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Motorcycle Handling and Chassis Design the art and science

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The book "Motorcycle Handling and Chassis Design" presents an updated and comprehensive exploration of motorcycle chassis and handling, building on foundational knowledge from the original 1984 publication. It emphasizes the evolution and improvement in motorcycle handling over decades, detailing advancements in chassis design, suspension technologies, and the critical role of tyres and aerodynamics. The text aims to provide readers with not only the theoretical background required for understanding motorcycle dynamics but also practical insights for design, modifications, and setup, encouraging an engaging and thorough study of the subject.

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10 Structural considerations
10 Structural considerations
9 Squat and dive
9 Squat and dive
12 Braking
12 Braking
Roll is probably the most familiar of the three and is the most obvious motion that occurs when we lean the bike over for cornering. Fig. 1.1 shows the roll axis passing through the CoG. However, as we shall see later the location of this axis depends on our frame of reference. If we are to study the behaviour of any type of vehicle we first need to consider just how it can move. The linear motions are easy to visualize, firstly the machine can move in a forward direction and the engine and brakes are responsible for controlling this. Road undulations and hills cause motion in a vertical direction and sidewinds can result in sideways movement. It is the angular motions that are somewhat less familiar to most people. The overall angular movements can be completely described by considering the motions about three separate axis. These axis are at right angles to one another and are known as roll, pitch and yaw.
Roll is probably the most familiar of the three and is the most obvious motion that occurs when we lean the bike over for cornering. Fig. 1.1 shows the roll axis passing through the CoG. However, as we shall see later the location of this axis depends on our frame of reference. If we are to study the behaviour of any type of vehicle we first need to consider just how it can move. The linear motions are easy to visualize, firstly the machine can move in a forward direction and the engine and brakes are responsible for controlling this. Road undulations and hills cause motion in a vertical direction and sidewinds can result in sideways movement. It is the angular motions that are somewhat less familiar to most people. The overall angular movements can be completely described by considering the motions about three separate axis. These axis are at right angles to one another and are known as roll, pitch and yaw.
Cradle frame, successor to the diamond bicycle pattern . The cradle tubes are extended rearward to the wheel spindle lugs. In the duplex cradle frame the cradle tubes are also extended upward to the steering head. bottom ends of the single front and seat tubes were spaced farther apart and rigidly connected by « brazed-in engine cradle, from the rear of which the tubes reached upward to the wheel-spindle lugs.
Cradle frame, successor to the diamond bicycle pattern . The cradle tubes are extended rearward to the wheel spindle lugs. In the duplex cradle frame the cradle tubes are also extended upward to the steering head. bottom ends of the single front and seat tubes were spaced farther apart and rigidly connected by « brazed-in engine cradle, from the rear of which the tubes reached upward to the wheel-spindle lugs.
In the design of these early frames, torsional and lateral stiffness seem to have been given a low priority. However, there were some commendable efforts between the wars to ensure the all-important torsional and lateral stiffness through triangulation of the frame structure. In the Cotton, the four long tubes connecting the steering head directly to the rear spindle lugs were triangulated in both plan view and elevation, and the machine was renowned for its excellent steering.
In the design of these early frames, torsional and lateral stiffness seem to have been given a low priority. However, there were some commendable efforts between the wars to ensure the all-important torsional and lateral stiffness through triangulation of the frame structure. In the Cotton, the four long tubes connecting the steering head directly to the rear spindle lugs were triangulated in both plan view and elevation, and the machine was renowned for its excellent steering.
A 1960s. picture of the author racing a Greeves Silverstone with cast aluminium H-section front down member, incorporating the steering head. The under-frame was an open channel section fabricated from steel sheet, a bent steel tube backbone completed the frame loop.
A 1960s. picture of the author racing a Greeves Silverstone with cast aluminium H-section front down member, incorporating the steering head. The under-frame was an open channel section fabricated from steel sheet, a bent steel tube backbone completed the frame loop.
The legendary Norton “featherbed” frame. The crossing over of the tubes at the steering head facilitated the use of a flat-bottom tank but impaired stiffness. A head steady was used which connected the steering head and top of the engine to overcome this. This shows just how easy it was to fit plunger springing in place of the then standard unsprung rigid rear end. This design reduced the amount of re-tooling required. A revolution was started in 1950 when the works Norton racers were supported by the McCandless Bros designed “featherbed” frame. It is hard to over-estimate the influence that this design has had or subsequent chassis development.
The legendary Norton “featherbed” frame. The crossing over of the tubes at the steering head facilitated the use of a flat-bottom tank but impaired stiffness. A head steady was used which connected the steering head and top of the engine to overcome this. This shows just how easy it was to fit plunger springing in place of the then standard unsprung rigid rear end. This design reduced the amount of re-tooling required. A revolution was started in 1950 when the works Norton racers were supported by the McCandless Bros designed “featherbed” frame. It is hard to over-estimate the influence that this design has had or subsequent chassis development.
Fully triangulated at the front but only vertically at the rear, this Francis-Barnett frame could be easily repaired by renewing any of the bolted-in tubes. Triangulated fully at the rear but only laterally at the front, this early Scott frame relied on the engine for some Of its stiffness.
Fully triangulated at the front but only vertically at the rear, this Francis-Barnett frame could be easily repaired by renewing any of the bolted-in tubes. Triangulated fully at the rear but only laterally at the front, this early Scott frame relied on the engine for some Of its stiffness.
In this unsprung beam frame for a moped the open channel section of the rear fork arms is strengthened by a welded-in U-shape strip. Welded from left and right hand pressings, this NSU beam frame was shaped to accommodate pivoted fork rear springing and to support the engine at the top and rear. Clearly, however, a plain beam could not connect the steering head directly to the rear-wheel spindle as did the top four tubes in the Cotton frame. Hence it was bifurcated at the rear to accommodate the wheel, and the resulting open channel section of the two arms was closed by welding-in a U-shape strip to restore strength. Welding the beam from two halves in this way made it possible to incorporate any necessary curvature in a vertical plane. NSU used curved beam frames not only on their moped but also on their Max roadster, their 250 cc world championship-winning Sportmax catalogue racing single, the works racing 250 cc Rennmax twin and 125 cc Rennfox single.
In this unsprung beam frame for a moped the open channel section of the rear fork arms is strengthened by a welded-in U-shape strip. Welded from left and right hand pressings, this NSU beam frame was shaped to accommodate pivoted fork rear springing and to support the engine at the top and rear. Clearly, however, a plain beam could not connect the steering head directly to the rear-wheel spindle as did the top four tubes in the Cotton frame. Hence it was bifurcated at the rear to accommodate the wheel, and the resulting open channel section of the two arms was closed by welding-in a U-shape strip to restore strength. Welding the beam from two halves in this way made it possible to incorporate any necessary curvature in a vertical plane. NSU used curved beam frames not only on their moped but also on their Max roadster, their 250 cc world championship-winning Sportmax catalogue racing single, the works racing 250 cc Rennmax twin and 125 cc Rennfox single.
Winner of the world 250 cc championship in 1953, this NSU Rennmax twin had a curved beam frame welded from left and right halves (MCW) Pivoted-fork or swinging-arm rear springing removed the need to bifurcate the rear end of a frame beam because with this type of suspension it is the swing-arm pivot, not the wheel spindle, to which the steering head has to be stiffly connected. (Naturally, the swing-arm itself should continue the torsional and lateral stiffness back to the spindle.) In the NSU Max and racers, the pressed-steel beam was curved downward at the rear to make a direct connection from steering head to rear pivot.
Winner of the world 250 cc championship in 1953, this NSU Rennmax twin had a curved beam frame welded from left and right halves (MCW) Pivoted-fork or swinging-arm rear springing removed the need to bifurcate the rear end of a frame beam because with this type of suspension it is the swing-arm pivot, not the wheel spindle, to which the steering head has to be stiffly connected. (Naturally, the swing-arm itself should continue the torsional and lateral stiffness back to the spindle.) In the NSU Max and racers, the pressed-steel beam was curved downward at the rear to make a direct connection from steering head to rear pivot.
To make a direct connection between steering head and rear fork pivot, the tubular beam on this early 125 cc Honda GP racer was curved through some 90 degrees. (Salmond) The Ariel Leader (and Arrow) frame derived great stiffness from the large cross-sectional area of the beam, which enclosed the 11 litre fuel tank.
To make a direct connection between steering head and rear fork pivot, the tubular beam on this early 125 cc Honda GP racer was curved through some 90 degrees. (Salmond) The Ariel Leader (and Arrow) frame derived great stiffness from the large cross-sectional area of the beam, which enclosed the 11 litre fuel tank.
Note also the leading link forks with Greeves type rubber pivot bushes, which also act as the springs.
Note also the leading link forks with Greeves type rubber pivot bushes, which also act as the springs.
A substantially similar layout was used on Norton’s unfinished Moto Guzzi-inspired experimental 500 cc flat single in the mid-1950s. In that case the oil was contained in a 114 mm.-diameter main tube and a welded-on underslung box that also supported the crankcase, while the fork pivot was bolted between é light-alloy gearbox plate on the left and an aluminium casting on the engine. When a tall, bulky engine (such as a 1-litre twin-cam four abreast) has to be accommodated, an even bigger gap has to be spanned. A self-defeating scheme adopted by some frame builders was to bridge the gap with a pair of bolted-on light-alloy plates, which could make nonsense of the tube’s torsional stiffness, depending on detail design.
A substantially similar layout was used on Norton’s unfinished Moto Guzzi-inspired experimental 500 cc flat single in the mid-1950s. In that case the oil was contained in a 114 mm.-diameter main tube and a welded-on underslung box that also supported the crankcase, while the fork pivot was bolted between é light-alloy gearbox plate on the left and an aluminium casting on the engine. When a tall, bulky engine (such as a 1-litre twin-cam four abreast) has to be accommodated, an even bigger gap has to be spanned. A self-defeating scheme adopted by some frame builders was to bridge the gap with a pair of bolted-on light-alloy plates, which could make nonsense of the tube’s torsional stiffness, depending on detail design.
Note how the head stock is supported by the two side beams of this 1997 NSR 250 racing Honda. Frame material is aluminium alloy.
Note how the head stock is supported by the two side beams of this 1997 NSR 250 racing Honda. Frame material is aluminium alloy.
The 1920s Ner-a-Car chassis comprised left and right channel-section pressings, cross-braced front and rear. The engine was slung low and hub-centre steering was a standard feature.
The 1920s Ner-a-Car chassis comprised left and right channel-section pressings, cross-braced front and rear. The engine was slung low and hub-centre steering was a standard feature.
This Foale frame used a large backbone above the engine, with triangulated support for the swing-arm pivot.
This Foale frame used a large backbone above the engine, with triangulated support for the swing-arm pivot.
Aseries C Vincent-HRD Rapide with hydraulically damped Girdraulic front fork and twin rear shock absorbers. The rear swinging fork was triangulated and it pivoted from two aluminium plates bolted to the rear of the gearbox.
Aseries C Vincent-HRD Rapide with hydraulically damped Girdraulic front fork and twin rear shock absorbers. The rear swinging fork was triangulated and it pivoted from two aluminium plates bolted to the rear of the gearbox.
Girder, telescopic, leading-link and trailing-link front forks, showing the path travelled by the wheel spindle throughout suspension movement.
Girder, telescopic, leading-link and trailing-link front forks, showing the path travelled by the wheel spindle throughout suspension movement.
Manfred Schweiger’s 1929 Scott TT Replica, showing one of the earliest telescopic front forks, in which the bottom bushes were supported by diamond-shaped girders. The lug at the top of the sliders actuated a central spring. (Margrholz) Yet BMW were by no means first with a telescopic front fork. Several decades earlier Scott employec the principle, though without damping. Nevertheless, their fork was well engineered, with the sliders firmly braced by being brazed into a lug above the wheel (where they operated a central spring) and with the bottom bushes well supported, latterly by diamond-shaped girders.
Manfred Schweiger’s 1929 Scott TT Replica, showing one of the earliest telescopic front forks, in which the bottom bushes were supported by diamond-shaped girders. The lug at the top of the sliders actuated a central spring. (Margrholz) Yet BMW were by no means first with a telescopic front fork. Several decades earlier Scott employec the principle, though without damping. Nevertheless, their fork was well engineered, with the sliders firmly braced by being brazed into a lug above the wheel (where they operated a central spring) and with the bottom bushes well supported, latterly by diamond-shaped girders.
One of the most famous examples of leading link front forks — that of the 1956 version of the 500 cc Moto Guzzi V-eight grand-prix racer. The externally mounted suspension struts allowed easy setup. (MCW) Where cost had a lower priority — i.e. in grand-prix racing — or where a manufacturer of roadsters set more store by high quality than low price and sleek looks, some engineers rejected the telescopic fork for its dynamic and structural shortcomings and preferred the leading-link variety. These offered greater torsional and lateral stiffness, lower unsprung weight, better damping (through proprietary struts) and < choice of steering geometry between virtually constant trail and constant wheelbase, depending on the inclination of the links.
One of the most famous examples of leading link front forks — that of the 1956 version of the 500 cc Moto Guzzi V-eight grand-prix racer. The externally mounted suspension struts allowed easy setup. (MCW) Where cost had a lower priority — i.e. in grand-prix racing — or where a manufacturer of roadsters set more store by high quality than low price and sleek looks, some engineers rejected the telescopic fork for its dynamic and structural shortcomings and preferred the leading-link variety. These offered greater torsional and lateral stiffness, lower unsprung weight, better damping (through proprietary struts) and < choice of steering geometry between virtually constant trail and constant wheelbase, depending on the inclination of the links.
A variant of the leading-link fork was the Ernie Earles type, with long links combined into an integral fork pivoted behind the wheel. BMW standardized this arrangement for practically two decades between abandoning their original telescopic fork and introducing a long-travel version in 1969. Essentially though, their Earles-type fork was a compromise for solo and sidecar duty. Although its lack of sliding friction was a boon in sidecar cornering, its high steering inertia blunted steering response in a solo. As an aid to this dual purpose, BMW provided a choice of two pivot mounting locations so that trail could be easily set for solo or sidecar use. Basic layout of the Difazio system. The wheel hub spins on large diameter bearings, and the non-rotating axle is integral with the steering “king-pin”. A forward facing swinging fork connects the axle to the main chassis, whilst the upper frame-work controls steering and takes care of braking reaction. (Motorcycle Mechanics)
A variant of the leading-link fork was the Ernie Earles type, with long links combined into an integral fork pivoted behind the wheel. BMW standardized this arrangement for practically two decades between abandoning their original telescopic fork and introducing a long-travel version in 1969. Essentially though, their Earles-type fork was a compromise for solo and sidecar duty. Although its lack of sliding friction was a boon in sidecar cornering, its high steering inertia blunted steering response in a solo. As an aid to this dual purpose, BMW provided a choice of two pivot mounting locations so that trail could be easily set for solo or sidecar use. Basic layout of the Difazio system. The wheel hub spins on large diameter bearings, and the non-rotating axle is integral with the steering “king-pin”. A forward facing swinging fork connects the axle to the main chassis, whilst the upper frame-work controls steering and takes care of braking reaction. (Motorcycle Mechanics)
Sketch of the BMW “telelever”. The forward facing link pivots at the rear and connects to the unsprung part of the front forks, suspension loads are thence fed into the suspension unit. The upper mounting of the non-sliding stanchions allows for the back and forth movement necessary in this design. Many years ago the OEC factory produced a design that had no directly defined steering axis, it featured a series of links which determined a “virtual” steering axis, the position of which varied with steering angle. This machine was often praised for its good steering.
Sketch of the BMW “telelever”. The forward facing link pivots at the rear and connects to the unsprung part of the front forks, suspension loads are thence fed into the suspension unit. The upper mounting of the non-sliding stanchions allows for the back and forth movement necessary in this design. Many years ago the OEC factory produced a design that had no directly defined steering axis, it featured a series of links which determined a “virtual” steering axis, the position of which varied with steering angle. This machine was often praised for its good steering.
Plunger springing was first to gain wide acceptance — partly because of racing successes by BMW anc Norton, and partly because, for the manufacturer, it was the easiest type to which existing rigid frame: could be adapted. However, its limitations were clear from the start. Firstly, the incorporation of sprinc boxes at the rear destroyed the triangulation of the seatstays and chainstays, thus permitting each side to flex independently in a vertical plane, even to the extent of fatigue failure. Secondly, resistance tc wheel tilting depended also on very stiff clamping of the spindle in the sliders (as it does with telescopic front forks and some leading links). Finally, the straight-line vertical movement of the wheel tightenec the chain considerably at the extremes of travel, so limiting the range of wheel movement anc necessitating a slack chain setting in the static load position.
Plunger springing was first to gain wide acceptance — partly because of racing successes by BMW anc Norton, and partly because, for the manufacturer, it was the easiest type to which existing rigid frame: could be adapted. However, its limitations were clear from the start. Firstly, the incorporation of sprinc boxes at the rear destroyed the triangulation of the seatstays and chainstays, thus permitting each side to flex independently in a vertical plane, even to the extent of fatigue failure. Secondly, resistance tc wheel tilting depended also on very stiff clamping of the spindle in the sliders (as it does with telescopic front forks and some leading links). Finally, the straight-line vertical movement of the wheel tightenec the chain considerably at the extremes of travel, so limiting the range of wheel movement anc necessitating a slack chain setting in the static load position.
In a bid to provide ample stiffness without triangulation the Velocette swing-arm was made from taper-section, taper: gauge tubing. Moto Guzzi also chose to triangulate their swing-arm when they introduced rear springing in 1935 and scored spectacular victories in the Senior and Lightweight TTs with Stanley Woods aboard. Later on, however, they changed to a plain fork welded from very large diameter tubing and claimed it to be equally stiff torsionally and stiffer laterally (because the earlier fork was triangulated only vertically, not laterally). Another swing-arm designed to provide ample stiffness without triangulation was introduced by Velocette on the works racers of the mid-1930s and standardized on the Mark VIIl KTT in 1939. In this, the arms comprised taper-diameter, taper-gauge tubes.
In a bid to provide ample stiffness without triangulation the Velocette swing-arm was made from taper-section, taper: gauge tubing. Moto Guzzi also chose to triangulate their swing-arm when they introduced rear springing in 1935 and scored spectacular victories in the Senior and Lightweight TTs with Stanley Woods aboard. Later on, however, they changed to a plain fork welded from very large diameter tubing and claimed it to be equally stiff torsionally and stiffer laterally (because the earlier fork was triangulated only vertically, not laterally). Another swing-arm designed to provide ample stiffness without triangulation was introduced by Velocette on the works racers of the mid-1930s and standardized on the Mark VIIl KTT in 1939. In this, the arms comprised taper-diameter, taper-gauge tubes.
This view of the so-called garden-gate Manx Norton shows how the plunger type rear springing destroyed the triangulation of the chainstays and seatstays in the unsprung frame that it replaced. At the front, wheel location was by means of the highly regarded “Roadholder” hydraulically damped telescopic forks. Many plain swing-arms lacked adequate torsional stiffness and this gave rise to a vogue for precisely matched left and right suspension struts to minimize one of the causes of twisting. Girling actually sold so-called matched pairs as an after-market fitment option.
This view of the so-called garden-gate Manx Norton shows how the plunger type rear springing destroyed the triangulation of the chainstays and seatstays in the unsprung frame that it replaced. At the front, wheel location was by means of the highly regarded “Roadholder” hydraulically damped telescopic forks. Many plain swing-arms lacked adequate torsional stiffness and this gave rise to a vogue for precisely matched left and right suspension struts to minimize one of the causes of twisting. Girling actually sold so-called matched pairs as an after-market fitment option.
On the Doug Hele designed 1952 BSA 250 cc grand-prix racer, the top apex of the triangulated rear fork actuated a horizontal suspension strut in a tunnel through the oil tank. Front fork was long leading link. Note the unusual steered head stock arrangement. (MCW)
On the Doug Hele designed 1952 BSA 250 cc grand-prix racer, the top apex of the triangulated rear fork actuated a horizontal suspension strut in a tunnel through the oil tank. Front fork was long leading link. Note the unusual steered head stock arrangement. (MCW)
A fairly modern example of a rear fork braced below pivot level, on the Suzuki RG500 racer. Unlike the Moto Guzzi layout of this type, two gas suspension struts are fitted, in the conventional position.
A fairly modern example of a rear fork braced below pivot level, on the Suzuki RG500 racer. Unlike the Moto Guzzi layout of this type, two gas suspension struts are fitted, in the conventional position.
One of the earliest rocker-arm rear-suspension layouts, on Kork Ballington’s Kawasaki GP machine. The A-bracket on the left was attached to the pivoted fork while the suspension strut was anchored to the frame at the bottom. (Watts)
One of the earliest rocker-arm rear-suspension layouts, on Kork Ballington’s Kawasaki GP machine. The A-bracket on the left was attached to the pivoted fork while the suspension strut was anchored to the frame at the bottom. (Watts)
Velocette used slotted top anchorages on the rear suspension, as a means of compensating for different loads, this altered both springing and damping effective rates. Wheel movement was also increased on the soft setting. Although rubber can be endowed with varying degrees of inherent damping, this was never sufficient tc preclude the use of separate friction or hydraulic dampers. A significant advantage of air springing (as also of rubber) is its progressive rate, which can be demonstrated by operating a tyre pump with the outlet sealed by one finger. A possible disadvantage, for really extreme conditions, is that the air heats up, so increasing its pressure, hence its effective rate. Velocette introduced this form of springing, made by Dowty, on their works racing machines in the mid-1930s, subsequently extending its use to the real end of the KTT in 1939 and the front end of some of their post-war roadsters. In production, leakage o both air and damping oil was a problem. Improvements in sealing materials and surface finishes enabled the Japanese to revive air springing in the 1970 / 80s, first as an auxiliary to steel springs, ther on its own. On the whole, however, single or multi-rate coil springs remain the first choice for mos designers. The French Fournales units being a notable exception.
Velocette used slotted top anchorages on the rear suspension, as a means of compensating for different loads, this altered both springing and damping effective rates. Wheel movement was also increased on the soft setting. Although rubber can be endowed with varying degrees of inherent damping, this was never sufficient tc preclude the use of separate friction or hydraulic dampers. A significant advantage of air springing (as also of rubber) is its progressive rate, which can be demonstrated by operating a tyre pump with the outlet sealed by one finger. A possible disadvantage, for really extreme conditions, is that the air heats up, so increasing its pressure, hence its effective rate. Velocette introduced this form of springing, made by Dowty, on their works racing machines in the mid-1930s, subsequently extending its use to the real end of the KTT in 1939 and the front end of some of their post-war roadsters. In production, leakage o both air and damping oil was a problem. Improvements in sealing materials and surface finishes enabled the Japanese to revive air springing in the 1970 / 80s, first as an auxiliary to steel springs, ther on its own. On the whole, however, single or multi-rate coil springs remain the first choice for mos designers. The French Fournales units being a notable exception.
The loaded tyre has a half width of W2 at its widest section (X2) and so the normal force is 2.P.W, . Therefore, the extra force over this section, when loaded, is 2.P.(W2— W4,) but as the tyre is only widened over a small portion of the bottom part of the circumference, this force supports the load F. Fig. 2.2 The in-line components (F; & F2) of the side- wall tension are reduced by factors equal to the cosines of the angles of the side-wall. This reduction is greater with the loaded tyre resulting in a greater compressive force on the lower part of the rim. Fig. 2.1 The left hand side shows half of an inflated but unloaded tyre, a tension (T) is created in the carcass by the internal pressure. To the right, the compressed and widened shape of the loaded tyre is shown.
The loaded tyre has a half width of W2 at its widest section (X2) and so the normal force is 2.P.W, . Therefore, the extra force over this section, when loaded, is 2.P.(W2— W4,) but as the tyre is only widened over a small portion of the bottom part of the circumference, this force supports the load F. Fig. 2.2 The in-line components (F; & F2) of the side- wall tension are reduced by factors equal to the cosines of the angles of the side-wall. This reduction is greater with the loaded tyre resulting in a greater compressive force on the lower part of the rim. Fig. 2.1 The left hand side shows half of an inflated but unloaded tyre, a tension (T) is created in the carcass by the internal pressure. To the right, the compressed and widened shape of the loaded tyre is shown.
Whereas, regardless of the sophistication of the conventional suspension system, it would be quite impractical to use wheels without pneumatic tyres, or some other form of tyre that allowed considerable bump deflection. The loads fed into the wheels without such tyres would be enormous at all but slow speeds, and continual wheel failure would be the norm. With no tyre the wheel would then be subject to an average vertical acceleration of approximately 1000 G. (the peak value would be higher than this). This means than if the wheel and brake assembly had a mass of 25 kg. then the average point load on the rim would be 245 KN. (about 25 tons). What wheel could stand that? If the wheel was shod with a normal tyre, then this would have at ground level, an initial spring rate, to a sharp edge, of approx. 17-35 N/mm., but rising as the tyre wraps around the step. In performing this function the pneumatic tyre is the first object that feels any road shocks and so acts as the most important element in the machine’s suspension system. To the extent that, whilst uncomfortable, it would be quite feasible to ride a bike around the roads, at reasonable speeds with no other form of bump absorption. In fact rear suspension was not at all common until the 1940s or 50s.
Whereas, regardless of the sophistication of the conventional suspension system, it would be quite impractical to use wheels without pneumatic tyres, or some other form of tyre that allowed considerable bump deflection. The loads fed into the wheels without such tyres would be enormous at all but slow speeds, and continual wheel failure would be the norm. With no tyre the wheel would then be subject to an average vertical acceleration of approximately 1000 G. (the peak value would be higher than this). This means than if the wheel and brake assembly had a mass of 25 kg. then the average point load on the rim would be 245 KN. (about 25 tons). What wheel could stand that? If the wheel was shod with a normal tyre, then this would have at ground level, an initial spring rate, to a sharp edge, of approx. 17-35 N/mm., but rising as the tyre wraps around the step. In performing this function the pneumatic tyre is the first object that feels any road shocks and so acts as the most important element in the machine’s suspension system. To the extent that, whilst uncomfortable, it would be quite feasible to ride a bike around the roads, at reasonable speeds with no other form of bump absorption. In fact rear suspension was not at all common until the 1940s or 50s.
Fig 2.3. This shows the vertical displacement of the front wheel. There is little difference between the maximum displacements for the two cases with a normal tyre, for a small step the front tyre absorbs most of the shock. However, in the case of a very stiff tyre, the wheel movement is increased by a factor of about 8 times. The tyre leaves the ground in this case and the landing bounces can be seen around 0.4 seconds. The sudden change of slope at the point marked (BS) is when the forks bottom out against the bump stops.
Fig 2.3. This shows the vertical displacement of the front wheel. There is little difference between the maximum displacements for the two cases with a normal tyre, for a small step the front tyre absorbs most of the shock. However, in the case of a very stiff tyre, the wheel movement is increased by a factor of about 8 times. The tyre leaves the ground in this case and the landing bounces can be seen around 0.4 seconds. The sudden change of slope at the point marked (BS) is when the forks bottom out against the bump stops.
Fig 2.4 These curves show the vertical movement of the CoG of the bike and rider. As in Fig 2.3 it is clear that the stiff tyre causes much higher bike movements, to the obvious detriment of comfort.
Fig 2.4 These curves show the vertical movement of the CoG of the bike and rider. As in Fig 2.3 it is clear that the stiff tyre causes much higher bike movements, to the obvious detriment of comfort.
Fig 2.5 Demonstrating the different accelerations transmitted to the bike and rider, these curves show the vertical accelerations at the CoG. The sudden step in the stiff tyre case at 0.012 seconds is due to the forks bottoming against the bump stops. (In this case, doubling the spring rate (curves not shown), prevented the bottoming out and the maximum acceleration was reduced to 12 Gs. from 32.)
Fig 2.5 Demonstrating the different accelerations transmitted to the bike and rider, these curves show the vertical accelerations at the CoG. The sudden step in the stiff tyre case at 0.012 seconds is due to the forks bottoming against the bump stops. (In this case, doubling the spring rate (curves not shown), prevented the bottoming out and the maximum acceleration was reduced to 12 Gs. from 32.)
Fig 2.6 Front wheel vertical acceleration for the two cases with a normal tyre. The early part is similar for the two cases, the suspension has little effect here, it is tyre deflection that is the most important for this height of step.
Fig 2.6 Front wheel vertical acceleration for the two cases with a normal tyre. The early part is similar for the two cases, the suspension has little effect here, it is tyre deflection that is the most important for this height of step.
Fig 2.7 As in Fig 2.6 this curve is of the wheel acceleration, but for the stiff tyre case, the values of the normal case are overwhelmed by the stiff tyre case, with a peak value of close to 600 G compared with less than 20 G. above. This high acceleration would cause very high structural loading.. Again note the effects of the landing bounces at 0.4 seconds.
Fig 2.7 As in Fig 2.6 this curve is of the wheel acceleration, but for the stiff tyre case, the values of the normal case are overwhelmed by the stiff tyre case, with a peak value of close to 600 G compared with less than 20 G. above. This high acceleration would cause very high structural loading.. Again note the effects of the landing bounces at 0.4 seconds.
As the tyre is so good at removing most of the road shocks, right at the point of application, perhaps it would be worth while to consider designing it to absorb even more and eliminate the need for other suspension. Unfortunately we would run into other problems. We have all seen large construction machinery bouncing down the road on their balloon tyres, sometimes this gets so violent that the wheels actually leave the ground. A pneumatic tyre acts just like a spring, and the rubber acts as a damper when it flexes, but when the tyre is made bigger the springing effect overwhelms the damping and we then get the uncontrolled bouncing. So there are practical restraints to the amount of cushioning that can be built into a tyre for any given application. Tyre stiffness or spring rate The springing characteristics mentioned above are largely affected by the tyre inflation pressure, bu there are other influences also. Carcass material and construction and the properties and tread patterr of the outer layer of rubber all have an effect on both the springing properties and the area in contac with the ground (contact patch). Under and over inflation both allow the tyre to assume non-optimun cross-sectional shapes, additionally the inflation pressure exerts an influence over the lateral flexibility o a tyre and this is a property of the utmost importance to motorcycle stability. Manufacturers recommendations should always be adhered to.
As the tyre is so good at removing most of the road shocks, right at the point of application, perhaps it would be worth while to consider designing it to absorb even more and eliminate the need for other suspension. Unfortunately we would run into other problems. We have all seen large construction machinery bouncing down the road on their balloon tyres, sometimes this gets so violent that the wheels actually leave the ground. A pneumatic tyre acts just like a spring, and the rubber acts as a damper when it flexes, but when the tyre is made bigger the springing effect overwhelms the damping and we then get the uncontrolled bouncing. So there are practical restraints to the amount of cushioning that can be built into a tyre for any given application. Tyre stiffness or spring rate The springing characteristics mentioned above are largely affected by the tyre inflation pressure, bu there are other influences also. Carcass material and construction and the properties and tread patterr of the outer layer of rubber all have an effect on both the springing properties and the area in contac with the ground (contact patch). Under and over inflation both allow the tyre to assume non-optimun cross-sectional shapes, additionally the inflation pressure exerts an influence over the lateral flexibility o a tyre and this is a property of the utmost importance to motorcycle stability. Manufacturers recommendations should always be adhered to.
Fig. 2.9. Vertical stiffness of a standard road tyre against a flat surface at different inflation pressures. This data is from an Avon Azaro Sport Il 170/60 ZR17. The upward arrows indicate the compression of the tyre and the 2™ line with the downward arrow (shown only at 1.9 bar for clarity) shows the behaviour of the tyre when the load is released. The shaded area between the two lines represents a loss of energy called hysteresis. This acts as a source of suspension damping and also heats the tyre. (From data supplied by Avon Tyres.) This spring rate acts in series with the suspension springs and is an important part of the overall suspension system. An interesting property of rubber is that when compressed and released it doesn't usually return exactly to its original position, this is known as hysteresis. This effect is shown only for the 1.9 bar. case, the curve drawn during the loading phase is not followed during the unloading phase. The area between these two curves represents a loss of energy which results in tyre heating and also acts as a form of suspension damping. In this particular case the energy lost over one loading and unloading cycle is approximately 10% of the total stored energy in the compressed tyre, and is a significant parameter controlling tyre bounce. These damping characteristics vary with loading frequency.
Fig. 2.9. Vertical stiffness of a standard road tyre against a flat surface at different inflation pressures. This data is from an Avon Azaro Sport Il 170/60 ZR17. The upward arrows indicate the compression of the tyre and the 2™ line with the downward arrow (shown only at 1.9 bar for clarity) shows the behaviour of the tyre when the load is released. The shaded area between the two lines represents a loss of energy called hysteresis. This acts as a source of suspension damping and also heats the tyre. (From data supplied by Avon Tyres.) This spring rate acts in series with the suspension springs and is an important part of the overall suspension system. An interesting property of rubber is that when compressed and released it doesn't usually return exactly to its original position, this is known as hysteresis. This effect is shown only for the 1.9 bar. case, the curve drawn during the loading phase is not followed during the unloading phase. The area between these two curves represents a loss of energy which results in tyre heating and also acts as a form of suspension damping. In this particular case the energy lost over one loading and unloading cycle is approximately 10% of the total stored energy in the compressed tyre, and is a significant parameter controlling tyre bounce. These damping characteristics vary with loading frequency.
Fig. 2.10. Lateral stiffness of the same tyre shown in fig. 2.9. The vertical load was constant at 355 kgf. and the wheel was kept vertical. As expected the tyre is somewhat stiffer with the higher inflation pressure but loses grip or saturates at the lower lateral load of 460 kgf. compared to 490 kgf. at the lower pressure. (From data supplied by Avon Tyres.) Saturation is indicated when the curve more or less becomes horizontal, this is when the tyre cannot support any more lateral force and it displaces or slides sideways, with an approximately constant force. The contact patch area and pressure produced at the lower air pressure has allowed more static grip. However, these tests were done with the artificial case of an upright and non-rotating wheel on a serrated steel surface and hence it would be risky to extrapolate this grip characteristic to a moving machine on a normal road. Although not shown, the lateral deformation would also be subject to some hysteresis and this damping and the lateral flexibility exert an important influence over the weave stability of the motorcycle.
Fig. 2.10. Lateral stiffness of the same tyre shown in fig. 2.9. The vertical load was constant at 355 kgf. and the wheel was kept vertical. As expected the tyre is somewhat stiffer with the higher inflation pressure but loses grip or saturates at the lower lateral load of 460 kgf. compared to 490 kgf. at the lower pressure. (From data supplied by Avon Tyres.) Saturation is indicated when the curve more or less becomes horizontal, this is when the tyre cannot support any more lateral force and it displaces or slides sideways, with an approximately constant force. The contact patch area and pressure produced at the lower air pressure has allowed more static grip. However, these tests were done with the artificial case of an upright and non-rotating wheel on a serrated steel surface and hence it would be risky to extrapolate this grip characteristic to a moving machine on a normal road. Although not shown, the lateral deformation would also be subject to some hysteresis and this damping and the lateral flexibility exert an important influence over the weave stability of the motorcycle.
Measurement setup. Various weights were placed on the end of a beam, which loaded the tyre via a thick plate of glass. The beam was arranged to apply the load to the tyre with a 4:1 leverage. So a 25 kgf. weight would load the tyre with 100 kgf. By tracing over the glass the contact area was determined.
Measurement setup. Various weights were placed on the end of a beam, which loaded the tyre via a thick plate of glass. The beam was arranged to apply the load to the tyre with a 4:1 leverage. So a 25 kgf. weight would load the tyre with 100 kgf. By tracing over the glass the contact area was determined.
Fig. 2.12 The top plot shows the measured contact patch pressure at various wheel loads for a constant inflation pressure of 2.4 bar. The lower curves show the contact pressure at various inflation pressures for a fixed load of 1210 N. The numbers at the data points correspond with the contact area tracings in the previous sketch, fig. 2.11. The plain line on each plot shows the case of the contact patch pressure being equal to the inflation pressure.
Fig. 2.12 The top plot shows the measured contact patch pressure at various wheel loads for a constant inflation pressure of 2.4 bar. The lower curves show the contact pressure at various inflation pressures for a fixed load of 1210 N. The numbers at the data points correspond with the contact area tracings in the previous sketch, fig. 2.11. The plain line on each plot shows the case of the contact patch pressure being equal to the inflation pressure.
Because the forces fed into the front tyre are offset from the steering axis, steering forces are introduced which have to be counteracted by trail effects and the rider, usually requiring more physical effort. As figure 2.14 shows, a road disturbance such as a stone can cause a couple tending to steer the wheel to that side. A wide tyre is found by more stones and produces a higher couple. The profiles and lateral flexing characteristics also need careful balancing front to rear. Indeed, both the tyre and bike manufacturers undertake considerable testing to determine the best blend. A combination of tyres that gives excellent results on one model may prove lethal on another. Relative to their weight and power, motorcycles have smaller tyre sections than cars. This derives chiefly from our need to bank while cornering, in which case wide tyres may impair handling and stability, even though grip may be improved. Figure 2.13 shows how the contact patch moves away from the centre plane of the wheel or the steering axis as the machine is banked. A greater angle of lean is necessary to balance centrifugal force, and this may require a slightly higher centre of gravity to restore cornering clearance.
Because the forces fed into the front tyre are offset from the steering axis, steering forces are introduced which have to be counteracted by trail effects and the rider, usually requiring more physical effort. As figure 2.14 shows, a road disturbance such as a stone can cause a couple tending to steer the wheel to that side. A wide tyre is found by more stones and produces a higher couple. The profiles and lateral flexing characteristics also need careful balancing front to rear. Indeed, both the tyre and bike manufacturers undertake considerable testing to determine the best blend. A combination of tyres that gives excellent results on one model may prove lethal on another. Relative to their weight and power, motorcycles have smaller tyre sections than cars. This derives chiefly from our need to bank while cornering, in which case wide tyres may impair handling and stability, even though grip may be improved. Figure 2.13 shows how the contact patch moves away from the centre plane of the wheel or the steering axis as the machine is banked. A greater angle of lean is necessary to balance centrifugal force, and this may require a slightly higher centre of gravity to restore cornering clearance.
As the brake is applied a torque is transmitted through the wheel to the contact patch, where it manifests itself as a linear horizontal force at the road surface. The road pushes backward on the tyre and with equal intensity the tyre pushes forward on the road. Under power the situation is similar except that the directions of the forces are reversed.
As the brake is applied a torque is transmitted through the wheel to the contact patch, where it manifests itself as a linear horizontal force at the road surface. The road pushes backward on the tyre and with equal intensity the tyre pushes forward on the road. Under power the situation is similar except that the directions of the forces are reversed.
Fig. 2.15 This curve is the general shape to be expected when considering the frictional force of a tyre compared to some slip parameter. For braking and driving this slip parameter is expressed as a percentage of the non-slipping wheel speed, and the peak in the graph occurs around 10 - 20%. For the steering case the slip is expressed in terms of a slip angle. With a motorcycle we also have to consider the effects of lean generated force. At all but the slowest of speeds the principal mechanism of grip between a tyre and a normal paved road involves some form of slippage between the tyre and the road. Moto-X machines driving through a corner in a deep rut naturally derive at least part of their cornering force in a different way. This idea of the need for slip may come as a surprise to many people as it is commonly thought that slip only takes place when riding on the limit and “drifting” the machine. The truth is, that at even moderate braking, accelerating or cornering loads the tyre is experiencing some slip over the surface. For braking and accelerating, tests show that we get the maximum amount of grip when the tyre is slipping by about 10% or more. That means that under hard acceleration the surface velocity of the tyre will be over 10% higher than the road speed of the bike. Conversely, under braking the wheel speed will be some 10% less than the road speed. If we plot a graph of the degree of grip against some slip parameter we can see that the relationship becomes highly non-linear as the tyre gets close to the limit, but the overall shape of the curve will be similar for both traction and cornering.
Fig. 2.15 This curve is the general shape to be expected when considering the frictional force of a tyre compared to some slip parameter. For braking and driving this slip parameter is expressed as a percentage of the non-slipping wheel speed, and the peak in the graph occurs around 10 - 20%. For the steering case the slip is expressed in terms of a slip angle. With a motorcycle we also have to consider the effects of lean generated force. At all but the slowest of speeds the principal mechanism of grip between a tyre and a normal paved road involves some form of slippage between the tyre and the road. Moto-X machines driving through a corner in a deep rut naturally derive at least part of their cornering force in a different way. This idea of the need for slip may come as a surprise to many people as it is commonly thought that slip only takes place when riding on the limit and “drifting” the machine. The truth is, that at even moderate braking, accelerating or cornering loads the tyre is experiencing some slip over the surface. For braking and accelerating, tests show that we get the maximum amount of grip when the tyre is slipping by about 10% or more. That means that under hard acceleration the surface velocity of the tyre will be over 10% higher than the road speed of the bike. Conversely, under braking the wheel speed will be some 10% less than the road speed. If we plot a graph of the degree of grip against some slip parameter we can see that the relationship becomes highly non-linear as the tyre gets close to the limit, but the overall shape of the curve will be similar for both traction and cornering.
Clearly then, something else must be done in order to persuade the tyre/road junction to generate the required cornering or centripetal force, to direct the tyre toward the centre of the curve. Ona car or other self balancing vehicle this is done by turning the wheel in more than the line of the curve. The angle between the plane of the tyre and the direction of the tangent to the curve is known as the ‘slip angle’. Consider a wheel, held upright as on a car, following a curved path as in fig. 2.16. If this wheel is aligned with the direction of the curve at any particular point on it (i.e. pointing in the direction of travel) then the wheel will tend to go straight on along a tangent to the curve and will generate no cornering force. Slip angle
Clearly then, something else must be done in order to persuade the tyre/road junction to generate the required cornering or centripetal force, to direct the tyre toward the centre of the curve. Ona car or other self balancing vehicle this is done by turning the wheel in more than the line of the curve. The angle between the plane of the tyre and the direction of the tangent to the curve is known as the ‘slip angle’. Consider a wheel, held upright as on a car, following a curved path as in fig. 2.16. If this wheel is aligned with the direction of the curve at any particular point on it (i.e. pointing in the direction of travel) then the wheel will tend to go straight on along a tangent to the curve and will generate no cornering force. Slip angle
Fig. 2.17 This shows how a slip angle results in a lateral sliding velocity, which causes a force lateral to the wheel’: direction. This force can be resolved in two components. One, normal to the direction of travel which balances thi cornering force and a second which is aligned with the direction of travel. This force acts against the forward motion o the vehicle and hence is a drag force. With a car this drag force, for a given corner, sets a limit to the maximum possibl speed achievable from a given engine power, even if the grip of the tyre is not exceeded.
Fig. 2.17 This shows how a slip angle results in a lateral sliding velocity, which causes a force lateral to the wheel’: direction. This force can be resolved in two components. One, normal to the direction of travel which balances thi cornering force and a second which is aligned with the direction of travel. This force acts against the forward motion o the vehicle and hence is a drag force. With a car this drag force, for a given corner, sets a limit to the maximum possibl speed achievable from a given engine power, even if the grip of the tyre is not exceeded.
The previous section explains how steering a wheel generates the force necessary to force a vehicle to turn around a bend. However, bicycles and motorcycles must lean when taking a corner and this leaning also creates a lateral cornering force. In fact at all but the slowest of speeds and cornering accelerations this force will likely be the major contributor to the total cornering force, and the steering effects will just make up for the difference between the required cornering force and that provided by the lean. Hence, the degree of steering necessary on a motorcycle is much less than that required by a car. The lateral tyre force due to the tyre camber angle is known as camber thrust or camber force. Let’s look at Fig. 2.18 to see how this force is created. Fig. 2.18 The top left sketch shows how the contact patch of the tyre flattens at an angle and effectively becomes a slice of a cone which tends to turn around the geometric apex of the cone. The other two diagrams show how this cone tries to turn a tighter circle than the actual bend radius.
