sudden increase in the resultant force causes a two wheeled motorcycle to be initially flicked towards the upright position before falling down on its side. A highside always starts from a leaning position.
  • vehiclel's swing arms are transversely pivoted they are ideally positioned to connect with vehicle l 's main structure at a very low point without requiring a two wheeled motorcycle style headstock.
  • the headstock is the part of a conventional motorcycle's frame where the telescopic front forks are rotationaly connected, to provide steering.
  • the low centre of mass that can be achieved from the absence of a headstock and exponentially widening track can keep vehicle 1 from falling over even beyond a 45 degree angle of lean and even with the addition of a rider without becoming unduly heavy or wide.
  • Vehicle 1 's transversely pivoted swing arms can also promote a very low centre of mass by being connected to vehiclel's main structure, or sprung mass, at a very low point. This negates the need for (i) a separate chassis because the swing arms and associated components can be connected directly to, for instance, the
  • the Figure 8 calculation determines the overall centre of mass height com between vehiclel and its rider to be 459.2mm from the ground when; vehiclel's 190kg mass has its centre (Ml) at 240mm from the ground and the rider's 90kg mass has its centre (M2) at 925mm from the ground.
  • Figure 9 shows the Figure 8 com result applied to Figure 1 wherein the arrow represents gravity acting through com to a midway position P between the respective contact patches r and q of wheels 1 and 2 with the ground 3.
  • Figure 10 shows Figure 9 but with vehiclel leaning by 45 degrees and generating the Figure 3 track dimension for 45 degrees wherein; the line com-P is parallel to and midway between the 45 degree inclined wheels 1 and 2.
  • the distance 459.2mm of com-P is the same as in Figure 9 because as vehiclel leans, wheel 1 moves upward by the same amount as wheel 2 moves downward, or visa- versa if vehiclel was leaning in the other direction.
  • the arrow in Figure 10 represents gravity acting through com to form a right angle with ground 3 at position S.
  • a right angled triangle can be formed between com-S and P-S by which P-S can be calculated, as shown in Figure 11.
  • the moving of the tyre's contact patch away from the tyre's centre line provides a useful margin of stability especially if the ground is not completely flat.
  • a static track would have to measure 650mm (650.54mm) when vehiclel was upright to generate the same amount of stability as the exponentially widening track measuring 460mm when vehiclel was upright. Thereby a static track would add 190mm to vehicle l's width.
  • vehicle 1 is 770mm wide when upright. This is 190mm more than vehicle 1 with the said 580mm width.
  • the width that determines how easy it is for a motorcycle to filter through congested traffic hereinafter referred to as the limiting width
  • the limiting width is usually taken from the outside of one of the rider's knees to the outside of the rider's other knee.
  • Limiting widths below 610mm makes a little difference in the associated motorcycle's ability to filter through congested traffic because the average rider's width is still the same.
  • the 610mm maximum limiting width conclusion is based on a study of a major European capital's motorcycle couriers, and their companies, who business it is to continually filter through congested traffic to quickly deliver important packages.
  • Vehiclel's exponentially increasing and decreasing track maximises stability whilst leaning, which is when stability is most needed, and minimises vehicle l 's width when upright which is when stability is less needed but where the narrower width is more beneficial for traffic filtering when leaning is negligible.
  • Vehiclel's transversely pivoted swing arms shortens the wheelbase when cornering to enhance cornering agility, and lengthens the wheelbase when upright to enhance straight line stability.
  • the wheelbase measurement is largely a compromise between the need for straight line stability and the need for cornering agility for any particular purpose such as touring, sport, trials etc. This is because the wheelbase measurement on two wheeled motorcycles remains virtually unchanged between straight line travelling and cornering.
  • vehiclel's average wheelbase exponentially shortens as vehicle 1 progressively leans away from the upright position, and exponentially lengthens as vehicle 1 progressively leans towards the upright position.
  • the reference to average wheelbase means the wheelbase as measured between the mid front track and mid rear track.
  • Figure 12 shows that the up and down movement of vehiclel's paired wheels 1 and 2 generated by a 30 degree angle of lean is 133mm when vehiclel 's upright track, as in Figure 1, measures 460mm.
  • FIG. 13 depicts vehicle 1 as seen from the direction of Figure 12's arrow.
  • the upwards and downwards path taken by each front wheel 1, 2, and back wheel 8, 9 is determined by their 310mm long transversely pivoted swing arms 10, 11, 12, 13, respectively.
  • the front pair of the transversely pivoted swing arms rotate about a first transverse axis 14 and the rear pair of the transversely pivoted swing arms rotate about a second transverse axis 15.
  • the first and second transverse axes 14 and 15 respectively are separated from each other by a non limiting measurement example of 850mm of vehiclel's main body which in Figure 13 is represented by shortened box 16.
  • the wheelbase measures 1410mm.
  • Vehicle I's transversely pivoted swing arms can be positioned sufficiently low and are ideally aligned to be connected directly to the engine crankcase. This enables vehicle 1 not to need a separate chassis which reduces mass. The reduction of mass has many advantages.
  • vehicle I's transversely pivoted swing arms do not require a separate chassis or frame. Therefore a considerable saving in mass can be made by comparison to vehicle 1 having a separate chassis or frame.
  • the mass reduction improves acceleration, fuel consumption, tyre wear, braking, handling,
  • Vehicle I's fourth wheel doubles the lowside and highside control of advantage A above, doubles the stability of advantage B above, increases the gyroscopic wheel stability by a third, and improves braking and cornering, by comparison to a three wheel version of vehicle 1.
  • Vehicle 1 's transversely pivoted swing arms are ideally aligned to be directly connected to the, for instance, engine/transmission unit which is why a separate chassis is not required and hence the great saving in weight.
  • the transversely pivoted swing arms of vehicle 1 can be directly connected to any known engine/transmission configuration with substantially upright cylinder/s.
  • the transversely pivoted swing arms of vehicle 1 can be directly connected to the engine/transmission unit where at least one cylinder is horizontal and in close horizontal proximity to the other engine and transmission components as per EP 10250390.1. £. Has an anti catapult effect
  • first transverse axis 14 of the front swing arms 10 and 11 will tend to ride up and over the front wheels 1 and 2. This is because first transverse axis 14 would not offer any resistance, unlike conventional headstock mounted forks or similar, combined with vehicle l's low centre of mass. Consequently the front of vehicle 1 will lift instead of the rear resulting in the rider not being catapulted off, which also makes vehicle l 's use of an airbag more effective than a conventional motorcycle.
  • Vehicle l's suspension linkages can be so arranged to allow vehicle 1 to lean when cornering at a reduced angle from the vertical than a two wheeled motorcycle for an equivalent, resultant force, trail, wheelbase, tyre section, centre of mass height, and environmental conditions.
  • the reduced leaning angle has many advantages.
  • Vehiclel's arrangement of its interconnecting suspension linkages and rotational axis for paired wheels can be designed to allow vehicle 1 to lean when cornering at a reduced angle from the vertical than a two wheeled motorcycle when the other factors that influence the angle of lean, as detailed above in the brief explanation, are the same. This is achieved by the said arrangement progressively shortening the leverage from the resilience means (e.g. suspension spring) to the associated wheel of each pair laying on the outside of a bend as vehicle 1 progressively leans. Consequently the suspension acting on the wheels laying on the outside of a bend becomes
  • suspension biasing The progressive biasing of the suspension as vehicle! leans is hereinafter referred to as suspension biasing.
  • vehicle 1 does not have to lean so far to negotiate a corner for a given resultant force than otherwise would be the case.
  • This has the following benefits; improved agility, promotes rider confidence, allows for flatter section tyres which improves grip and reduces tyre wear, less need for angular ground clearance which allows for some of vehiclel's components to be positioned lower resulting in a lower centre of mass with the previously mentioned benefits thereof.
  • a further advantage is as follows: As is well known different types of tyres interact with the ground with sufficient difference that their individual characteristics are discemable to the rider.
  • the suspension biasing could be designed to avoid the offending steep slip angles and yet still allow vehicle 1 to comer at a speed normally associated with the said offending slip angle.
  • a side effect of suspension biasing is that the overall friction is reduced during cornering. This is because with suspension biasing the vehicle's weight is
  • suspension biasing may need any of the, rake, trail, Akermann angle, steering offset, camber, settings to be altered from their normal non suspension biasing settings. Different amounts of suspension biasing could be applied a different ends of vehicle 1.
  • the suspension could be biased towards each wheel of a pair lying on the inside of a bend.
  • This can be achieved by designing vehiclel's interconnecting suspension linkages and rotational axis for paired wheels to progressively shorten the leverage from the resilience means (e.g. suspension spring) to the associated wheel of each pair lying on the inside of a bend as vehicle 1 progressively leans. Consequently the suspension acting on the wheels lying on the inside of a bend becomes
  • the advantage F is not unique to transversely pivoted swing arms as the same could apply to longitudinally pivoted swing arms although not to the same extent; the static track of longitudinally pivoted swing arms being a handicap to the scope of suspension biasing.
  • Another advantage of paired transversely pivoted swing arms is the pitch, not track, between the associated paired wheels remains constant irrespective of the independent up and down movement of the associate wheels which makes the drive train to typically the rear wheels easy.
  • the said constant pitch also ensures that suspension deflections do not cause the unsprung mass to generate a destabilising transverse force across vehicle 1. Additional components required to deliver the advantages A, B, C, D, E, F
  • transversely pivoted swing arms provides many unique benefits to vehicle 1.
  • transversely pivoted swing arms for paired wheels requires additional components to make provision for:
  • the leaning movement of vehicle 1 causes one wheel of a pair to move upwards and causes the other wheel of the same pair to move downwards relative to vehicle 1 's main structure.
  • the actual amount of the said upward and said downward movement is mainly determined by vehicle l's upright track measurement and angle of lean.
  • the 133mm upward movement and 133mm downward movement in Figure 12 is derived from vehiclel leaning by 30 degrees and having a 460mm upright track, because Figure 12 is based on the 30 degree angle of lean of Figure 2 which in turn is based on the 460mm upright track of Figure 1.
  • FIG. 16 shows, wheel 1, swing arm 10, and the first transverse axis 14, from Figure 15 combined with a steering axis 17 and ground (e.g. road surface) 3.
  • the steering axis 17 is generating a typically average and positive steering trail value of +75mm.
  • the radius of wheel 1 is 235mm.
  • the angle of the steering axis from the vertical is +17.7 degrees which is derived from Figure 16's right angled triangle formed between, the centre 18 of wheel 1, the centre 19 of wheel l's contact patch with the ground 3, and the intersection 20 of the steering axis 17 with the ground 3.
  • Figure 17 shows Figure 16 but with wheel 1 raised upwards by the said 227mm which causes swing arm 10 to rotate upwards by 47.1 degrees about the first transverse axis 14.
  • Figure 18 shows Figure 16 but with wheel 1 lowered by the said 164mm which causes swing arm 10 to rotate downward by 31.9 degrees about the first transverse axis 14.
  • the steering trail sometimes known as the castor effect, is the distance between where the projected steering axis would intersect the ground and the contact patch centre of the associated tyre with the ground. As vehicle 1 travels, the said contact patch will always try to trail behind the said steering axis intersection away from the direction of travel like a castor.
  • the steering trail is the major factor that causes a steered wheel to self centre or self align. Therefore, vehicle 1 is designed to travel with a positive steering trail, examples of which are shown in Figures 16, 17, 19, 20.
  • vehicle 1 needs a means to keep the trail of both steered wheels within workable positive parameters despite the up and down movement of the associated paired wheels. To achieve this, the angular relationship between vehicle l's paired steering axes and their associated swing arms is not fixed.
  • each steering axis can rotate about a third transverse axis located towards or at the front of the associated swing arm in the form of a predominately upright strut.
  • each strut can rotate about a fourth transverse axis located towards or at the front end of its own dedicated upper swing arm.
  • the rear ends of the two upper swing arms rotate about a common fifth transverse axis laying above the swing arm's common first transverse axis.
  • first transverse axis, swing arms, third transverse axes, struts, fourth transverse axes, upper swing arm, fifth transverse axis, and straight line distance between the first and fifth transverse axes form a quadrilateral or parallelogram for their associated wheel by which the associated steering trail can remain within workable positive parameters.
  • FIG 19 shows Figure 17 but with the addition of the above mentioned
  • first transverse axis 14 and fifth transverse axis 25 are joined to vehicle l 's sprung mass which is represented by distance 26.
  • Figure 20 shows Figure 19 but with swing arm 10 moved to the Figure 18 position.
  • Figures 19 and 20 show that the addition of the said parallelogram mechanism maintains the +75mm Figure 16 positive steering trail despite swing arm 10 moving either upward by 227mm, as per Figure 17, or downward by 164mm, as per Figure 18.
  • the parallelogram mechanism also ensures the trail remains constant throughout the said upward and said downward movement of the swing arm 10 and beyond these extremities if required.
  • the important aspect is to keep the steering trail within workable parameters. Therefore a strict parallelogram is not absolutely necessary as some very small variance in the trail would be tolerable.
  • Figure 21 shows an oblique and raised view of the Figure 19 and Figure 20 parallelogram mechanism for wheel 1 but where vehicle 1 is in the upright position as in Figure 16.
  • wheel 1 is paired with Figure l 's wheel 2.
  • wheel 2 has its own parallelogram mechanism comprised of, first transverse axis 14, swing arm 11, third transverse axis 27, strut 28, fourth transverse axis 29, upper swing arm 30, fifth transverse axis 25, straight line distance 26, which is a mirrored version of wheel 1 's parallelogram mechanism.
  • the wheels 1 and 2 are positioned away from their respective parallelograms so the two parallelograms can be easier seen, and wheel 2's steering axis 31 has been added.
  • the above described parallelogram mechanism also maintains a constant rake.
  • the rake is the angle of the steering axis from the vertical when vehicle 1 is upright.
  • the rake greatly determines the trail and determines the angle the associated tyre interfaces with the ground when steered.
  • the rake also has other benefits.
  • a vertical rake would ensure that the steered wheel remains upright when steered, excluding the influence of vehicle l's angle of lean. If the rake was inclined from the vertical in a direction parallel to vehicle l's longitudinal vertical plane, the associated wheel would also lean sideways, excluding vehicle l's angle of lean, as it is steered. The greater the said incline the greater the wheel would lean sideways in relation to the amount it was steered. If the said rake's incline was horizontal then any steering inputs would only lean the wheel sideways without steering the wheel at all. A said incline to one side of the vertical would cause the associated wheel to lean when steered in an opposite direction to an opposite said incline from the vertical.
  • vehicle 1 had a fixed angular relationship between each of its steering axes and associated swing arm it would cause the two steering axes to be raked differently from each other when vehicle 1 was leaning. Whereby, any steering actions would cause one wheel of a pair to steer less and lean more than the other wheel of the same pair. Furthermore the said steer less and lean more scenario would vary as vehicle 1 leaned by varying angles and steered by varying angles. These said variable steer less and lean more differences between paired steered wheels are not likely to coincide with a workable steering geometry. Hence the need to maintain a near constant rake irrespective of vehicle 1 's angle of lean to ether side.
  • the trail and rake can be kept within workable parameters by the above described parallelogram type steering system.
  • Figure 21 shows vehiclel's steering axis 17 being in-line with strut 22 and steering axis 31 being in-line with strut 28.
  • Figure 22 shows Figure 21 but with the addition of stub axle assembly 40 rotationally connected to strut 22 and thereby steering axis 17 too via kingpin 41, and stub axle assembly 42 rotationally connected to strut 28 and thereby steering axis 31 too via kingpin 43.
  • Steering movements are delivered to stub axle assembly 40 by steering rod 44 via joint 45.
  • Steering movements are delivered to stub axle assembly 42 by steering rod 46 via joint 47.
  • the length of the steering rods 44 and 46 have to be compatible with their associated said parallelogram to ensure the upward and downward movement of wheels 1 and 2 does not independently change the steering angle selected by the rider.
  • the free ends of steering rods 44 and 46 would be interconnected with a handlebar steering mechanism but said mechanism is not relevant to the main purpose of this application and is therefore not shown.
  • Figure 23 shows Figure 22 but with wheels 1 and 2 joined to their associated stub axle assemblies 40 and 42.
  • Two wheeled motorcycle type vehicles that employ a swing arm for the rear wheel normally have a suspension spring acting between the swing arm structure and the sprung mass, typically chassis.
  • two wheeled motorcycle type vehicles that employ a swing arm for the front wheel normally have a suspension spring acting between the swing arm structure and the sprung mass, typically chassis.
  • the cornering balance of vehicle 1 is mainly maintained by gravity countering the centrifugal force. Because gravity is constant its ability to counter varying centrifugal forces depends on varying the gravitational leverage between the combined centre of mass of vehicle 1 and its rider and the distance between the contact patches of paired tyres where the resultant force intersects the ground to maintain balance. This is achieved by varying vehiclel's angle of lean just like a two wheeled motorcycle.
  • vehicle 1 employed the above normal suspension arrangement then additional leverage would be required to counter the force of the said compressed suspension spring in addition to the centrifugal force to maintain cornering balance. Thereby vehicle 1 would have to lean by a greater angle by comparison to employing a suspension arrangement that did not compress the suspension on one side when vehicle 1 leaned.
  • vehiclel's suspension system is not arranged in the above normal manner. Instead vehicle l 's suspension system distributes the force from its resilience means to a position midway, or close to midway if suspension biasing is used, between the associated paired wheels throughout vehiclel 's range of lean in both directions. This is achieved via a balance beam and associated components.
  • Figure 24 shows Figure 21 but without showing wheels 1 and 2 or axes 14, 21, 23, 25, 27, 29 to help clarity by de-cluttering but with the following additions, lug 48 rigidly connected to swing arm 10, connecting link 49 jointed between lug 48 via joint 50 and one end of balance beam 51 via joint 52, lug 53 rigidly connected to swing arm 11, connecting link 54 jointed between lug 53 via joint 55 and the other end of balance beam 51 via joint 56, the balance beam 51 centrally pivoted to carrier 57 about axis 58, carrier 57 linearly guided by guides rods 59.