The previous section explains how steering a wheel generates the force necessary to force a vehicle to turn around a bend. However, bicycles and motorcycles must lean when taking a corner and this leaning also creates a lateral cornering force. In fact at all but the slowest of speeds and cornering accelerations this force will likely be the major contributor to the total cornering force, and the steering effects will just make up for the difference between the required cornering force and that provided by the lean. Hence, the degree of steering necessary on a motorcycle is much less than that required by a car. The lateral tyre force due to the tyre camber angle is known as camber thrust or camber force. Let’s look at Fig. 2.18 to see how this force is created. Fig. 2.18 The top left sketch shows how the contact patch of the tyre flattens at an angle and effectively becomes a slice of a cone which tends to turn around the geometric apex of the cone. The other two diagrams show how this cone tries to turn a tighter circle than the actual bend radius.
To reiterate, as the wheel leans over, cambers, it can be considered to behave as part of a cone which tries to turn a very tight corner, but the centripetal force required is impossibly high, and forces a sideways slip which causes this tight curve to straighten out and follow the turn radius that the rider desires. Thus cornering force can be generated without the slip angle so necessary with a car. If we were to plot a graph of camber force against camber angle then we would get a very similar shape to that shown in fig. 2.15. This is not surprising because, although not immediately apparent, the lateral force is accompanied by a similar side slipping mechanism, let’s have another look at what’s happening as in fig. 2.19.
To reiterate, as the wheel leans over, cambers, it can be considered to behave as part of a cone which tries to turn a very tight corner, but the centripetal force required is impossibly high, and forces a sideways slip which causes this tight curve to straighten out and follow the turn radius that the rider desires. Thus cornering force can be generated without the slip angle so necessary with a car. If we were to plot a graph of camber force against camber angle then we would get a very similar shape to that shown in fig. 2.15. This is not surprising because, although not immediately apparent, the lateral force is accompanied by a similar side slipping mechanism, let’s have another look at what’s happening as in fig. 2.19.
Fig. 2.20 This plot is broadly representative of the lateral tyre force produced by a combination of leaning and steering. Super-imposed is a graph showing the tyre force required at any lean angle to maintain a balanced turn.
Fig. 2.20 This plot is broadly representative of the lateral tyre force produced by a combination of leaning and steering. Super-imposed is a graph showing the tyre force required at any lean angle to maintain a balanced turn.
Normalized cornering force Fig. 2.21 Tyre characteristics similar to those shown in Fig. 2.20, except that the camber stiffness is increased.
Normalized cornering force Fig. 2.21 Tyre characteristics similar to those shown in Fig. 2.20, except that the camber stiffness is increased.
braking separately, it is also capable of supporting a 0.7 g turn and 0.7 g braking at the same time. Ir practice the friction circle is not truly circular, depending on the tyre characteristics it might be slightly elliptical. This effect is likely to be small. The “steering angle” of the handlebars depends on the required slip angle, but the “steering torque” to be applied by the rider to achieve this position depends on many other factors. Wheel diameter, tyre width, rake and trail are but some of the influencing parameters. On some machines it may even be essential to apply a negative torque to maintain a positive slip angle at the required level, the geometry may try to force the bike on to a tighter line if left to its own devices. A bike set up like this, tends to flop over easily, and may give the feeling that the rider needs to hold it up. Whereas another type of machine may need a positive torque to maintain the desired steering angle and this bike will feel as if it needs to be held down. A few numbers may help to show how different speeds and bend radii affect the lean angle and hence the effective cone radius, and camber thrust.
braking separately, it is also capable of supporting a 0.7 g turn and 0.7 g braking at the same time. Ir practice the friction circle is not truly circular, depending on the tyre characteristics it might be slightly elliptical. This effect is likely to be small. The “steering angle” of the handlebars depends on the required slip angle, but the “steering torque” to be applied by the rider to achieve this position depends on many other factors. Wheel diameter, tyre width, rake and trail are but some of the influencing parameters. On some machines it may even be essential to apply a negative torque to maintain a positive slip angle at the required level, the geometry may try to force the bike on to a tighter line if left to its own devices. A bike set up like this, tends to flop over easily, and may give the feeling that the rider needs to hold it up. Whereas another type of machine may need a positive torque to maintain the desired steering angle and this bike will feel as if it needs to be held down. A few numbers may help to show how different speeds and bend radii affect the lean angle and hence the effective cone radius, and camber thrust.
If we relate this to tyre characteristics as in fig. 2.20 then we can say that an under-steering vehicle has a lower normalized cornering stiffness for the front tyres than does the rear. Fig. 2.23 shows an example case of different characteristics between front and rear tyres. Actually the strict definitions of under-/over-steering are a little more complex but this visualization will suffice for our present purposes. The accurate definition of this characteristic effectively relates the
If we relate this to tyre characteristics as in fig. 2.20 then we can say that an under-steering vehicle has a lower normalized cornering stiffness for the front tyres than does the rear. Fig. 2.23 shows an example case of different characteristics between front and rear tyres. Actually the strict definitions of under-/over-steering are a little more complex but this visualization will suffice for our present purposes. The accurate definition of this characteristic effectively relates the
Fig. 2.24 Showing the attitude of the bike and wheels during high levels of drifting, as for example in speedway
Fig. 2.24 Showing the attitude of the bike and wheels during high levels of drifting, as for example in speedway
So where is the benefit in drifting? --- Well, as any drag racer knows maximum traction is achieved whe the tyre is spinning slightly. In fact the tyre companies tell us that approximately 10 - 20% slip is abot optimum on tarmac, as shown earlier. So if sufficient power is available to cause this degree of slip th total friction force on the tyre can be divided between the normal lateral cornering force and that due t driving force, some of which gets resolved towards the bend centre, thus part of the total cornering forc is under control of the throttle. Referring to the friction circle in fig. 2.22 this means that the resultar force moves rearward around the circle. In racing, feel and control over what the tyre is doing can b enhanced by skilled use of this technique. When cornering in the classic style it is hard to feel the limit c adhesion and there are minimal control options left to deal with the situation when that limit is exceedec On the other hand when the bike has been setup with significant drift the rider has more control over th total tyre force through the throttle. Therefore when he feels that he is about to lose it, he can just clos sliding at all the lean angle of the combined bike and rider would be 45° , but when sliding at 12° the required camber angle reduces to 44.4° , hardly a big difference! To return to speedway cornerin styles, where a more typical slide angle may be, say 50° , then again for the 1G. case the lean angle reduces to 33°. So we can see that to reduce the lean angle by a significant amount requires a hig] degree of drifting, more in fact than is usual in road racing.
So where is the benefit in drifting? --- Well, as any drag racer knows maximum traction is achieved whe the tyre is spinning slightly. In fact the tyre companies tell us that approximately 10 - 20% slip is abot optimum on tarmac, as shown earlier. So if sufficient power is available to cause this degree of slip th total friction force on the tyre can be divided between the normal lateral cornering force and that due t driving force, some of which gets resolved towards the bend centre, thus part of the total cornering forc is under control of the throttle. Referring to the friction circle in fig. 2.22 this means that the resultar force moves rearward around the circle. In racing, feel and control over what the tyre is doing can b enhanced by skilled use of this technique. When cornering in the classic style it is hard to feel the limit c adhesion and there are minimal control options left to deal with the situation when that limit is exceedec On the other hand when the bike has been setup with significant drift the rider has more control over th total tyre force through the throttle. Therefore when he feels that he is about to lose it, he can just clos sliding at all the lean angle of the combined bike and rider would be 45° , but when sliding at 12° the required camber angle reduces to 44.4° , hardly a big difference! To return to speedway cornerin styles, where a more typical slide angle may be, say 50° , then again for the 1G. case the lean angle reduces to 33°. So we can see that to reduce the lean angle by a significant amount requires a hig] degree of drifting, more in fact than is usual in road racing.
“Ecca” Erik Andersson appears to have forgotten that he is on a road circuit and seems to be practicing speedway at Alastaro circuit. Such extreme attitudes in road racing are not normally considered a stable situation. However, in this case Ecca retained control. (Esso Gunnerarsson) Racers often drift their machines for additional reasons. With a lot of power available the classic large radius line through a corner is not always the best and it is often beneficial to turn as sharply as possible at the bend entrance, and leave the straightest possible exit path to get the maximum acceleration out of the bend. Drifting the bike into the bend can sometimes help get the machine better lined up for this quick exit. Control over the yaw attitude is also useful to setup the exit angle from a corner, for example if the bike is steering outward too much, it can be made to turn inward by throttle application.
“Ecca” Erik Andersson appears to have forgotten that he is on a road circuit and seems to be practicing speedway at Alastaro circuit. Such extreme attitudes in road racing are not normally considered a stable situation. However, in this case Ecca retained control. (Esso Gunnerarsson) Racers often drift their machines for additional reasons. With a lot of power available the classic large radius line through a corner is not always the best and it is often beneficial to turn as sharply as possible at the bend entrance, and leave the straightest possible exit path to get the maximum acceleration out of the bend. Drifting the bike into the bend can sometimes help get the machine better lined up for this quick exit. Control over the yaw attitude is also useful to setup the exit angle from a corner, for example if the bike is steering outward too much, it can be made to turn inward by throttle application.
Fig. 2.26 Showing the principal constructional aspects of cross-ply (left) and radial-ply (right) tyres. The cross-ply is made up of successive layers or plies with the cord angled across the tyre at approximately 45 deg. each layer applied at opposite angles to the previous. Hence the name of cross-ply. A true radial tyre has its cord running from one bead radially over the tyre to the opposite bead, a circumferential belt is added to reduce tyre fling at speed, generally stabilize the construction and enhance load capacity. Even long after cars had almost exclusively changed over to radial tyres, those fitted to motorcycles retained the cross-ply form of construction and many tyre engineers claimed that radial tyres would never be suitable for motorcycles. It is true that cars and motorcycles have very different tyre requirements but recent history has shown that radials have benefits to offer motorcycles as well.
Fig. 2.26 Showing the principal constructional aspects of cross-ply (left) and radial-ply (right) tyres. The cross-ply is made up of successive layers or plies with the cord angled across the tyre at approximately 45 deg. each layer applied at opposite angles to the previous. Hence the name of cross-ply. A true radial tyre has its cord running from one bead radially over the tyre to the opposite bead, a circumferential belt is added to reduce tyre fling at speed, generally stabilize the construction and enhance load capacity. Even long after cars had almost exclusively changed over to radial tyres, those fitted to motorcycles retained the cross-ply form of construction and many tyre engineers claimed that radial tyres would never be suitable for motorcycles. It is true that cars and motorcycles have very different tyre requirements but recent history has shown that radials have benefits to offer motorcycles as well.
When the first radials started to appear most manufacturers opted for a compromise approach and dic not produce “true” radials with a 90 degree bias on the main cords and used a cross-ply layout with steeper bias angles, together with the angled belt cords of the bias belted tyre. Some tyres of that perioc that were called radials were obviously done so by the marketing departments as many couldn't justif the description on technical grounds. Many radials still use a biased belt but high performance tyres are moving to a 0 degree wind on belt, particularly for rears, which further restricts tyre growth at speed enhancing tyre stability. An interesting development along these lines is the Variable Belt Density tyre In this case the belt is wound closer in the centre of the tyre and wider apart at the edges as shown ir fig. 2.27. Prior to the cautious introduction of radial tyres in 1984, the bias belted tyre was developed. This followed the normal cross-ply construction but with the addition of circumferential belt plies. These belts werent truly circumferential and the cord was set at a slight angle, often around 25°. Nevertheless, these angled cords effectively triangulated the main plies and so stabilized the overall structure with the results of less tyre fling, better stability, longer life and they were capable of better handling the ever increasing power outputs of the contemporary engines.
When the first radials started to appear most manufacturers opted for a compromise approach and dic not produce “true” radials with a 90 degree bias on the main cords and used a cross-ply layout with steeper bias angles, together with the angled belt cords of the bias belted tyre. Some tyres of that perioc that were called radials were obviously done so by the marketing departments as many couldn't justif the description on technical grounds. Many radials still use a biased belt but high performance tyres are moving to a 0 degree wind on belt, particularly for rears, which further restricts tyre growth at speed enhancing tyre stability. An interesting development along these lines is the Variable Belt Density tyre In this case the belt is wound closer in the centre of the tyre and wider apart at the edges as shown ir fig. 2.27. Prior to the cautious introduction of radial tyres in 1984, the bias belted tyre was developed. This followed the normal cross-ply construction but with the addition of circumferential belt plies. These belts werent truly circumferential and the cord was set at a slight angle, often around 25°. Nevertheless, these angled cords effectively triangulated the main plies and so stabilized the overall structure with the results of less tyre fling, better stability, longer life and they were capable of better handling the ever increasing power outputs of the contemporary engines.
The elements of this are shown in figure 3.1. The primary function of trail is to build in a certain amount of steering stability, and it also is of great importance to the lean-in phase when cornering. We can see that both front and rear tyres contact the ground behind the point where the steering axis meets it, this gives rise to a castor (self-centring) effect on both wheels. The linear measurement of this castor along the ground (steering axis to centre of contact patch) is usually called the trail.
The elements of this are shown in figure 3.1. The primary function of trail is to build in a certain amount of steering stability, and it also is of great importance to the lean-in phase when cornering. We can see that both front and rear tyres contact the ground behind the point where the steering axis meets it, this gives rise to a castor (self-centring) effect on both wheels. The linear measurement of this castor along the ground (steering axis to centre of contact patch) is usually called the trail.
How the trail causes a self-centring effect can be understood from figure 3.3 which is a plan view of a wheel displaced from the straight-ahead position.
How the trail causes a self-centring effect can be understood from figure 3.3 which is a plan view of a wheel displaced from the straight-ahead position.
From this, we can see that increasing the trail as a means of increasing the restoring tendency on the wheels is subject to the law of diminishing returns. It must also be emphasized that the disturbance to a machine’s direction of travel, due to sideways displacement of the tyre contact patch, is less from the rear wheel than the front because of the much smaller angle to the direction of travel that the displacement causes. To summarize we can say that, while the large trail of the rear wheel has a relatively small restoring effect, the effect of rear-wheel displacement on directional stability is also small and so compensates. One may be forgiven for thinking that, because the positive trail of the rear wheel is much greater than that of the front (typically 50-100 mm. front, 1300-1500 mm. rear), the rear wheel would be the more important in this respect. The reverse is the case and there are several reasons for this. See figure 3.4. Imagine that the contact patch of each wheel is in turn displaced sideways by the same amount, say 12 mm. The front wheel will be turned approximately 7-10 degrees about the steering axis; this gives rise to a slip angle of the same amount and generates a sideways force that has only the relatively small steering inertia of the front wheel and fork to accelerate back to the straight-ahead position. However, the slip angle of the displaced rear wheel will be much less (approximately 1/2 degree) and so will the restoring force. Not only do we have a smaller force but it also has to act on the inertia of a major proportion of the machine and rider, hence the response is much weaker than in the case of the front wheel.
From this, we can see that increasing the trail as a means of increasing the restoring tendency on the wheels is subject to the law of diminishing returns. It must also be emphasized that the disturbance to a machine’s direction of travel, due to sideways displacement of the tyre contact patch, is less from the rear wheel than the front because of the much smaller angle to the direction of travel that the displacement causes. To summarize we can say that, while the large trail of the rear wheel has a relatively small restoring effect, the effect of rear-wheel displacement on directional stability is also small and so compensates. One may be forgiven for thinking that, because the positive trail of the rear wheel is much greater than that of the front (typically 50-100 mm. front, 1300-1500 mm. rear), the rear wheel would be the more important in this respect. The reverse is the case and there are several reasons for this. See figure 3.4. Imagine that the contact patch of each wheel is in turn displaced sideways by the same amount, say 12 mm. The front wheel will be turned approximately 7-10 degrees about the steering axis; this gives rise to a slip angle of the same amount and generates a sideways force that has only the relatively small steering inertia of the front wheel and fork to accelerate back to the straight-ahead position. However, the slip angle of the displaced rear wheel will be much less (approximately 1/2 degree) and so will the restoring force. Not only do we have a smaller force but it also has to act on the inertia of a major proportion of the machine and rider, hence the response is much weaker than in the case of the front wheel.
Although the primary purpose of front-wheel trail is to provide a degree of directional stability, there are various side effects too. Let us consider two of them.
Although the primary purpose of front-wheel trail is to provide a degree of directional stability, there are various side effects too. Let us consider two of them.
The basic reason for rake is less easy to explain than that for trail because it is just possible that we don't need it. Why then, do all current production machines have the steering axis raked back between about 23 and 30 degrees from the vertical — 27 degrees often used to be considered as a magic figure? There is no simple answer and several factors are probably relevant. There is no denying the greater convenience of construction (see figure 3.6). In most designs of hub-centre steering, simply for space reasons, the wheel spindle is not offset from the steering axis, hence trail is wholly dependent on rake angle, which is typically between 10 and 15 degrees to give the required result. This is a much steeper angle than usual, yet hub-centre layouts are renowned for their stability and steering. This reputation may stem from reasons other than the steep rake angle, but it certainly seems that this departure from the norm is not harmful and may indeed be beneficial. There is an appendix to this book that details some of the author's experiments with steep rake angles. Basically, these experiments showed that there were distinct benefits from the use of small rake angles, which were not generally offset by compensating disadvantages. It is interesting to note that since the first edition of this book in 1984 detailing those experiments, there has been a gradual trend toward the use of steeper rake angles, particularly for racing and sports machines. Considered as unstable values a couple of decades ago these machines now commonly use castor angles of 20 to 23 deaorees. However. some of the reason for change nrohably comes hack to ease of construction aqain.
The basic reason for rake is less easy to explain than that for trail because it is just possible that we don't need it. Why then, do all current production machines have the steering axis raked back between about 23 and 30 degrees from the vertical — 27 degrees often used to be considered as a magic figure? There is no simple answer and several factors are probably relevant. There is no denying the greater convenience of construction (see figure 3.6). In most designs of hub-centre steering, simply for space reasons, the wheel spindle is not offset from the steering axis, hence trail is wholly dependent on rake angle, which is typically between 10 and 15 degrees to give the required result. This is a much steeper angle than usual, yet hub-centre layouts are renowned for their stability and steering. This reputation may stem from reasons other than the steep rake angle, but it certainly seems that this departure from the norm is not harmful and may indeed be beneficial. There is an appendix to this book that details some of the author's experiments with steep rake angles. Basically, these experiments showed that there were distinct benefits from the use of small rake angles, which were not generally offset by compensating disadvantages. It is interesting to note that since the first edition of this book in 1984 detailing those experiments, there has been a gradual trend toward the use of steeper rake angles, particularly for racing and sports machines. Considered as unstable values a couple of decades ago these machines now commonly use castor angles of 20 to 23 deaorees. However. some of the reason for change nrohably comes hack to ease of construction aqain.
1) Reduction of castor effect For a given amount of ground trail, the self-aligning torque on the front wheel and fork depends on the length of the lever arm from the centre of the tyre contact patch to the steering axis, measured at right angles to that axis, real trail in other words. As can clearly be seen in figure 3.8, this lever arm is shortened as the rake is increased. In practice, this means that to maintain the same real trail we need more ground trail as the rake angle is increased. Let us now examine the main effects of rake angle. Figure 3.7 shows three different rake angles, all giving the same amount of ground trail. The real trail will be reduced in the first two cases. The reduction will be by about 10% for the normal range of rake angle, and around 3% at 15 degrees of rake.
1) Reduction of castor effect For a given amount of ground trail, the self-aligning torque on the front wheel and fork depends on the length of the lever arm from the centre of the tyre contact patch to the steering axis, measured at right angles to that axis, real trail in other words. As can clearly be seen in figure 3.8, this lever arm is shortened as the rake is increased. In practice, this means that to maintain the same real trail we need more ground trail as the rake angle is increased. Let us now examine the main effects of rake angle. Figure 3.7 shows three different rake angles, all giving the same amount of ground trail. The real trail will be reduced in the first two cases. The reduction will be by about 10% for the normal range of rake angle, and around 3% at 15 degrees of rake.
Fig. 3.9 Reduction in ground trail for several values of castor angle plotted against steering angle. See fig 3.10. This reduction in real trail is emphasized even more as we apply some steering angle, fig. 3.9 shows the reduction in ground trail at different castor angles for various degrees of handle-bar steering angle up to an extreme value of 80 degrees.
Fig. 3.9 Reduction in ground trail for several values of castor angle plotted against steering angle. See fig 3.10. This reduction in real trail is emphasized even more as we apply some steering angle, fig. 3.9 shows the reduction in ground trail at different castor angles for various degrees of handle-bar steering angle up to an extreme value of 80 degrees.
Fig. 3.10 Similar graphs to those in fig. 3.9 except that the curves show the effects on real trail Road bikes will seldom have more than about 45 degrees of steering lock, but these curves show that at a typical castor angle of 27 degrees the ground trail will have reduced from 89 mm. to 46mm. at a steering angle of 45 degrees. Fig. 3.10 shows the same data plotted for the real trail. Compare the difference between the two sets of curves at low steering angles. Both of these sets of graphs are plotted for a bike held vertically and with a 305 mm. radius front tyre and 89 mm. of ground trail.
Fig. 3.10 Similar graphs to those in fig. 3.9 except that the curves show the effects on real trail Road bikes will seldom have more than about 45 degrees of steering lock, but these curves show that at a typical castor angle of 27 degrees the ground trail will have reduced from 89 mm. to 46mm. at a steering angle of 45 degrees. Fig. 3.10 shows the same data plotted for the real trail. Compare the difference between the two sets of curves at low steering angles. Both of these sets of graphs are plotted for a bike held vertically and with a 305 mm. radius front tyre and 89 mm. of ground trail.
With a normal motorcycle (i.e. with positive trail and positive rake angle) held vertically, the steering head would drop as we turn the handlebar to either side. (With negative trail, which is not normal, it would rise.) The greater the rake angle, the greater the drop. This can be best appreciated by visualizing an extreme rake angle, as in figure 3.12. 3) Steering-head drop
With a normal motorcycle (i.e. with positive trail and positive rake angle) held vertically, the steering head would drop as we turn the handlebar to either side. (With negative trail, which is not normal, it would rise.) The greater the rake angle, the greater the drop. This can be best appreciated by visualizing an extreme rake angle, as in figure 3.12. 3) Steering-head drop
Fig. 3.14 Front end drop for a vertical bike at various rake angles over a range of steering angles. While this effect is detrimental to balance (hence another reason why trials bikes have steep head angles), and to directional stability while travelling in a straight line, it helps to steer the wheel into a corner when banked over. However, it is important not to over-dramatize such effects. The following figs. 3.14 and 3.15. put the actual values into perspective. The first of which shows the front end drop and the second shows the steering torque needed to lift the bike back up. At 27 degrees rake and 45 degrees of steer the front end drop is only about 8.2 mm. and the restoring torque needed is below 6 Nm. for each 50 kgf. of front end weight. At less extreme handlebar angles the effect is much less, up to 10 degrees of steer the drop is no more than 0.5 mm. with a 27 degree rake angle.
Fig. 3.14 Front end drop for a vertical bike at various rake angles over a range of steering angles. While this effect is detrimental to balance (hence another reason why trials bikes have steep head angles), and to directional stability while travelling in a straight line, it helps to steer the wheel into a corner when banked over. However, it is important not to over-dramatize such effects. The following figs. 3.14 and 3.15. put the actual values into perspective. The first of which shows the front end drop and the second shows the steering torque needed to lift the bike back up. At 27 degrees rake and 45 degrees of steer the front end drop is only about 8.2 mm. and the restoring torque needed is below 6 Nm. for each 50 kgf. of front end weight. At less extreme handlebar angles the effect is much less, up to 10 degrees of steer the drop is no more than 0.5 mm. with a 27 degree rake angle.
Fig. 3.15 Average additional torque needed to restore the steering to the straight ahead position due to front end drop. Note that this torque is approximately only 1% of that due to the slip angle at low steering angles.
Fig. 3.15 Average additional torque needed to restore the steering to the straight ahead position due to front end drop. Note that this torque is approximately only 1% of that due to the slip angle at low steering angles.
It is sometimes stated that, because the offset places the centre of gravity of the wheel and fork ahead of the steering axis, this produces a torque tending to steer the wheel into the curve while the machine is banked over. This is true only when the bike is stationary. Fig. 3.16 shows what happens during cornering, centrifugal force acts through the offset to steer the wheel out of the turn, but this effect is almost completely balanced by that of the gravitational force tending to steer into the turn. Actually with zero width tyres and the rider in line with the bike these two effects balance exactly. Hence, the offset has little nett effect on the self-steering characteristics of the machine, except when using very wide tyres.
It is sometimes stated that, because the offset places the centre of gravity of the wheel and fork ahead of the steering axis, this produces a torque tending to steer the wheel into the curve while the machine is banked over. This is true only when the bike is stationary. Fig. 3.16 shows what happens during cornering, centrifugal force acts through the offset to steer the wheel out of the turn, but this effect is almost completely balanced by that of the gravitational force tending to steer into the turn. Actually with zero width tyres and the rider in line with the bike these two effects balance exactly. Hence, the offset has little nett effect on the self-steering characteristics of the machine, except when using very wide tyres.
Fig 3.18 True ground level steering angle plotted against lean angle for a range of handlebar angles. The upper curve of each pair is for a rake angle of 27 degrees whereas the lower line is for zero castor angle. Note for example that for zero lean , with 27 degrees of rake, the handlebar angle must be increased to achieve the required true steering angle. To have a ground level steering of 10 degrees needs approx. 11 degrees of handlebar angle. On the other hand, increasing lean angle tends to increase the effective steering angle, although this is relatively unimportant within the normal range of use. The shaded area is a rough guide to combinations of lean and steer which we are unlikely to use. Fig. 3.17 This extreme example shows how steering rake reduces the effective steering angle at ground level.
Fig 3.18 True ground level steering angle plotted against lean angle for a range of handlebar angles. The upper curve of each pair is for a rake angle of 27 degrees whereas the lower line is for zero castor angle. Note for example that for zero lean , with 27 degrees of rake, the handlebar angle must be increased to achieve the required true steering angle. To have a ground level steering of 10 degrees needs approx. 11 degrees of handlebar angle. On the other hand, increasing lean angle tends to increase the effective steering angle, although this is relatively unimportant within the normal range of use. The shaded area is a rough guide to combinations of lean and steer which we are unlikely to use. Fig. 3.17 This extreme example shows how steering rake reduces the effective steering angle at ground level.
Fig. 3.19 This representation of a bike, as two planes which intersect along the steering axis, can help visualize the various geometric aspects caused by both leaning and steering. One of the planes is that of the main part of the frame which includes the rear wheel, the other is the centre plane of the front wheel. The front plane can rotate about the steering axis contained on the rear plane and the rear plane can lean relative to the vertical.
Fig. 3.19 This representation of a bike, as two planes which intersect along the steering axis, can help visualize the various geometric aspects caused by both leaning and steering. One of the planes is that of the main part of the frame which includes the rear wheel, the other is the centre plane of the front wheel. The front plane can rotate about the steering axis contained on the rear plane and the rear plane can lean relative to the vertical.
Figure 3.20 shows how, for a given bend, a long-wheelbase machine needs the front wheel to be turned farther into the bend. Consequently more effort is required for cornering; also, a given deflection of the front wheel (say, from bumps) will have less effect on its directional stability.
Figure 3.20 shows how, for a given bend, a long-wheelbase machine needs the front wheel to be turned farther into the bend. Consequently more effort is required for cornering; also, a given deflection of the front wheel (say, from bumps) will have less effect on its directional stability.
3) Inertia effects The wheelbase has an effect on load transfer under braking and acceleration: for a given centre-of- gravity height, the longer the wheelbase the smaller the load transfer. In addition, the moments of inertia in the pitch and yaw planes are increased, which makes a long-wheelbase machine more sluggish and stable. It is also clear from figure 3.21 that, for a given sideways deflection, the angle of the rear wheel to the direction of travel is smaller with a longer wheelbase, thus improving directional stability.
3) Inertia effects The wheelbase has an effect on load transfer under braking and acceleration: for a given centre-of- gravity height, the longer the wheelbase the smaller the load transfer. In addition, the moments of inertia in the pitch and yaw planes are increased, which makes a long-wheelbase machine more sluggish and stable. It is also clear from figure 3.21 that, for a given sideways deflection, the angle of the rear wheel to the direction of travel is smaller with a longer wheelbase, thus improving directional stability.
Fig. 3.23 The smaller wheel travels a smaller distance “a” to mount the step than the larger wheel which travels “b”. Fig. 3.22 Asmaller wheel drops farther into holes than a larger wheel.
Fig. 3.23 The smaller wheel travels a smaller distance “a” to mount the step than the larger wheel which travels “b”. Fig. 3.22 Asmaller wheel drops farther into holes than a larger wheel.
Fig. 3.26 The long lever arm causes high bending moments in the fork legs and steering head, which can give rise to juddering and wheel hop. Fig. 3.25 Lateral displacement due to fork and wheel flexure.
Fig. 3.26 The long lever arm causes high bending moments in the fork legs and steering head, which can give rise to juddering and wheel hop. Fig. 3.25 Lateral displacement due to fork and wheel flexure.
Keeping the rear wheel aligned with the steering axis involves not only the lateral stiffness of the wheel but also the torsional and lateral stiffness of the main frame and pivoted rear suspension.
Keeping the rear wheel aligned with the steering axis involves not only the lateral stiffness of the wheel but also the torsional and lateral stiffness of the main frame and pivoted rear suspension.
Load transfer. Under braking, vertical load is transferred from the rear wheel to the front, under acceleration, the transfer is in the opposite direction. Lengthening the wheelbase decreases the load transfer, as also does lowering the mass centre height and reducing the mass. Load transfer is not affected by the longitudinal location of the centre of gravity, though this controls the static weight supported by each wheel. Traction. Since the driving force that the rear wheel can deliver to the ground is limited by the load carried by the wheel, a rearward weight distribution improves traction. However, we must balance this requirement against the need to keep the front wheel on the ground for steering. A forward mass bias also helps directional stability, as it does in a dart or an arrow.
Load transfer. Under braking, vertical load is transferred from the rear wheel to the front, under acceleration, the transfer is in the opposite direction. Lengthening the wheelbase decreases the load transfer, as also does lowering the mass centre height and reducing the mass. Load transfer is not affected by the longitudinal location of the centre of gravity, though this controls the static weight supported by each wheel. Traction. Since the driving force that the rear wheel can deliver to the ground is limited by the load carried by the wheel, a rearward weight distribution improves traction. However, we must balance this requirement against the need to keep the front wheel on the ground for steering. A forward mass bias also helps directional stability, as it does in a dart or an arrow.
Pitch. Pitch inertia controls how rapidly the bike pitches forward or backward under the various inputs due to braking, accelerating and road bumps. Except in trials and motocross, there is no great need for a fast pitch response and so a large amount of pitch inertia is not generally harmful. Indeed, it may contribute to comfort when braking or when traversing bumpy surfaces, as shown in the chapter or suspension. It is not easy to geometrically define the axis about which a machine pitches because it varies with the configuration of the bike. For example, if a machine is sprung at the front but not at the rear, it will pitch about the rear-wheel axle, while a machine with the opposite arrangement (i.e. sprung rear, rigid front) would pitch about the front-wheel centre. In the case of a conventional machine, sprung at both ends, the pitch axis depends on suspension geometry and spring rates. Of the three angular motions, pitch is unique in that it comprises mainly of movement of the sprung part of the machine witk respect to the wheels, whereas roll and yaw are global movements of the bike including wheels. As far as linear motions are concerned, it is the amount of the machine’s mass that is important. But when it comes to the angular motions of pitch (about a transverse axis), yaw (about a vertical axis) and roll (about a longitudinal axis), the distribution of the mass is all-important because that governs what are called the moments of inertia. These are a measure of the inertia effect about the particular axis and its value determines the ease with which we can apply angular acceleration to the machine about that axis.
Pitch. Pitch inertia controls how rapidly the bike pitches forward or backward under the various inputs due to braking, accelerating and road bumps. Except in trials and motocross, there is no great need for a fast pitch response and so a large amount of pitch inertia is not generally harmful. Indeed, it may contribute to comfort when braking or when traversing bumpy surfaces, as shown in the chapter or suspension. It is not easy to geometrically define the axis about which a machine pitches because it varies with the configuration of the bike. For example, if a machine is sprung at the front but not at the rear, it will pitch about the rear-wheel axle, while a machine with the opposite arrangement (i.e. sprung rear, rigid front) would pitch about the front-wheel centre. In the case of a conventional machine, sprung at both ends, the pitch axis depends on suspension geometry and spring rates. Of the three angular motions, pitch is unique in that it comprises mainly of movement of the sprung part of the machine witk respect to the wheels, whereas roll and yaw are global movements of the bike including wheels. As far as linear motions are concerned, it is the amount of the machine’s mass that is important. But when it comes to the angular motions of pitch (about a transverse axis), yaw (about a vertical axis) and roll (about a longitudinal axis), the distribution of the mass is all-important because that governs what are called the moments of inertia. These are a measure of the inertia effect about the particular axis and its value determines the ease with which we can apply angular acceleration to the machine about that axis.
Fig. 3.33 shows how a centrifugal cornering force can be considered as acting through the centre of gravity which is normally above the roll centre. The distance between the CoG. and the roll axis produces a torque (usually called the roll couple) which causes the car body to lean. It is ironic then and unfortunate also, that these terms have recently become somewhat fashionable in the motorcycle world without any corresponding relevance. Motorcycles, being single track vehicles don't have any similar side to side suspension mechanisms and so the relevance of this term to motorcycles is hard to justify and consequently is most often completely misunderstood.
Fig. 3.33 shows how a centrifugal cornering force can be considered as acting through the centre of gravity which is normally above the roll centre. The distance between the CoG. and the roll axis produces a torque (usually called the roll couple) which causes the car body to lean. It is ironic then and unfortunate also, that these terms have recently become somewhat fashionable in the motorcycle world without any corresponding relevance. Motorcycles, being single track vehicles don't have any similar side to side suspension mechanisms and so the relevance of this term to motorcycles is hard to justify and consequently is most often completely misunderstood.
It is easy to see that there is no corresponding analogous concept between cars and bikes in this regard. Unlike a car the roll moment generated by steady cornering force on a motorcycle is exactly balanced by the opposite roll moment created by gravity acting on the leant over machine. It is only during the transient lean-in phase that any unbalanced roll torques exist and these are not caused by the same mechanisms as those shown above for a car. This will be explained elsewhere. If we were to apply the car criterion for determining the roll axis to a motorcycle then it would have to be near ground level, through the tyre contact patches, to produce the appropriate roll couple.
It is easy to see that there is no corresponding analogous concept between cars and bikes in this regard. Unlike a car the roll moment generated by steady cornering force on a motorcycle is exactly balanced by the opposite roll moment created by gravity acting on the leant over machine. It is only during the transient lean-in phase that any unbalanced roll torques exist and these are not caused by the same mechanisms as those shown above for a car. This will be explained elsewhere. If we were to apply the car criterion for determining the roll axis to a motorcycle then it would have to be near ground level, through the tyre contact patches, to produce the appropriate roll couple.
Fig 3.35 Adding the two motions produces a combined motion with a larger radius as shown. Fig 3.34 Shows both the angular and lateral motions
Fig 3.35 Adding the two motions produces a combined motion with a larger radius as shown. Fig 3.34 Shows both the angular and lateral motions
In the explanation above, we only considered the effects on the front wheel, gyroscopic and steering forces are at work on the rear also, but it is much harder to steer the rear wheel independently, as the whole bike must yaw, rather than just the wheel and forks, as on the front. Hence, only a small contribution is made to the auto-balance mechanism by the rear. We have now considered balance in a straight line, but as we lean when cornering, there must be other factors at work to maintain equilibrium under these conditions. precession of the front wheel tends to steer it to the left. Also, as shown in chapter 3, trail will also cause a left steering effect when the bike leans left. This sets the machine on a curved path (to the left), sc creating a centrifugal force (to the right), which counters the lean and tends to restore the machine to the vertical, the gyroscopic reactions are thus reversed tending to restore the steering to the straight ahead position. In practice, that which we regard as riding in a straight line, is really a series of balance correcting curves, if we could look at the actual paths taken by the centre-lines of the wheels, we should see that the front wheel path continually crosses that of the rear.
In the explanation above, we only considered the effects on the front wheel, gyroscopic and steering forces are at work on the rear also, but it is much harder to steer the rear wheel independently, as the whole bike must yaw, rather than just the wheel and forks, as on the front. Hence, only a small contribution is made to the auto-balance mechanism by the rear. We have now considered balance in a straight line, but as we lean when cornering, there must be other factors at work to maintain equilibrium under these conditions. precession of the front wheel tends to steer it to the left. Also, as shown in chapter 3, trail will also cause a left steering effect when the bike leans left. This sets the machine on a curved path (to the left), sc creating a centrifugal force (to the right), which counters the lean and tends to restore the machine to the vertical, the gyroscopic reactions are thus reversed tending to restore the steering to the straight ahead position. In practice, that which we regard as riding in a straight line, is really a series of balance correcting curves, if we could look at the actual paths taken by the centre-lines of the wheels, we should see that the front wheel path continually crosses that of the rear.
Fig. 4.4 This graph shows the turn radius that corresponds to various road speeds for different lean angles.
Fig. 4.4 This graph shows the turn radius that corresponds to various road speeds for different lean angles.
Fig. 4.5 Relationship between lateral cornering acceleration in Gs. and the required lean angle to maintain balance. The lean angle is measured through the tyre contact patch and the CoG of the combined bike and rider.
Fig. 4.5 Relationship between lateral cornering acceleration in Gs. and the required lean angle to maintain balance. The lean angle is measured through the tyre contact patch and the CoG of the combined bike and rider.
Fig. 4.6 Lean-in performance with no lateral tyre forces. Note how the leaning velocity or roll rate builds up to a maximum and then decreases to zero as the final lean angle is reached. Note also that the steering or handlebar angle does not exactly follow the shape of the input torque. In this simulation the tyre parameters are set to produce no lateral tyre forces. This is a situation that is difficult to achieve in practice, the closest we might manage would be to ride on an ice-rink with slick tyres. However, this simulation is very useful to test the “gyroscopic only” proposition above.
Fig. 4.6 Lean-in performance with no lateral tyre forces. Note how the leaning velocity or roll rate builds up to a maximum and then decreases to zero as the final lean angle is reached. Note also that the steering or handlebar angle does not exactly follow the shape of the input torque. In this simulation the tyre parameters are set to produce no lateral tyre forces. This is a situation that is difficult to achieve in practice, the closest we might manage would be to ride on an ice-rink with slick tyres. However, this simulation is very useful to test the “gyroscopic only” proposition above.
Not drawn on the graphs but the computer calculations showed only a very minimal sideways movemen of the bike and so we're not going to get very far around the turn despite a reasonable lean angle Obviously there is something very wrong or deficient with the idea that gyroscopic effects alone are a that’s necessary to start a bike turning. Or is there.....? Clearly what we’ve just described is the case of a moto-X rider in mid-air using the precessional effects of the still spinning front wheel to lay his bike over by turning the handlebars. Our simulation fits this situation well. The steering is turned by a relatively large amount and there is little sideways movement of the bike as a whole, but the machine can be leant over quite quickly to any desired angle. Just like our calculations this is done with no forces acting through the tyres.
Not drawn on the graphs but the computer calculations showed only a very minimal sideways movemen of the bike and so we're not going to get very far around the turn despite a reasonable lean angle Obviously there is something very wrong or deficient with the idea that gyroscopic effects alone are a that’s necessary to start a bike turning. Or is there.....? Clearly what we’ve just described is the case of a moto-X rider in mid-air using the precessional effects of the still spinning front wheel to lay his bike over by turning the handlebars. Our simulation fits this situation well. The steering is turned by a relatively large amount and there is little sideways movement of the bike as a whole, but the machine can be leant over quite quickly to any desired angle. Just like our calculations this is done with no forces acting through the tyres.
Fig. 4.7 Torques about the steering axis. For simplicity the damping is all assumed to be viscous and thus proportional to the velocity of steering motions. Note how the residual or total torque follows the input torque in the initial stage before the opposite damping and gyroscopic torques have built up. The calculations show that tyre feedback and steering torque from the weight of the machine acting about the steering axis are both zero. This is hardly surprising as we eliminated tyre forces from the system and our air-bound motoX’er only experiences gravity in so far as its effect on bringing him back to earth. Of the remaining graphs the Residual torque is the sum of all the other contributing torques, some cancelling others out. This total torque is what’s left to accelerate the steered inertia about the steering axis, we'll see later when considering other cases that it is normally lower in value than in this case. This is quite simply explained because in this case there have been very large steering movements and so the torque necessary to accelerate up to these large steering angles needs to be correspondingly larger also. Note too the magnitude of the damping torque, we'll also see that this is much higher than normal, the reason also being linked to the large and hence rapid steering motions.
Fig. 4.7 Torques about the steering axis. For simplicity the damping is all assumed to be viscous and thus proportional to the velocity of steering motions. Note how the residual or total torque follows the input torque in the initial stage before the opposite damping and gyroscopic torques have built up. The calculations show that tyre feedback and steering torque from the weight of the machine acting about the steering axis are both zero. This is hardly surprising as we eliminated tyre forces from the system and our air-bound motoX’er only experiences gravity in so far as its effect on bringing him back to earth. Of the remaining graphs the Residual torque is the sum of all the other contributing torques, some cancelling others out. This total torque is what’s left to accelerate the steered inertia about the steering axis, we'll see later when considering other cases that it is normally lower in value than in this case. This is quite simply explained because in this case there have been very large steering movements and so the torque necessary to accelerate up to these large steering angles needs to be correspondingly larger also. Note too the magnitude of the damping torque, we'll also see that this is much higher than normal, the reason also being linked to the large and hence rapid steering motions.
Fig. 4.8 Similar to fig. 4.6 except that the lateral displacement is added, note how this sideways movement starts to build up. The maximum steering angle applied is similar to the previous case, although the detail shape of the curve is different. Experience tells us that such steering angles are not realistic.
Fig. 4.8 Similar to fig. 4.6 except that the lateral displacement is added, note how this sideways movement starts to build up. The maximum steering angle applied is similar to the previous case, although the detail shape of the curve is different. Experience tells us that such steering angles are not realistic.
Fig. 4.9 Resultant roll torques for both of the two cases mentioned. In case 1 all the roll torque is due to gyroscopic effects, but in case 2, although the gyroscopic torque has the major influence the tyre and gravitational torques begin to have an effect also. Compare this with fig. 4.11.
Fig. 4.9 Resultant roll torques for both of the two cases mentioned. In case 1 all the roll torque is due to gyroscopic effects, but in case 2, although the gyroscopic torque has the major influence the tyre and gravitational torques begin to have an effect also. Compare this with fig. 4.11.
Fig. 4.10 These results have been calculated using realistic tyre properties and allowing for gyroscopic forces. This simulation now appears to represent the observed real world behaviour reasonably well. Note the scaling on the handlebar angle.
Fig. 4.10 These results have been calculated using realistic tyre properties and allowing for gyroscopic forces. This simulation now appears to represent the observed real world behaviour reasonably well. Note the scaling on the handlebar angle.
Fig. 4.11 These two sets of roll torque curves are the same data, the time scale has just been expanded in the second to show more detail during the important initial period. Note the almost insignificant contribution of the gyroscopic moment to the overall roll torque, in complete contrast to the two previous cases, in fig. 4.9.
Fig. 4.11 These two sets of roll torque curves are the same data, the time scale has just been expanded in the second to show more detail during the important initial period. Note the almost insignificant contribution of the gyroscopic moment to the overall roll torque, in complete contrast to the two previous cases, in fig. 4.9.
Fig. 4.12 The main four steering torques are shown here. We can see that for the first 20 or 30 msecs the resultant torque is due mainly to the rider input, but as the countersteering tyre force builds up the resultant begins to reduce. As the roll velocity begins to increase a matching gyroscopic torque is produced. After the initial torque swings have occurred a pattern emerges as the machine steadies. The gravitational and tyre induced torques tend to follow similar but opposite shapes, thus cancelling each other, but also the gyroscopic torque and rider input tend to mirror one another. In other words the rider input mainly goes to balance the gyroscopic steering torque. This explains why steering requires less effort with lighter wheels. As the roll torque is little influenced by gyroscopic effects it starts to look as if precession plays little or no part in the lean-in process when we simulate it with a reasonable tyre model. However, the precessional steering torque has the single largest value of all contributions to the total steering torque, as shown in fig. 4.12.