  • Acting on carrier 57 is resilience means 60 represented by the arrow. The other end of resilience means 60 connects to vehicle 1 's main structure which is the sprung mass.
  • Figure 25 shows Figure 24 but with the addition of suspension dampers 61 and 62 positioned at opposite ends of balance beam 51 from each other.
  • the suspension dampers 61 and 62 can also be locked to prevent tilting of the vehicle as per
  • suspension dampers 61 and 62 The components relevant to the main purpose of this application that are required to provide the suspension damping and anti tilt brake, are the suspension dampers 61 and 62.
  • Figure 26 combines Figure 23, 24, 25 to show the total amount of components that are required to provide the advantages A, B, C, D, E, F, are as follows:
  • Figure 26 typifies the amount of components employed by the various embodiments of vehiclel's front suspension system for paired transversely pivoted swing arms as described in EP01998472.3, EP10003570.8, and EP03253106.3.
  • the resilience means 60 normally a suspension spring, has to control the momentum of many unsprung components which, proportional to their combined mass, limits road holding which hereinafter will be referred to as traction.
  • the unsprung mass is the combined mass of the components that are not supported by the resilience means.
  • Figure 26 which typifies EP01998472.3, EP10003570.8, and EP03253106.3, has the following unsprung components: Wheels 1 and 2, stub axle assemblies 40 and 42, kingpins 41 and 43, joints 45 and 47, steering rods 44 and 46, struts 22 and 28, joints 34 and 35, joints 36 and 37, swing arms 10 and 11, joints 32 and 33, upper swing arms 24 and 30, joints 38 and 39, lugs 48 and 53, joints 50 and 55, connecting links 49 and 54, joints 52 and 56, balance beam 51, axis mechanism 58, carrier 57, suspension dampers 61 and 62 (unsprung proportion thereof).
  • sundry parts All the other associated unsprung components such as, brake parts, wiring, cables, mudguards, bearings, seals, washers, bolts, nuts, circlips, studs, split pins, fluids, etc. hereinafter referred to as sundry parts.
  • the unsprung portion of resilience means 60 and suspension dampers 61 and 62 would also contribute toward the unsprung mass where applicable.
  • Each of the above combinations of unsprung components has an unsprung mass. As the unsprung components move, their unsprung mass generates force which hereinafter will be referred to as momentum. The greater the unsprung mass the greater the momentum for a given speed of movement.
  • a resisting force is supplied by the resilience means which pushes against both the unsprung mass and the sprung mass.
  • the sprung mass being the combined mass of the components that are supported by the resilience means, including people and any other carried load.
  • the effectiveness of the resilience means in resisting the momentum of the unsprung mass greatly depends on the ratio between the unsprung mass and the sprung mass. For instance, if the unsprung mass was greater than the sprung mass the resilience means would move the sprung mass more, when the associated wheel/s encountered a bump on the ground, than moving the unsprung mass after the associated wheel/s encountered a bump on the ground. This would not be very useful in resisting the momentum to maintain traction.
  • the unsprung mass To ensure that the resilience means moves the unsprung mass more than the sprung mass the unsprung mass must be less than the sprung mass. The less the unsprung mass is in relation to the sprung mass the more the resilience means is able to resist the momentum and maintain traction
  • the ratio between the unsprung mass and sprung mass could be increased by adding to the sprung mass. But this would increase manufacturing costs, tyre wear, fuel consumption, and reduce acceleration, braking, handling, and manoeuvrability. It is much better to increase the ratio between the unsprung mass and sprung mass by reducing the unsprung mass. This can be achieved by reducing the mass of the unsprung components or/and changing some of the unsprung mass into sprung mass.
  • the swing arms could slope downwards from the first transverse axis 14 to wheel centre 18 and likewise at the rear of vehiclel as shown in the Figure 27 variation of Figure 15.
  • the front and back wheels that lay towards the inside of the comer whilst cornering generate a longer wheelbase than those lying toward the outside of the corner as shown in the Figure 28 variation of Figure 13 where vehiclel is cornering left.
  • the inside wheels 1 and 8 are producing longer levers L2 than the LI levers associated with the outside wheels 2 and 9. Consequently the force from resilience means 60 to the said inside wheels travels through the longer leverage L2 than to the said outside wheels via leverage LI. This means that the resultant force would act closer to the said outside wheels whilst cornering than otherwise would be the case which can be a desirable characteristic as described in F.
  • the said swing arm's downward slope also shortens the swing arm in any of the EP01998472.3 and EP03253106.3 upwardly looped swing arm variations of vehiclel, which makes the swing arm stiffer or/and saves weight and manufacturing costs.
  • the said swing arm's downward slope more equally distributes the suspension's compression and extension movements to each side of the horizontal plane when the suspension's compression potential exceeds the suspension's extension potential, as in Figuresl9 and 20.
  • the upper diagram in Figure 29 shows the said downwards slope of swing arm 10 with wheel 1 on ground 3.
  • the lower diagram in Figure 29 shows the said upper diagram but where the bump causes wheel 1 to move upwards as represented by vector X and to move forwards as represented by vector W.
  • the resultant of vectors X and W lies between them as represented by vector V.
  • the invention is a new suspension system for the paired front wheels of motorcycle type vehicles employing transversely pivoted swing arms for said paired wheels such as those described in but not limited to the prior art.
  • the invention can provide all of the above advantages A, B, C, D, E, F, of paired transversely pivoted swing arms and can also make provision for G, H, I, J, above.
  • the invention differs from the prior art in that it provides a much greater potential for enhancing traction than the prior art can deliver for an equivalent vehicle mass and thereby redresses the First Problem.
  • the invention can also provide a much better suspension deflection than the prior art can deliver without compromising the associated vehicle's angle of lean which redresses the Second Problem.
  • the enhanced traction of the invention is achieved by increasing the ratio between the sprung mass and unsprung mass by comparison to the prior art by converting some of the unsprung components associated with the prior art into sprung components.
  • the resilience means acts on the associated wheels with fewer components interconnected between the resilience means and the associated wheels than in the prior art.
  • each front transversely pivoted swing arm of a pair is a first variation of the invention.
  • each front transversely pivoted swing arm of a pair is a first front transversely pivoted swing arm of a pair:
  • the resilience means in both variations of the invention acts closer to the associated wheels than in the prior art. This reduces the unsprung components, which permits an increase in the ratio between the unsprung mass and sprung mass, which ensures the associated tyres will maintain traction with the ground for more of the time, which improves, stability, braking, and steering, and in the case of front wheel drive or four wheel drive also improves acceleration, beyond the limitations of the prior art for an equivalent vehicle mass.
  • each of the embodiments of the invention can allow the front wheels of the associated vehicle to deflect in closer proximity to the resulting force generated by the associated front wheel/s
  • Figure 1 Represents vehicle 1 's paired wheels with a 460mm track example.
  • Figure 2 Represents Figure 1 but leaning to a 30 degree example to demonstrate the increased track.
  • Figure 3 Represents the exponentially increasing and decreasing track
  • Figure 4 Represents a plan view of Figure 1 combined with the exponentially increasing track dimensions of Figure 3 to show the resulting slip angle.
  • Figure 5 Represents the increased track of Figure 4 with Figure 3's
  • Figure 6 Represents Figure 4 with a faster rate of lean or/and slower vehicle speed.
  • Figure 7 Represents Figure 5 with a faster rate of lean or/and slower vehicle speed.
  • Figure 8 Represents a visual calculation to determine a combined centre of mass height generated between vehicle 1 and its rider.
  • FIG. 9 Represents Figure 1 combined with the Figure 8 result.
  • Figure 10 Represents Figure 9 but with vehicle 1 leaning by 45 degrees.
  • Figure 11 Represents a visual stability calculation derived from figure 10.
  • Figure 12 Represents how far the paired wheels in Figure 2 move up and down.
  • Figure 13 Represents vehiclel viewed from the direction of Figure 12's arrow.
  • Figure 14 Represents a visual calculation to prove vehiclel 's shortening
  • Figure 15 Represents Figure 13 but when vehiclel is upright.
  • Figure 16 Represents a trail for vehicle 1.
  • Figure 17 Represents the effect to Figure 16's trail as the swing arm rises if the steering axis and swing arm were in a fixed angular relationship.
  • Figure 18 Represents the effect to Figure 16's trail as the swing arm lowers if the steering axis and swing arm were in a fixed angular relationship.
  • Figure 19 Represents Figure 17 but with the inclusion of a parallelogram and without the steering axis and swing arm being in a fixed angular relationship.
  • Figure 20 Represents Figure 19 but in the Figure 18 position.
  • Figure 21 Represents an oblique view of Figure 19 and 20 but in the Figure 16 position and for paired wheels which are shown away from their locations.
  • FIG. 22 Represents Figure 21 with the addition of relevant steering
  • Figure 23 Represents Figure 22 with the wheels in place.
  • Figure 24 Represents Figure 21 with the addition of suspension components and without the wheels.
  • FIG 25 Represents Figure 24 with the addition of the suspension dampers.
  • Figure 26 Represents the combination of Figures, 23, 24, 25.
  • Figure 27 Represents Figure 15 but with downward sloping swing arms.
  • Figure 28 Represents Figure 27 combined with Figure 13.
  • Figure 29 Represents a part of Figure 27 encountering a bump.
  • Figure 30 Is a vector diagram incorporating Figure 29 vector conclusions.
  • Figure 31 represents the outward facing side of wheel 2 comprising tyre 201, rim 202, rim mounted brake disc 203, and identifies cross section AA.
  • Figure 32 Represents the cross section AA of Figure 31 to reveal the top view of embodimentl/1 comprising stub axle assembly 42, upper frame 63R, upright 28, resilience/damping means 64R comprising compression coil spring and suspension damper - typically a shock absorber, bracket 65R, locating fasteners 66R, kingpin 43, guides 67R, transversely pivoted swing arm 11 (truncated in Figure 32).
  • R right side
  • Figure 33 Represents Figure 32 from the direction of arrow B but without wheel
  • Figure 41 Represents a common hydraulic fluid displacement device.
  • Figure 42 Shows section BB of Figure 41.
  • Figure 43 Represents Figure 42 with the addition of a damping means.
  • Figure 44 Shows the remotely positioned resilience means from the displacement devices.
  • Figure 45 Represents the use of the king pin as the piston rod.
  • Figure 46 Shows a one guide rod version of Figure 33
  • Figure 47 Shows view C 1 of Figure 46
  • Figure 49 Shows an embodiment where the guide rod is also the piston rod and shows an assembly to transfer steering torque
  • Figure 50 Shows view E of Figure 49
  • FIG 52 Shows Figure 49 where the steering torque assembly is also an anti brake dive mechanism
  • Figure 53 Shows a leading arm arrangement
  • Figure 54 Shows view C2 of Figure 53
  • Figure 56 Shows Figure 53 but where the brake torque is isolated from the
  • Figure 57 Shows a trailing arm arrangement
  • Figure 58 Shows view El of Figure 57
  • Figure 59 Shows Figure 57 but with the suspension more compressed
  • Figure 60 Shows Figure 57 but where the brake torque is isolated from the
  • Figure 61 Shows a trailing arm arrangement where the brake torque is used to generate an amount of anti brake dive.
  • Figure 62 Shows how one balance beam can be used for paired front and back wheels for the first variation of the invention
  • Figure 63 Shows Figure 62 but relating to the second variation of the invention
  • Figure 64 Shows a development of Figure 62 where the connecting links do not cross
  • Figure 65 Shows Figure 64 but relating to the second variation of the invention
  • Figure 66 Shows how an offset balance bean can be applied to Figures 64 and 65
  • a first embodiment of the first variation of the invention (embodimentl/1) will now be described with reference to Figures 31 to 40 and in relation to the prior art as typified by Figures 26:
  • wheel 2 is rigidly connected to stub axle assembly 42
  • stub axle assembly 42 is rotationally connected to housing 68R to provide wheel 2 with its rotational movement
  • housing 68R can linearly slide along guides 67R
  • guides 67R are statically fixed to upper frame 63R and lower frame 69R
  • the fixed positional relationship of the two guides 67R with each other is determined by upper frame 63R and by lower frame 69R
  • bracket 65R is connected to housing 68R
  • the lower end of resilience/damping means 64R is rotationally connected to bracket 65R via locating fastener 66R
  • the upper end of resilience/damping means 64R is rotationally connected to upper frame 63R via the locating fastener 66R
  • upper frame 63R is connected to upright 28
  • rotational joint 37 (which in Figure 37 can be one of the joint types described in GB2435021) interconnects the top of upright 28 with the front of upper swing arm 30, the rear of upper swing arm 30 is connected to joint 39
  • Suspension actions occur when deflections of tyre 201 in response to the ground surface's dips and bumps cause wheel 2, stub axle assembly 42, housing 68R, and bracket 65R, to move upwards and downwards as one unit in relation to upper frame 63R and lower frame 69R via guides 67R.
  • the said upward and downward movements which is sometimes referred to as bump and rebound, are controlled by resilience/damping means 64R.
  • embodimentl/1 only has four completely unsprung components associated with wheel 2 whose numbers are, 2, 42, 68R, 65R.
  • components 2 and 42 are common to both the prior art as typified by Figure 26 and embodimentl/1 they can be removed from the comparison of unsprung components.
  • the comparison of unsprung components is between the two components 68R and 65R of embodiment/ 1 and the seventeen components 43, 47, 46, 28, 35, 37, 11, 33, 30, 39, 53, 55, 54, 56, 51, 58, 57, of the prior art as typified by Figure 26.
  • the actual amount of unsprung components in some instances of prior art may be slightly less than those in Figure 26 the amount of unsprung components in the prior art are still considerably more than in any of the embodiments of the invention.
  • the ratio between the unsprung mass and sprung mass is the important aspect and not the amount of unsprung components. For instance, a vehicle with a 1 : 10 ratio between its unsprung mass and sprung mass and having a total of four unsprung components would have worse traction than a vehicle with a 1: 15 ratio between its unsprung mass and its sprung mass but having a total of say nineteen unsprung components (assuming the tyres, tyre pressure, suspension, and any other relevant parameters were optimised for both 1 :10 and 1 : 15 ratio scenarios).
  • embodiment 1/1 also only has four completely unsprung components associated with wheel 1 whose numbers are, 1, 40, 68L, 65L, and the prior art, as typified by Figure 26, has nineteen completely unsprung components associated with wheel 1 , whose numbers are, 1, 40, 41, 45, 44, 22, 34, 36, 10, 32, 24, 38, 48, 50, 49, 52, 51, 58, 57. Consequently embodimentl/1 also improves the unsprung mass to sprung mass ratio on the left side of the associated vehicle by the same extent as on the right side.
  • embodimentl/1 and prior art Therefore the omission of the unsprung portion of the resilience/damping means does not undermine the advantage of embodimentl/1 over the prior art in terms of increasing the ratio between the unsprung mass and sprung mass to improve traction.
  • the above comparisons also omit the sundry parts for the same reason of clarity as for the omission of the partly unsprung components.
  • Embodimentl/1 does not employ the unsprung rotational axis 58, carrier 57, guides 59, resilience means 60, dampers 61 and 62, of Figure 26. Instead, balance beam 51 in embodimentl/1 rotates about axis 58 that is fixed to a sprung part of the associated vehicle as shown in Figure 38. In embodimentl/1 resilience means 60 and dampers 61 and 62 associated with vehicle 1 as per Figure 26 have been replaced by resilience/damping means 64R/L as shown in Figures 37 and 38.
  • embodimentl/1 is positioned within the internal space of rim 202. Consequently embodiment 1/1's size is limited by the internal space of rim 202.
  • resilience/damping means 64 can instead be angled to the guides 67R (angle/1) as shown in Figure 34. Angle/1 also compensates for the greater increase in resistance of a shorter spring by comparison to a longer spring for the same rate of compression.
  • resilience/damping means 64R such as shown in Figure 34
  • resilience means 64R in Figure 34 had a distance of 150mm between the centres of the two location fasteners 66R, was angled by 45 degrees to guides 67R, had the above 80mm free length spring compressed to 60mm, was generating 40kg of resistance, then a 30mm upward movement of housing 68R would increase the spring's resistance to 79kg instead of thelOOkg in the above example,
  • angle/ 1 broadens the scope for the associated spring's design in two ways. Firstly, by providing a longer distance for the spring within the limitations of rim 202 and secondly, by progressively reducing the effective spring rate as the suspension compresses. Consequently, the greater the angle of angle/1, within workable parameters, the broader is the scope for the spring's design.
  • Resilience/damping means 64R could alternatively be angled (angle/2) in relation to guides 67R as in Figure 39.
  • the lower end of resilience/damping means 64R is located to housing 68R via a spacer or similar (unseen) that is long enough for resilience/damping means 64R to clear kingpin 43.
  • resilience/damping means 64R is located to the right side of upper frame 63R. To further increase the length and therefore the design scope of resilience/damping means 64R within the confines of rim 202 resilience/damping means 64R can be positioned to combine angle/1 and angle/2 to form a compound angle/3.
  • Figure 40 shows angle/3 where resilience/damping means 64R locates to a bracket 65R (or similar) extending from housing 68R as in Figure 34 but to the other side of kingpin 43 with the upper end of resilience/damping means 64R located to upper frame 63R in a similar position to that shown in Figure 34. In Figure 40 the location fasteners of Figure 34 have been replaced with ball joints 70R.
  • embodimentl/1 There will also be other design considerations associated with embodimentl/1 such as but not limited to: The amount of steering lock, steering momentum, lowering the centre of mass, further reduction of the unsprung mass, material costs, tooling costs, side force on guides 67R, camber, offset of wheel 2's rotational axis from steering axis 31, offset of steering axis 31 from the centre of tyre 201 's contact patch with ground 3, rake, trail, position and type of the braking mechanism, drive shaft if there is a drive to wheel 2, weather protection of parts, electric or hydraulic drive of wheel 2, and combinations thereof.
  • the position of resilience/damping means 64R can depart from those already discussed to further broaden the design scope without departing from the theme of embodimentl/1 as per the following non limiting examples and combinations thereof.