Fig. 4.12 The main four steering torques are shown here. We can see that for the first 20 or 30 msecs the resultant torque is due mainly to the rider input, but as the countersteering tyre force builds up the resultant begins to reduce. As the roll velocity begins to increase a matching gyroscopic torque is produced. After the initial torque swings have occurred a pattern emerges as the machine steadies. The gravitational and tyre induced torques tend to follow similar but opposite shapes, thus cancelling each other, but also the gyroscopic torque and rider input tend to mirror one another. In other words the rider input mainly goes to balance the gyroscopic steering torque. This explains why steering requires less effort with lighter wheels. As the roll torque is little influenced by gyroscopic effects it starts to look as if precession plays little or no part in the lean-in process when we simulate it with a reasonable tyre model. However, the precessional steering torque has the single largest value of all contributions to the total steering torque, as shown in fig. 4.12.
Fig. 4.13 Behaviour with no gyroscopic reactions. Compare with fig. 4.10. The main difference is in the shape of the curve of the rider input torque. In the previous case this input has a higher magnitude and remains negative for much longer. Without gyroscopic torque to oppose the rider’s input, the rider himself must modify his response so that the resultant steering torque remains within acceptable limits.
Fig. 4.13 Behaviour with no gyroscopic reactions. Compare with fig. 4.10. The main difference is in the shape of the curve of the rider input torque. In the previous case this input has a higher magnitude and remains negative for much longer. Without gyroscopic torque to oppose the rider’s input, the rider himself must modify his response so that the resultant steering torque remains within acceptable limits.
Fig. 4.14 Roll torques due to lateral rider weight shift without any input steering torque, (hands-off case). Note that the maximum roll torque from gravity is much lower that in previous which is simply due to the smaller lean angle developed without rider steering control.
Fig. 4.14 Roll torques due to lateral rider weight shift without any input steering torque, (hands-off case). Note that the maximum roll torque from gravity is much lower that in previous which is simply due to the smaller lean angle developed without rider steering control.
It is interesting to compare the shapes of the steer/slip angle curve with the applied torque curve. Until the steady state is reached the applied torque always remains negative, however, the steering movement shows a more complex shape. Initially it is negative and this, (through the tyre steering stiffness) gives the initial negative tyre lateral force that starts the roll process. During the final part of the manoeuvre, prompted by the reduction in applied torque the slip angle shows a positive bump in its curve, which in turn through the tyre steering stiffness will slow the roll. So if we talk about applied torque then counter-steering lasts for the whole duration of the transient phase, whereas, talking about steering movements the counter-steering only lasts for a brief initial period. The roll torque is little affected by precession as we have seen but this is far from the case with the steering torque. The roll velocity must be accompanied by a corresponding steering torque. This gyroscopically induced steering torque can easily be a major component of the total steering torque during lean-in. We have seen that these precessional effects are of great importance to the maintenance of auto-stability. Looking at the graphs of steering torques from case 3, we can see that the overall torque available for angular acceleration of the steering is much smaller than any of the
It is interesting to compare the shapes of the steer/slip angle curve with the applied torque curve. Until the steady state is reached the applied torque always remains negative, however, the steering movement shows a more complex shape. Initially it is negative and this, (through the tyre steering stiffness) gives the initial negative tyre lateral force that starts the roll process. During the final part of the manoeuvre, prompted by the reduction in applied torque the slip angle shows a positive bump in its curve, which in turn through the tyre steering stiffness will slow the roll. So if we talk about applied torque then counter-steering lasts for the whole duration of the transient phase, whereas, talking about steering movements the counter-steering only lasts for a brief initial period. The roll torque is little affected by precession as we have seen but this is far from the case with the steering torque. The roll velocity must be accompanied by a corresponding steering torque. This gyroscopically induced steering torque can easily be a major component of the total steering torque during lean-in. We have seen that these precessional effects are of great importance to the maintenance of auto-stability. Looking at the graphs of steering torques from case 3, we can see that the overall torque available for angular acceleration of the steering is much smaller than any of the
Fig. 4.16 Two examples of the front tyre forces on entry to a corner. The curves (1) are for a relatively normal lean-in and (2) are for a rapid turn in with the trail reduced from 90 mm. to 60 mm. and the wheel moment of inertia reduced from 0.41 to 0.25 Kg.m? . In both examples we can see how the available grip is reduced due to centrifugal force, but in case (1) this is relatively minor and the lateral tyre force required is well below the available grip until the turn is nearly fully established. On the other hand, the rapid roll rate in case (2) both reduces the available grip and also requires more grip during this period. The result being that at about 0.25 secs. the required grip exceeds that available and the front slides out. This is the final limitation on the maximum turn-in performance possible.
Fig. 4.16 Two examples of the front tyre forces on entry to a corner. The curves (1) are for a relatively normal lean-in and (2) are for a rapid turn in with the trail reduced from 90 mm. to 60 mm. and the wheel moment of inertia reduced from 0.41 to 0.25 Kg.m? . In both examples we can see how the available grip is reduced due to centrifugal force, but in case (1) this is relatively minor and the lateral tyre force required is well below the available grip until the turn is nearly fully established. On the other hand, the rapid roll rate in case (2) both reduces the available grip and also requires more grip during this period. The result being that at about 0.25 secs. the required grip exceeds that available and the front slides out. This is the final limitation on the maximum turn-in performance possible.
Reynolds, an early investigator into fluid flow, discovered that the turbulence characteristics of simila shaped but different sized objects was dependent on their size and air speed. In fact, if you halve the size of your object you need to double the air speed to achieve similar air flow patterns. This effect wa: formalised into a mathematical expression known as the Reynolds number. In order to achieve simila flow patterns over different size models we need to try and keep this number constant, but in man cases this is very difficult. Assume that we wish to investigate the aerodynamics of a motorcycle at 151 km/h. by using a sixth scale model in a wind tunnel. This would then demand (for a similar Reynold No.) that the tunnel air speed be approximately 900 km/h., which immediately introduces yet mor problems. Immense power would be needed to create this flow rate, and the forces on the mode would be extremely high, creating mounting difficulties. In fact, at the same Reynolds No. the drag force on the model would be equal to the drag force on the real thing. As if this was not bad enough, 90\ km/h. is very close to sonic velocity, the speed of sound. As we get closer to this speed compressibilit problems take over and disrupt the whole flow pattern. The end result of this is that small scale mode tests have to be done at low Reynolds numbers and much skill and care is needed in thei interpretation. In the future it is likely that computer based mathematical models will become goo enough to substitute for a wind tunnel in many cases. This is development is known as CFD o Computational Fluid Dynamics.
Reynolds, an early investigator into fluid flow, discovered that the turbulence characteristics of simila shaped but different sized objects was dependent on their size and air speed. In fact, if you halve the size of your object you need to double the air speed to achieve similar air flow patterns. This effect wa: formalised into a mathematical expression known as the Reynolds number. In order to achieve simila flow patterns over different size models we need to try and keep this number constant, but in man cases this is very difficult. Assume that we wish to investigate the aerodynamics of a motorcycle at 151 km/h. by using a sixth scale model in a wind tunnel. This would then demand (for a similar Reynold No.) that the tunnel air speed be approximately 900 km/h., which immediately introduces yet mor problems. Immense power would be needed to create this flow rate, and the forces on the mode would be extremely high, creating mounting difficulties. In fact, at the same Reynolds No. the drag force on the model would be equal to the drag force on the real thing. As if this was not bad enough, 90\ km/h. is very close to sonic velocity, the speed of sound. As we get closer to this speed compressibilit problems take over and disrupt the whole flow pattern. The end result of this is that small scale mode tests have to be done at low Reynolds numbers and much skill and care is needed in thei interpretation. In the future it is likely that computer based mathematical models will become goo enough to substitute for a wind tunnel in many cases. This is development is known as CFD o Computational Fluid Dynamics.
The value of reduced drag on everyday motorcycles was clearly shown back in 1956 when "Motor Cycle" magazine commissioned Laurie Watts, the technical artist, to design a semi-enclosing streamlined body shell to be built around a then five year old 350 cc. Royal Enfield. This resulted in a 45 Ibf. (20.4 kgf.) increase in weight, but, such was the value of lower drag that even acceleration in the 15 — 40 mph. (24 — 64 km/h.) range was significantly improved, as the table of performance in fig. 5.4 shows. These tests were carried out by the late Vic Willoughby at MIRA using electronic timing gear. Trying to reconcile these results into a percentage reduction in drag has its problems, largely because all tests were done with unchanged gearing. Thus, the engine RPM. and hence the BHP. produced would have been different during the maximum speed runs, depending on whether the bodywork was fitted or not. The engine power being produced at top speed may have been more with the fairing fitted as the machine was normally over-geared. However, by making one or two assumptions, | have
The value of reduced drag on everyday motorcycles was clearly shown back in 1956 when "Motor Cycle" magazine commissioned Laurie Watts, the technical artist, to design a semi-enclosing streamlined body shell to be built around a then five year old 350 cc. Royal Enfield. This resulted in a 45 Ibf. (20.4 kgf.) increase in weight, but, such was the value of lower drag that even acceleration in the 15 — 40 mph. (24 — 64 km/h.) range was significantly improved, as the table of performance in fig. 5.4 shows. These tests were carried out by the late Vic Willoughby at MIRA using electronic timing gear. Trying to reconcile these results into a percentage reduction in drag has its problems, largely because all tests were done with unchanged gearing. Thus, the engine RPM. and hence the BHP. produced would have been different during the maximum speed runs, depending on whether the bodywork was fitted or not. The engine power being produced at top speed may have been more with the fairing fitted as the machine was normally over-geared. However, by making one or two assumptions, | have
Left: Vic Willoughby testing the Motor Cycle magazine sponsored Royal Enfield “Dreamliner”.
Left: Vic Willoughby testing the Motor Cycle magazine sponsored Royal Enfield “Dreamliner”.
Fig 5.4 Table of performance for Royal Enfield “Dreamliner”. A very low performance machine by today’s standards but still dramatically illustrates the performance benefits of lower drag. Despite the extra weight the acceleration figures also show an improvement. calculated that the drag with the body fitted was probably in the range of 75 -- 85% of the unfaired machine. So, if this order of performance improvement is possible with such a body, just imagine what could be achieved with a purpose built device, -- a well streamlined FF. for example, with a smaller frontal area as well as the possibility of a better shape.
Fig 5.4 Table of performance for Royal Enfield “Dreamliner”. A very low performance machine by today’s standards but still dramatically illustrates the performance benefits of lower drag. Despite the extra weight the acceleration figures also show an improvement. calculated that the drag with the body fitted was probably in the range of 75 -- 85% of the unfaired machine. So, if this order of performance improvement is possible with such a body, just imagine what could be achieved with a purpose built device, -- a well streamlined FF. for example, with a smaller frontal area as well as the possibility of a better shape.
H.P. (Happy) Mueller streaks across the Utah salt at 240 km/h in the 125 cc NSU “flying hammock” in 1956. Probably one of the most aerodynamic motorcycles built. The tail fin was only fitted for runs above about 240 km/h. (MCW) So what shape is best from the drag view point? Well, the classic tear drop shape as shown in fig. 5.1 would be hard to beat, and an approximation to this was used very successfully by NSU to take many world speed records in the smaller classes. There were plans to use shorter versions of these streamliners in GP racing, and later on road machines, but the FIM changed the rules in 1957 and outlawed all but the most rudimentary of fairings. A legacy we are still left with today. The classic tear drop has a circular frontal cross section but when placed near the ground as for a road vehicle, the ground proximity interferes with the symmetrical airflow pattern and so the ideal shape would need modifying to account for this. This interference with the airflow is also most likely to introduce a lift force.
H.P. (Happy) Mueller streaks across the Utah salt at 240 km/h in the 125 cc NSU “flying hammock” in 1956. Probably one of the most aerodynamic motorcycles built. The tail fin was only fitted for runs above about 240 km/h. (MCW) So what shape is best from the drag view point? Well, the classic tear drop shape as shown in fig. 5.1 would be hard to beat, and an approximation to this was used very successfully by NSU to take many world speed records in the smaller classes. There were plans to use shorter versions of these streamliners in GP racing, and later on road machines, but the FIM changed the rules in 1957 and outlawed all but the most rudimentary of fairings. A legacy we are still left with today. The classic tear drop has a circular frontal cross section but when placed near the ground as for a road vehicle, the ground proximity interferes with the symmetrical airflow pattern and so the ideal shape would need modifying to account for this. This interference with the airflow is also most likely to introduce a lift force.
Although it never set a fashion for racing, this 1953 kneeler Norton was used to increase the one hour record to 134 miles (216 km.) at the Montlhery circuit with Ray Amm aboard. It is not hard to see how the term “dolphin fairing” originated. This 1954 NSU Rennmax is a direct ancestor of today’s streamlining. There can be little doubt that the nose gave rise to a lot of lift. Evolution of the racing fairing.
Although it never set a fashion for racing, this 1953 kneeler Norton was used to increase the one hour record to 134 miles (216 km.) at the Montlhery circuit with Ray Amm aboard. It is not hard to see how the term “dolphin fairing” originated. This 1954 NSU Rennmax is a direct ancestor of today’s streamlining. There can be little doubt that the nose gave rise to a lot of lift. Evolution of the racing fairing.
With their own full size wind tunnel the Carcano designed Moto Guzzis of the 1950s. probably had the most efficient “dustbin” fairings of the period. Shown here is a 1957 500 cc. single cylinder at the TT with Keith Campbell aboard. Its 240 km/h. top speed from just 47 bhp. being testament to the quality of the aerodynamics. Whilst full frontal fairings were popular in all capacity classes in the 1950s. Tail fairings were rarer, generally being fitted to the 125 cc. machines, which needed all the help they could get. Slipstreaming by following competitors was rendered less effective because of the reduced wake. This is multi-world-champion Carlo Ubbiali winning the 1956 Belgian GP on a 125 cc MV Agusta.
With their own full size wind tunnel the Carcano designed Moto Guzzis of the 1950s. probably had the most efficient “dustbin” fairings of the period. Shown here is a 1957 500 cc. single cylinder at the TT with Keith Campbell aboard. Its 240 km/h. top speed from just 47 bhp. being testament to the quality of the aerodynamics. Whilst full frontal fairings were popular in all capacity classes in the 1950s. Tail fairings were rarer, generally being fitted to the 125 cc. machines, which needed all the help they could get. Slipstreaming by following competitors was rendered less effective because of the reduced wake. This is multi-world-champion Carlo Ubbiali winning the 1956 Belgian GP on a 125 cc MV Agusta.
The aerodynamics of this 1999 125 cc. Aprilia are only different in detail to that of the older Honda on the left. The belly pan is narrower and fills almost completely the gap between the two wheels. Note also the front mudguard which is faired in to blend the air over the front forks. At the rear the seat back is larger and higher to help smooth the flow as it leaves Jeronimo Vidal’s back. After the 1957 FIM ban on certain types of fairing, the “dolphin” shape quickly became the norm by the end of the 50s and early 60s. This shot of the great Bob Mcintyre on a 250 Honda in the 1961 TT shows the basic form that has continued until the present time.
The aerodynamics of this 1999 125 cc. Aprilia are only different in detail to that of the older Honda on the left. The belly pan is narrower and fills almost completely the gap between the two wheels. Note also the front mudguard which is faired in to blend the air over the front forks. At the rear the seat back is larger and higher to help smooth the flow as it leaves Jeronimo Vidal’s back. After the 1957 FIM ban on certain types of fairing, the “dolphin” shape quickly became the norm by the end of the 50s and early 60s. This shot of the great Bob Mcintyre on a 250 Honda in the 1961 TT shows the basic form that has continued until the present time.
Fig. 5.5 Traces from the data acquisition system of a 250 GP racer. The horizontal axis shows elapsed time in seconds the upper curve shows velocity in km/h and the lower curve shows the airbox pressure in milli-bars. The inset shows th same data re-plotted against speed. Note how the real data is less than the theoretical maximum.
Fig. 5.5 Traces from the data acquisition system of a 250 GP racer. The horizontal axis shows elapsed time in seconds the upper curve shows velocity in km/h and the lower curve shows the airbox pressure in milli-bars. The inset shows th same data re-plotted against speed. Note how the real data is less than the theoretical maximum.
Fig 5.7 The drag force Fd acts at height h to create a moment Fd.h tending to unload the front wheel and load the rear. If the lift CP is in front of the CoG then this lift force will cause a greater reduction of load on the front compared to the back, and vice versa. It is not always realised, but even with a body shape that produces no nett lifting force (only drag) nor any direct aerodynamic pitching moment there is still a tendency for load to transfer off the front wheel and onto the rear due entirely to the effect of the drag force, as fig. 5.7 shows.
Fig 5.7 The drag force Fd acts at height h to create a moment Fd.h tending to unload the front wheel and load the rear. If the lift CP is in front of the CoG then this lift force will cause a greater reduction of load on the front compared to the back, and vice versa. It is not always realised, but even with a body shape that produces no nett lifting force (only drag) nor any direct aerodynamic pitching moment there is still a tendency for load to transfer off the front wheel and onto the rear due entirely to the effect of the drag force, as fig. 5.7 shows.
Fig 5.8 Drag induced reduction of load on front wheel at different air speeds. Shown for 2 static front wheel weights, for an example motorcycle with a Cd of 0.7 and a frontal area of 4000 sq.cm.
Fig 5.8 Drag induced reduction of load on front wheel at different air speeds. Shown for 2 static front wheel weights, for an example motorcycle with a Cd of 0.7 and a frontal area of 4000 sq.cm.
Originally led by BMW. with their R100 RS. fairing, and also on some racing Suzukis, this problem of cancelling the lift is now receiving more attention from the manufacturers.
Originally led by BMW. with their R100 RS. fairing, and also on some racing Suzukis, this problem of cancelling the lift is now receiving more attention from the manufacturers.
Fig. 5.9 Small winglets on this GP Suzuki were designed to provide some measure of downforce near the front to counter the tendency to lift the front end at speed. (MCW)
Fig. 5.9 Small winglets on this GP Suzuki were designed to provide some measure of downforce near the front to counter the tendency to lift the front end at speed. (MCW)
Fig 5.10 A fixed wing across the bike just increases the horizontal and vertical components acting on the tyres in the same proportion and hence there is no nett gain in cornering performance. Probably the reverse. Fig 5.11 In theory, a tilting wing that remained horizontal could enhance braking, traction and cornering power. As a side effect, a smaller angle of lean would be required for a given cornering speed.
Fig 5.10 A fixed wing across the bike just increases the horizontal and vertical components acting on the tyres in the same proportion and hence there is no nett gain in cornering performance. Probably the reverse. Fig 5.11 In theory, a tilting wing that remained horizontal could enhance braking, traction and cornering power. As a side effect, a smaller angle of lean would be required for a given cornering speed.
Whilst ultimate aerodynamic efficiency was not the prime goal of the author’s QL, it was desired to improve upon a standard machine within the bounds of practicality. The photos above right and lower left show a 1/6 scale model in a wind-tunnel. Measurements from this model were compared with a 1/6 scale model of a_ standard R90s BMW. This comparison indicated that the QL had 65% of the drag of the standard bike. Whilst the low Reynold’s numbers may have caused relative differences in the separation points it was clear that drag was significantly lower. Road tests later confirmed this as well as_ excellent weather protection and stability.
Whilst ultimate aerodynamic efficiency was not the prime goal of the author’s QL, it was desired to improve upon a standard machine within the bounds of practicality. The photos above right and lower left show a 1/6 scale model in a wind-tunnel. Measurements from this model were compared with a 1/6 scale model of a_ standard R90s BMW. This comparison indicated that the QL had 65% of the drag of the standard bike. Whilst the low Reynold’s numbers may have caused relative differences in the separation points it was clear that drag was significantly lower. Road tests later confirmed this as well as_ excellent weather protection and stability.
As can be seen from fig. 5.11 the moment due to the downforce helps counteract the centrifugal moment. Hence the moment due to the machine’s weight must be reduced — which means a smaller angle of lean. However, any benefits from such a device might well be outweighed by practical and stability problems. Ina straight line, any aerodynamic features that create extra down force will load the tyres more and make available more traction. Provided that our brakes are good enough and that we have sufficient power, this greater grip will allow for better acceleration and braking.
As can be seen from fig. 5.11 the moment due to the downforce helps counteract the centrifugal moment. Hence the moment due to the machine’s weight must be reduced — which means a smaller angle of lean. However, any benefits from such a device might well be outweighed by practical and stability problems. Ina straight line, any aerodynamic features that create extra down force will load the tyres more and make available more traction. Provided that our brakes are good enough and that we have sufficient power, this greater grip will allow for better acceleration and braking.
The NSU record breaker being tested in the full size wind tunnel at the Stuttgart Technical College. The size of such installations is quite apparent from this photograph. Compare this to the much smaller tunnel used by the author to test the 1/6". scale model of the QL, pictured earlier.
The NSU record breaker being tested in the full size wind tunnel at the Stuttgart Technical College. The size of such installations is quite apparent from this photograph. Compare this to the much smaller tunnel used by the author to test the 1/6". scale model of the QL, pictured earlier.
CFD plots of the pressure distribution on bike and rider. The shading indicates the static pressure, as per the central scale. Such plots are normally in colour which highlights the differences much better than these grey scale images. The right hand sketch is also marked up with iso-bars, or lines of constant pressure. (Advantage CFD, Reynard Motorsport Ltd) Large cost savings and rapid development times are the promise of such techniques. There is no need to build physical models, these are just defined from the CAD drawing of the vehicle. To iterate through a wide range of alternative designs needs only a few surfaces on a drawing to be changed, and the computer run repeated.
CFD plots of the pressure distribution on bike and rider. The shading indicates the static pressure, as per the central scale. Such plots are normally in colour which highlights the differences much better than these grey scale images. The right hand sketch is also marked up with iso-bars, or lines of constant pressure. (Advantage CFD, Reynard Motorsport Ltd) Large cost savings and rapid development times are the promise of such techniques. There is no need to build physical models, these are just defined from the CAD drawing of the vehicle. To iterate through a wide range of alternative designs needs only a few surfaces on a drawing to be changed, and the computer run repeated.
A 2D representation of the CFD grid of cells around the bike and rider. In reality the cells are 3D and fill the whole volume around the machine, literally millions of cells are necessary to model the flow successfully. (Advantage CFD, Reynard Motorsport Ltd) At the moment computer technology is at a level which still demands that quite extensive computer facilities are necessary for large models. The current high-end stand alone desktop computer being suitable for smaller problems such as the air flow inside air-boxes. Future generations of desktop computers will have more power and undoubtedly that will make CFD techniques much more accessible as has been the case with FEA (Finite Element Analysis) for structural design. Nearly forty years ago the author worked with FEA, but at that time such techniques were solely available for specialists with access to large main-frame computers, and required that the user had a high level of skill in the techniques. Nowadays, such methods are available to any designer with a desktop computer and any mid to high level CAD programme. CFD is mathematically similar to structural FEA, in fact it is really the FEA of fluid flow, but the big differences are the vastly greater number of elements or cells needed to adequately model a fluid flow problem, and more importantly the greater complexity of the non-linear equations necessary.
A 2D representation of the CFD grid of cells around the bike and rider. In reality the cells are 3D and fill the whole volume around the machine, literally millions of cells are necessary to model the flow successfully. (Advantage CFD, Reynard Motorsport Ltd) At the moment computer technology is at a level which still demands that quite extensive computer facilities are necessary for large models. The current high-end stand alone desktop computer being suitable for smaller problems such as the air flow inside air-boxes. Future generations of desktop computers will have more power and undoubtedly that will make CFD techniques much more accessible as has been the case with FEA (Finite Element Analysis) for structural design. Nearly forty years ago the author worked with FEA, but at that time such techniques were solely available for specialists with access to large main-frame computers, and required that the user had a high level of skill in the techniques. Nowadays, such methods are available to any designer with a desktop computer and any mid to high level CAD programme. CFD is mathematically similar to structural FEA, in fact it is really the FEA of fluid flow, but the big differences are the vastly greater number of elements or cells needed to adequately model a fluid flow problem, and more importantly the greater complexity of the non-linear equations necessary.
These CFD output illustrations show the size of the disturbance at two imaginary vertical slices through the air. The first is some distance behind the bike and the other is immediately behind the seat. Simply put these show the size of the hole punched in the air by the bike and rider. (Advantage CFD, Reynard Motorsport Ltd)
These CFD output illustrations show the size of the disturbance at two imaginary vertical slices through the air. The first is some distance behind the bike and the other is immediately behind the seat. Simply put these show the size of the hole punched in the air by the bike and rider. (Advantage CFD, Reynard Motorsport Ltd)
Two different types of CFD flow visualization. On the left we see overall streamlines and on the right we see what’s happening right on the surface of some parts of the rider and bike. Visualizations like this can be produced for any viewing angle. (Advantage CFD, Reynard Motorsport Ltd)
Two different types of CFD flow visualization. On the left we see overall streamlines and on the right we see what’s happening right on the surface of some parts of the rider and bike. Visualizations like this can be produced for any viewing angle. (Advantage CFD, Reynard Motorsport Ltd)
Fig 5.12 Adding the effects of the forward speed of 120 km/h and the side wind of 25 km/h results in a total equivalent air flow of 123 km/h acting at an angle of about 12 degrees. Fig 5.13 Shows how a well streamlined shape such as an aircraft's wing produces a side or lift force, whereas the brick shaped object is less affected by the inclined air stream. airflow at an angle to the machine. e.g. If the bike is travelling at 120 km/h. with a side wind of 25 km/h. then the effective breeze will be at 123 km/h. acting at an angle of about 12°.
Fig 5.12 Adding the effects of the forward speed of 120 km/h and the side wind of 25 km/h results in a total equivalent air flow of 123 km/h acting at an angle of about 12 degrees. Fig 5.13 Shows how a well streamlined shape such as an aircraft's wing produces a side or lift force, whereas the brick shaped object is less affected by the inclined air stream. airflow at an angle to the machine. e.g. If the bike is travelling at 120 km/h. with a side wind of 25 km/h. then the effective breeze will be at 123 km/h. acting at an angle of about 12°.
Let’s now consider how some of the realities, of motorcycle steering, control the reactions to a side force such as that produced by a side wind. (The author wishes to thank Douglas Milliken and Dr. Andreas Fuchs for providing the stimulus to look closer into these aspects.)
Let’s now consider how some of the realities, of motorcycle steering, control the reactions to a side force such as that produced by a side wind. (The author wishes to thank Douglas Milliken and Dr. Andreas Fuchs for providing the stimulus to look closer into these aspects.)
To prevent the side wind from gradually blowing the bike off course, it must be steered into the wind to compensate. Fig. 5.16. The purpose of this experiment is to demonstrate that a side force, such as wind, can introduce steering tendencies and that their magnitude and direction depend on the point of application of the side force. This has important implications for the behaviour of a motorcycle subject to a side wind. Steady state means riding along with a uniform side wind, which thus creates a constant side force acting through the lateral centre of pressure. Clearly, this causes a roll couple, which must be balanced by leaning into the wind. To minimize the angle of lean required to balance a given side force, we need a low centre of pressure, a high centre of gravity and a large weight. Fig. 5.15.
To prevent the side wind from gradually blowing the bike off course, it must be steered into the wind to compensate. Fig. 5.16. The purpose of this experiment is to demonstrate that a side force, such as wind, can introduce steering tendencies and that their magnitude and direction depend on the point of application of the side force. This has important implications for the behaviour of a motorcycle subject to a side wind. Steady state means riding along with a uniform side wind, which thus creates a constant side force acting through the lateral centre of pressure. Clearly, this causes a roll couple, which must be balanced by leaning into the wind. To minimize the angle of lean required to balance a given side force, we need a low centre of pressure, a high centre of gravity and a large weight. Fig. 5.15.
The main difference to note in this case is that the position for zero steer is higher up the bar than in the previous case. If you do this again with the bar mounted towards the rear you'll find that the rope location for zero steer is much lower.
The main difference to note in this case is that the position for zero steer is higher up the bar than in the previous case. If you do this again with the bar mounted towards the rear you'll find that the rope location for zero steer is much lower.
Fig. 5.18 With a given side wind force (Fw) and weight (Wt) the angle of lean is less when the CP is below the CoG. As the lean angle increases the component of the weight acting at right angles to the bike also increases but the normal component of the wind force will decrease. When the CP and CoG are at the same height these normal forces balance.
Fig. 5.18 With a given side wind force (Fw) and weight (Wt) the angle of lean is less when the CP is below the CoG. As the lean angle increases the component of the weight acting at right angles to the bike also increases but the normal component of the wind force will decrease. When the CP and CoG are at the same height these normal forces balance.
Fig. 5.19 shows how any imbalance in these normal forces will cause a steering effect, this is completely in accord with the results of our experiment which showed that this effect varied with where we applied the horizontal force. Figs. 5.18 & 5.19 implicitly assume that the fore and aft location of the CP is the same as that of the CoG, but we also noticed with our tests that this longitudinal position has an effect or the CP height required to give zero steering action. Basically it is the yaw moments that we must balance not the lateral forces. Some simple calculations show that the point of application of a side force which gives no steering action must lie on a line drawn through the rear tyre contact patch and the CoG. as shown in fig. 5.20.
Fig. 5.19 shows how any imbalance in these normal forces will cause a steering effect, this is completely in accord with the results of our experiment which showed that this effect varied with where we applied the horizontal force. Figs. 5.18 & 5.19 implicitly assume that the fore and aft location of the CP is the same as that of the CoG, but we also noticed with our tests that this longitudinal position has an effect or the CP height required to give zero steering action. Basically it is the yaw moments that we must balance not the lateral forces. Some simple calculations show that the point of application of a side force which gives no steering action must lie on a line drawn through the rear tyre contact patch and the CoG. as shown in fig. 5.20.
This is the Honda Gold Wing mentioned in the above text. Note the small 16” wheel at the front. For a number of reasons this machine had improved stability in gusting wind conditions. (photo: Kerry Dunlop) Zero offset between the steering axis and wheel axle. With normal steering geometry with about 25 — 50 mm. of offset, the bulk of the wheel side area is forward of the steering axis, this combined with the effect of today's large tyres and discs means that a considerable steering force can be generated by the action of a side wind on the wheel. But with the zero offset geometry used on the double arrangement, this wind force is balanced about the steering axis and no turning effect is produced.
This is the Honda Gold Wing mentioned in the above text. Note the small 16” wheel at the front. For a number of reasons this machine had improved stability in gusting wind conditions. (photo: Kerry Dunlop) Zero offset between the steering axis and wheel axle. With normal steering geometry with about 25 — 50 mm. of offset, the bulk of the wheel side area is forward of the steering axis, this combined with the effect of today's large tyres and discs means that a considerable steering force can be generated by the action of a side wind on the wheel. But with the zero offset geometry used on the double arrangement, this wind force is balanced about the steering axis and no turning effect is produced.
Coil Springs Coil springs in steel are the most common by a long way. They may be evenly wound (constant pitch) t give a linear rate, or they may be wound with a varying pitch to give a progressive rate. In this case a the spring is gradually compressed, the closer wound coils become "coil-bound" (i.e. touch each othe and act as a solid mass) and so the rate rises.
Coil Springs Coil springs in steel are the most common by a long way. They may be evenly wound (constant pitch) t give a linear rate, or they may be wound with a varying pitch to give a progressive rate. In this case a the spring is gradually compressed, the closer wound coils become "coil-bound" (i.e. touch each othe and act as a solid mass) and so the rate rises.
In some Hagon telescopic forks, designed for grass track racing, rubber bands provided the springing medium. This provided easy adjustment by adding or removing bands as required. Rubber has a natural progressive rate, and it also provides a small amount of inherent damping, though this generates heat Rubber has many properties that make it an interesting spring material, and it can and has been used in a variety ways. It is not suitable as a direct replacement material to make a coil spring. It must be used in an appropriate design. Greeves used it in the form of large bonded bushes that served as both the pivot bearings and the springing medium in their leading link front forks. The rubber itself was loaded in shear, though the bush as a whole was loaded in torsion.
In some Hagon telescopic forks, designed for grass track racing, rubber bands provided the springing medium. This provided easy adjustment by adding or removing bands as required. Rubber has a natural progressive rate, and it also provides a small amount of inherent damping, though this generates heat Rubber has many properties that make it an interesting spring material, and it can and has been used in a variety ways. It is not suitable as a direct replacement material to make a coil spring. It must be used in an appropriate design. Greeves used it in the form of large bonded bushes that served as both the pivot bearings and the springing medium in their leading link front forks. The rubber itself was loaded in shear, though the bush as a whole was loaded in torsion.
Fig. 6.3 At full extension and with 10 bar pressure in each, the two units can support the same load. However at full displacement the one with a compression ratio of 5:1 now has a pressure of 50 bar and can support more load than the other which has only risen to 30 bar within the same suspension stroke. The only difference between these otherwise identical struts is that the lower compression one has an additional 25 cc. of internal volume. The extent of the progression in rate is determined by the compression ratio of the unit (i.e. the ratio of the gas volume at the two extremes of travel). Fig 6.3 shows this variation between two units that start off by supporting the same load.
Fig. 6.3 At full extension and with 10 bar pressure in each, the two units can support the same load. However at full displacement the one with a compression ratio of 5:1 now has a pressure of 50 bar and can support more load than the other which has only risen to 30 bar within the same suspension stroke. The only difference between these otherwise identical struts is that the lower compression one has an additional 25 cc. of internal volume. The extent of the progression in rate is determined by the compression ratio of the unit (i.e. the ratio of the gas volume at the two extremes of travel). Fig 6.3 shows this variation between two units that start off by supporting the same load.
Fournales gas suspension strut. Despite the apparent simplicity of construction, shocks like this must be made to very high standards in order to work at all. The preload/ride-height adjustment found on normal coil spring units, does not have any effect on the spring rate, only on the initial load capacity. Such suspension will show an increased tendency to bottom out when heavily laden, unlike the adjusted gas shock.
Fournales gas suspension strut. Despite the apparent simplicity of construction, shocks like this must be made to very high standards in order to work at all. The preload/ride-height adjustment found on normal coil spring units, does not have any effect on the spring rate, only on the initial load capacity. Such suspension will show an increased tendency to bottom out when heavily laden, unlike the adjusted gas shock.
Fig. 6.4 Load vs. displacement curves for a typical Fournales gas spring unit. The “natural” air spring graph (line 3) has been modified up to a stroke of Te by the addition of a “top-out” spring as shown by line 4. increased too. If he took the opposite approach and added some gas, then the rate might well be decreased as desired but ride height will be increased. If too much gas is added then the preload may become so high that a bump is needed to even begin to compress the suspension, so negating the desired effect. Other than buying another unit with different characteristics, there is little that can be done. Some units allow for the addition or removal of small quantities of oil, this alters the internal volume and hence the effective spring rate, but the degree of progression will be changed also, because the compression ratio will have been altered.
Fig. 6.4 Load vs. displacement curves for a typical Fournales gas spring unit. The “natural” air spring graph (line 3) has been modified up to a stroke of Te by the addition of a “top-out” spring as shown by line 4. increased too. If he took the opposite approach and added some gas, then the rate might well be decreased as desired but ride height will be increased. If too much gas is added then the preload may become so high that a bump is needed to even begin to compress the suspension, so negating the desired effect. Other than buying another unit with different characteristics, there is little that can be done. Some units allow for the addition or removal of small quantities of oil, this alters the internal volume and hence the effective spring rate, but the degree of progression will be changed also, because the compression ratio will have been altered.
Fig. 6.5 From a 1996 patent by Ohlins and Yamaha this sketch shows an idea for a composite (fibre-glass) leaf spring fitted inside the actual swing-arm. As the swing-arm rises the actuating roller forces the leaf to bend over the support cam. The detail of the shape of the support cam surface controls the active length of the spring and hence the overall spring rate throughout the range of movement.
Fig. 6.5 From a 1996 patent by Ohlins and Yamaha this sketch shows an idea for a composite (fibre-glass) leaf spring fitted inside the actual swing-arm. As the swing-arm rises the actuating roller forces the leaf to bend over the support cam. The detail of the shape of the support cam surface controls the active length of the spring and hence the overall spring rate throughout the range of movement.
Fig. 6.7 Diagrammatic representation of a damper piston with shim stack for the control of the mid to high speed damping. The left hand sketch is simplified to show only the rebound shim stack. The figure is split in two to show the shim positions for the low and high speed cases. At low damper rod velocities the main flow is forced through the bypass passages, but as the rod velocity increases the pressure builds up enough to open the shim flow valves preventing excessive damper forces. The sketch to the right shows how two shim stacks can be used on the same damper piston to control both bump and rebound damping. At a certain point the shim stack will reach its maximum opening area and the damping will become dependent on the velocity squared again (V2 in fig. 6.6). The flow rate through the shims may also be limited by the serial flow passages leading to the stack with similar, but not identical, results. Sometimes an adjustable valve is put in series with the shim stack to control high speed damping.
Fig. 6.7 Diagrammatic representation of a damper piston with shim stack for the control of the mid to high speed damping. The left hand sketch is simplified to show only the rebound shim stack. The figure is split in two to show the shim positions for the low and high speed cases. At low damper rod velocities the main flow is forced through the bypass passages, but as the rod velocity increases the pressure builds up enough to open the shim flow valves preventing excessive damper forces. The sketch to the right shows how two shim stacks can be used on the same damper piston to control both bump and rebound damping. At a certain point the shim stack will reach its maximum opening area and the damping will become dependent on the velocity squared again (V2 in fig. 6.6). The flow rate through the shims may also be limited by the serial flow passages leading to the stack with similar, but not identical, results. Sometimes an adjustable valve is put in series with the shim stack to control high speed damping.
Fig 6.8 The effects of different adjustment possibilities on the shape of the damping characteristics.
Fig 6.8 The effects of different adjustment possibilities on the shape of the damping characteristics.
A hydraulic steering damper, with the piston rod extended to obviate the displacement problem. A typical installation. The MV Agusta F4 mounts the steering damper across the frame just behind the top fork yoke. Unusually the damper rod is mounted to the bike at both ends and the fork clamp causes the cylinder to move sideways through a ball jointed linkage. Many ingenious solutions to this problem have been tried over the years. One simple way to bypass the difficulty is to extend the rod through both ends of the cylinder so that there is no change in internal volume with piston movement. This does nothing to compensate for volume alterations due to temperature change, or just plain old fashioned fluid leakage, and this method is not currently used for suspension damping but is almost universally used for steering dampers.
A hydraulic steering damper, with the piston rod extended to obviate the displacement problem. A typical installation. The MV Agusta F4 mounts the steering damper across the frame just behind the top fork yoke. Unusually the damper rod is mounted to the bike at both ends and the fork clamp causes the cylinder to move sideways through a ball jointed linkage. Many ingenious solutions to this problem have been tried over the years. One simple way to bypass the difficulty is to extend the rod through both ends of the cylinder so that there is no change in internal volume with piston movement. This does nothing to compensate for volume alterations due to temperature change, or just plain old fashioned fluid leakage, and this method is not currently used for suspension damping but is almost universally used for steering dampers.
Current design trends have been influenced by the DeCARBON principal, where the oil and gas are separated by a floating piston, either in the unit body or in a separate remote chamber. The gas pressure (up to 300 psi.) keeps all the seals pressurized to reduce leakage and to stabilize the damping and prevent cavitation, however rapid the recoil stroke. On some dampers of this type there is provision for adjusting the pressure as a fine tuning aid.
Current design trends have been influenced by the DeCARBON principal, where the oil and gas are separated by a floating piston, either in the unit body or in a separate remote chamber. The gas pressure (up to 300 psi.) keeps all the seals pressurized to reduce leakage and to stabilize the damping and prevent cavitation, however rapid the recoil stroke. On some dampers of this type there is provision for adjusting the pressure as a fine tuning aid.
In this DeCarbon “gas shock” the displacement of the piston rod is accommodated by a _ high-pressure nitrogen chamber separated from the hydraulic fluid by a free-floating piston. Spring preload and damping are both adjustable To prevent aeration of the hydraulic fluid, the compressible element (Arcton gas) in this Girling strut was contained in a nylon cell
In this DeCarbon “gas shock” the displacement of the piston rod is accommodated by a _ high-pressure nitrogen chamber separated from the hydraulic fluid by a free-floating piston. Spring preload and damping are both adjustable To prevent aeration of the hydraulic fluid, the compressible element (Arcton gas) in this Girling strut was contained in a nylon cell
Fig. 6.10 Yamaha-Ohlins rotary damper. As the oscillating piston or vane moves one way or the other then oil must be forced through one of the two damper valves, which can incorporate all the sophistication appropriate for the expected use. Leakage between the vane and the housing would need to be kept to an absolute minimum, to avoid uncontrolled influence over the damping forces.
Fig. 6.10 Yamaha-Ohlins rotary damper. As the oscillating piston or vane moves one way or the other then oil must be forced through one of the two damper valves, which can incorporate all the sophistication appropriate for the expected use. Leakage between the vane and the housing would need to be kept to an absolute minimum, to avoid uncontrolled influence over the damping forces.
This rotary actuated steering damper was used by Yamaha on some early racing models. An orifice plate was moved linearly by means of a quick start thread. Suzuki fitted this rotary damper to the rear of the TL1000 L road bike. Aftermarket suppliers offered normal linear struts as a substitution.
This rotary actuated steering damper was used by Yamaha on some early racing models. An orifice plate was moved linearly by means of a quick start thread. Suzuki fitted this rotary damper to the rear of the TL1000 L road bike. Aftermarket suppliers offered normal linear struts as a substitution.
Fig. 6.12 With equal damping in both directions the CoG oscillates about its static ride height when travelling rapidly over a corrugated road. The lower curve shows how, in this simulation with all the damping in rebound, the suspension jacks down. In some cases this effect can be severe enough to compress right down to the bump stops. Note also that the magnitude of the jacking down seems out of all proportion to the small magnitude of the oscillatory motion. This approach has several disadvantages, the reasoning may have been sound from the comfort aspec' when negotiating a single bump but could give rise to a serious problem when crossing a corrugated surface. Each bump compresses the suspension quickly (because of the absence of damping) while the subsequent recoil stroke is slowed by the heavy damping. This may prevent the wheel from returning tc its static position before the next bump. The rapid repetition of this action may soon ratchet the suspension into a fully compressed state, so giving the effect of a solid frame, fig. 6.12.
Fig. 6.12 With equal damping in both directions the CoG oscillates about its static ride height when travelling rapidly over a corrugated road. The lower curve shows how, in this simulation with all the damping in rebound, the suspension jacks down. In some cases this effect can be severe enough to compress right down to the bump stops. Note also that the magnitude of the jacking down seems out of all proportion to the small magnitude of the oscillatory motion. This approach has several disadvantages, the reasoning may have been sound from the comfort aspec' when negotiating a single bump but could give rise to a serious problem when crossing a corrugated surface. Each bump compresses the suspension quickly (because of the absence of damping) while the subsequent recoil stroke is slowed by the heavy damping. This may prevent the wheel from returning tc its static position before the next bump. The rapid repetition of this action may soon ratchet the suspension into a fully compressed state, so giving the effect of a solid frame, fig. 6.12.