  • resilience/damping means 64R is shown far to the right of kingpin 43 to provide a large steering lock. Where a particular application of embodimentl/1, for instance when applied to a sports variation of vehicle 1, does not need as great a steering lock as Figure 36 would provide then resilience/damping means 64R can be positioned closer to kingpin 43. This would better centralise resilience/damping means 64R with the suspension deflections caused by wheel 2 which reduces the side loading on guide rods 67R and also reduces the steering momentum.
  • Resilience/damping means 64R can also be positioned to the left side of kingpin 43 when viewed from the perspective of Figure 36.
  • resilience/damping means 64R is positioned towards the inside of vehicle/ 1 from kingpin 43.
  • resilience/damping means 64R could be positioned towards the outside of vehicle/1 from kingpin 43 to provide a greater steering lock before the minimal desired clearance between resilience/damping means 64R and lower swing arm 11 is reached.
  • Figure 39 shows the lower end of resilience/damping means 64R connected to housing 68R to the left of kingpin 43 and the upper end of
  • resilience/damping means 64R connected to the far right of upper frame 63R.
  • the lower end of resilience means 64R can instead be mounted to housing 68R to the right of kingpin 43 or/and the upper end of resilience/damping means 64R can be mounted closer to kingpin 43 rather than as far away as possible.
  • angle/1 could be reversed in relation to guides 67R so that the lower end of resilience/damping means 64R connects instead directly to housing 68R via a location fastener 66R with the upper end of resilience/damping means 64R still connecting to the upper frame 63R via a location fastener 66R but instead closer to upright 28 and not as far to the right as possible as shown in Figure 34.
  • This would eliminate the need for bracket 65R which would further reduce the unsprung mass and associated material/tooling costs as well as reducing the steering momentum and increasing the room for other nearby components.
  • resilience/damping means 64R can be the other way up than shown in the Figures, and either location fasteners 66R or/and ball joints 70R can be used.
  • a stub axle for wheel 2 can be integral with or connected to housing 68R with at least one associated wheel bearing housed in stub axle assembly 42.
  • a stub axle for wheel 2 can be integral or connected to the stub axle assembly 42 with at least one associated wheel bearing housed in housing 68R.
  • a stub axle for wheel 2 could interconnect between at least one associated wheel bearing housed in housing 68R and at least one associated wheel bearing housed in stub axle assembly 42.
  • the stub axle, or the rotational axis of wheel 2 if different from the stub axle's longitudinal centre line, does not have to be perpendicular to the length of the guide rods 67R or kingpin 43.
  • Rim 202 can either bolt to the flange of stub axle assembly 42 or screw onto stub axle assembly 42, the latter with or without a knock-off type hub or similar.
  • the flange of stub axle assembly 42 can be any feasible diameter with the corresponding widening of the central hole of rim 202.
  • An extreme example of the central hole of rim 202 being widened would be akin to that used on Lambrettor scooters and Volkswagen cars and vans during the 1950/60s.
  • Rim 202 could also be an integral unit with stub axle assembly 42 with or without the associated said integral or connected stub axle.
  • brake disc 203 is mounted via its outer circumference to rim 202.
  • Brake disc 203 can either be solidly connected to rim 202 or be connected to rim 202 but allowed to float sideways to the rotation of wheel 2.
  • the associated brake calliper is not shown but would be either solidly or floatingly connected, such as provided by the invention described in GB2436672, to an unsprung component such as housing 68R.
  • Brake torque can be either resisted by upright 28 or by the calliper being rotationally and concentrically connected around wheel 2's rotational axis via a plate or similar that has an arm directly or indirectly connected, the latter via a joint, to housing 68R or indirectly connected via a joint to a sprung component such as upper frame 63R.
  • the brake disc can instead be of a conventional type, and thereby located by its inner circumference or inner flange, either connected to or integral with stub axle assembly 42 or associated sub axle. Alternatively a drum brake can be used. Where wheel 2 has an electric drive then braking can be applied electrically by reclaiming the kinetic energy or/and reverse drive. Where wheel 2 has a hydraulic drive then braking can be applied be restricting the flow of hydraulic fluid.
  • a hydraulic drive would not need a separate brake as a complete shut off of the hydraulic flow would prevent the wheel from turning.
  • an electric drive may or may not need a separate brake.
  • housing 68R slides along guides 67R which are circular in cross section.
  • guides 67R could have any other feasible cross sectional shape that is uniform along their length.
  • a single suitably designed non circular in cross section guide 67R could be used instead.
  • the guides 67R can also be, more than two, of different cross sectional shapes and areas from each other, positioned at different distances from the central plane of wheel 2 to each other.
  • lower frame 69R provides positional consistency between the lower ends of guides 67R and kingpin 43.
  • the rotational joint between kingpin 43 and upper frame 63R could be designed to negate the need for the rotational joint between lower frame 69R and the lower end of kingpin 43. This would then allow kingpin 43 to be much shorter than implied in the Figures which can save weight and reduces manufacturing costs.
  • lower frame 69R would provide positional consistency just between guides 67R. However, if each guide 67R had sufficient rigidity and was adequately secured to upper frame 63R the need for lower frame 69R could be negated which saves weight and reduces manufacturing costs and reduces steering momentum.
  • guides 67R may benefit from a stop to prevent housing 68L from sliding off.
  • the said stops could be integral (or only removable by the manufacturer or similar) with their guide 67R or detachable with a, screw, circlip, split pin, cotter pin, etc fixing.
  • Kingpin 43 does not have to be parallel to guides 67R because an angular relationship between kingpin 43 and guides 67R is also envisaged. Therefore in relation to Figures 34, 35, 40, the top end of kingpin 43 can be further away from guides 67R than the lower end of kingpin 43 or visa-versa, or/and in relation to Figures 33, 36, 39, the top end of kingpin 43 can be further away from guides 67R than the lower end of kingpin 43 or visa- versa.
  • Rotational axis 35 between lower swing arm 11 and kingpin 43 does not have to be perpendicular. Instead the angular relationship can be greater than ninety degrees or less than ninety degrees. With reference to Figure 39, the steering axis defined by kingpin 43 does not have to be parallel to guides 67R or/and perpendicular to the ground.
  • More than one resilience means 64R may be used for wheel 2 whereby the associated suspension spring's wire cross sectional area can be smaller than the wire cross sectional area of a single resilience means 64R.
  • the wire cross sectional area of the two springs can be approximately half that of the one spring for the same overall spring rate when other associated parameters are the same. The smaller the wire's cross sectional area the less is the room taken up by the associated spring's coils. This then allows each spring's movement to be greater than for a single spring for a given free spring length which can allow the angles of angle/ 1 angle/2 angle/3 to be reduced or maybe negated or/and to provide more suspension movement.
  • a similar advantage can be achieved by two or more springs being inside of each other within one resilience/damping means 64R.
  • Multiple resilience/damping means 64R can each be smaller in diameter than a single resilience/damping means 64R which may allow them to be positioned in places too small for a single larger resilience/damping means 64R. Unlike a single
  • resilience/damping means 64R multiple resilience/damping means 64R do not have to be centrally aligned with either wheel 2's rotational axis or/and guides 67R to minimise the twisting force of housing 68R acting on guides 67R caused by suspension deflections.
  • Multiple resilience/damping means 64R can employ different spring wire cross sectional areas from each other. Where multiple resilience/damping means 64R are used one or more may just contain the suspension spring/s and one or more may just contain the damping mechanism.
  • resilience/damping means 64R was instead just a suspension damper then a compression or bump suspension coil spring could encompass one or each guide 67R between housing 68R and upper frame 63R with or without an extension or rebound suspension coil spring encompassing one or each guide 67R between housing 68R and lower frame 69R, or said stop.
  • the resilience/damping means associated with wheel 1 and wheel 2 would turn with the steering and contribute to the steering momentum as well as contributing to the unsprung mass.
  • balance beam 51 the role of balance beam 51 is to control the up and down relationship between lower swing arms 10 and 11 as the associated vehicle leans, and also to balance the force from the suspension deflections from one side of the associated vehicle to the other side.
  • An anti tilt brake can act between balance beam 51 and a sprung part of the associated vehicle in the form of a hydraulic cylinder where an automatically or/and manually activated valve intermittently interrupts the hydraulic flow in the said hydraulic cylinder to further slow down destabilising events like lowsides and highsides in addition to that described in A or to stop the said hydraulic flow for parking.
  • the anti tilt brake can comprise a disc and brake calliper or brake drum and brake shoes acting between balance beam 51 and a sprung part of the associated vehicle.
  • the damping resilience means 64R/L are replaced in any of their previously described positions and by any of their previously described multiples with hydraulic displacement devices (displacement device).
  • Figure 41 shows a well known hydraulic displacement device 71L.
  • Figure 42 shows Figure 41 's cross section BB of cylinder 72L to reveal piston 73L connected to piston rod 74L.
  • Within cylinder 72L and above piston 73L is cavity 75L which contains hydraulic fluid.
  • the volume of cavity 75L varies according to the up and down movement of piston 73L in cylinder 72L in response to suspension deflections moving piston rod 74L up and down.
  • Opening 76L in cylinder 72L allows hydraulic fluid to be displaced to and from cavity 75L via hydraulic line 77L in response to the associated suspension deflections.
  • Hydraulic line 77L shown in Figure 41 in truncated form, interconnects the displacement devices 71L R with their associated resilience means.
  • Figure 43 shows Figure 42 but with the addition of damping device 78L fixed in cavity 75L to control the bump and rebound, initiated by the associated suspension deflections activating the resilience means, via through- holes in damping device 78 L that may or may not have valves.
  • a piston seal or/and close manufacturing tolerances ensures that the hydraulic fluid does not bypass piston 73L, and typically a linear bearing will interface between piston rod 74L and cylinder 72L.
  • resilience/damping means 64R/L each displacement device 71L/R is held in position by location fasteners 66R/L or/and ball joints 70R/L.
  • resilience means 79R/L are remotely positioned from their associated displacement devices 71L/R as shown in Figure 44.
  • resilience means 79R/L could be connected to any sprung part of the associated vehicle.
  • steering momentum resilience means 79R/L can also be remotely positioned from any steering components.
  • resilience means 79R/L are by the well known method of compressed gas contained in a spherical chamber wherein the compressed gas is separated from the hydraulic fluid by a membrane, bladder, sealed piston, etc.
  • the combined force from the compressed gas in resilience means 79R/L via hydraulic lines 77R/L is equally divided between their associated displacement devices 71R/L by balance beam 51, so that suspension deflections on say the right side of the associated vehicle are deterred from causing the associated vehicle to tilt to the left by the equal pressure in displacement device 71L on the left side of the same vehicle, and visa versa.
  • Embodiment2/l employs the same balance beam 51 arrangement as embodiment 1/1 and can incorporate the same anti tilt brake as embodiment 1/1.
  • the damping means 78L is positioned within cylinder 72L.
  • the damping means could alternatively be positioned along hydraulic lines 77R/L or within resilience means 79R L or combinations thereof.
  • single hydraulic lines 77R/L are shown but alternatively two or more hydraulic lines can be used. Where one or more hydraulic line is used for bump and one or more hydraulic line is used for rebound the bore diameters of the hydraulic lines for the bump and rebound can be different from each other to provide different flow rates which may negate the need for a separate suspension damper.
  • spherical resilience means 79R/L are shown but any known shape of a similar functioning resilience means can be used and not exclusive to just one resilience means per displacement device.
  • one compressed gas resilience means per set of paired wheels can be used where the compressed gas fills a cavity in a cylinder between two, pistons, bladders, or diaphragms, etc.
  • the hydraulic fluid from the displacement device/s associated with one wheel of a pair acts on one of the said pistons etc and the hydraulic fluid from the displacement device/s associated with the other wheel of the same pair acts on the other said piston etc.
  • the resilience means is via compressed gas a separate accumulator may not be required to compensate for a change in the volume of the hydraulic fluid due to a change in the said fluid's temperature.
  • a closed loop hydraulic system is shown but an open loop system can be used with a recirculation pump, oil tank, and associated valves.
  • a suspension brake could be added via a valve interrupting the hydraulic fluid flow along either or both the hydraulic lines 77R/L.
  • resilience means 79R/L can be replaced with hydraulic cylinders akin to 72L to form a master/slave cylinder arrangement to activate a mechanical resilience means such as a coil or leaf spring etc.
  • Embodiment2/l can make a similar improvement in traction as embodiment 1/1 by comparison to the prior art because embodiment2/l has a similar number of unsprung components as embodiment 1/1.
  • FIG. 45 A third embodiment of the first variation of the invention (embodiments/ 1) will now be described with reference to Figure 45.
  • stub axle assembly 40 is connected to bridging portion 80L in the same way as to housing 68L.
  • Lower frame 169L is rigidly connected to, or is integral with, the lower end of bridging portion 80L and upper frame 163L is rigidly connected to, or is integral with, the upper end of bridging portion 80L.
  • Upright 122 is rigidly connected to, or is integral with, upper frame 163L. Upright 122 connects to upper swing arm 24 in the same way as upright 22.
  • the king pin for steering is provided by piston rod 174L which rotates between upper frame 163L and lower frame 169L via suitable bearings.
  • piston 173L Fitted to piston rod 174L is piston 173L which can slide up and down the inside of cylinder 172L. Cylinder 172L is connected to swing arm 10 via rotational joint 34. Above piston 173L is cavity 175L which contains hydraulic fluid. The hydraulic fluid can enter or exit cavity 175L via opening 176L to and from resilience means 79L in the same way as the hydraulic fluid in displacement device 71L via opening 76L in Figure 44.
  • Embodiment3/l is a more compact design than the previous embodiments but by comparison the following components, lower frame 169L, upper frame 163L, upright 122, upper swing arm 24, steering rod 44, joint 45, become unsprung and thereby contribute to the unsprung mass.
  • embodiment3/l has far less unsprung components than the prior art which improves traction. Furthermore, the unsprung components in embodiment3/l do not include the comparatively heavy swing arms 10 and 11. Consequently the overall unsprung mass of embodiment3/l would not have to be much more than in embodiment 1/1 and embodiment2/l.
  • Embodiment/ 1 includes any practical variations associated with embodimentl/1 and embodiment2/l.
  • coil springs could be encompass piston rod 174L between displacement device 171L and upper frame 163L and lower frame 169L.
  • Each component indentified by L for the left side of the associated vehicle has a counterpart component for the right side.
  • the suspension movement occurs without axis 58 of balance beam 51 needing to move in relation to the sprung mass.
  • axis 58 is fixed to the sprung mass.
  • swing arms 10 and 11 are sprung components, i.e. supported by the resilience means.
  • space within the associated wheel rim is limited.
  • angle 1 angle2 and angle3 are ways where the use of the limited space within the associated wheel rim can be maximised.
  • Another way of using the said limited space is to limit the movement of the suspension within the said limited space with the remaining suspension movement being provided elsewhere.
  • a first embodiment of the second variation of the invention delivers some of the required suspension movement by any of the means described in the previous embodiments, hereinafter referred to as the primary suspension movement, combined with the rest of the required suspension movement being delivered by the method shown in Figure 26 where a resilience means 60 acts on carrier 57 which linearly slides on guides 59, hereinafter referred to the secondary suspension movement.
  • a resilience means 60 acts on carrier 57 which linearly slides on guides 59, hereinafter referred to the secondary suspension movement.
  • axis 58 in embodimentl/2 is not fixed to the sprung mass and swing arms 10 and 11 are sprung in relation to the primary suspension movement but unsprung in relation to the secondary suspension movement.
  • axis 58 can rotate about an axis that is parallel to first axis 14 or/and second axis 15 in a similar way as disclosed in EP0606191.
  • Side loadings on guide rods cause friction which has to be overcome before any suspension movement can occur. Thereby such side loading has a similar undesirable effect as increasing the unsprung weight.
  • the primary suspension movement is provided by a lever arm that causes the associated wheel to move through a curved path according to the length of the said lever arm when the primary suspension operates.
  • This differs from the previous embodiments where the suspension movement causes the associated wheel to move in a linear path.
  • the radially moving end of the lever arm would hold the stub axle assembly which may or may not include part of the steering mechanism such as a king pin.
  • An advantage of a lever arm is the almost complete absence of friction from side loadings which allows the suspension to work just as well under braking as when not braking.
  • Figures 46 to 61 show further developments of the invention.
  • Figure 47 is view CI of Figure 46.
  • a single guide 167R having a circular cross sectional area is possible when combined with torque arm assembly 81R to prevent housing 168R from rotating about steering axis 31 independently from the steering means controlled by the rider and without limiting the linear suspension movement of housing 168R on guide 167R.
  • torque arm assembly 81R comprises at least two links.
  • torque arm assembly 81R comprises upper link 82R which is rotationally connected at 83R to a sprung part of the vehicle which as a non limiting example in Figure 46 is bracket 84R extended from and fixed to upper frame 163R.
  • the other end of upper link 82R is rotationally connected at 85R to an end of lower link 86R.
  • the other end of lower link 86R is rotationally connected at 87R to a bracket 88R that is extended from and fixed to housing 168R.
  • at least two of the three rotational connections 83R, 85R, 87R would each have a single axis of rotation.
  • Torque arm assembly 81R transfers steering torque to housing 168R.
  • guide 167R is rotationally connected to lower frame 169R and fixed to upper frame 163R which is linked to the steering means controlled by the rider.
  • guide 1 7R could be fixed to lower frame 169R and rotationally connected to upper frame 163R.
  • guide 167R could be rotationally connected to the lower and upper frames 169R and 163R respectively.
  • Stub axle assembly 42 is rotationally connected to housing 168R.
  • Resilience/damping means 64R is at one end rotationally connected to bracket 89R extended from and fixed to housing 168R.
  • the other end of resilience/damping means 64R is rotationally connected to bracket 90R extended from and fixed to upper frame 1 3R.
  • steering movements can be imparted by a suitably designed
  • resilience/damping means without the need for a torque arm assembly like 81R.
  • steering torque imparted by the vehicle's operator to upper frame 163R could be transferred to housing 168R by just the interconnecting resilience/damping means 64R to negate the need for torque arm assembly 81R.
  • the said suitably designed resilient/damping means can be of a through-piston rod design as per residual/damping means 371R in Figure 56 to cope with the steering torque.
  • the steering torque can be shared between resilience/damping means 64R and torque arm assembly 81R.