Fig. 6.14 Simulation of the vertical front tyre / ground force as a bike passes over a 0.05 metre drop at 100 km/h. The dark line is with rebound damping only and the other is with bump damping only. Ultimate road holding requires the minimum variation in this force. This force can never go negative, because zero represents the point at which the tyre leaves the ground. Fig. 6.14 shows the response to a drop or step down, which is the opposite case to that of fig. 6.13. In this case the ground to tyre contact force is plotted as this shows some road holding implications. A force of zero indicates that the tyre has lost contact with the ground. Studying the first part of the curves, between 0.1 seconds ( when the wheel reaches the step ) and 0.16 seconds, shows that in the case with only bump damping, the wheel is airborne for a shorter period, as expected. However, when we look farther along the time axis we see that the overall force fluctuations are greater and that the front wheel becomes airborne once again, the total time spent with no contact is actually greater than in the other case, with rebound damping only. With only bump damping, when the wheel hits the ground at about 0.14 secs. the higher bump suspension forces cause a heavier landing, shown by the higher peak tyre force, which in turn has caused more tyre bounce thereby lifting the wheel off the ground again at 0.23 seconds.
Fig. 6.14 Simulation of the vertical front tyre / ground force as a bike passes over a 0.05 metre drop at 100 km/h. The dark line is with rebound damping only and the other is with bump damping only. Ultimate road holding requires the minimum variation in this force. This force can never go negative, because zero represents the point at which the tyre leaves the ground. Fig. 6.14 shows the response to a drop or step down, which is the opposite case to that of fig. 6.13. In this case the ground to tyre contact force is plotted as this shows some road holding implications. A force of zero indicates that the tyre has lost contact with the ground. Studying the first part of the curves, between 0.1 seconds ( when the wheel reaches the step ) and 0.16 seconds, shows that in the case with only bump damping, the wheel is airborne for a shorter period, as expected. However, when we look farther along the time axis we see that the overall force fluctuations are greater and that the front wheel becomes airborne once again, the total time spent with no contact is actually greater than in the other case, with rebound damping only. With only bump damping, when the wheel hits the ground at about 0.14 secs. the higher bump suspension forces cause a heavier landing, shown by the higher peak tyre force, which in turn has caused more tyre bounce thereby lifting the wheel off the ground again at 0.23 seconds.
The following table gives a broad overview of the human perception of different vertical motion frequencies insofar as they affect vehicle suspension.
The following table gives a broad overview of the human perception of different vertical motion frequencies insofar as they affect vehicle suspension.
Fig 6.15 Response of a damped mass spring system to a forced oscillation, such as the vertical movement of a bike compared to the road surface height variation. The vertical scale on the graph is known as the transmission factor and shows the peak amplitude of the movement of the mass in relation to the movement of the input. A value of 1 means that the two movements are equal. The horizontal scale is the frequency ratio, a value of 1 is the resonant or natural frequency, a value of 2 means that the input movement has a frequency of double the natural frequency. The individual curves show the effect of damping ratio, which is the degree of damping compared to critical damping. A value of zero is with no damping and a value of 1.0 is critical damping.
Fig 6.15 Response of a damped mass spring system to a forced oscillation, such as the vertical movement of a bike compared to the road surface height variation. The vertical scale on the graph is known as the transmission factor and shows the peak amplitude of the movement of the mass in relation to the movement of the input. A value of 1 means that the two movements are equal. The horizontal scale is the frequency ratio, a value of 1 is the resonant or natural frequency, a value of 2 means that the input movement has a frequency of double the natural frequency. The individual curves show the effect of damping ratio, which is the degree of damping compared to critical damping. A value of zero is with no damping and a value of 1.0 is critical damping.
Fig. 6.16 Valid for any vehicle, this graph shows the spring rate required for any suspension natural frequency — Hz., per 100 kgs. of sprung mass.
Fig. 6.16 Valid for any vehicle, this graph shows the spring rate required for any suspension natural frequency — Hz., per 100 kgs. of sprung mass.
Fig. 6.17 Applicable to any vehicle, this graph shows how the statically loaded spring sag, from the length of a free spring, is related to the suspension frequency in the range most likely used for suspension. Thus 6 is directly related to frequency. Fig. 6.17 shows this in graphical form
Fig. 6.17 Applicable to any vehicle, this graph shows how the statically loaded spring sag, from the length of a free spring, is related to the suspension frequency in the range most likely used for suspension. Thus 6 is directly related to frequency. Fig. 6.17 shows this in graphical form
Fig. 6.18 The relationship between velocity and suspension forcing frequency for various bump wavelengths. Note that the frequency is plotted on a logarithmic scale. The numbers on the various curves are the bump wavelength in metres. The shaded area is the most comfortable frequency range of between 1 and 2 Hz. We can see that for speeds between 30 and 300 km/h the most comfortable frequencies occur with wavelengths between 0.5 and 10 metres. Fig. 6.18 shows this in graphical form over a wide range of speed and wavelengths. However, this only tells us the range of frequencies to be expected, another very important aspect of ride comfort is the actual amplitude of the displacements that occur at different frequencies. this varies from road to road but in general we can say from experience that the amplitude reduces as the wavelength reduces. Standards organizations such as ISO have developed standardized models of various types of road surface for use in suspension analysis. These define the amplitudes to be expected at different
Fig. 6.18 The relationship between velocity and suspension forcing frequency for various bump wavelengths. Note that the frequency is plotted on a logarithmic scale. The numbers on the various curves are the bump wavelength in metres. The shaded area is the most comfortable frequency range of between 1 and 2 Hz. We can see that for speeds between 30 and 300 km/h the most comfortable frequencies occur with wavelengths between 0.5 and 10 metres. Fig. 6.18 shows this in graphical form over a wide range of speed and wavelengths. However, this only tells us the range of frequencies to be expected, another very important aspect of ride comfort is the actual amplitude of the displacements that occur at different frequencies. this varies from road to road but in general we can say from experience that the amplitude reduces as the wavelength reduces. Standards organizations such as ISO have developed standardized models of various types of road surface for use in suspension analysis. These define the amplitudes to be expected at different
Fig. 6.20 Here the sprung mass is grouped around the centre of gravity. A bump at the front will tend to rotate the machine backwards about the CoG, thus also compressing the rear. This is somewhat closer to reality but would predict an over quick response as there is no provision for pitch inertia. Fig. 6.19 Model of the suspension system of a motorcycle. In this simple layout the sprung mass is assumed to be divided into two separate front and rear parts. This is the implied model used when calculating simple suspension frequencies. The rear suspension will not be affected by disturbances at the front.
Fig. 6.20 Here the sprung mass is grouped around the centre of gravity. A bump at the front will tend to rotate the machine backwards about the CoG, thus also compressing the rear. This is somewhat closer to reality but would predict an over quick response as there is no provision for pitch inertia. Fig. 6.19 Model of the suspension system of a motorcycle. In this simple layout the sprung mass is assumed to be divided into two separate front and rear parts. This is the implied model used when calculating simple suspension frequencies. The rear suspension will not be affected by disturbances at the front.
Using models like this we can evaluate suspension performance against time or frequency for a wide range of loading conditions as shown in some previous and following examples, for some purposes it may be more useful to plot the frequency response of the complete motorcycle, rather than just a simplified single wheel model as in fig. 6.15. There are many resonances that will give various peaks in the curve rather than just the simple single peak shown in fig. 6.15.
Using models like this we can evaluate suspension performance against time or frequency for a wide range of loading conditions as shown in some previous and following examples, for some purposes it may be more useful to plot the frequency response of the complete motorcycle, rather than just a simplified single wheel model as in fig. 6.15. There are many resonances that will give various peaks in the curve rather than just the simple single peak shown in fig. 6.15.
Fig. 6.22 Four sets of frequency response curves with the same bike parameters plotted from simulations at different bump wavelengths (A), as shown. (A) is for A=20 m. which is several times the wheel base of 1.5 m. (B) is when the wave- length is equal to the wheelbase and (C) & (D) show the cases where the bump length is 2 and 3 times the wheelbase respectively. Note that the vertical scales differ in some cases. It is clear just how much the relationship between wheelbase and bump length affects the transmissibility of the bump displacement to the sprung part of the bike. However, these curves don’t tell all the story by any means. Compare (C) with the upper plot of fig. 6.32 , both show that there is little vertical movement of the CoG, but the front and rear parts are subject to quite high peak amplitudes but out of phase with each other. The front moves up as the rear moves down. In other words there is considerable pitching motion which is generally more uncomfortable than simple vertical heave motions.
Fig. 6.22 Four sets of frequency response curves with the same bike parameters plotted from simulations at different bump wavelengths (A), as shown. (A) is for A=20 m. which is several times the wheel base of 1.5 m. (B) is when the wave- length is equal to the wheelbase and (C) & (D) show the cases where the bump length is 2 and 3 times the wheelbase respectively. Note that the vertical scales differ in some cases. It is clear just how much the relationship between wheelbase and bump length affects the transmissibility of the bump displacement to the sprung part of the bike. However, these curves don’t tell all the story by any means. Compare (C) with the upper plot of fig. 6.32 , both show that there is little vertical movement of the CoG, but the front and rear parts are subject to quite high peak amplitudes but out of phase with each other. The front moves up as the rear moves down. In other words there is considerable pitching motion which is generally more uncomfortable than simple vertical heave motions.
Fig. 6.23 In these plots the velocity is kept constant at 100 km/h and the stimulating frequency was varied by changing the bump wavelength. (A), (B) & (C) differ only in the pitch inertia as shown. (A) & (B) cover the range of realistic values but (C) is an unrealistic test case with a pitch moment of inertia 10 times that of (A). This shows that with a large pitch inertia there will be much less pitch displacement and so we’d expect the front, rear and CoG to move up and down in concert. This is shown clearly in (C). The fourth set of curves show the transmissibility in terms of pitch movement for the three cases. Note the low transmissibility for case (C) with the very high pitch inertia.
Fig. 6.23 In these plots the velocity is kept constant at 100 km/h and the stimulating frequency was varied by changing the bump wavelength. (A), (B) & (C) differ only in the pitch inertia as shown. (A) & (B) cover the range of realistic values but (C) is an unrealistic test case with a pitch moment of inertia 10 times that of (A). This shows that with a large pitch inertia there will be much less pitch displacement and so we’d expect the front, rear and CoG to move up and down in concert. This is shown clearly in (C). The fourth set of curves show the transmissibility in terms of pitch movement for the three cases. Note the low transmissibility for case (C) with the very high pitch inertia.
Imagine that a wheel leaves the ground or travels over a hollow in the road. When the front or rear suspension is partly or fully compressed, it exerts equal and opposite forces on both the sprung and unsprung parts of the machine — i.e. it tries simultaneously to raise the bulk of the machine and return the wheel to the ground. We want a system that returns the wheel quickly to the ground, for good road- holding, without unduly disturbing the sprung mass. The acceleration of any object depends on its mass and the force acting on it. Since, in our case, the same force is acting on both the sprung and unsprung masses, their relative accelerations depend on the ratio of their masses, while their absolute accelerations are determined by the suspension force and the absolute values of the two masses. Fig. 6.24 These plots are the result of a computer simulation showing the effect of various sprung to unsprung mass ratios. The motorcycle is travelling at 100 km/h and encounters a step of 25mm. (0.025 m.). The front suspension frequency is 1.9 Hz. with the rear a bit higher at 2.2. The graphs show the vertical acceleration of the CoG expressed in Gs. for three values of sprung to unsprung mass ratio. This ratio is changed by a factor of five between each curve. The heavy line is for a typical ratio of 4.5, the solid thin line has 1/5 of the previous unsprung mass with a ratio of 22.5, and the dotted line is for an unsprung mass of about 5 times normal with a ratio of 0.9.
Imagine that a wheel leaves the ground or travels over a hollow in the road. When the front or rear suspension is partly or fully compressed, it exerts equal and opposite forces on both the sprung and unsprung parts of the machine — i.e. it tries simultaneously to raise the bulk of the machine and return the wheel to the ground. We want a system that returns the wheel quickly to the ground, for good road- holding, without unduly disturbing the sprung mass. The acceleration of any object depends on its mass and the force acting on it. Since, in our case, the same force is acting on both the sprung and unsprung masses, their relative accelerations depend on the ratio of their masses, while their absolute accelerations are determined by the suspension force and the absolute values of the two masses. Fig. 6.24 These plots are the result of a computer simulation showing the effect of various sprung to unsprung mass ratios. The motorcycle is travelling at 100 km/h and encounters a step of 25mm. (0.025 m.). The front suspension frequency is 1.9 Hz. with the rear a bit higher at 2.2. The graphs show the vertical acceleration of the CoG expressed in Gs. for three values of sprung to unsprung mass ratio. This ratio is changed by a factor of five between each curve. The heavy line is for a typical ratio of 4.5, the solid thin line has 1/5 of the previous unsprung mass with a ratio of 22.5, and the dotted line is for an unsprung mass of about 5 times normal with a ratio of 0.9.
Fig. 6.26 shows the degree of increased loading for different values of cornering acceleration. In ice racing, where machines are banked through 60-70 degrees, the static loadings could be increased two or threefold, which partially explains why rear springing has not been adopted in that sport. Another consideration, often overlooked, is the effect of cornering forces, particularly on road-racing and sports machines, where banking angles may exceed 45 degrees. This represents a 41 percent increase in the static suspension loading. Fig. 6.25 shows how the weight and cornering force combine to increase suspension loading.
Fig. 6.26 shows the degree of increased loading for different values of cornering acceleration. In ice racing, where machines are banked through 60-70 degrees, the static loadings could be increased two or threefold, which partially explains why rear springing has not been adopted in that sport. Another consideration, often overlooked, is the effect of cornering forces, particularly on road-racing and sports machines, where banking angles may exceed 45 degrees. This represents a 41 percent increase in the static suspension loading. Fig. 6.25 shows how the weight and cornering force combine to increase suspension loading.
Fig. 6.27 The effect of changing from a 45 degree bank to a similar lean angle towards the opposite side (centre part of an “S” bend). The dark line shows how the lean angle has changed and the other lines show the suspension compression and extension for two cases with different values of damping. The dotted line has less damping than the solid line, allowing the overshoot during the early part of the suspension extension. It might seem that the stronger damping would be preferable as the total suspension movement is slightly less but it is likely that the centrifugal force will then reduce the tyre load by a greater factor. Fig. 6.26 Showing the percentage increase in both front and rear suspension loading when cornering.
Fig. 6.27 The effect of changing from a 45 degree bank to a similar lean angle towards the opposite side (centre part of an “S” bend). The dark line shows how the lean angle has changed and the other lines show the suspension compression and extension for two cases with different values of damping. The dotted line has less damping than the solid line, allowing the overshoot during the early part of the suspension extension. It might seem that the stronger damping would be preferable as the total suspension movement is slightly less but it is likely that the centrifugal force will then reduce the tyre load by a greater factor. Fig. 6.26 Showing the percentage increase in both front and rear suspension loading when cornering.
Fig. 6.28 Front and rear suspension movement, showing the variation in the under-damped case. Note also the suspension extension during the first 0.5 second, this is due to the centrifugal force from the initial build up of roll velocity before the cornering force has overcome it and caused the compression. The damping value also affects the amount of this extension.
Fig. 6.28 Front and rear suspension movement, showing the variation in the under-damped case. Note also the suspension extension during the first 0.5 second, this is due to the centrifugal force from the initial build up of roll velocity before the cornering force has overcome it and caused the compression. The damping value also affects the amount of this extension.
Fig. 6.29 Lateral tyre forces. Even though it is the rear that is under-damped there is more variation on the front
Fig. 6.29 Lateral tyre forces. Even though it is the rear that is under-damped there is more variation on the front
Fig. 6.31 Lean angle, the initial peak variation is about 5% of the steady lean angle of 44 degrees. The weave has nearly died out after 6 seconds.
Fig. 6.31 Lean angle, the initial peak variation is about 5% of the steady lean angle of 44 degrees. The weave has nearly died out after 6 seconds.
Fig. 6.30 Rear wheel slip angle which is similar to the yaw angle of the bike.
Fig. 6.30 Rear wheel slip angle which is similar to the yaw angle of the bike.
Sprayson, designer of the famous Reynolds racing frames, used to specify one-third of the available wheel travel for extension (sometimes called sag or static sag) and two-thirds for compression, which seems as good a starting point as any for most purposes. See the end of chapter 9 for more details of some additional effects of spring preload.
Sprayson, designer of the famous Reynolds racing frames, used to specify one-third of the available wheel travel for extension (sometimes called sag or static sag) and two-thirds for compression, which seems as good a starting point as any for most purposes. See the end of chapter 9 for more details of some additional effects of spring preload.
Fig. 6.32 These two sets of curves demonstrate wheel base effects on a corrugated road. In each case the middle curve is the vertical displacement of the CoG, the other two lines are the vertical displacements at the front and the rear.
Fig. 6.32 These two sets of curves demonstrate wheel base effects on a corrugated road. In each case the middle curve is the vertical displacement of the CoG, the other two lines are the vertical displacements at the front and the rear.
Fig. 6.33 Boge Nivomat self-levelling suspension strut. Left is a twin chamber version with a single chamber to the right.
Fig. 6.33 Boge Nivomat self-levelling suspension strut. Left is a twin chamber version with a single chamber to the right.
Braking and acceleration Discussed in more detail in the chapter on anti-squat, these longitudinal accelerations cause a loac transfer from one end to the other. Depending on geometric aspects of the suspension design, whict affect anti-dive and squat, it is usual for these load transfers to extend or compress the suspensior medium. As shown in chapter 9 it is not unusual for the front forks to be subject to a load under braking of three times the static load. Loading variations like this place heavy demands on the suspensior system, which as we've already seen has to also cater for increased compression when cornering as well as its more obvious function of bump absorption and trying to maintain an even tyre force on the road. With such diverse requirements it is a wonder that suitable compromises can be made to enable the suspension system to function as well as it does.
Braking and acceleration Discussed in more detail in the chapter on anti-squat, these longitudinal accelerations cause a loac transfer from one end to the other. Depending on geometric aspects of the suspension design, whict affect anti-dive and squat, it is usual for these load transfers to extend or compress the suspensior medium. As shown in chapter 9 it is not unusual for the front forks to be subject to a load under braking of three times the static load. Loading variations like this place heavy demands on the suspensior system, which as we've already seen has to also cater for increased compression when cornering as well as its more obvious function of bump absorption and trying to maintain an even tyre force on the road. With such diverse requirements it is a wonder that suitable compromises can be made to enable the suspension system to function as well as it does.
Fig. 6.35 At 45 degrees lean a vertical bump force will create equal forces , one in line with the suspension movement and the other at right angles to it. These forces will each be 71% of the vertical bump force. At the moment this seems to be a little understood and very controversial topic, without any real consensus about whether a chassis should be as stiff as possible or have some built in compliance designed to reduce the tyre force variation. Some laboratory and track testing has indicated a reduction in this variation, to the benefit of traction in general, by the introduction of some extra flexibility. However, races are not won by technical considerations alone, the rider has to be accounted for also. It is a fact that some riders feel more at home on a bike with a rock solid feel to it, whereas others are happiest with a machine that appears to “give” a little. What is certain though, is that any deliberately introduced compliance must be below a level that would introduce other instability problems, such as wobbles or weaves. Prior to about the 1980s many chassis fell below that level and benefited greatly from increased chassis stiffness, however beyond a certain point it is doubtful if increased stiffness would be noticeable. It appears that some modern chassis may have exceeded that point possibly to the detriment of traction.
Fig. 6.35 At 45 degrees lean a vertical bump force will create equal forces , one in line with the suspension movement and the other at right angles to it. These forces will each be 71% of the vertical bump force. At the moment this seems to be a little understood and very controversial topic, without any real consensus about whether a chassis should be as stiff as possible or have some built in compliance designed to reduce the tyre force variation. Some laboratory and track testing has indicated a reduction in this variation, to the benefit of traction in general, by the introduction of some extra flexibility. However, races are not won by technical considerations alone, the rider has to be accounted for also. It is a fact that some riders feel more at home on a bike with a rock solid feel to it, whereas others are happiest with a machine that appears to “give” a little. What is certain though, is that any deliberately introduced compliance must be below a level that would introduce other instability problems, such as wobbles or weaves. Prior to about the 1980s many chassis fell below that level and benefited greatly from increased chassis stiffness, however beyond a certain point it is doubtful if increased stiffness would be noticeable. It appears that some modern chassis may have exceeded that point possibly to the detriment of traction.
Fig. 6.36 The sketches to the left show how the tyre movement needed to surmount a bump is changed by introducing lateral movement. “ig. 6.36 The sketches to the left show how the tyre movement needed to surmount a bump is changed by introducing ateral movement. |. Shows the case where all movement is along the path “c” which is co-linear with “a” the suspension compression jirection. 2. Shows how a small amount of lateral displacement “b” changes the tyre path, and the expanded view shows that the n-line movement reduces by the same amount. 3. When the lateral motion is equal to the in-line movement ( “a” = “b” ) then the wheel path is vertical when at a lean
Fig. 6.36 The sketches to the left show how the tyre movement needed to surmount a bump is changed by introducing lateral movement. “ig. 6.36 The sketches to the left show how the tyre movement needed to surmount a bump is changed by introducing ateral movement. |. Shows the case where all movement is along the path “c” which is co-linear with “a” the suspension compression jirection. 2. Shows how a small amount of lateral displacement “b” changes the tyre path, and the expanded view shows that the n-line movement reduces by the same amount. 3. When the lateral motion is equal to the in-line movement ( “a” = “b” ) then the wheel path is vertical when at a lean
Substituting motorcycle values for these parameters we find that the lateral motion could typically be around five times that of the vertical, even more when chassis flex is accounted for. So it would seem that the problem is not so much one of needing to introduce more lateral displacement as often suggested, but one of restoring lost damping and thus in-line movement. (B) shows how enough in-line movement converts the wheel motion to purely vertical. This in-line motion is of course provided by the normal suspension system. The ratio of the suspension movement to lateral motion is affected by many parameters, e.g. M, I, r as before but also the unsprung mass and compliance of the suspension and chassis. In most practical cases, at high lean angles, the lateral motion will exceed that due to suspension movement. We have already seen that this lateral motion will reduce the suspension movement, for a given size of bump. If we consider the basic requirements of dissipating the effects of a bump it will become clear that reducing the suspension movement is highly detrimental.
Substituting motorcycle values for these parameters we find that the lateral motion could typically be around five times that of the vertical, even more when chassis flex is accounted for. So it would seem that the problem is not so much one of needing to introduce more lateral displacement as often suggested, but one of restoring lost damping and thus in-line movement. (B) shows how enough in-line movement converts the wheel motion to purely vertical. This in-line motion is of course provided by the normal suspension system. The ratio of the suspension movement to lateral motion is affected by many parameters, e.g. M, I, r as before but also the unsprung mass and compliance of the suspension and chassis. In most practical cases, at high lean angles, the lateral motion will exceed that due to suspension movement. We have already seen that this lateral motion will reduce the suspension movement, for a given size of bump. If we consider the basic requirements of dissipating the effects of a bump it will become clear that reducing the suspension movement is highly detrimental.
Fig. 6.38 Tyre force response to a small sinusoidal bump (shown in the inset) for various values of damping. The light grey curve is with tyre damping only, the darkest is with typical suspension damping and tyre damping, and the other is for a quarter of the previous suspension damping. The oscillation visible prior to 0.5 seconds is due to the unsprung mass bouncing on the tyre (wheel hop) and the much lower frequency motion after 0.5 seconds is due to the sprung mass moving on the suspension springs.
Fig. 6.38 Tyre force response to a small sinusoidal bump (shown in the inset) for various values of damping. The light grey curve is with tyre damping only, the darkest is with typical suspension damping and tyre damping, and the other is for a quarter of the previous suspension damping. The oscillation visible prior to 0.5 seconds is due to the unsprung mass bouncing on the tyre (wheel hop) and the much lower frequency motion after 0.5 seconds is due to the sprung mass moving on the suspension springs.
Fig. 6.39 These force-displacement curves are measured from actual motorcycles. At left is a complete loading and unloading cycle covering positive and negative values. This shows the torsional characteristics from axle to axle. The second curve to the right show the stiffness of a complete wheel and swingarm assembly from a used sport bike. Despite the very irregular shape these characteristics are very repeatable from one loading cycle to the next. Slop and stiction etc. are causes of the strange shape. In both of these cases the main point to note in the current context is that they exhibit a hysteresis effect as explained in the tyre chapter. The shaded area represents a loss of energy that occurs during each complete loading cycle, damping in other words. (data courtesy of Dr. Robin Tuluie and MTS)
Fig. 6.39 These force-displacement curves are measured from actual motorcycles. At left is a complete loading and unloading cycle covering positive and negative values. This shows the torsional characteristics from axle to axle. The second curve to the right show the stiffness of a complete wheel and swingarm assembly from a used sport bike. Despite the very irregular shape these characteristics are very repeatable from one loading cycle to the next. Slop and stiction etc. are causes of the strange shape. In both of these cases the main point to note in the current context is that they exhibit a hysteresis effect as explained in the tyre chapter. The shaded area represents a loss of energy that occurs during each complete loading cycle, damping in other words. (data courtesy of Dr. Robin Tuluie and MTS)
Fig. 7.1 With such a large lever arm these types of deflection are present to some degree in all head stock mounted forks.
Fig. 7.1 With such a large lever arm these types of deflection are present to some degree in all head stock mounted forks.
Fig. 7.2 When a telescopic fork is fully extended the sliders are poorly supported because of a reduced overlap on the stanchions. On full bump, the overlap is at a maximum. Offsetting the axle allows a longer stanchion to extend farther into the slider for better support.
Fig. 7.2 When a telescopic fork is fully extended the sliders are poorly supported because of a reduced overlap on the stanchions. On full bump, the overlap is at a maximum. Offsetting the axle allows a longer stanchion to extend farther into the slider for better support.
Two examples of offset wheel spindle forks. The one on the left has the spindle offset behind the slider whereas on the right we see the more usual system with the axle in front. Surprisingly both of these designs are from contemporary models by the same manufacturer, Cagiva. The first design allows the fork to angle back more, and so is better able to deal with the rearward component of bump forces. (See fig. 7.6) This must be set against the increased wheelbase variation, more brake dive and probably greater steering inertia.
Two examples of offset wheel spindle forks. The one on the left has the spindle offset behind the slider whereas on the right we see the more usual system with the axle in front. Surprisingly both of these designs are from contemporary models by the same manufacturer, Cagiva. The first design allows the fork to angle back more, and so is better able to deal with the rearward component of bump forces. (See fig. 7.6) This must be set against the increased wheelbase variation, more brake dive and probably greater steering inertia.
Fig. 7.5 By moving the tyre contact patch forward, a sharp bump reduces trail, it may even go negative as shown. Fig. 7.6 The resultant wheel load from a bump acts radial inward from the contact point. Fv is the vertical component and FI the longitudinal component of the tyre force.
Fig. 7.5 By moving the tyre contact patch forward, a sharp bump reduces trail, it may even go negative as shown. Fig. 7.6 The resultant wheel load from a bump acts radial inward from the contact point. Fv is the vertical component and FI the longitudinal component of the tyre force.
Fig 7.4 On the left we can see the conditions needed for the trail to remain constant under the action of fork compression. Nose dive as experienced in braking reduces the trail as shown on the right.
Fig 7.4 On the left we can see the conditions needed for the trail to remain constant under the action of fork compression. Nose dive as experienced in braking reduces the trail as shown on the right.
fore-and-aft flexure under heavy braking. (This also reduces the horizontal loads on the steering head, so reducing flexure in the frame and possibly prolonging its life.) A single suspension strut, anchored to the frame, is connected to the sliders through a linkage system at the lower end, in this way the strut forms no part of the steering inertia. Typical modern “upside-down” fork leg on the right. The large diameter upper part is subject to the greatest bending moment. On the “conventional” leg (left) this bending moment is less ably carried by the smaller diameter stanchion. The Altec fork. Note how the slider is supported at each end throughout the whole range of movement. The hydraulic cylinder for the fore-and—aft bracing is clearly visible.
fore-and-aft flexure under heavy braking. (This also reduces the horizontal loads on the steering head, so reducing flexure in the frame and possibly prolonging its life.) A single suspension strut, anchored to the frame, is connected to the sliders through a linkage system at the lower end, in this way the strut forms no part of the steering inertia. Typical modern “upside-down” fork leg on the right. The large diameter upper part is subject to the greatest bending moment. On the “conventional” leg (left) this bending moment is less ably carried by the smaller diameter stanchion. The Altec fork. Note how the slider is supported at each end throughout the whole range of movement. The hydraulic cylinder for the fore-and—aft bracing is clearly visible.
An unusual interpretation of the USD theme, made by Ceriani. Suspension is provided by a single spring and damper unit connected to a bridge piece connecting both legs together. Gilera produced this system. Essentially this is a single leg telescopic suspension. Steering torque is transmitted via a folding linkage. This whole design is very much like that used on large aircraft landing gear.
An unusual interpretation of the USD theme, made by Ceriani. Suspension is provided by a single spring and damper unit connected to a bridge piece connecting both legs together. Gilera produced this system. Essentially this is a single leg telescopic suspension. Steering torque is transmitted via a folding linkage. This whole design is very much like that used on large aircraft landing gear.
Leading link An alternative to the telescopic fork is the leading-link variety, which has the distinction of having beer fitted to what were probably the best-handling racing machines of their period — the world champior Moto Guzzis of the mid 1950s. The Earles fork with longer links, used on some early MV Agusta racers and for many years by BMW, is a variation on the same theme but has the disadvantage of considerably higher steered inertia.
Leading link An alternative to the telescopic fork is the leading-link variety, which has the distinction of having beer fitted to what were probably the best-handling racing machines of their period — the world champior Moto Guzzis of the mid 1950s. The Earles fork with longer links, used on some early MV Agusta racers and for many years by BMW, is a variation on the same theme but has the disadvantage of considerably higher steered inertia.
From 1955 until 1970 BMW used Earles leading link forks. Two pivot positions were available. The rear was for solo use and the forward one gave less trail for sidecar duty. The brake back- plate was attached to the link and there was enough antidive effect to lift the bike under front wheel braking only. A leading link fork which helped to create an enviable reputation for handling, as fitted to the 1954 NSU 125cc Rennfox world championship-winner. The stanchions were fabricated from steel pressings and the links had wide spindle clamps. (MCW) Fig 7.7 With leading link forks, the relative heights of wheel spindle and link pivots determine the path of wheel travel.
From 1955 until 1970 BMW used Earles leading link forks. Two pivot positions were available. The rear was for solo use and the forward one gave less trail for sidecar duty. The brake back- plate was attached to the link and there was enough antidive effect to lift the bike under front wheel braking only. A leading link fork which helped to create an enviable reputation for handling, as fitted to the 1954 NSU 125cc Rennfox world championship-winner. The stanchions were fabricated from steel pressings and the links had wide spindle clamps. (MCW) Fig 7.7 With leading link forks, the relative heights of wheel spindle and link pivots determine the path of wheel travel.
Fig 7.8 In a trailing link layout, steering inertia is increased by the relatively long distance of the mass of the fork from the steering axis. A rather complicated interpretation of the trailing link theme by Frenchman Eric Offenstadt. The wheel spindle clamps could be slid along the link tubes for trail adjustment. The long upright link outside of the main assembly was there to transfer suspension loads to a high mounted single spring/damper unit. A leading link fork design by the author. These were used on large capacity road bikes, the links were separate for styling reasons and relied on a relatively large diameter wheel spindle to provide sufficient stiffness.
Fig 7.8 In a trailing link layout, steering inertia is increased by the relatively long distance of the mass of the fork from the steering axis. A rather complicated interpretation of the trailing link theme by Frenchman Eric Offenstadt. The wheel spindle clamps could be slid along the link tubes for trail adjustment. The long upright link outside of the main assembly was there to transfer suspension loads to a high mounted single spring/damper unit. A leading link fork design by the author. These were used on large capacity road bikes, the links were separate for styling reasons and relied on a relatively large diameter wheel spindle to provide sufficient stiffness.
Trailing-link forks differ from the leading-link variety only insofar as the link pivots are ahead of the wheel spindle, not behind. Their disadvantage is higher steering inertia, since the bulk of the mass is relatively far from the steering axis — an effect that is only partly off-set by the smaller amount of material needed to reach the pivots, fig 7.8. Girder forks in their day were almost as widely used as telescopic are today, and some were known for their excellent steering. Structurally there is much to commend this type of fork if executed well. Performance was generally limited by friction dampers, very crude by current hydraulic standards. The links that allowed suspension movement were quite short and would not have been very suitable for anything but a small total amount of suspension movement.
Trailing-link forks differ from the leading-link variety only insofar as the link pivots are ahead of the wheel spindle, not behind. Their disadvantage is higher steering inertia, since the bulk of the mass is relatively far from the steering axis — an effect that is only partly off-set by the smaller amount of material needed to reach the pivots, fig 7.8. Girder forks in their day were almost as widely used as telescopic are today, and some were known for their excellent steering. Structurally there is much to commend this type of fork if executed well. Performance was generally limited by friction dampers, very crude by current hydraulic standards. The links that allowed suspension movement were quite short and would not have been very suitable for anything but a small total amount of suspension movement.
kingpin is bushed to rock on the axle. Swivelling on ball bearings at the top and bottom of the kingpin is a steerable but non-rotating drum supporting the large wheel bearings (130 mm outside diameter, 85 mm inside diameter); the drum is slotted for steering. Bolted to the outer flanks of the drum are twc upright A-frames (one each side) whose apexes are united by a bridge-piece above the wheel. Difazio hub-centre steering is shown on this 500 cc Suzuki-powered racer. Suspension is by pivoted fork and steering by a parallelogram linkage. The steering links had a double duty and were also needed for the upper location of the A- frames, and hence were subject to reactions against braking torque. Range of kingpin angles was from 16 to 22 degrees. (MCW)
kingpin is bushed to rock on the axle. Swivelling on ball bearings at the top and bottom of the kingpin is a steerable but non-rotating drum supporting the large wheel bearings (130 mm outside diameter, 85 mm inside diameter); the drum is slotted for steering. Bolted to the outer flanks of the drum are twc upright A-frames (one each side) whose apexes are united by a bridge-piece above the wheel. Difazio hub-centre steering is shown on this 500 cc Suzuki-powered racer. Suspension is by pivoted fork and steering by a parallelogram linkage. The steering links had a double duty and were also needed for the upper location of the A- frames, and hence were subject to reactions against braking torque. Range of kingpin angles was from 16 to 22 degrees. (MCW)
Bimota Tesi A snag with any suspension system (front or rear) giving approximately vertical wheel movement is that this does not absorb the rearward component of the force due to road bumps, as does a telescopic fork. But this is probably a small price to pay for the benefits of hub-centre steering. Certainly, overall rigidity is much superior to that of a fork mounted on a conventional steering head. Also, since the loads fed into the main frame are much smaller, that component can be made lighter. On the debit side, hub- centre layouts of the types described above are often criticized on the grounds of appearance, weight and limited steering lock. Probably the currently best known of the true hub-centre designs, there have been several versions of this over a long development period. The original was part of a college design project and was probably best remembered for its use of hydraulic steering. Heralded in many press reports at the time as a major innovation, the hydraulic steering appears to have been just another source of potential problems. As Bimota seemed to have discovered. It is difficult to see any benefits which outweigh the disadvantages. Anyone with experience of such closed circuit hydraulics knows that provision for expansion and piston rod displacement is necessary. The solution adopted was the provision of a pressurized gas chamber, separated from the oil by a flexible diaphragm, similar to many suspension units. Unfortunately, the compliance thus introduced may cause sponginess in the steering.
Bimota Tesi A snag with any suspension system (front or rear) giving approximately vertical wheel movement is that this does not absorb the rearward component of the force due to road bumps, as does a telescopic fork. But this is probably a small price to pay for the benefits of hub-centre steering. Certainly, overall rigidity is much superior to that of a fork mounted on a conventional steering head. Also, since the loads fed into the main frame are much smaller, that component can be made lighter. On the debit side, hub- centre layouts of the types described above are often criticized on the grounds of appearance, weight and limited steering lock. Probably the currently best known of the true hub-centre designs, there have been several versions of this over a long development period. The original was part of a college design project and was probably best remembered for its use of hydraulic steering. Heralded in many press reports at the time as a major innovation, the hydraulic steering appears to have been just another source of potential problems. As Bimota seemed to have discovered. It is difficult to see any benefits which outweigh the disadvantages. Anyone with experience of such closed circuit hydraulics knows that provision for expansion and piston rod displacement is necessary. The solution adopted was the provision of a pressurized gas chamber, separated from the oil by a flexible diaphragm, similar to many suspension units. Unfortunately, the compliance thus introduced may cause sponginess in the steering.
A 1983 Milan Show special — the student designed Bimota Tesi, with hydraulically operated steering. Geometrically similar to the Difazio, this design loaded the wheel spindle heavily in torsion when braking. The design of later Bimota front ends changed considerably from the first version, improving the areas o doubt mentioned above. For one, the hydraulic steering was replaced by mechanical links.
A 1983 Milan Show special — the student designed Bimota Tesi, with hydraulically operated steering. Geometrically similar to the Difazio, this design loaded the wheel spindle heavily in torsion when braking. The design of later Bimota front ends changed considerably from the first version, improving the areas o doubt mentioned above. For one, the hydraulic steering was replaced by mechanical links.
Fig 7.9 Three types of double link front suspension. Basically only different in the location of the links, although the layout on the left is usually made with the upright as a fork so that a conventional wheel with double sided mounting can be used. The other two require a single sided upright and special single sided dished wheels.
Fig 7.9 Three types of double link front suspension. Basically only different in the location of the links, although the layout on the left is usually made with the upright as a fork so that a conventional wheel with double sided mounting can be used. The other two require a single sided upright and special single sided dished wheels.
There were many Elf sponsored racers with alternative front suspension designs. This one was a Honda powered endurance racer. Scaling from the photograph indicates that the ground trail was more than average at about 127 mm. In the forward end of the top link the pivot bearing is mounted in an eccentric block which could be easily adjusted to change the trail. Because the lower ball-joint is close to the wheel axle this trail adjustment only changes the wheelbase by a small amount. Note also the single sided cast rear swing-arm, with external sprocket, setting a style followed by others.
There were many Elf sponsored racers with alternative front suspension designs. This one was a Honda powered endurance racer. Scaling from the photograph indicates that the ground trail was more than average at about 127 mm. In the forward end of the top link the pivot bearing is mounted in an eccentric block which could be easily adjusted to change the trail. Because the lower ball-joint is close to the wheel axle this trail adjustment only changes the wheelbase by a small amount. Note also the single sided cast rear swing-arm, with external sprocket, setting a style followed by others.
Original Hossack tubular construction, showing front fork, steering linkage, wishbones and actuation of suspension strut. The steering linkage allows for suspension movement without introducing bump-steer in the straight ahead position. A small amount of bump steer would have been introduced when the steering was turned by a large angle, but this is of little importance on such a machine. (Dave Jenner) This type is the antithesis of the above, within the double link family. Instead of a very small upright with large curved arms, this setup has a larger upright which is located by very light and stiff wishbones. These are mounted above the tyre. Although the upright is usually constructed in the form of a fork, which allows for the use of normal wheels and brakes, there is no reason in principle why this could not be a single-sided component, to give the quick wheel changing of the Elf.e. but the stiffness to weight ratio is probably better with the fork.
Original Hossack tubular construction, showing front fork, steering linkage, wishbones and actuation of suspension strut. The steering linkage allows for suspension movement without introducing bump-steer in the straight ahead position. A small amount of bump steer would have been introduced when the steering was turned by a large angle, but this is of little importance on such a machine. (Dave Jenner) This type is the antithesis of the above, within the double link family. Instead of a very small upright with large curved arms, this setup has a larger upright which is located by very light and stiff wishbones. These are mounted above the tyre. Although the upright is usually constructed in the form of a fork, which allows for the use of normal wheels and brakes, there is no reason in principle why this could not be a single-sided component, to give the quick wheel changing of the Elf.e. but the stiffness to weight ratio is probably better with the fork.
As in the Difazio and Elf layouts, the precise geometry of wheel movement is governed by the lengths and angular dispositions of the upper and lower wishbones, so allowing the designer much more flexibility in the geometric layout than with a telescopic fork or one with leading or trailing links. Indeed the Hossack design may almost be regarded as an Elf with the pivot axes moved from within the whee to above it. In this light, the Hossack has both advantages and disadvantages compared to the Elf. we'll call the upright. There is no steering head. Instead, the fork pivots on spherical bearings at the front apexes of upper and lower forward-facing wishbones. The lower wishbone is triangulated upward from its pivot to actuate the suspension strut. Rake and trail are easily altered by screwing the top spherical joint into or out of its lug at the top of the fork. A system of links designed to eliminate significant bump-steer connects the fork to the handlebar. The fork itself is rigid and well triangulated, so providing strong resistance to lateral deflection of the tyre contact patch from the steering axis.
As in the Difazio and Elf layouts, the precise geometry of wheel movement is governed by the lengths and angular dispositions of the upper and lower wishbones, so allowing the designer much more flexibility in the geometric layout than with a telescopic fork or one with leading or trailing links. Indeed the Hossack design may almost be regarded as an Elf with the pivot axes moved from within the whee to above it. In this light, the Hossack has both advantages and disadvantages compared to the Elf. we'll call the upright. There is no steering head. Instead, the fork pivots on spherical bearings at the front apexes of upper and lower forward-facing wishbones. The lower wishbone is triangulated upward from its pivot to actuate the suspension strut. Rake and trail are easily altered by screwing the top spherical joint into or out of its lug at the top of the fork. A system of links designed to eliminate significant bump-steer connects the fork to the handlebar. The fork itself is rigid and well triangulated, so providing strong resistance to lateral deflection of the tyre contact patch from the steering axis.
A 1983 proposal by John Wright-Bailey. Similar to the Hossack genre this differs (like the McKagen dirt bikes) in using a suspended head stock, which allows the fitting of taper roller bearings, or similar, for the steering pivot, thus reducing steering friction. (MCN)
A 1983 proposal by John Wright-Bailey. Similar to the Hossack genre this differs (like the McKagen dirt bikes) in using a suspended head stock, which allows the fitting of taper roller bearings, or similar, for the steering pivot, thus reducing steering friction. (MCN)
The author’s interpretation of the double link theme. Shown here fitted to a Suzuki powered Q2. The lower arm is a simple aluminium casting, a shallow section for this was made possible by mounting the suspension unit above the wheel and so the curved arm only had in-plane stresses to deal with. The upper-link is a fabricated tubular “A-arm”. Steering is via a drag link just visible above the top link. made thinner and lighter. The price paid for this, is that the upright must be longer, and hence heavier and/or less stiff, than the Elf.e. On balance it is my opinion that it is a better compromise, as it also allows the steering drag link to be above the wheel where it is closer to the handle-bars, thus simplifying linkages, and need not be mounted wide to clear the wheel on full lock. The Parker system differs from the author’s in that the steering mechanism does not employ a drag link, instead it uses a telescoping torsion tube to pass the rider's messages on to the wheel.
The author’s interpretation of the double link theme. Shown here fitted to a Suzuki powered Q2. The lower arm is a simple aluminium casting, a shallow section for this was made possible by mounting the suspension unit above the wheel and so the curved arm only had in-plane stresses to deal with. The upper-link is a fabricated tubular “A-arm”. Steering is via a drag link just visible above the top link. made thinner and lighter. The price paid for this, is that the upright must be longer, and hence heavier and/or less stiff, than the Elf.e. On balance it is my opinion that it is a better compromise, as it also allows the steering drag link to be above the wheel where it is closer to the handle-bars, thus simplifying linkages, and need not be mounted wide to clear the wheel on full lock. The Parker system differs from the author’s in that the steering mechanism does not employ a drag link, instead it uses a telescoping torsion tube to pass the rider's messages on to the wheel.
Yamaha GTS in the flesh. The suspension components were of generous proportion to say the least, or in other words, downright heavy. It was a marketing risk for Yamaha to produce such a machine and they must be applauded for that. Presumably they didn’t want to take more risks with structural failure, with its attendant bad publicity and liability concerns. Yamaha GTS. The layout is very similar to Parker’s own bikes and_ retains’ the telescopic steering column to eliminate any possibility of bump steer.
Yamaha GTS in the flesh. The suspension components were of generous proportion to say the least, or in other words, downright heavy. It was a marketing risk for Yamaha to produce such a machine and they must be applauded for that. Presumably they didn’t want to take more risks with structural failure, with its attendant bad publicity and liability concerns. Yamaha GTS. The layout is very similar to Parker’s own bikes and_ retains’ the telescopic steering column to eliminate any possibility of bump steer.
Bump-steer More familiar to car enthusiasts, the term bump-steer refers to the results of a conflict between the suspension and steering geometries. With telescopic forks and indeed any head stock mountec suspension, the handlebars are connected directly and suspension movement has no effect on the steering angle.
Bump-steer More familiar to car enthusiasts, the term bump-steer refers to the results of a conflict between the suspension and steering geometries. With telescopic forks and indeed any head stock mountec suspension, the handlebars are connected directly and suspension movement has no effect on the steering angle.
The possibility of bump-steer has been completely eliminated by the telescopic steering column of the Parker design and GTS, but both the systems using a drag link and also those like the Hossack with a hinged linkage have the potential for such problems. In practice it is usually impossible to entirely eliminate bump steer over the full range of suspension and steering movements from these designs. Before constructing the Q2, shown above, the author did a series of tests that deliberately introduced large amounts of bump-steer to observe the practical riding effects. The results of these indicated that whilst it was desirable to minimize the bump-steer in the straight ahead position, it was much less important at large steering angles, which are only used at very low speed when manoeuvring. This compromise is not usually difficult to achieve.
The possibility of bump-steer has been completely eliminated by the telescopic steering column of the Parker design and GTS, but both the systems using a drag link and also those like the Hossack with a hinged linkage have the potential for such problems. In practice it is usually impossible to entirely eliminate bump steer over the full range of suspension and steering movements from these designs. Before constructing the Q2, shown above, the author did a series of tests that deliberately introduced large amounts of bump-steer to observe the practical riding effects. The results of these indicated that whilst it was desirable to minimize the bump-steer in the straight ahead position, it was much less important at large steering angles, which are only used at very low speed when manoeuvring. This compromise is not usually difficult to achieve.
Yet another approach on the Elf bikes (post Cortanze). This design owes much to the McPherson strut, common on modern cars. The typical lower arm is retained but the upper location is done entirely by the suspension unit which has a larger than normal sliding rod.
Yet another approach on the Elf bikes (post Cortanze). This design owes much to the McPherson strut, common on modern cars. The typical lower arm is retained but the upper location is done entirely by the suspension unit which has a larger than normal sliding rod.
At first site the Saxon/Motad system bears little resemblance to the Elf above, but it can still be considered as an adaptation of a "McPherson strut" for bike use. It uses just one pivoted link and a fixed pivot at the top. In place of a single strut the wheel is held in two sliding legs similar to a telescopic fork, although the suspension function is handled externally by a separate spring/damper unit. The steering axis is defined by the line drawn through the upper fixed pivot and the lower moving one, which is attached to the fork sliders, although the overhang from the wishbone to the wheel is much greater than is usual in the car version of strut suspension.
At first site the Saxon/Motad system bears little resemblance to the Elf above, but it can still be considered as an adaptation of a "McPherson strut" for bike use. It uses just one pivoted link and a fixed pivot at the top. In place of a single strut the wheel is held in two sliding legs similar to a telescopic fork, although the suspension function is handled externally by a separate spring/damper unit. The steering axis is defined by the line drawn through the upper fixed pivot and the lower moving one, which is attached to the fork sliders, although the overhang from the wishbone to the wheel is much greater than is usual in the car version of strut suspension.
Killeen BMW have also been strong advocates of ABS braking systems and this may have had an influence over the decision to ditch telescopic forks. The cycling frequency of the ABS system can be close to the longitudinal resonant frequency of the fork and wheel assembly, resulting in unwanted vibration. The telelever is stiffer in this direction and so its natural frequency would be somewhat higher and less likely to suffer similar problems. Additionally, as shown in the chapter on squat and dive, the geometry of the telelever layout can be selected to reduce nose-dive under braking. An interesting combination of telescopic fork and hub-centre steering was proposed and patented by Tom Killeen, who called it head-to-hub steering. “The telescopic fork”, according to Killeen’s description, “is mounted on a kingpin or spherical bearing in the hub centre and on a self-aligning or spherical bearing at the steering head on the main frame. The arms of the pivoted fork are bowed outward for tyre clearance on full lock and the pivot on the main frame is so positioned as to minimize the rocking of the telescopic fork at the steering head on bump and rebound.” Except in appearance, it is difficult to see what advantage this system offers compared with, say, the Difazio — but it is superior to the conventional telescopic fork in several ways. In all conditions other than braking, the sliders are relieved of fore-and-
Killeen BMW have also been strong advocates of ABS braking systems and this may have had an influence over the decision to ditch telescopic forks. The cycling frequency of the ABS system can be close to the longitudinal resonant frequency of the fork and wheel assembly, resulting in unwanted vibration. The telelever is stiffer in this direction and so its natural frequency would be somewhat higher and less likely to suffer similar problems. Additionally, as shown in the chapter on squat and dive, the geometry of the telelever layout can be selected to reduce nose-dive under braking. An interesting combination of telescopic fork and hub-centre steering was proposed and patented by Tom Killeen, who called it head-to-hub steering. “The telescopic fork”, according to Killeen’s description, “is mounted on a kingpin or spherical bearing in the hub centre and on a self-aligning or spherical bearing at the steering head on the main frame. The arms of the pivoted fork are bowed outward for tyre clearance on full lock and the pivot on the main frame is so positioned as to minimize the rocking of the telescopic fork at the steering head on bump and rebound.” Except in appearance, it is difficult to see what advantage this system offers compared with, say, the Difazio — but it is superior to the conventional telescopic fork in several ways. In all conditions other than braking, the sliders are relieved of fore-and-
It is sometimes claimed for hub-centre steering that the elimination of the conventional steering head reduces frontal area. In the case of a roadster, this is not so because of the upright riding position. But it is clearly true for a low record breaker such as the fully streamlined, 2 ft-diameter (610 mm.) Harley- Davidson projectile in which Cal Rayborn raised the world record beyond 426 km/h. (265 mph.) in 1970.
It is sometimes claimed for hub-centre steering that the elimination of the conventional steering head reduces frontal area. In the case of a roadster, this is not so because of the upright riding position. But it is clearly true for a low record breaker such as the fully streamlined, 2 ft-diameter (610 mm.) Harley- Davidson projectile in which Cal Rayborn raised the world record beyond 426 km/h. (265 mph.) in 1970.
aft loading, hence stiction is minimal. Even during braking, these longitudinal loads are only those resisting the caliper forces; the main loads are taken by the pivoted fork. The fork also resists differential movement of the telescopic legs and much enhances lateral stiffness.
aft loading, hence stiction is minimal. Even during braking, these longitudinal loads are only those resisting the caliper forces; the main loads are taken by the pivoted fork. The fork also resists differential movement of the telescopic legs and much enhances lateral stiffness.
This OEC is resident at the National Motorcycle Museum near Birmingham, England. The side links only moved for steering and were not pivoted vertically for suspension movement. These links held two upright steered “tubes” which held the wheel and suspension.
This OEC is resident at the National Motorcycle Museum near Birmingham, England. The side links only moved for steering and were not pivoted vertically for suspension movement. These links held two upright steered “tubes” which held the wheel and suspension.
Fig. 7.11 The geometric aspects of a four-bar linkage steering/suspension system. The sketch to the right represents a normal head stock steering design. Both drawings show the projection onto the ground. The grey coloured lines show the steering in the straight ahead position and the black shows it steered approximately 20° to the left. The tyre contact patches are shown as ellipse shapes. The inset shows an analogy to the directional stability characteristics of each type.
Fig. 7.11 The geometric aspects of a four-bar linkage steering/suspension system. The sketch to the right represents a normal head stock steering design. Both drawings show the projection onto the ground. The grey coloured lines show the steering in the straight ahead position and the black shows it steered approximately 20° to the left. The tyre contact patches are shown as ellipse shapes. The inset shows an analogy to the directional stability characteristics of each type.
The most weight-efficient way to eliminate these defects is to triangulate the fork and connect the apex to the springing medium, as patented by Vincent-H.R.D in 1928. Nearly half a century later, Yamahe ‘reinvented’ the idea (initially for motocross), which acquired a semblance of novelty through the terms monoshock and cantilever. Possibly to avoid a charge of copying, as well as to keep the weight low down, Suzuki triangulated the fork on the RG500 road racer below pivot level (as did Moto Guzzi in the mid-1930s), though retaining two shock absorbers. These were steeply inclined forward and fitted with either pneumatic suspension units or progressive-rate springs. (Refer to fig. 8.5 and illustrations in the first chapter.)
The most weight-efficient way to eliminate these defects is to triangulate the fork and connect the apex to the springing medium, as patented by Vincent-H.R.D in 1928. Nearly half a century later, Yamahe ‘reinvented’ the idea (initially for motocross), which acquired a semblance of novelty through the terms monoshock and cantilever. Possibly to avoid a charge of copying, as well as to keep the weight low down, Suzuki triangulated the fork on the RG500 road racer below pivot level (as did Moto Guzzi in the mid-1930s), though retaining two shock absorbers. These were steeply inclined forward and fitted with either pneumatic suspension units or progressive-rate springs. (Refer to fig. 8.5 and illustrations in the first chapter.)
At the time of writing there are more variations, in use, of the detail of rear suspension design than there are with front suspension. Basically they are all some form of pivoted arm or fork (swinging arm), but there are single sided and dual, there are single and dual suspension struts, a wide variety of rocker designs and some incorporate drive shafts whilst others have to accommodate the pull of chain drive. More recently there has been a trend towards rocker-type rear suspension, with a whole new jargon such as Kawasaki’s Uni-track, Honda’s Pro-link and Suzuki’s Full Floater. etc... An early example of a link system for racing, in the 1970s, was the French Godier-Genoud endurance-racing Kawasaki, with the pivoted fork triangulated downwards and connected to the suspension strut via a bell-crank. The aim of all these designs is normally to obtain progressive-rate springing and damping by geometric means. If progression is desirable, then this is maybe a good way to achieve it because the springing and damping rates vary together. To achieve a progressive effect, we need to turn a link or lever through a large angle for a given linear movement, and this necessitates a short lever. All the rocker systems have this in common. Provided they give similar changes in effective spring rate (measured at the wheel spindle), and due regard is paid to stiffness and weight, then none of the layouts has any particular advantage over the others, despite the makers’ claims. Thus the choice of design is best determined on structural and space considerations. It is also beneficial to use the minimum number of joints in the system. Another consideration for a designer of pivoted rear suspension is pivot-bearing loads. With a simple traditional swing-arm, controlled by a pair of struts mounted near upright at the rear, suspension forces place very little load on the pivot. However, the Vincent layout and all the rocker-arm systems considerably increase these loads. Fig. 8.3 shows what happens in some selected designs. In most cases, this may not constitute a serious problem, but bearing life may be shortened and the design of the pivot mounting on the main frame must take these increased loads into account. Generally, on a chain driven machine the chain pull creates more load on the pivot than the suspension, but both must be added to get the total effect.
At the time of writing there are more variations, in use, of the detail of rear suspension design than there are with front suspension. Basically they are all some form of pivoted arm or fork (swinging arm), but there are single sided and dual, there are single and dual suspension struts, a wide variety of rocker designs and some incorporate drive shafts whilst others have to accommodate the pull of chain drive. More recently there has been a trend towards rocker-type rear suspension, with a whole new jargon such as Kawasaki’s Uni-track, Honda’s Pro-link and Suzuki’s Full Floater. etc... An early example of a link system for racing, in the 1970s, was the French Godier-Genoud endurance-racing Kawasaki, with the pivoted fork triangulated downwards and connected to the suspension strut via a bell-crank. The aim of all these designs is normally to obtain progressive-rate springing and damping by geometric means. If progression is desirable, then this is maybe a good way to achieve it because the springing and damping rates vary together. To achieve a progressive effect, we need to turn a link or lever through a large angle for a given linear movement, and this necessitates a short lever. All the rocker systems have this in common. Provided they give similar changes in effective spring rate (measured at the wheel spindle), and due regard is paid to stiffness and weight, then none of the layouts has any particular advantage over the others, despite the makers’ claims. Thus the choice of design is best determined on structural and space considerations. It is also beneficial to use the minimum number of joints in the system. Another consideration for a designer of pivoted rear suspension is pivot-bearing loads. With a simple traditional swing-arm, controlled by a pair of struts mounted near upright at the rear, suspension forces place very little load on the pivot. However, the Vincent layout and all the rocker-arm systems considerably increase these loads. Fig. 8.3 shows what happens in some selected designs. In most cases, this may not constitute a serious problem, but bearing life may be shortened and the design of the pivot mounting on the main frame must take these increased loads into account. Generally, on a chain driven machine the chain pull creates more load on the pivot than the suspension, but both must be added to get the total effect.
Fig 8.3 The simple traditional design on the left balances the rear weight directly by the suspension struts and so there is no weight-induced load on the fork pivot. The centre sketch shows how the pivot load can be increased with a mono- shock triangulated fork, in this case the pivot is subject to 2.25 times the wheel load. On the right we see how there must be load on the pivot to balance the forces with a rocker system. Each layout must be considered separately for there are an infinite variety of possible designs. The chain pull on a large machine is likely to create much greater pivot forces.
Fig 8.3 The simple traditional design on the left balances the rear weight directly by the suspension struts and so there is no weight-induced load on the fork pivot. The centre sketch shows how the pivot load can be increased with a mono- shock triangulated fork, in this case the pivot is subject to 2.25 times the wheel load. On the right we see how there must be load on the pivot to balance the forces with a rocker system. Each layout must be considered separately for there are an infinite variety of possible designs. The chain pull on a large machine is likely to create much greater pivot forces.
Unless the suspension unit is mounted as in the left side sketch of fig. 8.3 the spring rate at the wheel will be different from the spring rate of the suspension unit, usually less. In fact the relationship between these two rates follows a square law, that is, if the leverage ratio is such that the wheel movement is double that of the suspension unit then the actual spring rate will have to be four times that of the required effective wheel rate. Fig. 8.4 shows why.
Unless the suspension unit is mounted as in the left side sketch of fig. 8.3 the spring rate at the wheel will be different from the spring rate of the suspension unit, usually less. In fact the relationship between these two rates follows a square law, that is, if the leverage ratio is such that the wheel movement is double that of the suspension unit then the actual spring rate will have to be four times that of the required effective wheel rate. Fig. 8.4 shows why.
In this age of the widespread use of rocker and linkage type suspensions it is often forgotten that the more traditional systems can produce progressive or regressive characteristics also. Using a 1970s. RG500 of the type used by Barry Sheene as an example in fig. 8.5, we can see that geometric changes lead to a progressive rate. When the suspension is compressed, the line of the suspension force acts at a greater distance from the swing-arm pivot and so has a greater effect on the wheel motion. The difference in effective wheel rates will obey the square law and, forgetting about the inaccuracies inherent in measuring from such a picture, is about 11%.
In this age of the widespread use of rocker and linkage type suspensions it is often forgotten that the more traditional systems can produce progressive or regressive characteristics also. Using a 1970s. RG500 of the type used by Barry Sheene as an example in fig. 8.5, we can see that geometric changes lead to a progressive rate. When the suspension is compressed, the line of the suspension force acts at a greater distance from the swing-arm pivot and so has a greater effect on the wheel motion. The difference in effective wheel rates will obey the square law and, forgetting about the inaccuracies inherent in measuring from such a picture, is about 11%.
Fig 8.7 Example rocker system. Because the rocker ratio L;, : Lz varies with wheel movement, this rear suspension system gives progressive rate springing and damping. The effective wheel spring rate varies as the square of the above ratio. In the extended position The side connected to the swing-arm has the most mechanical advantage and so the rate will be softer than in the compressed case.
Fig 8.7 Example rocker system. Because the rocker ratio L;, : Lz varies with wheel movement, this rear suspension system gives progressive rate springing and damping. The effective wheel spring rate varies as the square of the above ratio. In the extended position The side connected to the swing-arm has the most mechanical advantage and so the rate will be softer than in the compressed case.
Kawasaki were one of the first to use a rocker and link system for road racing. Illustrated here fitted to the KR500 with ‘monocoque’ chassis. The lower suspension mount is attached to the underside of the swing-arm, whilst the mainly horizontal movement here has some effect over the variable rate properties its main advantage is that it eliminates the need for additional frame structure to support the unit.
Kawasaki were one of the first to use a rocker and link system for road racing. Illustrated here fitted to the KR500 with ‘monocoque’ chassis. The lower suspension mount is attached to the underside of the swing-arm, whilst the mainly horizontal movement here has some effect over the variable rate properties its main advantage is that it eliminates the need for additional frame structure to support the unit.
On the Australian built Drysdale V8 of 1999, space was at such a premium that the suspension unit was mounted across the frame under the swinging arm. Viewed from underneath we see how two vertical links drop from the swing arm to operate the two cranks at each end of the spring unit, compressing it in concertina fashion. Compare this with the 1982 Yamaha OW61 illustrated in the first chapter. (photo: Greg Parish, ACS.)
On the Australian built Drysdale V8 of 1999, space was at such a premium that the suspension unit was mounted across the frame under the swinging arm. Viewed from underneath we see how two vertical links drop from the swing arm to operate the two cranks at each end of the spring unit, compressing it in concertina fashion. Compare this with the 1982 Yamaha OW61 illustrated in the first chapter. (photo: Greg Parish, ACS.)
A possible advantage of compressing a suspension unit from each end is a reduction in effective unsprung mass. Due to the square law relationship each end of the spring will only contribute one eighth of the unsprung mass, adding the two ends means that the effective unsprung mass of the spring wil only be one quarter that of a the same spring compressed equally, but from one end. The effective unsprung mass of the damper is harder to determine because each end does not have equal mass.
A possible advantage of compressing a suspension unit from each end is a reduction in effective unsprung mass. Due to the square law relationship each end of the spring will only contribute one eighth of the unsprung mass, adding the two ends means that the effective unsprung mass of the spring wil only be one quarter that of a the same spring compressed equally, but from one end. The effective unsprung mass of the damper is harder to determine because each end does not have equal mass.
This Yamaha RD500LC of 1984 is interesting because the swing-arm connects directly to the rocker without an intermediary link, this means that the rocker must be connected to the main chassis by way of a short link. As the swing-arm moves upward the lower end of the rocker moves forward compressing the unit. Although made in the same year as the example to the left this Yamaha design has a more usual vertical mounting for the unit. Used on several different models this layout has no rocker in the normal sense, basically it just uses two links to define the movement path of the suspension unit.
This Yamaha RD500LC of 1984 is interesting because the swing-arm connects directly to the rocker without an intermediary link, this means that the rocker must be connected to the main chassis by way of a short link. As the swing-arm moves upward the lower end of the rocker moves forward compressing the unit. Although made in the same year as the example to the left this Yamaha design has a more usual vertical mounting for the unit. Used on several different models this layout has no rocker in the normal sense, basically it just uses two links to define the movement path of the suspension unit.
Once the idea of using rockers and linkage became accepted, it seems that most permutations of the design were tried out within a fairly short period around the beginning of the 1980s. Few of the curren designs are significantly different, even on GP racing machinery. A notable exception can be seen or the Tul-aris designed by Dr. Robin Tuluie. This uses a titanium “flexure link” which eliminates one pivoted joint in the system, reducing possible backlash and friction. The following figure shows how suck a link can be used. The rear end of the link only moves vertically by less than 2 mm. throughout the ful range of suspension travel, if the link is sufficiently long this movement can be provided by flex along its length. The front end therefore does not need to have a rotating joint and can be bolted directly anc rigidly to the chassis.
Once the idea of using rockers and linkage became accepted, it seems that most permutations of the design were tried out within a fairly short period around the beginning of the 1980s. Few of the curren designs are significantly different, even on GP racing machinery. A notable exception can be seen or the Tul-aris designed by Dr. Robin Tuluie. This uses a titanium “flexure link” which eliminates one pivoted joint in the system, reducing possible backlash and friction. The following figure shows how suck a link can be used. The rear end of the link only moves vertically by less than 2 mm. throughout the ful range of suspension travel, if the link is sufficiently long this movement can be provided by flex along its length. The front end therefore does not need to have a rotating joint and can be bolted directly anc rigidly to the chassis.
It is only over the past couple of years that F1 cars have started to use one dimensional flexure links in their minimal movement suspension systems but the Tul-aris goes one better with the use of a two dimensional version. In other words the link can flex in a horizontal direction as well as a vertical one, this allows for any misalignment caused by swing-arm and/or chassis flex. As shown in chapter 10 flexure can cause failure by metal fatigue and so the design of the link is of extreme importance, as failure could have serious consequences. It is only with the use of modern testing and computer analysis methods that such components can now be designed with confidence. | suspect that we'll see increasing use of this type of link in bikes from the major manufacturers. For racing use there is the possibility of improved performance, and for road machines, manufacturing costs can be reduced. Constructors without the knowledge and facilities for data collection, testing, simulating and fatigue analysis are strongly advised not to attempt the use of designs which rely on the deliberate flexing of suspension components.
It is only over the past couple of years that F1 cars have started to use one dimensional flexure links in their minimal movement suspension systems but the Tul-aris goes one better with the use of a two dimensional version. In other words the link can flex in a horizontal direction as well as a vertical one, this allows for any misalignment caused by swing-arm and/or chassis flex. As shown in chapter 10 flexure can cause failure by metal fatigue and so the design of the link is of extreme importance, as failure could have serious consequences. It is only with the use of modern testing and computer analysis methods that such components can now be designed with confidence. | suspect that we'll see increasing use of this type of link in bikes from the major manufacturers. For racing use there is the possibility of improved performance, and for road machines, manufacturing costs can be reduced. Constructors without the knowledge and facilities for data collection, testing, simulating and fatigue analysis are strongly advised not to attempt the use of designs which rely on the deliberate flexing of suspension components.
Chain effects Fig. 8.8 Both the tyre force and the force in the chain act on the swing- arm and its pivot. To a first approximation, F;, the swing-arm force is equal to the expression shown in the sketch. Where :
Chain effects Fig. 8.8 Both the tyre force and the force in the chain act on the swing- arm and its pivot. To a first approximation, F;, the swing-arm force is equal to the expression shown in the sketch. Where :
Fig.8.9 The sketch on the left shows how the path of the wheel movement follows a smaller radius (R1) than that required to maintain constant chain slack (R2). If the swing-arm pivot is concentric with the front sprocket then the two radii are the same and the chain slack will not vary. In most cases this form of construction is not practical and another way to achieve the same end is by the use of a parallelogram double swing-arm system as shown on the right. If the two swing- arms have the same length as the centre distance between the two sprockets then the motion of the wheel axle will follow the same curve as a single swing-arm pivoted about the front sprocket. That is, a virtual swing-arm is created. the swing-arm pivot concentric with the front sprocket, however, this is easier said than done. It woulc normally entail a very wide swing-arm at the front end.
Fig.8.9 The sketch on the left shows how the path of the wheel movement follows a smaller radius (R1) than that required to maintain constant chain slack (R2). If the swing-arm pivot is concentric with the front sprocket then the two radii are the same and the chain slack will not vary. In most cases this form of construction is not practical and another way to achieve the same end is by the use of a parallelogram double swing-arm system as shown on the right. If the two swing- arms have the same length as the centre distance between the two sprockets then the motion of the wheel axle will follow the same curve as a single swing-arm pivoted about the front sprocket. That is, a virtual swing-arm is created. the swing-arm pivot concentric with the front sprocket, however, this is easier said than done. It woulc normally entail a very wide swing-arm at the front end.
This early Bimota featured a two-part chassis with the rear fork pivoted on the gearbox sprocket axis. The extra width and weight seem a high price to pay to eliminate relatively small variations in chain slack. Other Bimota models have used the conventional system of mounting the swing-arm pivot behind the gearbox.
This early Bimota featured a two-part chassis with the rear fork pivoted on the gearbox sprocket axis. The extra width and weight seem a high price to pay to eliminate relatively small variations in chain slack. Other Bimota models have used the conventional system of mounting the swing-arm pivot behind the gearbox.
The geometry of swing-arm mounting and its relationship to chain runs is also very important to the anti- squat behaviour and these aspects are covered in a separate chapter. It is important to remember that any design must satisfy many diverse requirements. It is on moto-X and enduro machines, with their long suspension travel, that chain slack variation is really likely to be a serious problem. Despite numerous patented designs of varying complexity to counter this, the most generally used method is also probably the simplest. That is, a spring loaded roller or idler sprocket on the underside of the chain run. This simply takes up the excess slack with the swing-arm at the extremes of its range. Of the alternative methods the most practical is probably the use of two chains and an intermediate double sprocket mounted on the swing-arm pivot. The first chain is short and connects the gearbox sprocket to the intermediate one which in turn connects via the second chain to the rear sprocket. The A-Trak system, described in the next chapter, fig. 9.14, has apparently been used with success also, and is easy to retro-fit to some existing machines. Wheel trajectory
The geometry of swing-arm mounting and its relationship to chain runs is also very important to the anti- squat behaviour and these aspects are covered in a separate chapter. It is important to remember that any design must satisfy many diverse requirements. It is on moto-X and enduro machines, with their long suspension travel, that chain slack variation is really likely to be a serious problem. Despite numerous patented designs of varying complexity to counter this, the most generally used method is also probably the simplest. That is, a spring loaded roller or idler sprocket on the underside of the chain run. This simply takes up the excess slack with the swing-arm at the extremes of its range. Of the alternative methods the most practical is probably the use of two chains and an intermediate double sprocket mounted on the swing-arm pivot. The first chain is short and connects the gearbox sprocket to the intermediate one which in turn connects via the second chain to the rear sprocket. The A-Trak system, described in the next chapter, fig. 9.14, has apparently been used with success also, and is easy to retro-fit to some existing machines. Wheel trajectory
Fig. 8.10 The sketch to the right shows a fairly normal swing-arm layout, clearly displaying the near vertical path of the wheel. The bump force (F) acts radially inward from the tyre contact point and requires that the wheel moves backward as well as up in order to minimize the impact. The high mounting of the pivot, as shown left, of the Track-Lever system gives this type of trajectory, rearward from the vertical. Because of their high build, moto-X machines do approach this sort of layout to a limited degree, but the Track-Lever system as described in the following chapter, follows a much more rearward path similar in angle to that of telescopic front forks. The bump force acts radially inward and the ideal wheel trajectory is in-line with that force, then the force is all directed into the suspension unit before being passed on to the rest of the machine. If the bump force direction is not aligned with the wheel motion then a rearward component of the force has to be reacted by horizontal forces fed directly into the bulk of the machine without first being cushioned by the suspension. Bumps come in different sizes and so it is not possible to design a layout that is optimum in all circumstances, but larger bumps create larger forces and so it is more important to consider these.
Fig. 8.10 The sketch to the right shows a fairly normal swing-arm layout, clearly displaying the near vertical path of the wheel. The bump force (F) acts radially inward from the tyre contact point and requires that the wheel moves backward as well as up in order to minimize the impact. The high mounting of the pivot, as shown left, of the Track-Lever system gives this type of trajectory, rearward from the vertical. Because of their high build, moto-X machines do approach this sort of layout to a limited degree, but the Track-Lever system as described in the following chapter, follows a much more rearward path similar in angle to that of telescopic front forks. The bump force acts radially inward and the ideal wheel trajectory is in-line with that force, then the force is all directed into the suspension unit before being passed on to the rest of the machine. If the bump force direction is not aligned with the wheel motion then a rearward component of the force has to be reacted by horizontal forces fed directly into the bulk of the machine without first being cushioned by the suspension. Bumps come in different sizes and so it is not possible to design a layout that is optimum in all circumstances, but larger bumps create larger forces and so it is more important to consider these.
The simple swing-arm on this 1956 GP Gilera 500-4 was typical of those used by most racers, road and moto-X cross machines during the 1950s. and 1960s. Simple triangulation could have improved rigidity with a minor addition of weight.
The simple swing-arm on this 1956 GP Gilera 500-4 was typical of those used by most racers, road and moto-X cross machines during the 1950s. and 1960s. Simple triangulation could have improved rigidity with a minor addition of weight.
Fabricated in aluminium this arm from a Yamaha Ré6 is typical of modern designs. Deep section main arms are braced heavily above. Many racing swing-arms have the whole top section boxed in with sheet material to achieve the maximum stiffness.
Fabricated in aluminium this arm from a Yamaha Ré6 is typical of modern designs. Deep section main arms are braced heavily above. Many racing swing-arms have the whole top section boxed in with sheet material to achieve the maximum stiffness.
Typical modern racing swing- arm. It is double sided, made of aluminium alloy and heavily braced. Constructions like this are many times stiffer than swing-arms of old.
Typical modern racing swing- arm. It is double sided, made of aluminium alloy and heavily braced. Constructions like this are many times stiffer than swing-arms of old.
However, the majority of manufacturers still retain the standard of the past 50 years, the dual sided swing arm, in one form or another. Is this because of tradition, product liability considerations, styling or does the symmetrical two sided arm have inherent technical advantages? This book is only concerned with the technical aspects of that question and to find the answer we need to look dispassionately at what characteristics make a good arm and how the two approaches meet those requirements. Although single sided swing arms have long been used on small mopeds and scooters, it wasn’t until the Cortanze designed, Elf sponsored endurance racing bikes of the 1970s. that the removal of one of the two traditional arms became generally considered as a serious possibility for large sport and racing machines. Since then there has been a mixed reaction to its introduction. Honda have used it on a variety of sports and racing vehicles, both chain and shaft drive, Ducati have succeeded on the racetrack whereas BMW have changed over exclusively to single sided suspension on their shaft drive models, firstly with a simple rigid arm and latterly with an articulated “paralever” system to control the rise and squat. MV have decided to use a cast single sided arm on their new F4 superbike. In fact the use of single sided arms is now much more widespread than commonly thought.
However, the majority of manufacturers still retain the standard of the past 50 years, the dual sided swing arm, in one form or another. Is this because of tradition, product liability considerations, styling or does the symmetrical two sided arm have inherent technical advantages? This book is only concerned with the technical aspects of that question and to find the answer we need to look dispassionately at what characteristics make a good arm and how the two approaches meet those requirements. Although single sided swing arms have long been used on small mopeds and scooters, it wasn’t until the Cortanze designed, Elf sponsored endurance racing bikes of the 1970s. that the removal of one of the two traditional arms became generally considered as a serious possibility for large sport and racing machines. Since then there has been a mixed reaction to its introduction. Honda have used it on a variety of sports and racing vehicles, both chain and shaft drive, Ducati have succeeded on the racetrack whereas BMW have changed over exclusively to single sided suspension on their shaft drive models, firstly with a simple rigid arm and latterly with an articulated “paralever” system to control the rise and squat. MV have decided to use a cast single sided arm on their new F4 superbike. In fact the use of single sided arms is now much more widespread than commonly thought.
Left, Early version of the BMW single sided swing arm for shaft drive. The arm was fabricated from large diameter steel tube.
Left, Early version of the BMW single sided swing arm for shaft drive. The arm was fabricated from large diameter steel tube.
Chain drive single sided arm, interpretation by Honda. Note the levers each side of the arm just in front of the sprocket and brake disk, these take the torque from a floating brake caliper back to the main frame. Just like a normal floating caliper this gives control over the degree of rise. In order to mount the wheel bearings between the sprocket and disk the arm is quite complex in shape. This design necessitates a rotating axle which is large in diameter and hollow. As real designs of both single and double sided arms vary so much, it is difficult to compare like with like. The single sided arms for chain drive machines as produced by Honda, Ducati and others, use quite complex shapes, due to the design feature that places the sprocket outboard, and the structural properties of these arms can be evaluated with the use of FEA (Finite Element Analysis) techniques. This is a computer based method that considers the structure as being composed of many small elements joined together, according to set rules, to form the whole. Fig. 8.11 shows an example of how a Ducati cast magnesium arm was divided up for such analysis. Single sided arms for shaft drive machines are generally of simpler construction. The deflection characteristics of the two sided swingarm depend greatly on the nature and rigidity of the wheel spindle and its clamping method. The worst, typified by arms with a thin plate at the end slotted for chain adjustment, approach the case of a pin jointed axle. The best are probably found on shaft drive machines with a substantial axle clamp. Fig. 8.12 shows the deflection pattern in bending of the two extremes. The first is with no stiffening effect from the spindle assembly and the second is when it Structural comparison
Chain drive single sided arm, interpretation by Honda. Note the levers each side of the arm just in front of the sprocket and brake disk, these take the torque from a floating brake caliper back to the main frame. Just like a normal floating caliper this gives control over the degree of rise. In order to mount the wheel bearings between the sprocket and disk the arm is quite complex in shape. This design necessitates a rotating axle which is large in diameter and hollow. As real designs of both single and double sided arms vary so much, it is difficult to compare like with like. The single sided arms for chain drive machines as produced by Honda, Ducati and others, use quite complex shapes, due to the design feature that places the sprocket outboard, and the structural properties of these arms can be evaluated with the use of FEA (Finite Element Analysis) techniques. This is a computer based method that considers the structure as being composed of many small elements joined together, according to set rules, to form the whole. Fig. 8.11 shows an example of how a Ducati cast magnesium arm was divided up for such analysis. Single sided arms for shaft drive machines are generally of simpler construction. The deflection characteristics of the two sided swingarm depend greatly on the nature and rigidity of the wheel spindle and its clamping method. The worst, typified by arms with a thin plate at the end slotted for chain adjustment, approach the case of a pin jointed axle. The best are probably found on shaft drive machines with a substantial axle clamp. Fig. 8.12 shows the deflection pattern in bending of the two extremes. The first is with no stiffening effect from the spindle assembly and the second is when it Structural comparison
Fig. 8.12 The lateral deformation modes of the two extreme possibilities for the dual arm design, depending on how rigidly the spindle assembly attaches to the arms. On the left is the equivalent of a pin-jointed connection, and centre is the case with a completely rigid fixing between the axle and side arms. For clarity the two examples are shown with equal lateral displacements, but in reality the pinned case will deform by about 3 or 4 times more. The swingarm with a second cross-tube is shown to the right. (David Sanchez) To get an idea if either the twin sided or mono-arm designs have any overwhelming structural advantage over the other let’s consider an FEA of simple examples of each, in steel, and look at how they compare. Each type is analyzed in two forms as shown in figs. 8.12 & 8.13. The twin sided example is used with and without a second cross bracing tube, and the mono-arm with and without the gusset behind the pivot cross tube. Both cases use 457 mm. long arms and both have identical pivot tubes of 203 mm. width, 44.4 mm. diameter and 3.17 mm. wall thickness. For the traditional two sided design the arms are 51 x 25 x 1.59 mm. This was a typical configuration long used by many aftermarket manufacturers of replacement arms. The single sided arm chosen is that used on the Q2 (pictured below), which employed a 76 mm. diameter round tube with a 1.59 mm. wall. Thus both are from practical examples.
Fig. 8.12 The lateral deformation modes of the two extreme possibilities for the dual arm design, depending on how rigidly the spindle assembly attaches to the arms. On the left is the equivalent of a pin-jointed connection, and centre is the case with a completely rigid fixing between the axle and side arms. For clarity the two examples are shown with equal lateral displacements, but in reality the pinned case will deform by about 3 or 4 times more. The swingarm with a second cross-tube is shown to the right. (David Sanchez) To get an idea if either the twin sided or mono-arm designs have any overwhelming structural advantage over the other let’s consider an FEA of simple examples of each, in steel, and look at how they compare. Each type is analyzed in two forms as shown in figs. 8.12 & 8.13. The twin sided example is used with and without a second cross bracing tube, and the mono-arm with and without the gusset behind the pivot cross tube. Both cases use 457 mm. long arms and both have identical pivot tubes of 203 mm. width, 44.4 mm. diameter and 3.17 mm. wall thickness. For the traditional two sided design the arms are 51 x 25 x 1.59 mm. This was a typical configuration long used by many aftermarket manufacturers of replacement arms. The single sided arm chosen is that used on the Q2 (pictured below), which employed a 76 mm. diameter round tube with a 1.59 mm. wall. Thus both are from practical examples.
Fig. 8.14 The results of the FEA comparison. The stiffness values are shown relative to the best in each case. The /ower right sketch illustrates the three modes of displacement considered in the analysis, when subject to a lateral force (F) at the road/tyre junction . Camber and steer are angular motions and the lateral displacement is linear. Fig. 8.13 FEA model of the single sided swing-arm. The illustration to the left shows how the structure is divided up into small elements for analysis. The detail of the internal construction is shown to the right. The stabilizing plate is an example of the attention to detail that should be taken when using large section thin walled members. This plate largely prevents localized deformation at the point where the gusset joins to the side arm, reducing lateral displacement and stress concentration. In this case the plate increased the lateral stiffness by about 30%. The analysis considered two cases — with and without the gusset. (David Sanchez)
Fig. 8.14 The results of the FEA comparison. The stiffness values are shown relative to the best in each case. The /ower right sketch illustrates the three modes of displacement considered in the analysis, when subject to a lateral force (F) at the road/tyre junction . Camber and steer are angular motions and the lateral displacement is linear. Fig. 8.13 FEA model of the single sided swing-arm. The illustration to the left shows how the structure is divided up into small elements for analysis. The detail of the internal construction is shown to the right. The stabilizing plate is an example of the attention to detail that should be taken when using large section thin walled members. This plate largely prevents localized deformation at the point where the gusset joins to the side arm, reducing lateral displacement and stress concentration. In this case the plate increased the lateral stiffness by about 30%. The analysis considered two cases — with and without the gusset. (David Sanchez)
arms. On the other hand, if the side arms are splayed out toward the rear, as in most practical cases then some steer deflection will occur, as can be seen in fig.8.12. Therefore, it is hard to generalize abou this characteristic of the double arm design as a type. It is necessary to look at the detail of specific examples.
arms. On the other hand, if the side arms are splayed out toward the rear, as in most practical cases then some steer deflection will occur, as can be seen in fig.8.12. Therefore, it is hard to generalize abou this characteristic of the double arm design as a type. It is necessary to look at the detail of specific examples.
The mono-arm example used above is similar to this single arm used on the author’s Q2 design. The gusseting behind the pivot tube added to the overall stiffness by quite a significant amount as shown in the analysis.
The mono-arm example used above is similar to this single arm used on the author’s Q2 design. The gusseting behind the pivot tube added to the overall stiffness by quite a significant amount as shown in the analysis.
Rear hub detail of the Q2. This rotated about a fixed stub axle, shown above, similar to that used on the front of many cars. In fact the axle was taken directly from an XJ6 Jaguar car.
Rear hub detail of the Q2. This rotated about a fixed stub axle, shown above, similar to that used on the front of many cars. In fact the axle was taken directly from an XJ6 Jaguar car.
Showing the above single sided arm fitted to the Q2. Suspension was by means of a single Fournales air shock attached close to the rear of the arm.
Showing the above single sided arm fitted to the Q2. Suspension was by means of a single Fournales air shock attached close to the rear of the arm.
Honda’s shaft drive Revere. When the implementation is this easy there seems little point in considering a double sided configuration. Chain drive designs are seldom so simple.
Honda’s shaft drive Revere. When the implementation is this easy there seems little point in considering a double sided configuration. Chain drive designs are seldom so simple.
We can see that steady state load transfer is proportional to the bike’s mass, CoC height and acceleration, and inversely proportional to the wheelbase. each wheel. For example, when the CoG is rearward there is less load on the front wheel to start with and so it will need less load transfer before it lifts.
We can see that steady state load transfer is proportional to the bike’s mass, CoC height and acceleration, and inversely proportional to the wheelbase. each wheel. For example, when the CoG is rearward there is less load on the front wheel to start with and so it will need less load transfer before it lifts.
Fig. 9.1. To avoid looping at a steady acceleration of 1 g the CoG must be in front of and below a line drawn at 45°, through the rear tyre contact patch as shown. A similar line drawn through the front tyre backwards at 45° will define the CoG height limit to avoid forward looping with 1 g braking. These lines should be drawn with the bike in the attitude adopted under these conditions. The following illustration, fig. 9.1, shows how we can construct a line on a side view of bike and rider to determine a limit position for the CoG height to avoid a backwards rotation for any given value of steady acceleration. To avoid looping at 1 g (32.2 ft. /sec’ or 9.8 m/sec” ) acceleration then the overall CoG must not be any higher than a line drawn at 45°, as shown. So we see that the CoG can be raised if it is also moved forward. In practice many modern bikes have laden weight distributions of approximately 50/50 with CoG heights close to half of the wheelbase and hence are very close to our 1 g limit line. No wonder that it is so easy to wheelie most sports bikes, these days. Under the right conditions modern rubber allows us a co-efficient of friction a bit higher than 1, and if we constructed our limit line for an acceleration of say 1.2 g then it would have to be drawn lower at 40°. Of course this assumes that the bike has sufficient power to produce accelerations of this level. Dynamic transient effects also have a large influence on looping etc..
Fig. 9.1. To avoid looping at a steady acceleration of 1 g the CoG must be in front of and below a line drawn at 45°, through the rear tyre contact patch as shown. A similar line drawn through the front tyre backwards at 45° will define the CoG height limit to avoid forward looping with 1 g braking. These lines should be drawn with the bike in the attitude adopted under these conditions. The following illustration, fig. 9.1, shows how we can construct a line on a side view of bike and rider to determine a limit position for the CoG height to avoid a backwards rotation for any given value of steady acceleration. To avoid looping at 1 g (32.2 ft. /sec’ or 9.8 m/sec” ) acceleration then the overall CoG must not be any higher than a line drawn at 45°, as shown. So we see that the CoG can be raised if it is also moved forward. In practice many modern bikes have laden weight distributions of approximately 50/50 with CoG heights close to half of the wheelbase and hence are very close to our 1 g limit line. No wonder that it is so easy to wheelie most sports bikes, these days. Under the right conditions modern rubber allows us a co-efficient of friction a bit higher than 1, and if we constructed our limit line for an acceleration of say 1.2 g then it would have to be drawn lower at 40°. Of course this assumes that the bike has sufficient power to produce accelerations of this level. Dynamic transient effects also have a large influence on looping etc..
Fig. 9.3. Balance of forces on sprung part of the bike. Line 1 shows the force line passing through the CoG, and line 2 has one half of the slope of the other. Fh is equal in both cases as this is equal to the driving force.
Fig. 9.3. Balance of forces on sprung part of the bike. Line 1 shows the force line passing through the CoG, and line 2 has one half of the slope of the other. Fh is equal in both cases as this is equal to the driving force.
Fig. 9.4. R75/5 BMW. The force line passes close to the estimated loaded CoG thereby causing a very high anti-squat effect. Later models featured a lengthened swing arm which reduced the anti-squat to about 84% of the former value.
Fig. 9.4. R75/5 BMW. The force line passes close to the estimated loaded CoG thereby causing a very high anti-squat effect. Later models featured a lengthened swing arm which reduced the anti-squat to about 84% of the former value.
This 1950 MV Agusta racer had shaft drive and a rather elaborate way of partially isolating driving torque from the rear suspension, using a parallel pair of pivoted forks and a floating crown-wheel housing.
This 1950 MV Agusta racer had shaft drive and a rather elaborate way of partially isolating driving torque from the rear suspension, using a parallel pair of pivoted forks and a floating crown-wheel housing.
Fig. 9.5 shows the torque link parallel to the swing-arm, but a parallelogram system like this has a severe disadvantage. The degree of anti-squat effect will change greatly over the range of suspension movement. If the arm and link point upward toward the front when lightly loaded then there will be some anti-squat, but when the arm is horizontal as shown in the sketch there will be zero anti-squat, just like the case of a normal shaft drive with the pivot at ground level. When the suspension is loaded some more such that the arm points downward the situation gets worse because then there will be a pro-squat tendency. This changing anti-squat characteristic has the same effect as a reduction in rear spring rate, and the more power that is applied the more will be this effective reduction. This would seem to be exactly the opposite to what we want at a time when more load is being transferred to the rear. As we shall see later a partial solution to this problem is to use non-parallel links converging towards the front. The crown wheel housing is mounted in such a way that it is free to rotate on the axle. In this way the swing-arm is isolated from the torque of the housing. Thus the driving torque cannot be passed to the frame by the swing-arm, which can now only pass forces along its length. To prevent the housing frorr spinning backward we need to attach a torque reaction link as shown. So the swing-arm pushes forward on the main part of the bike while the torque link pulls back. The forward push must be greater than the pull back or the bike will be going nowhere. As the angular motion of the housing is no longet the same as that of the swing-arm, it becomes necessary to have a universal joint at the rear end of the drive shaft as well as the front, and usually it is necessary to make provision for changes in drive shafi lanath.