  • Brake torque can be transferred from the associated brake (not shown) to housing 168R by an interconnecting bracket (not shown).
  • cylinder 272R of the resilience/damping means has a piston rod 274R which also provides linear guidance for the suspension movement.
  • Piston rod 274R is fixed to boss 92R.
  • Stub axle 93R is fixed to boss 92R.
  • Bracket 188R is fixed to boss 92R.
  • Cylinder 272R is fixed to upper frame 263R which is linked to the steering means controlled by the rider. To allow steering movements, cylinder 272R rotates in knuckle joint 94R about steering axis 31. Thereby cylinder 272R also acts as a kingpin.
  • Figure 50 is view E of Figure 49.
  • knuckle joint 94R rotates about axis 27 of rotational connection 35.
  • steering axis 31 is not confined to a fixed angular relationship with swing arm 11 as previously discussed under G.
  • the third rotational axis 27 of rotational connection 35 forms a corner of the parallelogram or quadrilateral that was also previously discussed under G.
  • piston rod 274R has a circular cross sectional area torque arm assembly 81R from Figures 46, 47, 48, can be used to prevent piston rod 274R from rotating within cylinder 272R to alter the steering independently from the steering means controlled by the rider and without limiting the linear suspension movement of piston rod 274R in cylinder 272R.
  • Torque arm assembly 81R is not required if piston rod 274R is non circular in cross section and is instead confined solely to linear movement by a matching female part that is fixed to cylinder 272R.
  • Brake torque can be transferred from the associated brake (not shown) to boss 92R by an interconnecting bracket (not shown).
  • the rotational axes of rotational connections 83R, 85R, 87R, are perpendicular to the suspension movement. When the said axes are also parallel to the rotational axis of their associated wheel the associated torque arm assembly can be used to limit brake dive.
  • Brake dive When braking, a resultant force proportional to the rate of deceleration combined with gravity acts closer to the front of the vehicle than just gravity alone when not decelerating. Consequently the front suspension is compressed more when braking, commonly known as brake dive, than when not braking. Brake dive can be reduced or eliminated as the manufacturer desires by an anti dive mechanism. There are several known anti dive mechanisms which are often built into the resilience/damping means. Any of the known anti dive mechanisms can be incorporated into the various embodiments of the invention.
  • Anti dive can be achieved by harnessing the braking torque. For instance, and with reference to Figures 46 to 51, by attaching the brake (not shown) to lower link 86R of torque arm assembly 81R where rotational connection 87R of lower link 86R has the same rotational axis as the associated brake drum or disc.
  • torque arm assembly 181R comprises an upper link 182R which has one end rotationally connected at 183R to a bracket fixed to cylinder 272R and the other end rotationally connected at 185R to one end of lower link 186R.
  • the other end of lower link 186R is rotationally connected at 187R about the same longitudinal axis as stub axle 93R.
  • the Figure 52 arrangement would transfer the braking torque, indicated by the arrow, to a sprung part of the vehicle via torque arm assembly 181R. Consequently the braking torque would try to increase the angle between rotational connections 183R, 185R, 187R, which in turn would extend the front suspension by extending piston rod 274R from cylinder 272R.
  • the desired amount or range of anti dive can be achieved by calculating the length and positions of the lower link 186R and upper link 182R in relation to each other and the resulting force generated by the deceleration caused by braking.
  • Torque arm assembly 81R in Figures 46 to 51 can be combined with at least one torsion spring where one leg of the torsion spring would bare on the upper link 82R and the other leg of the torsion spring would bare on the lower link 86R.
  • the said torsion spring could provide all of the required suspension resilience means, whereby piston rod 274R becomes a guide, or just some of the required suspension resilience.
  • a torsion spring could act between the upper link 82R and cylinder 272R or/and between the lower link 86R and bracket 188R, or at all of the rotational joints 83R, 85R, 87R.
  • At least one of the rotational connections 83R, 85R, 87R, of torque arm assembly 81R could include a friction suspension damper.
  • the associated friction in suspension compression can differ from the associated friction in suspension extension.
  • the said difference could be provided by each interfacing friction surface comprised from directional friction differences.
  • the said directional friction difference is provided by the associated friction surface comprised from, for instance, abrasions that are not concentric to themselves but are uniformly aligned with each other. When such interfacing friction surfaces have the said uniform alignment but in opposite directions to each other it would generate more friction in one direction that the other direction.
  • the said interfacing friction surfaces could be mounted to discs or quadrants.
  • Torque arm assembly 181R from Figure 52 can also be used with any of the embodiments of the invention shown in Figures 32 to 48 to transfer steering torque, and to provide anti dive if the brake was attached to lower arm 186.
  • any of the shafts that interface with a linear bearing or linear seal can be protected from the ingress of contaminants by a gaiter or bellows that are suitably flexible and suitably resilient to said containments.
  • Figure 53 shows king pin 95R rotationally connected to knuckle 94R to provide steering movement.
  • king pin 95R is fixed to upper frame 263R which is connected to the steering means operated by the rider.
  • the other end of king pin 95R is rotationally connected at 97R to one end of leading arm 96R.
  • the other end of leading arm 96R is rotationally connected at 98R to one end of resilience/damping means 64R.
  • the other end of resilience/damping means 64R is rotationally connected at 99R to king pin 95R via a bracket fixed to king pin 95R.
  • Stub axle 193R is connected to leading arm 96R.
  • Figure 54 shows Figure 53 from the direction of C2.
  • Figure 55 shows Figure 53 but with the suspension more compressed to show the movement of leading arm 96R about rotational connection 97R.
  • anti dive can be achieved by mounting the brake (not shown) directly to leading arm 96R.
  • the clockwise braking torque applied to leading arm 96R about rotational connection 97R would try to lengthen resilience/damping means 64R but the associated brake dive would try to shorten resilient/damping means 64R by applying an anti clockwise torque to leading arm 96R about rotational connection 97R.
  • the clockwise brake torque counters the anti clockwise brake dive torque.
  • the amount or range of anti dive depends on the position of rotational connection 97R from stub axle 193R in relation to the resulting force generated by braking deceleration which can be calculated by the manufacturer.
  • king pin 95R is dog legged but alternatively king pin 95R can be straight.
  • Stub axle 193R can be either lower than rotational axis 97R as depicted in Figure 53 or share the same horizontal plane as rotational axis 97R, or be above rotational axis 97R.
  • Figures 57 to 60 show the trailing arm alternative to the previously described leading arm arrangement in Figures 53 to 56.
  • the components in Figures 57 to 60 are the same as those in Figures 53 to 56 and share the same component numbers. However, the components in Figures 57 to 60 that are specific to the trailing arm are prefixed with T to identify trailing arm.
  • Figure 58 shows Figure 57 from the direction of El.
  • Figure 59 shows Figure 57 but with the suspension more compressed to show the movement of trailing arm T96R about rotational connection T97R.
  • the brake (not shown) is mounted to brake plate 1 0R as in Figure 56 to isolate the brake torque from trailing arm T196R.
  • resilient/damping means 371R is mounted in the same way to trailing arm T196R as resilience/damping means 371R is to leading arm 196R in Figure 56.
  • the brake (not shown) is connected to lower link 286R which has one end rotationally connected at 287R, to either stub axle T393R or trailing arm T296R, about the longitudinal axis of stub axle T393R.
  • the other end of lower link 286R is rotationally connected at 285R to one end of upper link 282R.
  • the other end of upper link 282R is rotationally connected at 283R to king pin T195R.
  • Stub axle T393R can be fixed to or rotationally connected to trailing arm T296R. In the latter case lower link 286R can be fixed to stub axle T393R.
  • Trailing arm T296R is rotationally connected at T197R to king pin T195R.
  • resilient/damping means T164R is rotationally connected at T198R to trailing arm T296R on the opposite side of rotational connection T197R from stub axle T393R.
  • the other end of resilience/damping means T164R is rotationally connected at T199R to kingpin T195R via a bracket fixed to king pin T195R.
  • lower link 286 can also provide the associated brake with side float by rotational joint 283 and 287 both being spherical bearings,
  • side float can also be provided by the brake plate 100 by being rotationally connected to piston rod 374 or by piston rod 374 being rotationally connected to lever arm 196 via a spherical bearing and where piston rod 374 can also rotate in cylinder 372.
  • EP1918187 has a front suspension and steering system akin to vehicle 1.
  • the front suspension is integrally positioned with the two components 38 that are equivalent to vehiclel 's struts or upright which herein are numbered 22 and 28.
  • the three previously described characteristics (1.1) (2.1) (3.1) of the first variation of the invention are equally true for EP1918187.
  • the said one embodiment of EP1918187 achieves a similar unsprung component reduction as the invention by comparison to vehicle 1.
  • EP1918187 does not give any reason for the said one embodiment's suspension position. Furthermore, the subject commonly known as the unsprung mass is not referred to at all in any form despite the applicant, Piaggio, being an established vehicle manufacturer and thereby surely knowing about the subject of unsprung mass. Evidently the minimising of unsprung mass, i.e. the previously described First Problem, was not a goal of EP1918187. Although the said one embodiment in EP1918187 and the invention both reduce the unsprung components by comparison to vehicle 1, it is only the invention that is aimed at overcoming the First Problem of minimising the unsprung mass.