Fig. 9.5 shows the torque link parallel to the swing-arm, but a parallelogram system like this has a severe disadvantage. The degree of anti-squat effect will change greatly over the range of suspension movement. If the arm and link point upward toward the front when lightly loaded then there will be some anti-squat, but when the arm is horizontal as shown in the sketch there will be zero anti-squat, just like the case of a normal shaft drive with the pivot at ground level. When the suspension is loaded some more such that the arm points downward the situation gets worse because then there will be a pro-squat tendency. This changing anti-squat characteristic has the same effect as a reduction in rear spring rate, and the more power that is applied the more will be this effective reduction. This would seem to be exactly the opposite to what we want at a time when more load is being transferred to the rear. As we shall see later a partial solution to this problem is to use non-parallel links converging towards the front. The crown wheel housing is mounted in such a way that it is free to rotate on the axle. In this way the swing-arm is isolated from the torque of the housing. Thus the driving torque cannot be passed to the frame by the swing-arm, which can now only pass forces along its length. To prevent the housing frorr spinning backward we need to attach a torque reaction link as shown. So the swing-arm pushes forward on the main part of the bike while the torque link pulls back. The forward push must be greater than the pull back or the bike will be going nowhere. As the angular motion of the housing is no longet the same as that of the swing-arm, it becomes necessary to have a universal joint at the rear end of the drive shaft as well as the front, and usually it is necessary to make provision for changes in drive shafi lanath.
To prevent driving torque from extending the rear suspension excessively on this Foale-framed shaft-drive Guzzi the crown-wheel housing was made free floating and was connected to the main frame by a pivoted link above the swing- arm. Note the second Hooke’s joint at rear of drive shaft. Guzzi themselves later adopted a similar solution, along with a very similar chassis, for their production sports machines, the V11 sport.
To prevent driving torque from extending the rear suspension excessively on this Foale-framed shaft-drive Guzzi the crown-wheel housing was made free floating and was connected to the main frame by a pivoted link above the swing- arm. Note the second Hooke’s joint at rear of drive shaft. Guzzi themselves later adopted a similar solution, along with a very similar chassis, for their production sports machines, the V11 sport.
Fig. 9.6. Current model BMWs (R1100S shown here) feature the “paralever” system. Which despite the name quite obviously does not use parallel links. Note that the torque link is under the swing-arm unlike the MotoGuzzi which places it above. Under static load the BMW link is in tension, which is reduced with driving and increased under braking. The anti-squat value would remain reasonably constant throughout the range of suspension movement if the forward link pivot was relocated close to point “P”. The angle shown as 0 is often referred to as the anti-squat angle. of both lines are almost identical. Therefore the newer BMWs have an anti-squat value close to that which was previously reasoned to be close to the ideal, much better than their earlier brethren.
Fig. 9.6. Current model BMWs (R1100S shown here) feature the “paralever” system. Which despite the name quite obviously does not use parallel links. Note that the torque link is under the swing-arm unlike the MotoGuzzi which places it above. Under static load the BMW link is in tension, which is reduced with driving and increased under braking. The anti-squat value would remain reasonably constant throughout the range of suspension movement if the forward link pivot was relocated close to point “P”. The angle shown as 0 is often referred to as the anti-squat angle. of both lines are almost identical. Therefore the newer BMWs have an anti-squat value close to that which was previously reasoned to be close to the ideal, much better than their earlier brethren.
Fig. 9.7. Compare this diagram of chain and swing-arm forces with those in figs. 9.2 & 9.5 for shaft drive. The chain pulls back and down on the main part of the bike whilst the swing arm forces work in the opposite direction. The horizontal component of the swing-arm pushing force must be greater than that of the chain pull-back by the amount equal to the overall driving force. Note how with this layout the anti-squat force from the swing-arm is considerably higher than the pro-squat force generated by the chain pull, giving an overall anti-squat effect. As we'll see later when the suspension compresses past the horizontal position both the chain and swing-arm will point downwards at the front, when this happens it will be the chain pull that causes an anti-squat effect with an opposing squat tendency from the swing-arm force. We can determine whether the overall effect is one of pro-squat or anti-squat by constructing a force line as shown for shaft drives in fig. 9.6 and for chain drive in fig. 9.8.
Fig. 9.7. Compare this diagram of chain and swing-arm forces with those in figs. 9.2 & 9.5 for shaft drive. The chain pulls back and down on the main part of the bike whilst the swing arm forces work in the opposite direction. The horizontal component of the swing-arm pushing force must be greater than that of the chain pull-back by the amount equal to the overall driving force. Note how with this layout the anti-squat force from the swing-arm is considerably higher than the pro-squat force generated by the chain pull, giving an overall anti-squat effect. As we'll see later when the suspension compresses past the horizontal position both the chain and swing-arm will point downwards at the front, when this happens it will be the chain pull that causes an anti-squat effect with an opposing squat tendency from the swing-arm force. We can determine whether the overall effect is one of pro-squat or anti-squat by constructing a force line as shown for shaft drives in fig. 9.6 and for chain drive in fig. 9.8.
Fig. 9.8 Using a similar geometric construction to that used to analyze the BMW paralever system in fig. 9.6 we can draw a force line through the tyre contact patch and the ‘Instantaneous Force Centre’ determined by the junction of lines through the swing-arm and chain line. We can’t really talk about virtual arms & pivots in this case of chain drive with a single swing-arm because the motion of the wheel etc. is determined purely by the location of the ‘real’ swing-arm pivot. In the case of chain drive we can determine an ‘Instantaneous Force Centre’ by drawing lines througt the chain and swing-arm. The meeting point of these lines can be considered as the point through whict the combined chain and swing-arm forces act, this force must be balanced by equal and opposite forces at the tyre contact patch. (Neglecting the rear wheel mass as before.)
Fig. 9.8 Using a similar geometric construction to that used to analyze the BMW paralever system in fig. 9.6 we can draw a force line through the tyre contact patch and the ‘Instantaneous Force Centre’ determined by the junction of lines through the swing-arm and chain line. We can’t really talk about virtual arms & pivots in this case of chain drive with a single swing-arm because the motion of the wheel etc. is determined purely by the location of the ‘real’ swing-arm pivot. In the case of chain drive we can determine an ‘Instantaneous Force Centre’ by drawing lines througt the chain and swing-arm. The meeting point of these lines can be considered as the point through whict the combined chain and swing-arm forces act, this force must be balanced by equal and opposite forces at the tyre contact patch. (Neglecting the rear wheel mass as before.)
Fig. 9.10 shows how the anti-squat percentage varies with different layouts of swing-arm and sprocket positioning. Example ‘A’ shows how, when the swing-arm and chain line diverge towards the front, the instantaneous force centre can actually be behind the wheel. This has no special significance because it is the line through this force centre and the rear tyre contact patch which is important. However, when the force centre is behind the wheel then if it is above ground level we'll have pro-squat performance and anti-squat when it drops below the ground. This is opposite to the situation when the force centre is in front of the wheel. suspension positions. Fig. 9.9 shows the locations of the force line for different values of anti-sque percentage. A line with a negative slope shows the case when the chain and swing-arm reactions add t the suspension compression instead of opposing it — pro-squat.
Fig. 9.10 shows how the anti-squat percentage varies with different layouts of swing-arm and sprocket positioning. Example ‘A’ shows how, when the swing-arm and chain line diverge towards the front, the instantaneous force centre can actually be behind the wheel. This has no special significance because it is the line through this force centre and the rear tyre contact patch which is important. However, when the force centre is behind the wheel then if it is above ground level we'll have pro-squat performance and anti-squat when it drops below the ground. This is opposite to the situation when the force centre is in front of the wheel. suspension positions. Fig. 9.9 shows the locations of the force line for different values of anti-sque percentage. A line with a negative slope shows the case when the chain and swing-arm reactions add t the suspension compression instead of opposing it — pro-squat.
Fig. 9.10 Sketch ‘A’ is drawn exaggerated to show the real possibility that the instantaneous force centre can be behind the wheel and not always in front, in this case the anti-squat effect is negative, or pro-squat. Cases ‘B’ and ‘C’ show the differences in anti-squat that can occur on the same machine throughout the range of suspension movement. ‘B’ shows the suspension in the extended state and the slope of the force line shows a high degree of anti-squat. ‘C’ on the other hand is with the suspension in the compressed state and has a greatly reduced anti-squat tendency.
Fig. 9.10 Sketch ‘A’ is drawn exaggerated to show the real possibility that the instantaneous force centre can be behind the wheel and not always in front, in this case the anti-squat effect is negative, or pro-squat. Cases ‘B’ and ‘C’ show the differences in anti-squat that can occur on the same machine throughout the range of suspension movement. ‘B’ shows the suspension in the extended state and the slope of the force line shows a high degree of anti-squat. ‘C’ on the other hand is with the suspension in the compressed state and has a greatly reduced anti-squat tendency.
Fig. 9.11 Cases ‘B’ and ‘C’ from fig. 9.10 are superimposed to determine the anti-squat percentage in each case. The detail of the chain lines and rear sprockets are not shown for clarity. It is clear how the base for determining the 100% value is different for the two suspension positions. Scaling from the sketches gives 30% anti-squat when the suspension is in the compressed state as against 133% when extended. Although not drawn for simplicity, the anti-squat when in mid compression is equal to 100%.
Fig. 9.11 Cases ‘B’ and ‘C’ from fig. 9.10 are superimposed to determine the anti-squat percentage in each case. The detail of the chain lines and rear sprockets are not shown for clarity. It is clear how the base for determining the 100% value is different for the two suspension positions. Scaling from the sketches gives 30% anti-squat when the suspension is in the compressed state as against 133% when extended. Although not drawn for simplicity, the anti-squat when in mid compression is equal to 100%.
Such adjustability though doesn’t really solve the problem of such a wide range of anti-squat variatior with suspension movement. It just allows us to adjust the anti-squat to more desirable levels in the most important area of the suspension range. For example, keeping with fig. 9.11, if the swing-arm pivot is lowered by a little over 25 mm. then at full extension the anti-squat will be reduced to a more desirable 100% from the original 133%. A heavy price has to be paid at the other end of the suspension range though, here the anti-squat will change from 30% to about — 34%. A strong pro-squat tendency in other words. The adjustment has helped at one extreme but made the situation much worse at the other limit of suspension movement.
Such adjustability though doesn’t really solve the problem of such a wide range of anti-squat variatior with suspension movement. It just allows us to adjust the anti-squat to more desirable levels in the most important area of the suspension range. For example, keeping with fig. 9.11, if the swing-arm pivot is lowered by a little over 25 mm. then at full extension the anti-squat will be reduced to a more desirable 100% from the original 133%. A heavy price has to be paid at the other end of the suspension range though, here the anti-squat will change from 30% to about — 34%. A strong pro-squat tendency in other words. The adjustment has helped at one extreme but made the situation much worse at the other limit of suspension movement.
Sprocket size also has an effect on anti-squat behaviour as fig. 9.12 shows.
Sprocket size also has an effect on anti-squat behaviour as fig. 9.12 shows.
Fig. 9.13 Showing how to determine the layout requirements for constant anti-squat behaviour, 100% in this case. Firstly draw in the force lines, for 100%, from the tyre contact patches to the intersection of the CoG height and the vertical through the front wheel spindle. Next draw in the forward extensions of the swing-arm inclination, the instantaneous force centres are defined where these lines meet, points 1 and 2. The required chain directions are then found by drawing the chain lines through the force centres tangent to the rear sprocket. The insets show two different methods by which the desired chain run may be achieved in practice. Before looking at whether any of these alternatives have merit or not, let’s investigate the theoretical possibility of achieving constant anti-squat characteristics, and then see how close some of the other ideas come to accomplishing it in practice. Fig. 9.13 maintains the same basic dimensions used in fig. 9.11 but let’s work backwards and start from the force lines drawn to give 100% anti-squat for the full bump and full rebound positions, we know from the earlier discussion that the instantaneous force centres must lie along these lines. These force centres must also lie along the lines drawn through the swing-arm and so their intersections with the force lines must define the actual force centres. These are points 1 and 2 in the sketch, 1 is for the bump case and 2 is for rebound.
Fig. 9.13 Showing how to determine the layout requirements for constant anti-squat behaviour, 100% in this case. Firstly draw in the force lines, for 100%, from the tyre contact patches to the intersection of the CoG height and the vertical through the front wheel spindle. Next draw in the forward extensions of the swing-arm inclination, the instantaneous force centres are defined where these lines meet, points 1 and 2. The required chain directions are then found by drawing the chain lines through the force centres tangent to the rear sprocket. The insets show two different methods by which the desired chain run may be achieved in practice. Before looking at whether any of these alternatives have merit or not, let’s investigate the theoretical possibility of achieving constant anti-squat characteristics, and then see how close some of the other ideas come to accomplishing it in practice. Fig. 9.13 maintains the same basic dimensions used in fig. 9.11 but let’s work backwards and start from the force lines drawn to give 100% anti-squat for the full bump and full rebound positions, we know from the earlier discussion that the instantaneous force centres must lie along these lines. These force centres must also lie along the lines drawn through the swing-arm and so their intersections with the force lines must define the actual force centres. These are points 1 and 2 in the sketch, 1 is for the bump case and 2 is for rebound.
e Mounting the drive sprocket coaxial with the swing-arm pivot. This can be accomplished in one of two ways. By extending the swing-arm forward and putting the pivot in line with the gearbox sprocket, or by mounting a secondary pair of sprockets on the swing-arm pivot and using a second chain to drive it. The first approach is normally considered too difficult mechanically as it requires a very wide Let’s consider some of the approaches that have been used in practice: Clearly, constant anti-squat characteristics are feasible from both a theoretical and practical standpoint. This was demonstrated only for the 100% anti-squat case but also holds true for other values.
e Mounting the drive sprocket coaxial with the swing-arm pivot. This can be accomplished in one of two ways. By extending the swing-arm forward and putting the pivot in line with the gearbox sprocket, or by mounting a secondary pair of sprockets on the swing-arm pivot and using a second chain to drive it. The first approach is normally considered too difficult mechanically as it requires a very wide Let’s consider some of the approaches that have been used in practice: Clearly, constant anti-squat characteristics are feasible from both a theoretical and practical standpoint. This was demonstrated only for the 100% anti-squat case but also holds true for other values.
Fig. 9.14 The A-Trak system features two idler sprockets mounted on the swing-arm above and below the pivot. Because these sprockets are fixed to the swing-arm anc not the main chassis, it is the chain line from the gearbox sprocket to the idler that is relevant for estimation of anti-squat performance. The idler sprockets are drawr unrealistically small above purely to aid clarity.
Fig. 9.14 The A-Trak system features two idler sprockets mounted on the swing-arm above and below the pivot. Because these sprockets are fixed to the swing-arm anc not the main chassis, it is the chain line from the gearbox sprocket to the idler that is relevant for estimation of anti-squat performance. The idler sprockets are drawr unrealistically small above purely to aid clarity.
If the two swing-arms are arranged in a non-equal length, non-parallel layout we have some degree of independent control over the characteristics. Rather than analysing various possible geometries, let's follow the example of fig. 9.13 and see if we can synthesize a layout that can provide constant 100% anti-squat throughout the full suspension range. If the two swing-arms form a parallelogram, symmetrically disposed, and laid out as in fig. 9.15, which is the usual practice with this design, then all the effort is wasted. The virtual swing-arm will have its virtual pivot at infinity and be of infinite length, but the direction of the virtual arm will be identical to that of a single arm and so its intersection with the chain line will also be identical. Thus the anti-squat performance will not be changed from that of the single arm.
If the two swing-arms are arranged in a non-equal length, non-parallel layout we have some degree of independent control over the characteristics. Rather than analysing various possible geometries, let's follow the example of fig. 9.13 and see if we can synthesize a layout that can provide constant 100% anti-squat throughout the full suspension range. If the two swing-arms form a parallelogram, symmetrically disposed, and laid out as in fig. 9.15, which is the usual practice with this design, then all the effort is wasted. The virtual swing-arm will have its virtual pivot at infinity and be of infinite length, but the direction of the virtual arm will be identical to that of a single arm and so its intersection with the chain line will also be identical. Thus the anti-squat performance will not be changed from that of the single arm.
Fig. 9.16 Determining the swing-arm pivot position to achieve 100% anti-squat at both extremes of the suspension range. The intersection of the respective chains lines and the force lines for 100% define the instantaneous force centres (1 & 2). The swing-arm lines must pass through these points and where the two swing-arms meet, P, is the only common point and hence is the required pivot position.
Fig. 9.16 Determining the swing-arm pivot position to achieve 100% anti-squat at both extremes of the suspension range. The intersection of the respective chains lines and the force lines for 100% define the instantaneous force centres (1 & 2). The swing-arm lines must pass through these points and where the two swing-arms meet, P, is the only common point and hence is the required pivot position.
From these figures it can be seen that (in common with the A-Trak) most of the improvement occurs between normal ride height and full bump, this is likely to be the most important part of the suspension range, partly because the suspension becomes more compressed under hard cornering. examples, equally disposed about the neutral position, the anti-squat percentages change to 131% (133%) on rebound, 100% (100%) in mid stroke and 60% (30%) on full bump. The figures in brackets are from the conventional system shown in fig. 9.11. The range of variation with the Tracklever is therefore 2.2 to 1 which compares favourably with the 4.4 to 1 of the standard design.
From these figures it can be seen that (in common with the A-Trak) most of the improvement occurs between normal ride height and full bump, this is likely to be the most important part of the suspension range, partly because the suspension becomes more compressed under hard cornering. examples, equally disposed about the neutral position, the anti-squat percentages change to 131% (133%) on rebound, 100% (100%) in mid stroke and 60% (30%) on full bump. The figures in brackets are from the conventional system shown in fig. 9.11. The range of variation with the Tracklever is therefore 2.2 to 1 which compares favourably with the 4.4 to 1 of the standard design.
If the CP and CoG heights are the same then the amount of anti-squat needed to remove driving effects from the suspension will also be the same. This relationship will depend on the type of bike anc positioning of the rider. A racer crouched over the tank behind a fairing may well have a CP height close to that of the CoG, but a normally seated rider on a big tourer will in all probability have a higher CP. on aerodynamics we saw how the drag force causes a load transfer from the front wheel to the rear in like manner to the acceleration case. Fig. 9.18 shows how we can determine the direction of the force line to isolate the rear suspension from aerodynamic drag influences.
If the CP and CoG heights are the same then the amount of anti-squat needed to remove driving effects from the suspension will also be the same. This relationship will depend on the type of bike anc positioning of the rider. A racer crouched over the tank behind a fairing may well have a CP height close to that of the CoG, but a normally seated rider on a big tourer will in all probability have a higher CP. on aerodynamics we saw how the drag force causes a load transfer from the front wheel to the rear in like manner to the acceleration case. Fig. 9.18 shows how we can determine the direction of the force line to isolate the rear suspension from aerodynamic drag influences.
sharp torque to the swing-arm, tending to compress the suspension, and since the unsprung mass (the wheel etc.) is considerably less than that of the sprung mass, the wheel tends to leave the ground more quickly than the bulk of the machine can move downward, mainly due to gravity. As the wheel leaves the ground and the tyre therefore loses traction, the brake locks and the compressing moment vanishes. When the wheel returns to the ground, still locked, it may skid, and the shock load on the tyre may cause sufficient sudden braking torque to set the whole process in motion again — and again. This effect is likely to be more serious than the chattering previously described for the acceleration case. There are two main reasons for this, firstly the wheel is being momentarily lifted away from the ground rather thar being pressed onto it. Secondly the largest portion of the braking effort comes from the front and so under heavy braking there is little load on the rear wheel to start with. Riding style can determine whether this becomes a problem in practice, a rider who applies the brake in a gentle fashion may never experience trouble whereas someone with a less delicate touch may find it totally uncontrollable. (This is one of the reasons that different riders often report widely differing opinions on the handling characteristics of the same machine. Handling is affected by riding style) .
sharp torque to the swing-arm, tending to compress the suspension, and since the unsprung mass (the wheel etc.) is considerably less than that of the sprung mass, the wheel tends to leave the ground more quickly than the bulk of the machine can move downward, mainly due to gravity. As the wheel leaves the ground and the tyre therefore loses traction, the brake locks and the compressing moment vanishes. When the wheel returns to the ground, still locked, it may skid, and the shock load on the tyre may cause sufficient sudden braking torque to set the whole process in motion again — and again. This effect is likely to be more serious than the chattering previously described for the acceleration case. There are two main reasons for this, firstly the wheel is being momentarily lifted away from the ground rather thar being pressed onto it. Secondly the largest portion of the braking effort comes from the front and so under heavy braking there is little load on the rear wheel to start with. Riding style can determine whether this becomes a problem in practice, a rider who applies the brake in a gentle fashion may never experience trouble whereas someone with a less delicate touch may find it totally uncontrollable. (This is one of the reasons that different riders often report widely differing opinions on the handling characteristics of the same machine. Handling is affected by riding style) .
The layout in Fig. 9.21. shows a floating brake mounting with two alternative locations for the torque stay pivot. With the torque stay mounted high up and well back, the virtual pivot for this arrangement is at the intersection of the swing-arm and the torque arm, we can see that this gives a very high anti-rise percentage. Although we can evaluate the anti-rise characteristic in a similar way to the anti-squat, by constructing < force line through either the real or virtual pivot position, actually deciding on the most desirable degrees of anti-rise is much harder than for anti-squat, because of the infinity of ways that front and rear braking can be combined. Fig.9.20 shows how the anti-rise depends on the swing arm length for the conventional system without a floating caliper mount (in this case we can use the actual swing-arm there is no need for a virtual one) , a short arm may give more than 100% anti-rise, i.e. the geometn may cause the anti-rise to exceed the rise due to load transfer and the rear may actually sink dow under the action of the rear brake alone. Any anti-rise such as this, relies on force and torque reaction: which come from rear braking, therefore these effects will be completely absent when using only the front brake.
The layout in Fig. 9.21. shows a floating brake mounting with two alternative locations for the torque stay pivot. With the torque stay mounted high up and well back, the virtual pivot for this arrangement is at the intersection of the swing-arm and the torque arm, we can see that this gives a very high anti-rise percentage. Although we can evaluate the anti-rise characteristic in a similar way to the anti-squat, by constructing < force line through either the real or virtual pivot position, actually deciding on the most desirable degrees of anti-rise is much harder than for anti-squat, because of the infinity of ways that front and rear braking can be combined. Fig.9.20 shows how the anti-rise depends on the swing arm length for the conventional system without a floating caliper mount (in this case we can use the actual swing-arm there is no need for a virtual one) , a short arm may give more than 100% anti-rise, i.e. the geometn may cause the anti-rise to exceed the rise due to load transfer and the rear may actually sink dow under the action of the rear brake alone. Any anti-rise such as this, relies on force and torque reaction: which come from rear braking, therefore these effects will be completely absent when using only the front brake.
The lower position of the pivot is situated to give about 100% anti-rise, thereby cancelling out any rear suspension reaction to rear wheel braking. As with the BMW paralever and chain drive constructions for anti-squat determination, we draw in a force line through the intersection of the torque arm and swing- arm (virtual pivot). Sometimes the high mounted torque stay is suggested as a way of keeping more pressure on the tyre during braking. This location will give a higher anti-rise effect than the conventional system with the caliper mounted to the swing-arm directly, and the possible disadvantages of that (mentioned above) will be worsened. The idea that the braking reaction will pull the bike down on to the road, was explained, in the first paragraph of this chapter, to be impossible under steady state conditions. The initial transient will in fact work toward reducing the dynamic tyre load.
The lower position of the pivot is situated to give about 100% anti-rise, thereby cancelling out any rear suspension reaction to rear wheel braking. As with the BMW paralever and chain drive constructions for anti-squat determination, we draw in a force line through the intersection of the torque arm and swing- arm (virtual pivot). Sometimes the high mounted torque stay is suggested as a way of keeping more pressure on the tyre during braking. This location will give a higher anti-rise effect than the conventional system with the caliper mounted to the swing-arm directly, and the possible disadvantages of that (mentioned above) will be worsened. The idea that the braking reaction will pull the bike down on to the road, was explained, in the first paragraph of this chapter, to be impossible under steady state conditions. The initial transient will in fact work toward reducing the dynamic tyre load.
Surprising as it may seem, we can also evaluate the dive characteristics of telescopic forks by using the concepts of virtual swing-arms, pivots and percentage anti-dive (as shown in fig. 9.24) in a similar manner to that applied to the squat and rise problem of the rear suspension . The axle follows a straight path in line with the forks, as shown, if the fork legs are replaced by an infinitely long virtual swing-arm mounted at the wheel spindle and at right angles to the sliders, then the wheel motion will be unchanged, and the line through the contact patch and the virtual pivot will again define the anti-dive percentage. This line can be compared to the line required for 100% anti-dive, defined in similar fashion to the anti- squat percentage shown previously, except that this time it is the CoG height over the rear axle that is the base for determining what constitutes 100%. These graphs are normalized such that the value of the displacement at a castor angle of zero is made equal to 1.0. They are drawn assuming single rate fork springs and ignoring the reduction in rake angle with braking, because this will depend on the actual spring rates used.
Surprising as it may seem, we can also evaluate the dive characteristics of telescopic forks by using the concepts of virtual swing-arms, pivots and percentage anti-dive (as shown in fig. 9.24) in a similar manner to that applied to the squat and rise problem of the rear suspension . The axle follows a straight path in line with the forks, as shown, if the fork legs are replaced by an infinitely long virtual swing-arm mounted at the wheel spindle and at right angles to the sliders, then the wheel motion will be unchanged, and the line through the contact patch and the virtual pivot will again define the anti-dive percentage. This line can be compared to the line required for 100% anti-dive, defined in similar fashion to the anti- squat percentage shown previously, except that this time it is the CoG height over the rear axle that is the base for determining what constitutes 100%. These graphs are normalized such that the value of the displacement at a castor angle of zero is made equal to 1.0. They are drawn assuming single rate fork springs and ignoring the reduction in rake angle with braking, because this will depend on the actual spring rates used.
Fig. 9.24 The forks move in a straight line and so could be replaced by an infinitely long swing-arm. The anti-dive force line can then be constructed between the tyre contact patch and the virtual swing-arm pivot and is therefore parallel to the virtual swing-arm. We can define 100% anti-dive by using similar reasoning to that for anti-squat.
Fig. 9.24 The forks move in a straight line and so could be replaced by an infinitely long swing-arm. The anti-dive force line can then be constructed between the tyre contact patch and the virtual swing-arm pivot and is therefore parallel to the virtual swing-arm. We can define 100% anti-dive by using similar reasoning to that for anti-squat.
Two very different approaches to the problem of controlling dive by opposing manufacturers. Above is a 1979 Suzuki RG500. Hydraulic pressure from the braking circuit is fed to valves on the forks which control the flow of damping fluid, slowing the rate of dive only. To the right is a Kawasaki GP racer, using a totally mechanical system. The calipers are attached to floating mounts which in turn are prevented from rotating by means of the approximately vertical links.
Two very different approaches to the problem of controlling dive by opposing manufacturers. Above is a 1979 Suzuki RG500. Hydraulic pressure from the braking circuit is fed to valves on the forks which control the flow of damping fluid, slowing the rate of dive only. To the right is a Kawasaki GP racer, using a totally mechanical system. The calipers are attached to floating mounts which in turn are prevented from rotating by means of the approximately vertical links.
Many of these alternatives are frequently described as having "natural anti-dive", but this is not really < helpful way of looking at their characteristics, because it ignores many of the subtleties and in many cases a small change to their geometry could in fact produce a pro-dive effect. However, the geometric layout often allows for a wide adjustment to their “anti-dive” properties.
Many of these alternatives are frequently described as having "natural anti-dive", but this is not really < helpful way of looking at their characteristics, because it ignores many of the subtleties and in many cases a small change to their geometry could in fact produce a pro-dive effect. However, the geometric layout often allows for a wide adjustment to their “anti-dive” properties.
Another problem here is loss of feel — some degree of dive seems necessary to aid the rider’s judgement. On forks with shorter leading links, it is usual to incorporate a floating brake anchorage to prevent full fork extension, then any desired degree of anti-dive can be designed in by manipulating the On old Earles-fork BMWs, where the brake shoe plate is directly clamped to the front pivoted fork, the action is similar to the effects of the shaft drive at the rear — i.e. brake torque tends to extend the suspension in opposition to the effect of forward weight transfer. Front braking alone would often cause the suspension to top-out losing sensitivity to road shocks.
Another problem here is loss of feel — some degree of dive seems necessary to aid the rider’s judgement. On forks with shorter leading links, it is usual to incorporate a floating brake anchorage to prevent full fork extension, then any desired degree of anti-dive can be designed in by manipulating the On old Earles-fork BMWs, where the brake shoe plate is directly clamped to the front pivoted fork, the action is similar to the effects of the shaft drive at the rear — i.e. brake torque tends to extend the suspension in opposition to the effect of forward weight transfer. Front braking alone would often cause the suspension to top-out losing sensitivity to road shocks.
As an example of short leading links with a floating brake plate fig. 9.28 shows a Moto-Guzzi racer Ignoring perspective distortion in the photo, the anti-dive can be seen be around 64%, this is a muct better value than the Earles forks above. Giving approximately 20% of the dive present with telescopics Note how the links slope downward toward the rear, this gives a wheel axle movement over bumps similar to telescopics, that is rearward as well as upward. In the front suspension chapter we saw tha this aids bump absorbsion and comfort. There is some variation in the anti-dive percentage, as the suspension loads the anti-dive reduces, but not by as much as many others. This can be a greater problem with short link forks in general, the shorter the link the greater is the range of angular movemen for a given suspension displacement. The virtual pivot will move about more with this greater range o movement. Double-link, hub-centre and similar
As an example of short leading links with a floating brake plate fig. 9.28 shows a Moto-Guzzi racer Ignoring perspective distortion in the photo, the anti-dive can be seen be around 64%, this is a muct better value than the Earles forks above. Giving approximately 20% of the dive present with telescopics Note how the links slope downward toward the rear, this gives a wheel axle movement over bumps similar to telescopics, that is rearward as well as upward. In the front suspension chapter we saw tha this aids bump absorbsion and comfort. There is some variation in the anti-dive percentage, as the suspension loads the anti-dive reduces, but not by as much as many others. This can be a greater problem with short link forks in general, the shorter the link the greater is the range of angular movemen for a given suspension displacement. The virtual pivot will move about more with this greater range o movement. Double-link, hub-centre and similar
Fig. 9.29 The GTS Yamaha, shown with the suspension links in loaded and unloaded positions. In the unloaded attitude the suspension links converge toward the front putting the virtual pivot ahead of the bike (not fully shown), as a result the unloaded anti-dive force-line has a negative slope giving approximately half of the pro-dive effect of telescopic forks (-44% compared to —93% at 25° rake). When loaded the relatively short top link angles more than the lower arm and the virtual pivot moves rearward giving an anti-dive of around 26%.
Fig. 9.29 The GTS Yamaha, shown with the suspension links in loaded and unloaded positions. In the unloaded attitude the suspension links converge toward the front putting the virtual pivot ahead of the bike (not fully shown), as a result the unloaded anti-dive force-line has a negative slope giving approximately half of the pro-dive effect of telescopic forks (-44% compared to —93% at 25° rake). When loaded the relatively short top link angles more than the lower arm and the virtual pivot moves rearward giving an anti-dive of around 26%.
Load transfer onto the front tyre is not instantaneous, it is passed to the tyre through the fork springs anc damping, and so it can only build up as the fork compresses. At the instant of applying the front brake < rearward horizontal force (Fb) will be created, and due to the rearward rake angle, this will produce < component of this force in the direction of the fork springs (Fs). Therefore until sufficient load transfer Telescopic forks, or any other type which has strong pro-dive characteristics, exhibit a strange and undesirable effect immediately after the application of the brake. The vertical load on the front wheel can actually decrease for a short period, which increases the chances of inadvertently locking the wheel. Conversely, when the braking force is removed or reduced, as for example when the wheel locks, this type of front suspension will cause a momentary increase in vertical tyre force which may help restore full braking. This may seem to be in direct contradiction to the known phenomenon of forward load transfer, but is quite easily explainable, consider figs. 9.22 and 9.30.
Load transfer onto the front tyre is not instantaneous, it is passed to the tyre through the fork springs anc damping, and so it can only build up as the fork compresses. At the instant of applying the front brake < rearward horizontal force (Fb) will be created, and due to the rearward rake angle, this will produce < component of this force in the direction of the fork springs (Fs). Therefore until sufficient load transfer Telescopic forks, or any other type which has strong pro-dive characteristics, exhibit a strange and undesirable effect immediately after the application of the brake. The vertical load on the front wheel can actually decrease for a short period, which increases the chances of inadvertently locking the wheel. Conversely, when the braking force is removed or reduced, as for example when the wheel locks, this type of front suspension will cause a momentary increase in vertical tyre force which may help restore full braking. This may seem to be in direct contradiction to the known phenomenon of forward load transfer, but is quite easily explainable, consider figs. 9.22 and 9.30.
Fig. 9.31 Computer simulation showing the vertical tyre to road loading during the application of hard braking and then at 1 second the sudden reduction of the coefficient of friction to 0.8 to represent a loss of adhesion. Note that to show the effects more clearly the vertical scale is expanded on the second graph. The dark line is representative of normal telescopic forks and the lighter line shows a system with 100% anti-dive.
Fig. 9.31 Computer simulation showing the vertical tyre to road loading during the application of hard braking and then at 1 second the sudden reduction of the coefficient of friction to 0.8 to represent a loss of adhesion. Note that to show the effects more clearly the vertical scale is expanded on the second graph. The dark line is representative of normal telescopic forks and the lighter line shows a system with 100% anti-dive.
Fig. 9.32 These plots show the effect of a step input of driving power on tyre force (upper) and suspension compression (lower), at different values of anti-squat percentage. The dark curves are for 0%, the lightest for 200% and the remaining curves show the results for 100%. variation in these values between different motorcycles is more likely to be just between different values of anti-squat, unlike the swing between pro-dive and anti-dive usual between different designs of front suspension. The largest differences at the rear are between chain drive and simple shaft drive designs, such as shown in fig. 9.4. Fig. 9.32 shows the results of some simulations of acceleration with different values of anti-squat.
Fig. 9.32 These plots show the effect of a step input of driving power on tyre force (upper) and suspension compression (lower), at different values of anti-squat percentage. The dark curves are for 0%, the lightest for 200% and the remaining curves show the results for 100%. variation in these values between different motorcycles is more likely to be just between different values of anti-squat, unlike the swing between pro-dive and anti-dive usual between different designs of front suspension. The largest differences at the rear are between chain drive and simple shaft drive designs, such as shown in fig. 9.4. Fig. 9.32 shows the results of some simulations of acceleration with different values of anti-squat.
Fig. 9.33 The plot to the left shows the velocity history for the three values of anti-squat used in fig. 9.32. starting from an initial velocity of 100km/h. The darkest line is for 0% anti-squat, the lightest for 200% and the other for 100%. In a racing context, the most significant aspect of the above simply boils down to the question of what speed we can reach in a given time. The left hand curves in fig.9.33 show the speeds obtained usinc the three values of anti-squat from the previous example. Only the first 0.3 second is shown as this is were the tyre load variation was greatest.
Fig. 9.33 The plot to the left shows the velocity history for the three values of anti-squat used in fig. 9.32. starting from an initial velocity of 100km/h. The darkest line is for 0% anti-squat, the lightest for 200% and the other for 100%. In a racing context, the most significant aspect of the above simply boils down to the question of what speed we can reach in a given time. The left hand curves in fig.9.33 show the speeds obtained usinc the three values of anti-squat from the previous example. Only the first 0.3 second is shown as this is were the tyre load variation was greatest.
Fig. 9.34 Tyre loading for rear wheel braking only. The dark line is for 0% anti-rise, typical of a floating brake design with parallelogram linkage in mid-position. The lighter line shows 200% which is typical of a fixed brake anchorage with a short swing-arm. The right side sketch clearly demonstrates the coupled natured of motorcycle suspension. Due to load transfer and pitch motion the dynamic loading on the front varies with changes to the rear under rear only braking. Unlike the previous two control requirements, front braking and acceleration, braking on the rear wheel causes a load transfer which unloads the rear suspension and braking tyre, i.e. causes rear rise. This greatly increases the chances that wheel hopping becomes bad enough for the tyre to intermittently leave the ground. We've seen in figs. 9.19 to 9.21 how we can alter the anti-rise percentage, by changing swing-arm length and/or fitting a swinging brake mount with a separate torque link to the chassis.
Fig. 9.34 Tyre loading for rear wheel braking only. The dark line is for 0% anti-rise, typical of a floating brake design with parallelogram linkage in mid-position. The lighter line shows 200% which is typical of a fixed brake anchorage with a short swing-arm. The right side sketch clearly demonstrates the coupled natured of motorcycle suspension. Due to load transfer and pitch motion the dynamic loading on the front varies with changes to the rear under rear only braking. Unlike the previous two control requirements, front braking and acceleration, braking on the rear wheel causes a load transfer which unloads the rear suspension and braking tyre, i.e. causes rear rise. This greatly increases the chances that wheel hopping becomes bad enough for the tyre to intermittently leave the ground. We've seen in figs. 9.19 to 9.21 how we can alter the anti-rise percentage, by changing swing-arm length and/or fitting a swinging brake mount with a separate torque link to the chassis.
Fig. 9.35 These two cases with both wheels braking are represented by the same line shading as in fig. 9.34 — dark represents 0% rear anti-rise and the lighter lines are for 200%. Note that the rear tyre leaves the ground around 0.1 seconds in the 0% case These simulations assume that both brakes are dynamically operated to achieve the optimum braking possible, even though this is very difficult for a real rider to do in practice. Like the case of rear wheel only braking, two cases are considered with 0% and 200% anti-rise at the rear, but in each case the front remains fixed at 100% pro-dive — telescopic forks in other words. Fig. 9.35 shows the tyre loading on both front and rear tyres and fig. 9.36 shows the pitch angle and suspension compression.
Fig. 9.35 These two cases with both wheels braking are represented by the same line shading as in fig. 9.34 — dark represents 0% rear anti-rise and the lighter lines are for 200%. Note that the rear tyre leaves the ground around 0.1 seconds in the 0% case These simulations assume that both brakes are dynamically operated to achieve the optimum braking possible, even though this is very difficult for a real rider to do in practice. Like the case of rear wheel only braking, two cases are considered with 0% and 200% anti-rise at the rear, but in each case the front remains fixed at 100% pro-dive — telescopic forks in other words. Fig. 9.35 shows the tyre loading on both front and rear tyres and fig. 9.36 shows the pitch angle and suspension compression.
Fig. 9.36 Showing the pitch angle to the front and the suspension compression for the same two cases used in fig. 9.35. Note the large difference in the initial slopes of the pitch curves around 0.1 seconds, the darker line shows that the 0% anti-rise case builds up a higher pitch velocity which in turn helps explain the rear wheel lifting off the ground in this case.
Fig. 9.36 Showing the pitch angle to the front and the suspension compression for the same two cases used in fig. 9.35. Note the large difference in the initial slopes of the pitch curves around 0.1 seconds, the darker line shows that the 0% anti-rise case builds up a higher pitch velocity which in turn helps explain the rear wheel lifting off the ground in this case.
Fig. 9.37 Rear wheel vertical displacement due to the action of front wheel only braking. Negative values indicate tyre compression and positive values show when the tyre has left the ground. The dark curve shows the motion with no rear spring preload, in this case the rear wheel stays on the ground. The light curve is when the same rear spring is preloaded by 25mm. This is enough to cause the rear wheel to start hopping with an increasing magnitude, which would probably lead to the rider easing off on the brake.
Fig. 9.37 Rear wheel vertical displacement due to the action of front wheel only braking. Negative values indicate tyre compression and positive values show when the tyre has left the ground. The dark curve shows the motion with no rear spring preload, in this case the rear wheel stays on the ground. The light curve is when the same rear spring is preloaded by 25mm. This is enough to cause the rear wheel to start hopping with an increasing magnitude, which would probably lead to the rider easing off on the brake.
Fig. 9.38 Front wheel motion under acceleration. The dark curve has no front preload and the lighter plot is with 25 mm. The acceleration is just below the level needed for a steady state wheelie, but in both of these cases there are small initial wheelies due to dynamic effects, as often seen when accelerating.
Fig. 9.38 Front wheel motion under acceleration. The dark curve has no front preload and the lighter plot is with 25 mm. The acceleration is just below the level needed for a steady state wheelie, but in both of these cases there are small initial wheelies due to dynamic effects, as often seen when accelerating.
An example of the use of idler sprockets to control the anti-squat behaviour as suggested in the text. This single cylinder racer was built by Chris Cosentino in the USA, and also features a Hossack style front suspension with a very steep steering axis of just a few degrees. The sprocket at (A) controls squat under power and (B) controls the overrun. The inset shows more detail of the sprocket layout. Chain-drive too, introduces some anti-squat, but with long wheel movement bikes this sometimes may turn to pro-squat when the suspension is compressed. In any event, without the use of some extra mechanical features the anti-squat effect will vary throughout the range of suspension movement. A graphical method was shown to determine either a swing-arm pivot or front sprocket location to give substantially constant anti-squat characteristics. Unfortunately, the required positions are unlikely to be practical in most cases but a method making use of idler sprockets to achieve the same end was suggested.
An example of the use of idler sprockets to control the anti-squat behaviour as suggested in the text. This single cylinder racer was built by Chris Cosentino in the USA, and also features a Hossack style front suspension with a very steep steering axis of just a few degrees. The sprocket at (A) controls squat under power and (B) controls the overrun. The inset shows more detail of the sprocket layout. Chain-drive too, introduces some anti-squat, but with long wheel movement bikes this sometimes may turn to pro-squat when the suspension is compressed. In any event, without the use of some extra mechanical features the anti-squat effect will vary throughout the range of suspension movement. A graphical method was shown to determine either a swing-arm pivot or front sprocket location to give substantially constant anti-squat characteristics. Unfortunately, the required positions are unlikely to be practical in most cases but a method making use of idler sprockets to achieve the same end was suggested.
A perfect demonstration of both dive and load transference off the rear wheel as this bike is set up for a turn. Refer to fig. 9.37. (SM30) feedback indicating the level of braking. There are other riders who feel that as the degree of dive varies from bike to bike, depending on front spring rate and various geometric parameters, that it is not a reliable indicator of how close the tyre is to reaching its maximum grip. After all, the drivers of F1 racing cars experience very little dive but seem to have appropriate feedback for the task.
A perfect demonstration of both dive and load transference off the rear wheel as this bike is set up for a turn. Refer to fig. 9.37. (SM30) feedback indicating the level of braking. There are other riders who feel that as the degree of dive varies from bike to bike, depending on front spring rate and various geometric parameters, that it is not a reliable indicator of how close the tyre is to reaching its maximum grip. After all, the drivers of F1 racing cars experience very little dive but seem to have appropriate feedback for the task.
Structural efficiency in nature. This section through a bone from a bird’s wing shows a mixture of triangulation and honeycomb techniques to achieve a high level of stiffness to weight. Fig 10.1 Under force F, the four-sided structure easily distorts to a lozenge shape, relying on corner stiffness alone to prevent complete collapse. For the triangular structure to distort as shown would require side A to lengthen and side B to shorten. To visualize the effect of triangulation, we need only consider the two simple structures illustrated igure 10.1. If their bases are fixed while a force is applied as shown, then the four-sided frame m jistort to a lozenge shape, with complete collapse prevented only by the tubes’ resistance to bending he corners. In contrast, the triangular frame can distort only by a change in length of any or all thr sides.
Structural efficiency in nature. This section through a bone from a bird’s wing shows a mixture of triangulation and honeycomb techniques to achieve a high level of stiffness to weight. Fig 10.1 Under force F, the four-sided structure easily distorts to a lozenge shape, relying on corner stiffness alone to prevent complete collapse. For the triangular structure to distort as shown would require side A to lengthen and side B to shorten. To visualize the effect of triangulation, we need only consider the two simple structures illustrated igure 10.1. If their bases are fixed while a force is applied as shown, then the four-sided frame m jistort to a lozenge shape, with complete collapse prevented only by the tubes’ resistance to bending he corners. In contrast, the triangular frame can distort only by a change in length of any or all thr sides.
Of course, the four-sided frame may be stiffened tremendously by adding one or two diagonals (known as bracing struts), so converting it to two or four triangles. If only one bracing strut is added, it must be of sufficient diameter to resist compression loads if the direction of the applied force can be reversed. But if two diagonals are added they can be in very thin material, even wire, because one or the other will A few experiments with a piece of wire or welding rod will show how easy it is to bend but how difficult it is to change in length, thus proving how much more rigid the triangular arrangement can be. A practical structure such as a motorcycle frame may comprise several such triangles, and, if designed correctly, should be very efficient. To test whether a design is fully triangulated or not, imagine that all the connections are by pin-joints — i.e. the joints have no resistance to bending. If the structure remains intact when loaded, then it is fully triangulated and may be regarded as a complete structure. If it collapses, however (as would our four-sided figure), then it is structurally unsound and is called a mechanism.
Of course, the four-sided frame may be stiffened tremendously by adding one or two diagonals (known as bracing struts), so converting it to two or four triangles. If only one bracing strut is added, it must be of sufficient diameter to resist compression loads if the direction of the applied force can be reversed. But if two diagonals are added they can be in very thin material, even wire, because one or the other will A few experiments with a piece of wire or welding rod will show how easy it is to bend but how difficult it is to change in length, thus proving how much more rigid the triangular arrangement can be. A practical structure such as a motorcycle frame may comprise several such triangles, and, if designed correctly, should be very efficient. To test whether a design is fully triangulated or not, imagine that all the connections are by pin-joints — i.e. the joints have no resistance to bending. If the structure remains intact when loaded, then it is fully triangulated and may be regarded as a complete structure. If it collapses, however (as would our four-sided figure), then it is structurally unsound and is called a mechanism.
In laying out a triangulated frame, simple considerations of space and shape may present difficulties, since some engines just don’t lend themselves to that type of construction. Nevertheless there are one or two ways of getting out of trouble. For example, three additional solutions to the problem of stiffening a four-sided structure where there is no room for a direct bracing strut are shown in figure 10.3. The Motorcycle frames comprising bolted triangular structures have been built and they resemble our theoretical pin-jointed example. Although this approach has some advantages, such as simplified repair, there are disadvantages, too. Movement may develop in the joints — as a result of wear, corrosion or even manufacturing clearances — and this will negate the stiffness benefits of triangulation. Also, the nuts and bolts at the joints may weigh more than welding. Indeed, it is more usual to weld the joints, though this too can cause problems. In practice, it is rarely possible to ensure that all the loads in the individual tubes are either compressive or tensile. The very thickness of the tubes introduces physical offsets from the perfect triangular form, which in turn creates bending moments that are sometimes impossible to calculate. To complicate the issue further, the welding process itself will leave indeterminate residual stresses in the chassis, while the joints themselves may cause stress concentrations (see later). All these considerations mean that appropriate safety factors have to be applied to the calculated stresses if fatigue life is to be adequate.
In laying out a triangulated frame, simple considerations of space and shape may present difficulties, since some engines just don’t lend themselves to that type of construction. Nevertheless there are one or two ways of getting out of trouble. For example, three additional solutions to the problem of stiffening a four-sided structure where there is no room for a direct bracing strut are shown in figure 10.3. The Motorcycle frames comprising bolted triangular structures have been built and they resemble our theoretical pin-jointed example. Although this approach has some advantages, such as simplified repair, there are disadvantages, too. Movement may develop in the joints — as a result of wear, corrosion or even manufacturing clearances — and this will negate the stiffness benefits of triangulation. Also, the nuts and bolts at the joints may weigh more than welding. Indeed, it is more usual to weld the joints, though this too can cause problems. In practice, it is rarely possible to ensure that all the loads in the individual tubes are either compressive or tensile. The very thickness of the tubes introduces physical offsets from the perfect triangular form, which in turn creates bending moments that are sometimes impossible to calculate. To complicate the issue further, the welding process itself will leave indeterminate residual stresses in the chassis, while the joints themselves may cause stress concentrations (see later). All these considerations mean that appropriate safety factors have to be applied to the calculated stresses if fatigue life is to be adequate.
This sketch of the Bimota KB2 chassis shows the considerable use of external triangulation and the multi-tube support for the steering head. The large amount of welding involved restricts such a layout to small-scale production. second and third methods are examples of what might be called external triangulation — and the second is the basis of the Vincent-type of pivoted rear fork. An example of the use of external triangulation is the Bimota KB2 frame, in which the steering head is supported by several tubes. As a complex chassis, involving many tubes and much welding, this would not be a viable proposition for a large manufacturer, but it is acceptable in the context of Bimota’s specialist output. Indeed, such is its visual impact that it may have marketing as well as technical merits.
This sketch of the Bimota KB2 chassis shows the considerable use of external triangulation and the multi-tube support for the steering head. The large amount of welding involved restricts such a layout to small-scale production. second and third methods are examples of what might be called external triangulation — and the second is the basis of the Vincent-type of pivoted rear fork. An example of the use of external triangulation is the Bimota KB2 frame, in which the steering head is supported by several tubes. As a complex chassis, involving many tubes and much welding, this would not be a viable proposition for a large manufacturer, but it is acceptable in the context of Bimota’s specialist output. Indeed, such is its visual impact that it may have marketing as well as technical merits.
For a machine with a conventional steering head it is possible to design a simple triangulated structure connecting the steering head to the rear-fork pivot, as shown in figure 10.4. However, this is seldom practicable because of the modifications necessary to accommodate the engine, seat and other components. Nevertheless the author managed this with frame for a 250cc Rotax-powered racer which was built on these lines.
For a machine with a conventional steering head it is possible to design a simple triangulated structure connecting the steering head to the rear-fork pivot, as shown in figure 10.4. However, this is seldom practicable because of the modifications necessary to accommodate the engine, seat and other components. Nevertheless the author managed this with frame for a 250cc Rotax-powered racer which was built on these lines.
This Foale frame for a 250 cc Rotax-powered racer is of a simple triangulated type, as shown in fig. 10.4. The narrowness of the engine allowed this frame design, which is normally hard to implement with other engines.
This Foale frame for a 250 cc Rotax-powered racer is of a simple triangulated type, as shown in fig. 10.4. The narrowness of the engine allowed this frame design, which is normally hard to implement with other engines.
If these tubes are subjected to a force (F) as shown, tending to bend them, the bending moment at the base of the segments is F x L. Imagine the bases to be pivoted at point O and the bending moment resisted by a force at the centre (f; and 6). To balance the applied moment, f; and f: are inversely proportional to the tube diameters — i.e. if the larger tube is twice the diameter of the other, then fo will have half the value of f;. See figure 10.5. Now let us assume that the deflection at the centre is proportional to this force — i.e. double for the smaller tube. But since this double deflection is at only half the radius, the net effect is that the angular deflection of the smaller tube is four times that of the larger tube. This angular deflection translates to a lateral movement at the top of the member, thus, for the same loading, the deflection about 0 of the top segment of the double-diameter tube is only a quarter of the deflection of the other. The same reasoning applies to all the segments of equal-length members, therefore the resistance to an applied lateral load is four times as great when tube diameter is doubled. Although the above is a simplified method of considering the problem, a rigid mathematical analysis gives the same results: the lateral stiffness of equal-length, equal-weight, thin-wall tubes of the same material is proportional to the square of the diameter. Torsional stiffness follows the same rule. And, although we have used round tubes to illustrate the principle, it holds for all other cross- trAmMH AA Shee
If these tubes are subjected to a force (F) as shown, tending to bend them, the bending moment at the base of the segments is F x L. Imagine the bases to be pivoted at point O and the bending moment resisted by a force at the centre (f; and 6). To balance the applied moment, f; and f: are inversely proportional to the tube diameters — i.e. if the larger tube is twice the diameter of the other, then fo will have half the value of f;. See figure 10.5. Now let us assume that the deflection at the centre is proportional to this force — i.e. double for the smaller tube. But since this double deflection is at only half the radius, the net effect is that the angular deflection of the smaller tube is four times that of the larger tube. This angular deflection translates to a lateral movement at the top of the member, thus, for the same loading, the deflection about 0 of the top segment of the double-diameter tube is only a quarter of the deflection of the other. The same reasoning applies to all the segments of equal-length members, therefore the resistance to an applied lateral load is four times as great when tube diameter is doubled. Although the above is a simplified method of considering the problem, a rigid mathematical analysis gives the same results: the lateral stiffness of equal-length, equal-weight, thin-wall tubes of the same material is proportional to the square of the diameter. Torsional stiffness follows the same rule. And, although we have used round tubes to illustrate the principle, it holds for all other cross- trAmMH AA Shee
Figure 10.6 shows a side view of a structural member subject to a bending load, this shortens the top surface (subject to compression) and stretches the lower surface (subject to tension). Somewhere between the two surfaces is a position of zero length change and this is known as the neutral axis. The bending produces no tension or compression stresses on this axis — unlike the outer surfaces, which are subject to stress that can be calculated by the following formula:
Figure 10.6 shows a side view of a structural member subject to a bending load, this shortens the top surface (subject to compression) and stretches the lower surface (subject to tension). Somewhere between the two surfaces is a position of zero length change and this is known as the neutral axis. The bending produces no tension or compression stresses on this axis — unlike the outer surfaces, which are subject to stress that can be calculated by the following formula:
Fig 10.7 Thin ribs or gussets may increase stresses. Unless | (second moment of area) is increased in proportion to c, the maximum fibre stress in the rib will be higher than in a plain tube. Fig 10.8 Two types of gusset failure: tearing under tensile stress (left) and buckling under compression. From this it follows that, to keep surface stresses to a minimum, we must keep c as small as possible consistent with a large value for |. This requirement highlights a danger inherent in the use of simple sheet-metal gussets to stiffen a tubular chassis, especially where the tubes are of large diameter.
Fig 10.7 Thin ribs or gussets may increase stresses. Unless | (second moment of area) is increased in proportion to c, the maximum fibre stress in the rib will be higher than in a plain tube. Fig 10.8 Two types of gusset failure: tearing under tensile stress (left) and buckling under compression. From this it follows that, to keep surface stresses to a minimum, we must keep c as small as possible consistent with a large value for |. This requirement highlights a danger inherent in the use of simple sheet-metal gussets to stiffen a tubular chassis, especially where the tubes are of large diameter.
Unless ’strengthening’ gussets or ribs are sized correctly, therefore, there is a risk that they will actually weaken the structure. The same applies to ribs on castings. Another danger in the use of ribs, brackets and other causes of sharp changes in section is stress concentration. To explain this, figure 10.9 shows a solid bar subject to tensile stress, which may be represented by a series of equally spaced ‘force lines’. When these force lines meet a sudden change in section, such as a notch, they bunch up in the vicinity to raise the local stress to two or three (or even more) times the average across the section.
Unless ’strengthening’ gussets or ribs are sized correctly, therefore, there is a risk that they will actually weaken the structure. The same applies to ribs on castings. Another danger in the use of ribs, brackets and other causes of sharp changes in section is stress concentration. To explain this, figure 10.9 shows a solid bar subject to tensile stress, which may be represented by a series of equally spaced ‘force lines’. When these force lines meet a sudden change in section, such as a notch, they bunch up in the vicinity to raise the local stress to two or three (or even more) times the average across the section.
Fig 10.10 Wrong (top) and better (lower) ways to weld a gusset on to a bent tube.
Fig 10.10 Wrong (top) and better (lower) ways to weld a gusset on to a bent tube.
Legendary designer Taglioni showing off a triangulated Ducati frame. Current versions of such triangulated frames made in steel tubing have proved very successful in SuperBike racing against the generally more fashionable aluminium twin spar machinery. Such frames have a high structural efficiency. This particular frame is structurally incomplete without the engine. (Cathcart)
Legendary designer Taglioni showing off a triangulated Ducati frame. Current versions of such triangulated frames made in steel tubing have proved very successful in SuperBike racing against the generally more fashionable aluminium twin spar machinery. Such frames have a high structural efficiency. This particular frame is structurally incomplete without the engine. (Cathcart)
Triangulation en extremis: a Krauser frame for a road going BMW transverse flat twin. Its high price reflected the large amount of labour involved in its construction.
Triangulation en extremis: a Krauser frame for a road going BMW transverse flat twin. Its high price reflected the large amount of labour involved in its construction.
Structural comparison We have seen that there are two basic structural approaches, some form of beam frame or a triangulated structure. Most real frames use either of these methods or some combination of both. Accordingly it is interesting to compare the stiffness characteristics of elementary versions of each type. Some designers have combined a curved backbone with use of the engine as a structural member. An example was the 125 cc Honda twin that spearheaded the Japanese invasion of the TT in 1959 — the small-diameter back-bone being curved through nearly 90 degrees and the engine bolted in to stiffen the assembly.
Structural comparison We have seen that there are two basic structural approaches, some form of beam frame or a triangulated structure. Most real frames use either of these methods or some combination of both. Accordingly it is interesting to compare the stiffness characteristics of elementary versions of each type. Some designers have combined a curved backbone with use of the engine as a structural member. An example was the 125 cc Honda twin that spearheaded the Japanese invasion of the TT in 1959 — the small-diameter back-bone being curved through nearly 90 degrees and the engine bolted in to stiffen the assembly.
Both types of frame were analysed for torsional and lateral loading. In it’s purest form the triangulated frame can be considered as pin jointed, and as such it is quite simple to calculate the stiffness properties manually, as it is with the simple backbone frame. However, we must remember that a real world triangulated frame is welded and not pin jointed. In this case it is easier to do a simple Finite Element Analysis and in the interests of uniformity this was used for all cases. The results in the following table are shown in a normalized form, comparing the properties to those of a tubular backbone of 100 mm. OD. and 1.0 mm. wall thickness. The wall thicknesses of the other examples were adjusted to give equal weight. Fig. 10.11 shows two such examples. The beam frame is a simple circular backbone similar to th frame shown in the upper photograph on page 1-10, and the triangulated example is as shown in fic 10.4. These represent the simplest and purest forms of both approaches to achieving structur: efficiency, even though they are seldom practical in real situations without modification.
Both types of frame were analysed for torsional and lateral loading. In it’s purest form the triangulated frame can be considered as pin jointed, and as such it is quite simple to calculate the stiffness properties manually, as it is with the simple backbone frame. However, we must remember that a real world triangulated frame is welded and not pin jointed. In this case it is easier to do a simple Finite Element Analysis and in the interests of uniformity this was used for all cases. The results in the following table are shown in a normalized form, comparing the properties to those of a tubular backbone of 100 mm. OD. and 1.0 mm. wall thickness. The wall thicknesses of the other examples were adjusted to give equal weight. Fig. 10.11 shows two such examples. The beam frame is a simple circular backbone similar to th frame shown in the upper photograph on page 1-10, and the triangulated example is as shown in fic 10.4. These represent the simplest and purest forms of both approaches to achieving structur: efficiency, even though they are seldom practical in real situations without modification.
The fabricated light-alloy back-bone on Kork Ballington’s Kawasaki KR500 GP machine comprised a 32 litre fuel tank with the steering head incorporated at the front. Aluminium side plates at the rear supported the engine and swing-arm pivot (MCW) Quite a range of designs is possible here — the most popular being a T-shape structure comprising lef and right steel pressings united by spot or electric-resistance welding, as in the Ariel Leader and the once highly popular Yamaha FS1E. This construction makes for rigidity and low production cost, thougt the high initial tooling outlay rules it out for small production runs and specials. Also, the end product is heavier than an equally rigid tubular backbone because of the inevitable excess metal in areas of low stress.
The fabricated light-alloy back-bone on Kork Ballington’s Kawasaki KR500 GP machine comprised a 32 litre fuel tank with the steering head incorporated at the front. Aluminium side plates at the rear supported the engine and swing-arm pivot (MCW) Quite a range of designs is possible here — the most popular being a T-shape structure comprising lef and right steel pressings united by spot or electric-resistance welding, as in the Ariel Leader and the once highly popular Yamaha FS1E. This construction makes for rigidity and low production cost, thougt the high initial tooling outlay rules it out for small production runs and specials. Also, the end product is heavier than an equally rigid tubular backbone because of the inevitable excess metal in areas of low stress.
On Santiago Herrero’s highly successful 1960s 250 cc Ossa GP racer the welded aluminium chassis doubled as the fuel tank. This Spanish machine came within a hair’s width of toppling the might of Honda for the world championship. (MCW)
On Santiago Herrero’s highly successful 1960s 250 cc Ossa GP racer the welded aluminium chassis doubled as the fuel tank. This Spanish machine came within a hair’s width of toppling the might of Honda for the world championship. (MCW)
The highly successful NSU “flying-hammock” record breaker of the mid-1950s had a true monocoque construction. The shell was in 1 mm thick aluminium alloy sheet, reinforced by bulkheads. Very similar to aircraft construction techniques. machines described as monocoques should more properly be said to have fabricated backbones. The original Honda NR500 racer was an exception, with the fairing an integral part of the bike. An unfortunate result, considering the frequent attention required in racing, was the fact that extensive dismantling was necessary for many routine maintenance tasks. (It is worth noting that the British Maxton concern was later commissioned to produce a more conventional frame for this machine.) A true candidate for the term monocoque was the NSU ‘flying hammock’ record breaker of the early 1950s. In this, the primary structural strength was provided by the fully streamlined shell, hand-beaten in
The highly successful NSU “flying-hammock” record breaker of the mid-1950s had a true monocoque construction. The shell was in 1 mm thick aluminium alloy sheet, reinforced by bulkheads. Very similar to aircraft construction techniques. machines described as monocoques should more properly be said to have fabricated backbones. The original Honda NR500 racer was an exception, with the fairing an integral part of the bike. An unfortunate result, considering the frequent attention required in racing, was the fact that extensive dismantling was necessary for many routine maintenance tasks. (It is worth noting that the British Maxton concern was later commissioned to produce a more conventional frame for this machine.) A true candidate for the term monocoque was the NSU ‘flying hammock’ record breaker of the early 1950s. In this, the primary structural strength was provided by the fully streamlined shell, hand-beaten in
Stripped down to the bare chassis this BMW boxer shows the simplicity that can be achieved by using the engine as the main structural member. The rear sub-frame to support the seat and rear suspension unit is the most complex part of the whole. The engine and gearbox unit form a very stiff and strong assembly quite capable of handling chassis loads. Potentially, this is the most efficient way to build a bike with a large engine — and the post-war Vincent V- twin was an outstanding example. The principle here is to use the inherent stiffness of the engine- gearbox unit to provide the major support between the steering head and the rear-suspension pivot. If that pivot is incorporated in the rear of the gearbox casting, then a simple lightweight structure will usually suffice to join the steering head to the top of the engine. On the Vincent the rear-fork pivot was clamped between aluminium-alloy plates at the back of the gearbox, while both cylinder heads were attached to a fabricated backbone that doubled as an oil tank and incorporated the bolted-on steering head. More recently, the Norton-Cosworth parallel-twin racer demonstrated the potential of this method. The rear-fork pivot was cast in the back of the gearbox while the steering head was supported by a triangulated frame bolted to the cylinder head and a simple structure reached rearward to support the seat. Unfortunately, development was halted by the Norton company collapse.
Stripped down to the bare chassis this BMW boxer shows the simplicity that can be achieved by using the engine as the main structural member. The rear sub-frame to support the seat and rear suspension unit is the most complex part of the whole. The engine and gearbox unit form a very stiff and strong assembly quite capable of handling chassis loads. Potentially, this is the most efficient way to build a bike with a large engine — and the post-war Vincent V- twin was an outstanding example. The principle here is to use the inherent stiffness of the engine- gearbox unit to provide the major support between the steering head and the rear-suspension pivot. If that pivot is incorporated in the rear of the gearbox casting, then a simple lightweight structure will usually suffice to join the steering head to the top of the engine. On the Vincent the rear-fork pivot was clamped between aluminium-alloy plates at the back of the gearbox, while both cylinder heads were attached to a fabricated backbone that doubled as an oil tank and incorporated the bolted-on steering head. More recently, the Norton-Cosworth parallel-twin racer demonstrated the potential of this method. The rear-fork pivot was cast in the back of the gearbox while the steering head was supported by a triangulated frame bolted to the cylinder head and a simple structure reached rearward to support the seat. Unfortunately, development was halted by the Norton company collapse.
Stiffness was enhanced in the Moto Guzzi V8 racer and record breaker by incorporating the rear swing-arm pivot lug in the back of the gearbox casting. (MCW) assembly largely bridging the gap between the wheels. At the rear the swing-arm mounts are completely integral with the gearbox castings, a construction method also used by a few past and present manufacturers of chain driven machines. At the front there is a simple structure to hold the dummy head-stock and the suspension “A” arm is attached directly to the engine. In some ways the use of the telelever front-end has made such a construction much easier as the loads fed back from the front suspension are spread over a wider area with lower local loads. Surprisingly, the most complicated part of the external structure is the subframe for mounting the seat and rear suspension unit.
Stiffness was enhanced in the Moto Guzzi V8 racer and record breaker by incorporating the rear swing-arm pivot lug in the back of the gearbox casting. (MCW) assembly largely bridging the gap between the wheels. At the rear the swing-arm mounts are completely integral with the gearbox castings, a construction method also used by a few past and present manufacturers of chain driven machines. At the front there is a simple structure to hold the dummy head-stock and the suspension “A” arm is attached directly to the engine. In some ways the use of the telelever front-end has made such a construction much easier as the loads fed back from the front suspension are spread over a wider area with lower local loads. Surprisingly, the most complicated part of the external structure is the subframe for mounting the seat and rear suspension unit.
The Parker inspired front suspension necessitates a different type of frame structure for the Yamaha GTS. It is hard tc classify this frame in any of the normal categories, the engine is structural but is not the principal structure. Aluminiurr castings at the front and rear support the suspension arms at each end. These castings are supported by a bolted or casting (not shown in this photo) that runs over the gearbox and to the side of the cylinders. A simple tubular frame extends from these casting up to support the top end of the steering column. Refer to chapter 7 for more details.
The Parker inspired front suspension necessitates a different type of frame structure for the Yamaha GTS. It is hard tc classify this frame in any of the normal categories, the engine is structural but is not the principal structure. Aluminiurr castings at the front and rear support the suspension arms at each end. These castings are supported by a bolted or casting (not shown in this photo) that runs over the gearbox and to the side of the cylinders. A simple tubular frame extends from these casting up to support the top end of the steering column. Refer to chapter 7 for more details.
With the engine installed, this Harris Magnum frame has a high degree of stiffness, helped by triangulation at the steering head, which is supported by eight tubes. The structure was formed with bronze welded 531 tubing. (Parker) This frame type was traditionally made in welded steel tube and most current examples still are, but there was a period in the 1980s and early 1990s when some manufacturers used aluminium often in square or rectangular section, until the lower stressed twin spar became more common. This is not the best way to use aluminium tubing and this form of construction seemed to be marketing led design.
With the engine installed, this Harris Magnum frame has a high degree of stiffness, helped by triangulation at the steering head, which is supported by eight tubes. The structure was formed with bronze welded 531 tubing. (Parker) This frame type was traditionally made in welded steel tube and most current examples still are, but there was a period in the 1980s and early 1990s when some manufacturers used aluminium often in square or rectangular section, until the lower stressed twin spar became more common. This is not the best way to use aluminium tubing and this form of construction seemed to be marketing led design.
Comparing this GSX1200 multi-tubular frame from 1999, with simple swing-arm and two rear suspension units, to the Hayabusa (pictured later) of the same period may lead one to believe that the designers at Suzuki were somewhat schizophrenic. In reality it just shows that frame designers are under fashion and image constraints as much as technical ones. In overall design this frame is little changed from those of twenty years before.
Comparing this GSX1200 multi-tubular frame from 1999, with simple swing-arm and two rear suspension units, to the Hayabusa (pictured later) of the same period may lead one to believe that the designers at Suzuki were somewhat schizophrenic. In reality it just shows that frame designers are under fashion and image constraints as much as technical ones. In overall design this frame is little changed from those of twenty years before.
A 1984 Yamaha FJ1100 displays an unusual variant of the multi-tubular frame. It was called the “Lateral Frame Concept”. The tubes surrounding the engine are kept wide at the front and the head- stock is supported in-side with multiple tubes. Some-what after the style set by the Bimota KB2 pictured earlier.
A 1984 Yamaha FJ1100 displays an unusual variant of the multi-tubular frame. It was called the “Lateral Frame Concept”. The tubes surrounding the engine are kept wide at the front and the head- stock is supported in-side with multiple tubes. Some-what after the style set by the Bimota KB2 pictured earlier.
This Suzuki GSX-R 1100 from 1992 is typical of the application of aluminium to a relatively conventional frame. This is not the best way to use this material and appears to have been a stop gap until the twin spar took over. Note that the rear sections that hold the swing-arm pivot and just above are open section castings. The rear sub frame is bolted on as are the engine support tubes that pass under and in front of the engine. This enables engine removal.
This Suzuki GSX-R 1100 from 1992 is typical of the application of aluminium to a relatively conventional frame. This is not the best way to use this material and appears to have been a stop gap until the twin spar took over. Note that the rear sections that hold the swing-arm pivot and just above are open section castings. The rear sub frame is bolted on as are the engine support tubes that pass under and in front of the engine. This enables engine removal.
From a structural viewpoint the twin-spar frame could be considered as a bifurcated backbone or a boxed in version of the simple triangulated structure shown in fig. 10.4. Experience shows that given enough metal it can achieve sufficient stiffness, even for racing duty, but from a structural efficiency basis (stiffness to weight ratio) this type of frame is not particularly good. Even made in aluminium it is not a construction method that gives a particularly light frame, if made in steel as most other frames are, it would seem positively obese. However, historically very few production motorcycle chassis seem to have been designed with structural efficiency high on the list of priorities, and other considerations often take precedence. Success on the race track and in the showroom is of prime interest to manufacturers and from that aspect the twin-spar frame must be considered as successful as the multi-tube type was a few decades ago. In all cases castings have often been used for the head stock area and the swing-arm and rear engine mounting plates. These castings are welded to the side spars. It is quite usual but not universal to have bolted on seat frames, which strangely have sometimes been made of steel tube.
From a structural viewpoint the twin-spar frame could be considered as a bifurcated backbone or a boxed in version of the simple triangulated structure shown in fig. 10.4. Experience shows that given enough metal it can achieve sufficient stiffness, even for racing duty, but from a structural efficiency basis (stiffness to weight ratio) this type of frame is not particularly good. Even made in aluminium it is not a construction method that gives a particularly light frame, if made in steel as most other frames are, it would seem positively obese. However, historically very few production motorcycle chassis seem to have been designed with structural efficiency high on the list of priorities, and other considerations often take precedence. Success on the race track and in the showroom is of prime interest to manufacturers and from that aspect the twin-spar frame must be considered as successful as the multi-tube type was a few decades ago. In all cases castings have often been used for the head stock area and the swing-arm and rear engine mounting plates. These castings are welded to the side spars. It is quite usual but not universal to have bolted on seat frames, which strangely have sometimes been made of steel tube.
Contemporary variations on the same theme. Above is the R1 from Yamaha and below is the Suzuki Hayabusa.
Contemporary variations on the same theme. Above is the R1 from Yamaha and below is the Suzuki Hayabusa.
Above: The 1998 Kawasaki Ninja used a twin-spar frame very little different from its main competitors shown above. Note that in common with the R1 and Hayabusa the front engine mounting is quite high up, rendering traditional down-tubes and bottom cradle unnecessary, thereby freeing up valuable space. Modern large engines and their voluminous air-boxes present a huge challenge to the frame designer and the twin-spar offers considerable packaging advantages. Especially in the racing context this type of frame allows much easier access to work on the engine, in particular, removal and replacement of carburettors and access to spark plugs. The elimination of down-tubes and the lower cradle also frees up valuable space in the area needed by exhaust and cooling systems.
Above: The 1998 Kawasaki Ninja used a twin-spar frame very little different from its main competitors shown above. Note that in common with the R1 and Hayabusa the front engine mounting is quite high up, rendering traditional down-tubes and bottom cradle unnecessary, thereby freeing up valuable space. Modern large engines and their voluminous air-boxes present a huge challenge to the frame designer and the twin-spar offers considerable packaging advantages. Especially in the racing context this type of frame allows much easier access to work on the engine, in particular, removal and replacement of carburettors and access to spark plugs. The elimination of down-tubes and the lower cradle also frees up valuable space in the area needed by exhaust and cooling systems.
Both the V-twin engined Suzuki TL1000s (left) and the Bimota DB4 have frames with a beam down each side of the engine and as such have similarities to the normal twin-spar. The principal difference being that the spars are formed as triangulated structures rather than box sections. Left: The twin-spar is no longer reserved solely for pavement bikes as this shot of the frame for a Honda CR250 moto-X machine shows. Note the ribbed castings used for the swing-arm and rear engine mounting. The head stock is also cast and is joined to the rear by bent extruded rectangular tube sections.
Both the V-twin engined Suzuki TL1000s (left) and the Bimota DB4 have frames with a beam down each side of the engine and as such have similarities to the normal twin-spar. The principal difference being that the spars are formed as triangulated structures rather than box sections. Left: The twin-spar is no longer reserved solely for pavement bikes as this shot of the frame for a Honda CR250 moto-X machine shows. Note the ribbed castings used for the swing-arm and rear engine mounting. The head stock is also cast and is joined to the rear by bent extruded rectangular tube sections.
Unusual carbon fibre frame made in Japan for the CBX750 Honda engine for endurance racing. The frame was reported to weigh less than 6 kgf. Called the Black Buffalo after the shape and colour of the frame. The backbone inspired design is very high above the engine and it would be very difficult to fit an airbox of the volume commonly used today, but perhaps the frame could double up as a substitute for that function.
Unusual carbon fibre frame made in Japan for the CBX750 Honda engine for endurance racing. The frame was reported to weigh less than 6 kgf. Called the Black Buffalo after the shape and colour of the frame. The backbone inspired design is very high above the engine and it would be very difficult to fit an airbox of the volume commonly used today, but perhaps the frame could double up as a substitute for that function.
The frame for this GP racing 500cc. Heron Suzuki from the late 1970s is difficult to fit into any category. Sometimes called monocoque but it is hardly that. The boxy appearance is a result of being constructed from aluminium sided honeycomb material like that shown in fig. 13.3. The inset shows details of how this material is bent. Reportedly this chassis was very rigid and reduced lap times, but for whatever reason, development was not continued for very long. The front brake is of a rim type, but unlike the stainless steel example illustrated in chapter 12, this disk was made of carbon. Some frames don’t fit into the descriptions above for a variety of reasons. The material used fo. construction dictates how the frame needs to be designed. This would be the case for example with any attempt to make a chassis from composite material such as carbon fibre. To use such material purely as a substitute for steel or aluminium in a similar design would be ill advised, each type of material needs é structural design to suit its own special properties. The Heron Suzuki pictured below is a good example. This was made from honeycomb material, as described in chapter 13, this material is not really suitable for bending, and so when this is necessary the inner sheet is cut away so that the bend only occurs on the outer sheet. To restore the strength of the inner sheet a folded strip is glued over the cut. This form of construction leads to the sharp edged boxy appearance shown.
The frame for this GP racing 500cc. Heron Suzuki from the late 1970s is difficult to fit into any category. Sometimes called monocoque but it is hardly that. The boxy appearance is a result of being constructed from aluminium sided honeycomb material like that shown in fig. 13.3. The inset shows details of how this material is bent. Reportedly this chassis was very rigid and reduced lap times, but for whatever reason, development was not continued for very long. The front brake is of a rim type, but unlike the stainless steel example illustrated in chapter 12, this disk was made of carbon. Some frames don’t fit into the descriptions above for a variety of reasons. The material used fo. construction dictates how the frame needs to be designed. This would be the case for example with any attempt to make a chassis from composite material such as carbon fibre. To use such material purely as a substitute for steel or aluminium in a similar design would be ill advised, each type of material needs é structural design to suit its own special properties. The Heron Suzuki pictured below is a good example. This was made from honeycomb material, as described in chapter 13, this material is not really suitable for bending, and so when this is necessary the inner sheet is cut away so that the bend only occurs on the outer sheet. To restore the strength of the inner sheet a folded strip is glued over the cut. This form of construction leads to the sharp edged boxy appearance shown.
This chassis for a Q2 was designed by the author to handle the totally different structural requirements of a machine fitted with a form of double link front suspension. The enlargement shows more detail of the upper suspension link mounting.
This chassis for a Q2 was designed by the author to handle the totally different structural requirements of a machine fitted with a form of double link front suspension. The enlargement shows more detail of the upper suspension link mounting.
In the traditional old singles and 360-degree vertical twins (pistons in step) vibration was a serious problem, exacerbated in many cases by relatively flexible frames that could very well vibrate in sympathy with the engine. If a frame resonance coincided with a particular engine speed, then the vibration transmitted to both rider and structure would be magnified. A common method of reducing this vibration (which, in a slowly rotating big single, was aggravated by pulsating torque reaction) was to fit a cylinder- head steady (i.e. a stay or bracket bracing the head to the frame). This worked by using the bulk of the engine to stiffen the frame, so raising its resonant frequency and reducing the magnitude of the vibration.
In the traditional old singles and 360-degree vertical twins (pistons in step) vibration was a serious problem, exacerbated in many cases by relatively flexible frames that could very well vibrate in sympathy with the engine. If a frame resonance coincided with a particular engine speed, then the vibration transmitted to both rider and structure would be magnified. A common method of reducing this vibration (which, in a slowly rotating big single, was aggravated by pulsating torque reaction) was to fit a cylinder- head steady (i.e. a stay or bracket bracing the head to the frame). This worked by using the bulk of the engine to stiffen the frame, so raising its resonant frequency and reducing the magnitude of the vibration.
Unfortunately, in the Triumph case the required pivot point proved to be within the rear-wheel area, which was clearly impractical. However, since the engine’s angular movement about the pivot was extremely small, virtually the same movement was obtained by attaching the engine to the frame by short links top and bottom, so arranged that if they were extended rearward they would meet at the theoretical pivot point, sometimes called a virtual pivot. Tests ona rolling road indicated an enormous reduction in the vibration felt by the rider, as in fig. 11.4.
Unfortunately, in the Triumph case the required pivot point proved to be within the rear-wheel area, which was clearly impractical. However, since the engine’s angular movement about the pivot was extremely small, virtually the same movement was obtained by attaching the engine to the frame by short links top and bottom, so arranged that if they were extended rearward they would meet at the theoretical pivot point, sometimes called a virtual pivot. Tests ona rolling road indicated an enormous reduction in the vibration felt by the rider, as in fig. 11.4.
This point is called the centre of gyration and is the point about which the engine assembly would tend to rotate if a sudden vertical force was applied at the crankshaft axis. No vertical force is transmitted to the centre of gyration and so is the perfect point to mount from.
This point is called the centre of gyration and is the point about which the engine assembly would tend to rotate if a sudden vertical force was applied at the crankshaft axis. No vertical force is transmitted to the centre of gyration and so is the perfect point to mount from.
Fig 11.4 Vibrations induced in the handlebars of a modified Bonneville are on average 25% of those in the handlebars of a Triumph Tiger of the same capacity, and over much of the range less than those of a 4 cylinder machine of just over half of the capacity.
Fig 11.4 Vibrations induced in the handlebars of a modified Bonneville are on average 25% of those in the handlebars of a Triumph Tiger of the same capacity, and over much of the range less than those of a 4 cylinder machine of just over half of the capacity.
In the Norton Cosworth Challenge parallel twin the engine was the main structural chassis member, with the rear-fork pivot cast in the gearbox casing. A simple structure atop the motor to carry the steering head was all that was necessary to complete the main parts. See chapter 10 for an illustration of the frame.
In the Norton Cosworth Challenge parallel twin the engine was the main structural chassis member, with the rear-fork pivot cast in the gearbox casing. A simple structure atop the motor to carry the steering head was all that was necessary to complete the main parts. See chapter 10 for an illustration of the frame.
The new head mounting points and the normal front cast mounting lug on the crankcase provided a strong and widely spaced base on which to bolt the front subframe to support the head stock. At the rear, a long swing-arm was used which pivoted concentrically with the gearbox drive sprocket. On the right hand side a new clutch cover was cast incorporating a pivot bearing for the swing-arm. There was another special casting made for the right hand side to carry the other bearing. The result of this was a very light but extremely rigid chassis. Engines not specifically designed to be used as the main frame member can often be adapted for such duty. There is often good potential to provide a rigid structure and lighter all-up weight, but care must be taken with the details in order to ensure connections that don’t work loose in use, hence negating any structural gain. During the 70s and 80s the Kawasaki Z900/1000 range was a popular donor of engines for those interested in a chassis change. The following picture shows how the author provided a strong and slop free frame mounting on the cylinder head.
The new head mounting points and the normal front cast mounting lug on the crankcase provided a strong and widely spaced base on which to bolt the front subframe to support the head stock. At the rear, a long swing-arm was used which pivoted concentrically with the gearbox drive sprocket. On the right hand side a new clutch cover was cast incorporating a pivot bearing for the swing-arm. There was another special casting made for the right hand side to carry the other bearing. The result of this was a very light but extremely rigid chassis. Engines not specifically designed to be used as the main frame member can often be adapted for such duty. There is often good potential to provide a rigid structure and lighter all-up weight, but care must be taken with the details in order to ensure connections that don’t work loose in use, hence negating any structural gain. During the 70s and 80s the Kawasaki Z900/1000 range was a popular donor of engines for those interested in a chassis change. The following picture shows how the author provided a strong and slop free frame mounting on the cylinder head.
This Honda CBX engine permitted a different approach to that employed for the Kawasaki above. Cylinder head lugs already existed that were suitable for mounting a similar front subframe.
This Honda CBX engine permitted a different approach to that employed for the Kawasaki above. Cylinder head lugs already existed that were suitable for mounting a similar front subframe.
A simple triangulated structure was bolted to the above brackets and the standard front mounting lugs on the crankcase. Specially made side castings replace the original parts to give support for the rear fork pivot bearings, which are coaxial with the gearbox sprocket.
A simple triangulated structure was bolted to the above brackets and the standard front mounting lugs on the crankcase. Specially made side castings replace the original parts to give support for the rear fork pivot bearings, which are coaxial with the gearbox sprocket.
The area just to the rear of the headstock was boxed-in and used as a secondary fuel tank. This added some damping to further reduce the vibration. In practice, this machine produced low amounts of vibration except at the handle-bars which vibrated badly in the first instance. Laboratory testing revealed that this was principally due to the handlebars themselves adopting a vibration mode. Subsequent work on tuning the vibration characteristics of the handlebars reduced this problem considerably.
The area just to the rear of the headstock was boxed-in and used as a secondary fuel tank. This added some damping to further reduce the vibration. In practice, this machine produced low amounts of vibration except at the handle-bars which vibrated badly in the first instance. Laboratory testing revealed that this was principally due to the handlebars themselves adopting a vibration mode. Subsequent work on tuning the vibration characteristics of the handlebars reduced this problem considerably.
Fig. 12.1 Showing the kinetic energy to be absorbed, to stop from different initial speeds. This example is drawn for a fictitious racing machine with 200 kg. of laden mass of the bike and rider and a value of 0.3 m’ for CdA. The total energy must be dissipated between the brakes and aerodynamic drag, which becomes more important at the higher speeds. The lower curve is the energy to be dissipated in the brakes alone.
Fig. 12.1 Showing the kinetic energy to be absorbed, to stop from different initial speeds. This example is drawn for a fictitious racing machine with 200 kg. of laden mass of the bike and rider and a value of 0.3 m’ for CdA. The total energy must be dissipated between the brakes and aerodynamic drag, which becomes more important at the higher speeds. The lower curve is the energy to be dissipated in the brakes alone.
Fig. 12.2 Aerodynamic drag adds significantly to the total braking potential particularly at high speeds. The brakes are assumed to give a constant tyre force equivalent to 1g. braking. At about 125 km/h the aerodynamic drag force is 11% of the braking force rising to 88% at 350 km/h. The total deceleration at 350 km/h can be as high as 1.88 G. The importance of aerodynamic effects on the degree of deceleration is shown clearly in fig. 12.2 Even at the relatively low speed of 125 km/h, the drag force adds 11% to the effect of the brakes alone. As the aerodynamic force acts directly on the machine, it does not use up any of the potential grip from the tyres, and so is additive to the deceleration provided by the brakes. In the rain the available braking through the tyres is reduced and so air drag becomes relatively even more important. Rolling resistance also adds to the total force slowing the bike but this has been ignored in these calculations because it is a smaller effect. Friction in the wheel bearings, chain and engine/gearbox has to act through the tyres tc produce a retarding force, just like the brakes, and so it is implicitly assumed that this is part of the tota force from the brakes. Engine braking is the most significant of these and in many cases is the principal braking mechanism at the rear.
Fig. 12.2 Aerodynamic drag adds significantly to the total braking potential particularly at high speeds. The brakes are assumed to give a constant tyre force equivalent to 1g. braking. At about 125 km/h the aerodynamic drag force is 11% of the braking force rising to 88% at 350 km/h. The total deceleration at 350 km/h can be as high as 1.88 G. The importance of aerodynamic effects on the degree of deceleration is shown clearly in fig. 12.2 Even at the relatively low speed of 125 km/h, the drag force adds 11% to the effect of the brakes alone. As the aerodynamic force acts directly on the machine, it does not use up any of the potential grip from the tyres, and so is additive to the deceleration provided by the brakes. In the rain the available braking through the tyres is reduced and so air drag becomes relatively even more important. Rolling resistance also adds to the total force slowing the bike but this has been ignored in these calculations because it is a smaller effect. Friction in the wheel bearings, chain and engine/gearbox has to act through the tyres tc produce a retarding force, just like the brakes, and so it is implicitly assumed that this is part of the tota force from the brakes. Engine braking is the most significant of these and in many cases is the principal braking mechanism at the rear.
Average measured braking decelerations of various road machines, expressed in Gs. All data refers to 60 mph to zero stops (96.6 km/h). At these speeds aerodynamic effects have minimal effect on stopping distances. These results were obtained with a sophisticated computer controlled radar system and probably have low error.
Average measured braking decelerations of various road machines, expressed in Gs. All data refers to 60 mph to zero stops (96.6 km/h). At these speeds aerodynamic effects have minimal effect on stopping distances. These results were obtained with a sophisticated computer controlled radar system and probably have low error.
Braking for the third bend at Misano. The shaded area represents when the bike is actually in the curve. The high initial deceleration is partly due to aerodynamic effects. Note how the deceleration then decreases as_ the speed reduces, some of which is due to the rider easing the brakes' on approach to the corner. The velocity is measured at the front wheel and so is lower than the true velocity under hard braking due to tyre slip, and so_ the calculated deceleration will initially appear higher than the real deceleration.
Braking for the third bend at Misano. The shaded area represents when the bike is actually in the curve. The high initial deceleration is partly due to aerodynamic effects. Note how the deceleration then decreases as_ the speed reduces, some of which is due to the rider easing the brakes' on approach to the corner. The velocity is measured at the front wheel and so is lower than the true velocity under hard braking due to tyre slip, and so_ the calculated deceleration will initially appear higher than the real deceleration.