  • the invention minimises the unsprung mass by also minimising the size of the unsprung components. This, results in an intersection between the unsprung and sprung mass being as close as possible to the associated wheel's rotational axis in the first and second variations of the invention.
  • EP1918187 with reference to Figure 7a therein has the intersection 76 between the unsprung components 38 and the sprung components (the other numbered components) laying outside of the said circular plane.
  • EP 1918187 does not indentify the provision of sufficient suspension travel within any limited space as a practical challenge to overcome.
  • the anti dive mechanism provided by torque arm assembliesl81 and 281 can be applied to the front suspension of a car which would make the ride more comfortable by reducing or by the total absence of brake dive in all braking situations. Without an anti dive mechanism a driver has to gradually ease off braking as the car slows until the very last motion is arrested by just the vehicle's rolling resistance in order to arrest the vehicle without brake dive. Not only is this practice difficult to master, usually only by best chauffeurs, it is not possible on a downwards slope or when having to brake rapidly or with automatic transmission.
  • the invention would be interconnected with a single balance beam or hydragas equivalent for the front paired wheels, and where paired back wheels are also used the associated vehicle would employ another balance beam or another hydragas equivalent for the rear paired wheels, thus requiring two balance beams or two hydragas gas equivalents or one of each.
  • the balance beam that interconnects with the invention can also be simultaneously used for paired rear wheels when the interconnections from the single balance beam to one pair of wheels crosses over.
  • the balance beam that interconnects with the invention can also be simultaneously used for paired rear wheels when the interconnections from the single balance beam to one pair of wheels crosses over.
  • one pair of wheels for instance at the front, are connected to the balance beam as herein described but with the right rear wheel connected to the left end of the balance beam and with the left rear wheel connected to the right end of the balance beam. So that the right front wheel and left rear wheel are connected to the right end of the balance beam and the left front wheel and rear right wheel are connected to the left end of the balance beam.
  • the single balance beam can rotate about a static axis.
  • the single balance beam axis can rotate about a linear moving balance beam axis as per vehicle 1 or rotate about a radially moving balance beam axis as per EP0606191.
  • the resilience means of the secondary suspension movement pushes the balance beam axis, or carrier thereof, towards a stop or where separate resilience means of the secondary suspension movement push the balance beam axis, or carrier thereof, in opposite directions to each other so that the movement of the balance beam axis is determined by the opposing resilience means of the secondary suspension.
  • Secondary suspension movement can also be achieved with a single static balance beam axis where the connection from each primary suspension movement includes means for the secondary suspension movement between the associated swing arms and balance beam's ends and where the rear suspension either employs the same primary and secondary suspension movement arrangement or where all of the rear suspension movement is provided by an arrangement akin to the associated secondary suspension movement.
  • the single balance beam cross over arrangement can take many forms and can be used without the invention being included.
  • Figure 62 shows the front right swing arm 11 and left rear swing arm 12 linked to the right end 56 of balance beam 51 and the front left swing arm 10 and rear right swing arm 13 linked to the left end 52 of balance beam 51.
  • the single balance beam rotates about axis 58 which is statically fixed in relation the vehicle's main frame and where connecting links 49, 54, 490, 540, are all under compression.
  • one or both connecting links may employ a portion that is offset from one end or from both ends to bridge the other connecting link.
  • one connecting link 490 or 540 can provide clearance for the other by employing opposing bridging portions as shown by, for instance, connecting link 540 in Figure 62.
  • connecting link 490 and 540 incorporate damping/resilience means 601 and 602 respectively.
  • Figure 62 shows the first variation of the invention where at the vehicle's front end the intersection between the unsprung mass and sprung mass is located within the circular plane dimensionally defined by the associated tyre's diameter and where said circular plane is positionally defined by its parallel centre line being aligned with the rotational axis of the associate wheel, and axis 58 of balance beam 51 is static in relation the associated vehicle's main frame.
  • Figure 63 shows Figure 62 but with the addition of a secondary suspension as per the second variation of the invention.
  • the secondary suspension can instead be accommodated by axis 58 of balance beam 51 moving linearly on a carrier as per Figure 24 or radially on a carrier as previously discussed where the secondary suspension can act on one side of the said carriers pushing them towards a stop or on opposite sides of the said carriers to resist their linear or radial movement in both directions.
  • Figure 64 shows a variation of Figure 62 where lug 480 is in a fixed angular relationship with rear swing arm 12 but projected downwards from transverse axis 15.
  • lug 530 is in a fixed angular relationship with rear swing arm 13 but is projected downwards from transverse axis 15.
  • the left connecting link 490 can be rotationally connected between lug 480 and the left end of balance beam 51 via rotational connections 500 and 52 respectively and the right connecting link 540 can be rotationally connected between lug 530 and the right end of balance beam 51 via rotational connections 550 and 56 respectively.
  • connecting links 540 and 490 do not cross over each other. With this arrangement connecting links 540 and 490 are under tension.
  • connecting links 540 and 490 incorporate resilience means 602 and 601 respectively 602 and 601 will work under tension rather than in compression.
  • transverse axes 14 and 15 are on the same horizontal plane as in Figure 62. Therefore, axis 58 of balance beam 51 can be perpendicular to connecting links 49, 54, 490, 540, if balance bean 51 was angled from the vertical with connecting link 49 being parallel to connecting link 490 and with connecting link 54 being parallel to connecting link 540.
  • Figure 65 shows Figure 64 but with the addition of the secondary suspension from Figure 63.
  • the secondary suspension can instead be accommodated by axis 58 of balance beam 51 moving linearly on a carrier as per Figure 24 or radially on a carrier as previously discussed where the secondary suspension acts on one side or on opposite sides of the said carriers to resist their linear or radial movement.
  • Suspension biasing can be achieved for the embodiments shown in Figures 64 and 65 by axis 58 of balance beam 51 being offset from the ends 52 and 56 of balance beam 51 as previously discussed under F and as shown by balance beam 151 in Figure 66. Although this could be applied to the embodiments shown in Figures 62 and 63 it would result in the suspension at one end of the vehicle being biased in an opposite direction to the other end of the vehicle. However suspension biasing for the embodiments shown in Figures 62 and 63 can be achieved in the same direction at both ends of the vehicle by making appropriate non perpendicular alignments between, each swing arm and associated lug, each lug and associated connecting link, each connecting link and associated end of the balance, or combinations thereof.
  • the arrangement of the rear connecting links 540 and 490 can swap positions with the arrangement of the front connecting links 54 and 49.
  • the rear suspension can employ the suspension mechanism for either the first or second variations of the invention.
  • a suspension system for the paired front wheels of motorcycle type vehicles where to cause the paired wheel's track to increase whilst cornering and to ensure the trail and rake of each wheel of the pair always remain within workable parameters each wheel of the pair has its own movable quadrilateral or parallelogram mechanism comprised from (i) a transversely pivoted swing arm (ii) a transversely pivoted upper swing arm (iii) a steerable strut that is jointed to the free ends of the swing arm and upper swing arm (iv) a main frame or prime move to which the said transverse pivots are connected; where to ensure suspension deflections do not unduly tilt the associated vehicle sideways there are also means to disperse the suspension force between paired wheels, characterised by the suspension being able to move independently of the swing arms.
  • a suspension system according to any of the preceding clauses where a piston rod in a hydraulic displacement device or/and damping means also acts as a kingpin for steering.
  • the resilience means is by one or more, coil spring, leaf spring, gas spring, electromagnetic spring, rubber or similar type material spring, pneumatic spring, or combination thereof.
  • the means to disperse the suspension force between paired wheels is by a balance beam or similar mechanism.
  • a suspension system for a motorcycle type vehicle having two front wheels, two stub axles one for each front wheel, two swing arms one for each stub axle, each swing arm confined for radial movement about a transverse axis of the vehicle, means to constrain each swing arm relative to each other for equal and opposite movements as the vehicle lean, each stub axle supporting a suspension means, the suspension means supporting each swing arm, the suspension movement of the suspension means being independent of the swing arms, characterised by the suspension movement being within the internal space of the associated front wheel's rim.
  • each suspension means is provided by a hydraulic cylinder that displaces hydraulic fluid to and from a remotely positioned resilient means via one or more hydraulic lines.
  • a motorcycle type vehicle having, two rear wheels, a rear swing arm for each rear wheel, each rear swing arm confined for radial movement about a transverse axis of the vehicle, where each rear swing arm is constrained for equal and opposite movements relative to each other as the vehicle leans by being interconnected by the same balance beam or hydragas equivalent that interconnects each front swing arm.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Transportation (AREA)
  • Axle Suspensions And Sidecars For Cycles (AREA)
EP12726822.5A 2011-05-20 2012-05-08 Vorderes aufhängungssystem Withdrawn EP2709862A2 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
GBGB1108598.2A GB201108598D0 (en) 2011-05-20 2011-05-20 Front suspension system
PCT/GB2012/000418 WO2012160323A2 (en) 2011-05-20 2012-05-08 Front suspension system

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EP2709862A2 true EP2709862A2 (de) 2014-03-26

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WO2012160323A2 (en) 2012-11-29
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