Fig. 12.4 Average power required to stop. The lower line shows the power that must be handled by the brake discs and engine/power train friction. Fig. 12.1 slows the total energy that must be dissipated but it is the braking power that determines how quickly it must be dispersed and is the main parameter that dictates the degree of cooling needed to maintain acceptable temperatures. Fig. 12.1 shows the total amount of energy to be dissipated by the brakes and engine friction etc.. This is the total amount of heat that must be got rid of. However, that gives no indication of how hot the brake system will get nor of how much cooling rate is necessary to maintain workable temperatures of these components. To calculate these things we must also consider how quickly we stop from a given speed. It is the time rate of dispersing this energy that is important and this is called “braking power’, fig. 12.4 shows this for the same case used earlier. Remember these figures are for a relatively light-weight racer, a heavier road bike would produce more power during emergency stops from the same speeds.
Fig. 12.4 Average power required to stop. The lower line shows the power that must be handled by the brake discs and engine/power train friction. Fig. 12.1 slows the total energy that must be dissipated but it is the braking power that determines how quickly it must be dispersed and is the main parameter that dictates the degree of cooling needed to maintain acceptable temperatures. Fig. 12.1 shows the total amount of energy to be dissipated by the brakes and engine friction etc.. This is the total amount of heat that must be got rid of. However, that gives no indication of how hot the brake system will get nor of how much cooling rate is necessary to maintain workable temperatures of these components. To calculate these things we must also consider how quickly we stop from a given speed. It is the time rate of dispersing this energy that is important and this is called “braking power’, fig. 12.4 shows this for the same case used earlier. Remember these figures are for a relatively light-weight racer, a heavier road bike would produce more power during emergency stops from the same speeds.
Fig. 12.5 Illustrating in general terms, the effect of the disc heat capacity on the temperature build up. The high capacity disc has double the mass of the other. The final working temperatures will be the same in each case. Heat capacity is the measure of how much heat we have to put into an object (brake disc) to raise its temperature, or in other words, the amount of heat stored in a disc at a certain temperature. For a given material the heat capacity depends on the mass of the disc. The heat capacity of a braking system doesn’t affect the ultimate steady state temperature reached, but it does affect how quickly the disc reaches that temperature, and how quickly it cools down after the braking has ended. Fig. 12.5 shows this effect for two discs, one with double the mass of the other.
Fig. 12.5 Illustrating in general terms, the effect of the disc heat capacity on the temperature build up. The high capacity disc has double the mass of the other. The final working temperatures will be the same in each case. Heat capacity is the measure of how much heat we have to put into an object (brake disc) to raise its temperature, or in other words, the amount of heat stored in a disc at a certain temperature. For a given material the heat capacity depends on the mass of the disc. The heat capacity of a braking system doesn’t affect the ultimate steady state temperature reached, but it does affect how quickly the disc reaches that temperature, and how quickly it cools down after the braking has ended. Fig. 12.5 shows this effect for two discs, one with double the mass of the other.
Fig. 12.6 Maximum possible braking for various ratios of CoG height to wheelbase length. At the higher ratios less load on the rear tyre means that the front must provide most of the braking effort. When there is no vertical load on the rear tyre the bike will loop forward if subject to increased braking. This case is for speeds low enough for air drag to be minimal. In chapter 9 on Squat and dive we saw that load transfer occurs under braking and how it is related to the ratio of CoG height to wheelbase. The higher the CoG the higher will be the load transfer to the extent that racing and sports bikes can easily lift their rear wheels when braking hard at lower speeds. Depending on the grip of the tyres this will typically occur if the CoG is higher than about 45 — 55% of the wheelbase. Under these conditions the rear wheel is incapable of helping with the braking effort and the front must take the full load. However, if the CoG was lower then some vertical load would still be present on the rear tyre, at maximum braking, which could then be used to relieve the front of some of this duty. Fig. 12.6 shows how the maximum possible braking forces vary between the front and rear tyres depending on the CoG to wheel base ratio, between zero (CoG at ground level) and 60% of the wheel base. The example bike has a laden 50/50 weight distribution and under static loading is assumed to have identical friction coefficients of 1.0 at each tyre.
Fig. 12.6 Maximum possible braking for various ratios of CoG height to wheelbase length. At the higher ratios less load on the rear tyre means that the front must provide most of the braking effort. When there is no vertical load on the rear tyre the bike will loop forward if subject to increased braking. This case is for speeds low enough for air drag to be minimal. In chapter 9 on Squat and dive we saw that load transfer occurs under braking and how it is related to the ratio of CoG height to wheelbase. The higher the CoG the higher will be the load transfer to the extent that racing and sports bikes can easily lift their rear wheels when braking hard at lower speeds. Depending on the grip of the tyres this will typically occur if the CoG is higher than about 45 — 55% of the wheelbase. Under these conditions the rear wheel is incapable of helping with the braking effort and the front must take the full load. However, if the CoG was lower then some vertical load would still be present on the rear tyre, at maximum braking, which could then be used to relieve the front of some of this duty. Fig. 12.6 shows how the maximum possible braking forces vary between the front and rear tyres depending on the CoG to wheel base ratio, between zero (CoG at ground level) and 60% of the wheel base. The example bike has a laden 50/50 weight distribution and under static loading is assumed to have identical friction coefficients of 1.0 at each tyre.
Fig. 12.7 At very high speeds drag forces can unload the front wheel and load the rear, completely changing the required braking balance compared to the slow speed case. Optimal possible braking occurs with a CoG height to wheelbase ratio of around 0.5 or 50%. In chapter 5 on Aerodynamics it was shown that air drag caused a load transfer off the front wheel onto the rear. At very high speeds this can be sufficient to almost unload the front wheel completely, and so the balance of vertical forces and hence available braking forces will be very different at low and high speeds. Fig. 12.7 shows this for the same example machine as above, except that it is assumed to be travelling at a speed with enough aerodynamic drag to just unload the front wheel completely. This is an extreme but not totally impossible case.
Fig. 12.7 At very high speeds drag forces can unload the front wheel and load the rear, completely changing the required braking balance compared to the slow speed case. Optimal possible braking occurs with a CoG height to wheelbase ratio of around 0.5 or 50%. In chapter 5 on Aerodynamics it was shown that air drag caused a load transfer off the front wheel onto the rear. At very high speeds this can be sufficient to almost unload the front wheel completely, and so the balance of vertical forces and hence available braking forces will be very different at low and high speeds. Fig. 12.7 shows this for the same example machine as above, except that it is assumed to be travelling at a speed with enough aerodynamic drag to just unload the front wheel completely. This is an extreme but not totally impossible case.
Generation of torque Fig. 12.8 The disc on the left has a smaller radial working depth than the one on the right and it can easily been seen that the effective or average radius of the area swept by the pads is greater. The disc on the left produces a higher torque than the other for a given line pressure and piston area. The heat capacity will be less also and so reach its operating temperature quicker. technically best layout for maximum braking would be a long and low machine to reduce the load transfer from both drag and braking, but to take advantage of this layout the riders would have to be capable of adjusting in an optimal manner the front to rear balance. Such a layout is similar to that used for F1 car racing, but here the driver has but one braking control to optimize, although front to rear brake balance is usually adjustable in-flight by means of a separate control.
Generation of torque Fig. 12.8 The disc on the left has a smaller radial working depth than the one on the right and it can easily been seen that the effective or average radius of the area swept by the pads is greater. The disc on the left produces a higher torque than the other for a given line pressure and piston area. The heat capacity will be less also and so reach its operating temperature quicker. technically best layout for maximum braking would be a long and low machine to reduce the load transfer from both drag and braking, but to take advantage of this layout the riders would have to be capable of adjusting in an optimal manner the front to rear balance. Such a layout is similar to that used for F1 car racing, but here the driver has but one braking control to optimize, although front to rear brake balance is usually adjustable in-flight by means of a separate control.
The peak of development of the drum brake is well illustrated here on the left, fitted to the 1968 works 250/4 Yamaha GP racer. A brake drum on each side of the wheel and each one with twin leading shoes, operated by cable with a balance roller fitted near the hand control to equalize the effort applied to each side. Air ducts at the front and rear were to try and get sufficient cooling. Just over a decade later disc brakes were firmly established in racing and on other performance orientated bikes. From the same marque on the right is a 1979 500 GP machine. Note the relatively large radial depth of the swept area, and how the disc is firmly riveted to the disc carriers, both features in contrast to current design ideas. Disc brakes are now so universal on motorcycles, with only a few exceptions even at the moped end of the market, that it is easy to forget that it is really only from the 1970s. that discs have been in common use.
The peak of development of the drum brake is well illustrated here on the left, fitted to the 1968 works 250/4 Yamaha GP racer. A brake drum on each side of the wheel and each one with twin leading shoes, operated by cable with a balance roller fitted near the hand control to equalize the effort applied to each side. Air ducts at the front and rear were to try and get sufficient cooling. Just over a decade later disc brakes were firmly established in racing and on other performance orientated bikes. From the same marque on the right is a 1979 500 GP machine. Note the relatively large radial depth of the swept area, and how the disc is firmly riveted to the disc carriers, both features in contrast to current design ideas. Disc brakes are now so universal on motorcycles, with only a few exceptions even at the moped end of the market, that it is easy to forget that it is really only from the 1970s. that discs have been in common use.
On the left is a modern high performance brake disc from AP racing. The working disc is made of iron and has a small radial depth. The centre carrier is of aluminium alloy for lightness. The two are held together by means of 8 top-hat fixings, which allow the disc some lateral clearance but transmit the braking torque in a secure manner. The floating nature of this disc fixing is a precaution against warpage due to uneven temperature distribution. The surface slots are to permit dust from pad wear to be cleaned off the working surface, these slots are “handed” for each side of the bike, it is important that they rotate in the correct direction, the disc shown is for RHS mounting as shown by the arrow.
On the left is a modern high performance brake disc from AP racing. The working disc is made of iron and has a small radial depth. The centre carrier is of aluminium alloy for lightness. The two are held together by means of 8 top-hat fixings, which allow the disc some lateral clearance but transmit the braking torque in a secure manner. The floating nature of this disc fixing is a precaution against warpage due to uneven temperature distribution. The surface slots are to permit dust from pad wear to be cleaned off the working surface, these slots are “handed” for each side of the bike, it is important that they rotate in the correct direction, the disc shown is for RHS mounting as shown by the arrow.
Although disc brakes had occasionally been used before, it was during the 1970s. that the real changeover began in earnest. Initially for racing by European and American riders. In the beginning design just followed car practice with the use of cast iron for the disc material. When the Japanese factories started to fit discs to road bikes they used stainless steel, although less suitable from a braking point of view it had the advantage that it didn’t rust in damp climates nor if the bike was unused for a while. Poor wet weather braking was another feature of the early stainless steel brakes and it was usual to see slots and holes used to try and break up the surface water. Italian manufacturers continued to use cast iron even on road bikes. As mentioned earlier, it is desirable to keep disc weight to a minimum and to this end there were many experiments with alternative materials such as aluminium, but due to the relatively soft surface these needed some form of hard coating. Hard chrome plating, plasma sprayed ceramic coatings and hard anodizing have all been tried with varying degrees of success but none of these have stood the test of time. More recently, for top level competition, discs made exclusively of carbon have become firmly established in those racing classes that don’t ban their use. Such bans are based on the high cost of such technology rather than on technical grounds. Carbon has many features on its side, importantly it can withstand high temperatures and is very light. It must be used with pads of the same material and early designs only gave good braking at high temperatures causing some problems with stopping during the initial cold application of the brakes. Different manufacturing methods can now produce carbon components that are much less sensitive to this orohlem.
Although disc brakes had occasionally been used before, it was during the 1970s. that the real changeover began in earnest. Initially for racing by European and American riders. In the beginning design just followed car practice with the use of cast iron for the disc material. When the Japanese factories started to fit discs to road bikes they used stainless steel, although less suitable from a braking point of view it had the advantage that it didn’t rust in damp climates nor if the bike was unused for a while. Poor wet weather braking was another feature of the early stainless steel brakes and it was usual to see slots and holes used to try and break up the surface water. Italian manufacturers continued to use cast iron even on road bikes. As mentioned earlier, it is desirable to keep disc weight to a minimum and to this end there were many experiments with alternative materials such as aluminium, but due to the relatively soft surface these needed some form of hard coating. Hard chrome plating, plasma sprayed ceramic coatings and hard anodizing have all been tried with varying degrees of success but none of these have stood the test of time. More recently, for top level competition, discs made exclusively of carbon have become firmly established in those racing classes that don’t ban their use. Such bans are based on the high cost of such technology rather than on technical grounds. Carbon has many features on its side, importantly it can withstand high temperatures and is very light. It must be used with pads of the same material and early designs only gave good braking at high temperatures causing some problems with stopping during the initial cold application of the brakes. Different manufacturing methods can now produce carbon components that are much less sensitive to this orohlem.
Aermacchi Harley-Davidson conducted an interesting experiment on their racing two-stroke twins in the mid-1970s. They mounted the rear brake disc on the gearbox sprocket so that the braking force was taken through the chain. Getting sufficient braking at the rear of a 250 was unlikely to be a grea problem and so we must look elsewhere to find any potential benefits, which might include:
Aermacchi Harley-Davidson conducted an interesting experiment on their racing two-stroke twins in the mid-1970s. They mounted the rear brake disc on the gearbox sprocket so that the braking force was taken through the chain. Getting sufficient braking at the rear of a 250 was unlikely to be a grea problem and so we must look elsewhere to find any potential benefits, which might include:
Calipers There are two main classes of caliper — single acting and double acting. The double acting have < piston or pistons on each side of the disc. The body of the caliper is fixed to the fork leg and only the pistons move to apply force to the pads. The single acting only has pistons on one side of the disc anc to balance the pad force the caliper body has to be free to move laterally. This movement has beer provided either by allowing the caliper to slide on two pins or by having an approximately vertical pivo giving the caliper a swinging motion, (early BMW and Honda). The pivoted design gives rise to unever pad wear, the pad on the inside of the disc wearing more at the front and the outside pad wearing more toward the rear. Both movement methods have a serious deficiency in every-day road use, particularly in countries prone to wet weather, and that is corrosion of the moving surfaces. This causes the calipel body to seize on the sliding or pivoting pins, thus losing the ability to properly balance pad pressure with a consequent reduction in braking ability. The principal reason for the use of single acting calipers is simply one of cost, although sometimes there are space saving benefits. The inside half of the single acting caliper can be thinner, allowing the discs to be mounted closer to the wheel, thus reducing fork width, but this is more of theoretical benefit than a practical one.
Calipers There are two main classes of caliper — single acting and double acting. The double acting have < piston or pistons on each side of the disc. The body of the caliper is fixed to the fork leg and only the pistons move to apply force to the pads. The single acting only has pistons on one side of the disc anc to balance the pad force the caliper body has to be free to move laterally. This movement has beer provided either by allowing the caliper to slide on two pins or by having an approximately vertical pivo giving the caliper a swinging motion, (early BMW and Honda). The pivoted design gives rise to unever pad wear, the pad on the inside of the disc wearing more at the front and the outside pad wearing more toward the rear. Both movement methods have a serious deficiency in every-day road use, particularly in countries prone to wet weather, and that is corrosion of the moving surfaces. This causes the calipel body to seize on the sliding or pivoting pins, thus losing the ability to properly balance pad pressure with a consequent reduction in braking ability. The principal reason for the use of single acting calipers is simply one of cost, although sometimes there are space saving benefits. The inside half of the single acting caliper can be thinner, allowing the discs to be mounted closer to the wheel, thus reducing fork width, but this is more of theoretical benefit than a practical one.
The old and the new. Nearly thirty years separate these two calipers. On the right is an old style twin piston type made from two separate aluminium dye castings, weight without pads was 900 grams. The advent of CNC machine tools has made it economically feasible to machine critical components from solid billet as typified by the modern racing caliper to the left. Weighing just 690 grams this has been made in one piece to give maximum rigidity, and six pistons are used to spread the load over the long thin pad. Different piston diameters are used to reduce the tendency of the pads to wear to a taper, four pistons are 28.6 mm and two are 25.4 mm in diameter. Total piston area of 35.8 cm* compared to 26.8 cm’ of the older model. The different pad shapes are, shown underneath, note the much smaller radial depth of the newer one which also has a smaller surface area of 19.8 cm’ in place of 22.4 cm’.
The old and the new. Nearly thirty years separate these two calipers. On the right is an old style twin piston type made from two separate aluminium dye castings, weight without pads was 900 grams. The advent of CNC machine tools has made it economically feasible to machine critical components from solid billet as typified by the modern racing caliper to the left. Weighing just 690 grams this has been made in one piece to give maximum rigidity, and six pistons are used to spread the load over the long thin pad. Different piston diameters are used to reduce the tendency of the pads to wear to a taper, four pistons are 28.6 mm and two are 25.4 mm in diameter. Total piston area of 35.8 cm* compared to 26.8 cm’ of the older model. The different pad shapes are, shown underneath, note the much smaller radial depth of the newer one which also has a smaller surface area of 19.8 cm’ in place of 22.4 cm’.
Fig. 12.9 Diagrammatic layout of the original Brembo hydraulic coupled brakes as used by Moto Guzzi. The object is to eliminate reliance on a high degree of expertise in apportioning front and rear effort. (Brembo)
Fig. 12.9 Diagrammatic layout of the original Brembo hydraulic coupled brakes as used by Moto Guzzi. The object is to eliminate reliance on a high degree of expertise in apportioning front and rear effort. (Brembo)
Although development of motorcycle ABS systems had been going on from the late 1960s it wasn’t until 1987 that they became available on production models. Until then, design was heading in two separate directions; electronically controlled and mechanically actuated. The operation of the two systems is basically similar, the elementary sequence of events for both is as follows:
Although development of motorcycle ABS systems had been going on from the late 1960s it wasn’t until 1987 that they became available on production models. Until then, design was heading in two separate directions; electronically controlled and mechanically actuated. The operation of the two systems is basically similar, the elementary sequence of events for both is as follows:
Lucas-Girling SCS (Stop Control System) mechanically actuated ABS. The SCS unit can be seen just ahead of the forks, a notched rubber belt, behind the brake disc, drives the internal flywheel at three times wheel speed. The graphs to the right follow a single lock-unlock cycle.
Lucas-Girling SCS (Stop Control System) mechanically actuated ABS. The SCS unit can be seen just ahead of the forks, a notched rubber belt, behind the brake disc, drives the internal flywheel at three times wheel speed. The graphs to the right follow a single lock-unlock cycle.
Yamaha ABS system. First used on the FJ1000A in 1991 and later miniaturized and silenced in 1998 for the 250 Majesty scooter. Although the SCS system was still in development at least until 1987, as with many other products it could not withstand the onslaught of modern electronics and as far as is known no mechanical systems actually made it into motorcycle production.
Yamaha ABS system. First used on the FJ1000A in 1991 and later miniaturized and silenced in 1998 for the 250 Majesty scooter. Although the SCS system was still in development at least until 1987, as with many other products it could not withstand the onslaught of modern electronics and as far as is known no mechanical systems actually made it into motorcycle production.
An interesting feature of BMW's linking system is that it features an "adaptive brake force distribution" system. This senses wheel speed and the amount of brake bias is also modulated according to the load condition of the bike. By sensing wheel-lock on either wheel it then transfers more braking effort to the opposite wheel. So each wheel will receive the maximum amount of braking torque that it can use. This system can handle up to 10 lock/unlock cycles a second. The first generation BMW systems were very heavy and added significantly to the selling price. Th current, third generation versions are better in both respects, and more sophisticated in concept an performance. The brakes are linked and both front and rear brakes have their ABS function controlle: individually.
An interesting feature of BMW's linking system is that it features an "adaptive brake force distribution" system. This senses wheel speed and the amount of brake bias is also modulated according to the load condition of the bike. By sensing wheel-lock on either wheel it then transfers more braking effort to the opposite wheel. So each wheel will receive the maximum amount of braking torque that it can use. This system can handle up to 10 lock/unlock cycles a second. The first generation BMW systems were very heavy and added significantly to the selling price. Th current, third generation versions are better in both respects, and more sophisticated in concept an performance. The brakes are linked and both front and rear brakes have their ABS function controlle: individually.
Fig 13.1 Top left shows the meaning of the term stress. The significance of strain is shown below that. A typical stress- strain curve for steel is on the right. The initial slope of the line is the Young’s Modulus.
Fig 13.1 Top left shows the meaning of the term stress. The significance of strain is shown below that. A typical stress- strain curve for steel is on the right. The initial slope of the line is the Young’s Modulus.
When we continue to apply sufficient load beyond the yield point we reach the point of ultimate failure and the material actually breaks. The amount of strain that occurs between the yield point and the failure point is a measure of the materials ductility.
When we continue to apply sufficient load beyond the yield point we reach the point of ultimate failure and the material actually breaks. The amount of strain that occurs between the yield point and the failure point is a measure of the materials ductility.
When we apply an outward stress on all three axis we prevent this lateral strain and the material has no choice but to fail in a brittle manner. longitudinal strain and this lateral strain is known as Poisson's ratio and for most metals is about 0.33. | is easy to calculate that this ratio must be between 0 and 0.5, and in fact rubber is around 0.5.
When we apply an outward stress on all three axis we prevent this lateral strain and the material has no choice but to fail in a brittle manner. longitudinal strain and this lateral strain is known as Poisson's ratio and for most metals is about 0.33. | is easy to calculate that this ratio must be between 0 and 0.5, and in fact rubber is around 0.5.
Examples of cast chassis components. On the left is a cast frame from Eric Offenstadt, the backbone design is hollow in the main section. To the right is the swing-arm for the same frame. To avoid the increased pattern making and casting costs inherent with hollow sections the swing-arm uses a ribbed design. This is easier to make but results in a less structurally efficient component.
Examples of cast chassis components. On the left is a cast frame from Eric Offenstadt, the backbone design is hollow in the main section. To the right is the swing-arm for the same frame. To avoid the increased pattern making and casting costs inherent with hollow sections the swing-arm uses a ribbed design. This is easier to make but results in a less structurally efficient component.
Then tubular aluminium frames started to appear on works racers, with Yamaha taking the initiative This trend started cautiously, when just the pivoted rear fork was made in light alloy, before spreading tc the complete chassis. In the development of aluminium frames, however, it is interesting to note tha tube sizes increased rapidly to compensate for the low Young’s Modulus, as explained earlier. A great help in this context would be the spread of proper triangulation. In GP racing now, the use of aluminiurr alloy fabricated twin-spar frames is almost universal, and is also widely featured on expensive sports models for the street. It must be remembered that the fatigue characteristics of aluminium are such tha failure is inevitable eventually in components subjected to alternating stress, hence limited life must be accepted. In the case of works racers, their natural rapid obsolescence makes this less of a serious problem.
Then tubular aluminium frames started to appear on works racers, with Yamaha taking the initiative This trend started cautiously, when just the pivoted rear fork was made in light alloy, before spreading tc the complete chassis. In the development of aluminium frames, however, it is interesting to note tha tube sizes increased rapidly to compensate for the low Young’s Modulus, as explained earlier. A great help in this context would be the spread of proper triangulation. In GP racing now, the use of aluminiurr alloy fabricated twin-spar frames is almost universal, and is also widely featured on expensive sports models for the street. It must be remembered that the fatigue characteristics of aluminium are such tha failure is inevitable eventually in components subjected to alternating stress, hence limited life must be accepted. In the case of works racers, their natural rapid obsolescence makes this less of a serious problem.
This trend setting twin-spar GP Kobas from the early 1980s. housed a 250 Rotax engine. Note that the swingarm anc main frame are constructed of flat aluminium alloy welded at the corners. Today’s frames may be neater in appearance due to different manufacturing methods but this early Kobas had structurally stiffer sections, and a more direc connection between head-stock and swing-arm pivot than many that have followed. For touring machines, where long life dictates lower stress levels, it may well be that aluminium’s weight advantage over steel is lost.
This trend setting twin-spar GP Kobas from the early 1980s. housed a 250 Rotax engine. Note that the swingarm anc main frame are constructed of flat aluminium alloy welded at the corners. Today’s frames may be neater in appearance due to different manufacturing methods but this early Kobas had structurally stiffer sections, and a more direc connection between head-stock and swing-arm pivot than many that have followed. For touring machines, where long life dictates lower stress levels, it may well be that aluminium’s weight advantage over steel is lost.
The John Britten designed V twins had very successful chassis largely built of carbon reinforced plastic, but to date there have probably been more failures than successes with this material.
The John Britten designed V twins had very successful chassis largely built of carbon reinforced plastic, but to date there have probably been more failures than successes with this material.
Off-road bikes have generally stayed with wire spoked wheels as shown on this Cagiva. Despite attempts to use cast wheels, experience has shown that the wire wheel is on average more robust for this kind of duty.
Off-road bikes have generally stayed with wire spoked wheels as shown on this Cagiva. Despite attempts to use cast wheels, experience has shown that the wire wheel is on average more robust for this kind of duty.
Peter Williams was a pioneer in the use of cast magnesium wheels for motorcycles, on the Tom Arter G50 Matchless. He later went into production with the wheels shown above. Nowadays 3 spokes are considered sufficient, but early wheels seldom had less than 5, some even having 9. An odd number of spokes is best as this reduces the chances of the casting cracking when cooling from the as cast state. Some manufacturers used an even number of spokes but avoided this problem by grouping them in pairs (3 pairs for a 6 spoke wheel) such that no two spokes were diametrically opposed. Where expense was of little concern, complete wheels have been machined from an Aluminium billet and this has become a more practical proposition for one-offs and show machines, with the spread of CNC machine tools.
Peter Williams was a pioneer in the use of cast magnesium wheels for motorcycles, on the Tom Arter G50 Matchless. He later went into production with the wheels shown above. Nowadays 3 spokes are considered sufficient, but early wheels seldom had less than 5, some even having 9. An odd number of spokes is best as this reduces the chances of the casting cracking when cooling from the as cast state. Some manufacturers used an even number of spokes but avoided this problem by grouping them in pairs (3 pairs for a 6 spoke wheel) such that no two spokes were diametrically opposed. Where expense was of little concern, complete wheels have been machined from an Aluminium billet and this has become a more practical proposition for one-offs and show machines, with the spread of CNC machine tools.
Considering the speed of development in plastics technology, it may soon be a practical proposition tc make wheels in some form of this material, thus considerably reducing unsprung mass. Such wheels are already well established on BMX cycle-type machines. Carbon fibre reinforced plastic was for a lonc time touted as the material of the future for wheels but some high profile and spectacular structura failures delayed further experimentation for quite a while. However, these earlier problems have now been overcome and carbon fibre wheels and rims are now a reality.
Considering the speed of development in plastics technology, it may soon be a practical proposition tc make wheels in some form of this material, thus considerably reducing unsprung mass. Such wheels are already well established on BMX cycle-type machines. Carbon fibre reinforced plastic was for a lonc time touted as the material of the future for wheels but some high profile and spectacular structura failures delayed further experimentation for quite a while. However, these earlier problems have now been overcome and carbon fibre wheels and rims are now a reality.
A novel way by BMW to allow the use of tubeless tyres with a spoked wheel. The spoke holes in the rim are outside of the air space. Honda Comstar wheel. The extruded and rolled rim has an inner flange to which the stamped spokes are riveted. The bearings and disk mountings are in a relatively normal cast hub which bolts to the inner end of the spokes. These Dymag wheels are of composite construction. The rim is made in carbon fibre but the central spider and hub is cast in magnesium. This has the advantages that various centre sections can be installed in a given rim, and the weight reduction is in the rim which is the most important for a reduction of the rotating moment of inertia of the wheel, thus improving acceleration and corner lean-in. Such wheels have been approved for road use.
A novel way by BMW to allow the use of tubeless tyres with a spoked wheel. The spoke holes in the rim are outside of the air space. Honda Comstar wheel. The extruded and rolled rim has an inner flange to which the stamped spokes are riveted. The bearings and disk mountings are in a relatively normal cast hub which bolts to the inner end of the spokes. These Dymag wheels are of composite construction. The rim is made in carbon fibre but the central spider and hub is cast in magnesium. This has the advantages that various centre sections can be installed in a given rim, and the weight reduction is in the rim which is the most important for a reduction of the rotating moment of inertia of the wheel, thus improving acceleration and corner lean-in. Such wheels have been approved for road use.
Whilst this is a simplified view of the phenomenon it is quite accurate for tyre loading in the linear range of characteristics and suitable for our comparison. Crossing a slope (motorcycle)
Whilst this is a simplified view of the phenomenon it is quite accurate for tyre loading in the linear range of characteristics and suitable for our comparison. Crossing a slope (motorcycle)
Fig. 14.4 shows the case of a motorcycle crossing a slope and just as in the car case the tyres must produce a lateral force, but this is where the two vehicles can differ greatly. The motorcycle must keep vertical attitude to retain balance, unlike the car which leans over to match the slope. In other words, the motorcycle adopts a camber angle relative to the road surface. Refer back to chapter 2 on tyre grip anc in particular figs. 2.20 & 2.21. If both tyres have the characteristics as in fig. 2.20 with normalizec camber stiffnesses equal to 1.0, then camber forces will exactly balance FI, the force down the slope.
Fig. 14.4 shows the case of a motorcycle crossing a slope and just as in the car case the tyres must produce a lateral force, but this is where the two vehicles can differ greatly. The motorcycle must keep vertical attitude to retain balance, unlike the car which leans over to match the slope. In other words, the motorcycle adopts a camber angle relative to the road surface. Refer back to chapter 2 on tyre grip anc in particular figs. 2.20 & 2.21. If both tyres have the characteristics as in fig. 2.20 with normalizec camber stiffnesses equal to 1.0, then camber forces will exactly balance FI, the force down the slope.
It would be too tedious to consider each of the thirty three possible groups of combinations of relative camber and steer stiffnesses, suffice it to say that; if the handle-bars are held in the straight ahead position, and the camber stiffnesses are less than 1.0 then the path of the bike will depend on the under- /over-steer characteristics in the same manner as a car. If the camber stiffnesses are both greater than 1.0 then the effect of under-/over-steer will be reversed, i.e. the neutral and under-steering cases will have paths leading uphill and the over-steer case will be down the slope. Which is completely at odds with our normal conception of these terms. without any slip angle being needed. Therefore there will be no slip velocity down the slope an hence the bike will just run straight on in its original direction. This lack of reliance on slip angles to produce the required tyre forces simply means that the car definition of under-/over-steering has no particular significance in this case. However, the under-/over-steer possibilities are more numerous in the case of motorcycles, depending on the relationship between the values of camber and steering stiffnesses at the two ends. In fact there are thirty three different combinations that we could consider. We have seen in this example of riding across a slope that a camber stiffness equal to 1.0 has a particular significance, therefore we need to consider three sets of values of camber stiffness, at each end. Viz. Less than 1.0, equal to 1.0 and greater than 1.0. Front and back tyres can have different values and within each possible combination we can have three different relationships between the front and rear steering stiffnesses: Front and rear equal, front greater than the rear and front less than the rear. Without the complicating effects of camber force these three combinations would define under-/over-steer, but the situation is not that simple.
It would be too tedious to consider each of the thirty three possible groups of combinations of relative camber and steer stiffnesses, suffice it to say that; if the handle-bars are held in the straight ahead position, and the camber stiffnesses are less than 1.0 then the path of the bike will depend on the under- /over-steer characteristics in the same manner as a car. If the camber stiffnesses are both greater than 1.0 then the effect of under-/over-steer will be reversed, i.e. the neutral and under-steering cases will have paths leading uphill and the over-steer case will be down the slope. Which is completely at odds with our normal conception of these terms. without any slip angle being needed. Therefore there will be no slip velocity down the slope an hence the bike will just run straight on in its original direction. This lack of reliance on slip angles to produce the required tyre forces simply means that the car definition of under-/over-steering has no particular significance in this case. However, the under-/over-steer possibilities are more numerous in the case of motorcycles, depending on the relationship between the values of camber and steering stiffnesses at the two ends. In fact there are thirty three different combinations that we could consider. We have seen in this example of riding across a slope that a camber stiffness equal to 1.0 has a particular significance, therefore we need to consider three sets of values of camber stiffness, at each end. Viz. Less than 1.0, equal to 1.0 and greater than 1.0. Front and back tyres can have different values and within each possible combination we can have three different relationships between the front and rear steering stiffnesses: Front and rear equal, front greater than the rear and front less than the rear. Without the complicating effects of camber force these three combinations would define under-/over-steer, but the situation is not that simple.
Fig. 14.5 The end views show the difference between the upright attitude of the car and motorcycle tyres. The plan view on the right shows how the motorcycle tyre forces at the contact patch can cause a self steering effect about the steering axis.
Fig. 14.5 The end views show the difference between the upright attitude of the car and motorcycle tyres. The plan view on the right shows how the motorcycle tyre forces at the contact patch can cause a self steering effect about the steering axis.
A motorcycle is a very integrated system, i.e. changes in just one parameter usually cause effects throughout the whole machine, and so chain pull and suspension characteristics will exert some influence over phenomenon such as high-siding. There are two broad groups of high-side; On the other hand, the limit over-steer situation is very similar for both vehicles, but the motorcycle has the additional problem that when the rear slides out the reduction in total lateral tyre force will probably result in a fall also. In the case of the motorcycle both limit under- and over-steer are unstable situations. In general the over-steer situation is preferable as a rear slide is sometimes possible to control and recover from. Limit over-steering is often brought on by the application of engine power and so according to the friction circle concept introduced in chapter 2, there will be a reduction in available lateral tyre force, simply reducing the throttle opening with some steering correction is often sufficient to restore stability and control. Set against this possibility of recovering from a rear slide is the more dangerous possibility of a “high-side”.
A motorcycle is a very integrated system, i.e. changes in just one parameter usually cause effects throughout the whole machine, and so chain pull and suspension characteristics will exert some influence over phenomenon such as high-siding. There are two broad groups of high-side; On the other hand, the limit over-steer situation is very similar for both vehicles, but the motorcycle has the additional problem that when the rear slides out the reduction in total lateral tyre force will probably result in a fall also. In the case of the motorcycle both limit under- and over-steer are unstable situations. In general the over-steer situation is preferable as a rear slide is sometimes possible to control and recover from. Limit over-steering is often brought on by the application of engine power and so according to the friction circle concept introduced in chapter 2, there will be a reduction in available lateral tyre force, simply reducing the throttle opening with some steering correction is often sufficient to restore stability and control. Set against this possibility of recovering from a rear slide is the more dangerous possibility of a “high-side”.
To complete this analysis let us consider four separate braking conditions. Suppose a bike, while braking, is deflected from its true direction of travel, as in fig. 14.7. The inertia force (F,) gives rise to an unsettling couple (F,x;) about the front tyre contact patch. At the same time, the rear braking force (Fr) creates a correcting torque (FrX2). For natural stability, the correcting torque must exceed the unbalancing one. This can occur only under low or moderate deceleration, when there is little load transfer and so the rear wheel can provide a relatively high proportion of the total braking. Heavy braking, on the other hand, involves considerable forward load transfer, so that the front wheel is required to do the lion’s share, if not all, of the work, and the unsettling couple then exceeds the correcting torque — an unstable condition. If we are to reduce this tendency, we need to minimize the load transfer by means of a long wheelbase and low mass centre. A forward centre of gravity also helps by reducing the destabilizing moment by a greater amount than it reduces the rear wheel stabilizing effect. However, a forward CoG position limits the degree of braking before the bike spins forward.
To complete this analysis let us consider four separate braking conditions. Suppose a bike, while braking, is deflected from its true direction of travel, as in fig. 14.7. The inertia force (F,) gives rise to an unsettling couple (F,x;) about the front tyre contact patch. At the same time, the rear braking force (Fr) creates a correcting torque (FrX2). For natural stability, the correcting torque must exceed the unbalancing one. This can occur only under low or moderate deceleration, when there is little load transfer and so the rear wheel can provide a relatively high proportion of the total braking. Heavy braking, on the other hand, involves considerable forward load transfer, so that the front wheel is required to do the lion’s share, if not all, of the work, and the unsettling couple then exceeds the correcting torque — an unstable condition. If we are to reduce this tendency, we need to minimize the load transfer by means of a long wheelbase and low mass centre. A forward centre of gravity also helps by reducing the destabilizing moment by a greater amount than it reduces the rear wheel stabilizing effect. However, a forward CoG position limits the degree of braking before the bike spins forward.
which, through the escapement mechanism, gives the pendulum a kick on each swing to ensure that it does not come to rest. A child’s garden swing is just a special case of pendulum. We have only to give it a gentle push each time it reaches a peak for the amplitude of the swing to increase very quickly; in extreme cases the swing may even be made to go over the top. If the gentle push is timed wrongly, however, it is surprising how quickly the swing can be brought to rest. (No apologies are made for this lengthy discussion of swings and pendulums because a basic knowledge of simple harmonic motion is essential to an understanding of the causes of several motorcycle problems and of the function of suspension systems. ) Now let us revisit the concept of resonance, first mentioned in chapter 6 on suspension. Experiments with our pendulum or swing will soon show that (except for extreme angles of swing) the number of complete oscillations in a given time is almost constant, regardless of the amplitude of the movement. This is called the resonant frequency (or natural frequency) and is mainly determined by the length of the pendulum. As we have seen, if our push (forcing function) to the child’s swing is correctly timed (i.e. ir phase) and at the natural frequency it is easy to build up high amplitudes of oscillation. However, if the forcing function is either out of phase or of a different frequency, then the oscillation will be much less inclined to build up and may even be heavily suppressed.
which, through the escapement mechanism, gives the pendulum a kick on each swing to ensure that it does not come to rest. A child’s garden swing is just a special case of pendulum. We have only to give it a gentle push each time it reaches a peak for the amplitude of the swing to increase very quickly; in extreme cases the swing may even be made to go over the top. If the gentle push is timed wrongly, however, it is surprising how quickly the swing can be brought to rest. (No apologies are made for this lengthy discussion of swings and pendulums because a basic knowledge of simple harmonic motion is essential to an understanding of the causes of several motorcycle problems and of the function of suspension systems. ) Now let us revisit the concept of resonance, first mentioned in chapter 6 on suspension. Experiments with our pendulum or swing will soon show that (except for extreme angles of swing) the number of complete oscillations in a given time is almost constant, regardless of the amplitude of the movement. This is called the resonant frequency (or natural frequency) and is mainly determined by the length of the pendulum. As we have seen, if our push (forcing function) to the child’s swing is correctly timed (i.e. ir phase) and at the natural frequency it is easy to build up high amplitudes of oscillation. However, if the forcing function is either out of phase or of a different frequency, then the oscillation will be much less inclined to build up and may even be heavily suppressed.
On the left are two different length displacement potentiometers typical of those used for suspension movement measuring. A typical installation on front forks is shown on the right. Note also the plastic ring around the slider tube, this is a low tech but very effective way of quickly detecting the maximum suspension compression that has taken place. The data derived from suspension monitoring is invaluable for tuning a bike to specific circuits etc.. The overall range of suspension movement can be a very good guide to required spring rate, for example if the range of actual displacement is much less than the available movement it is an indicator that the spring rate is too high. If the range of movement tends to be all near the limit of either full compression or full extension then it is likely that the spring preload is in need of adjustment. There are many features in the data display that experienced people can use to help tune damping settings too. If the dive under braking occurs too quickly then increased bump damping at the front might be a solution but this may have to be compromised against damping requirements in other parts of the track.
On the left are two different length displacement potentiometers typical of those used for suspension movement measuring. A typical installation on front forks is shown on the right. Note also the plastic ring around the slider tube, this is a low tech but very effective way of quickly detecting the maximum suspension compression that has taken place. The data derived from suspension monitoring is invaluable for tuning a bike to specific circuits etc.. The overall range of suspension movement can be a very good guide to required spring rate, for example if the range of actual displacement is much less than the available movement it is an indicator that the spring rate is too high. If the range of movement tends to be all near the limit of either full compression or full extension then it is likely that the spring preload is in need of adjustment. There are many features in the data display that experienced people can use to help tune damping settings too. If the dive under braking occurs too quickly then increased bump damping at the front might be a solution but this may have to be compromised against damping requirements in other parts of the track.
Fig. 15.1 This 15 seconds of track data with only 3 parameters shown can yield an enormous amount of information when you know where to look. (From data supplied by 2D)
Fig. 15.1 This 15 seconds of track data with only 3 parameters shown can yield an enormous amount of information when you know where to look. (From data supplied by 2D)
Fig. 15.2 Measured data from two different race bikes. Note how the average fork compression increases with speed in the upper chart, from 112 to 225 km/h the suspension compresses by an extra 16 mm. The lower machine appears to have an unchanging front ride height, in the same speed range. Load transfer, suspension settings and aerodynamic effects all play a part in this aspect of chassis performance. (From data supplied by 2D) km/h to 225 km/h and both show the machines accelerating away from a slow corner. It is interesting tc compare the superimposed line of the averaged fork compression during the acceleration period. The two bikes exhibit quite significant differences, the upper chart shows a gradually increasing fron suspension compression as speed rises, whereas the lower data shows no similar tendencies.
Fig. 15.2 Measured data from two different race bikes. Note how the average fork compression increases with speed in the upper chart, from 112 to 225 km/h the suspension compresses by an extra 16 mm. The lower machine appears to have an unchanging front ride height, in the same speed range. Load transfer, suspension settings and aerodynamic effects all play a part in this aspect of chassis performance. (From data supplied by 2D) km/h to 225 km/h and both show the machines accelerating away from a slow corner. It is interesting tc compare the superimposed line of the averaged fork compression during the acceleration period. The two bikes exhibit quite significant differences, the upper chart shows a gradually increasing fron suspension compression as speed rises, whereas the lower data shows no similar tendencies.
Fig. 15.3 The shaded areas define aspects of this data sample which are discussed in the following text. (Data from 2D) Load is not the only factor that determines fork movement, preload and spring rate are important as well and the lower chart in fact seems more indicative of excessive preload than it does of the load transfer and aerodynamic forces balancing each other so exactly.
Fig. 15.3 The shaded areas define aspects of this data sample which are discussed in the following text. (Data from 2D) Load is not the only factor that determines fork movement, preload and spring rate are important as well and the lower chart in fact seems more indicative of excessive preload than it does of the load transfer and aerodynamic forces balancing each other so exactly.
Fig. 15.4 is an indication of the frequency of occurrence of different magnitude loads as found in practice. Note that the middle range of load occurs most frequently, while both low and high values occur much less often. This is to be expected as low loading is produced at low speeds on smooth roads, while high loading is a result of severe conditions — both extremes being encountered for only a relatively smal proportion of the machine’s life. Mathematical techniques are available for the treatment of such test data and, together with stress analysis, they enable us to investigate the frame’s fatigue behaviour. Fatigue analysis is based on summing the microscopic damage done during each load cycle, more damage is done at higher values of the load cycle. Such sophisticated testing methods are usually beyond the scope of the small chassis maker and the enthusiast building a one-off. Indeed, it is unlikely that even large manufacturers make full use of available techniques. However, good results can usually be achieved by rule-of-thumb and simpler calculations based on previous practice and experience. If the designer himself has the experience, so much the better. If not — and all designers must start somewhere — then a review of existing hardware is a good basis to build on. will result in a fragile structure, intolerant of hard use. Since metal fatigue is the most common cause of failure, in non-extreme circumstances, a study of the high-load conditions gives little indication of the life of a structure in normal service. (Fatigue, as explained in Chapter 10, is a result of pulsating stress.) Since conditions of use may vary widely from machine to machine, it is impossible to predict the exact future loading history of any particular motorcycle. The best we can do is to average out the expected life/load records. Such load cycle information can be obtained by use of a bike resembling the new design as closely as possible. This must be fitted with strain gauges and other transducers, then ridden under a wide variety of road conditions, with records kept of the time history of the loading. Given sufficient testing, we can thus build up a picture of the load cycles of a typical machine of that type over its expected lifetime. Additionally, data can be recorded of axle accelerations, these results can then be used to form the basis of the inputs to laboratory testing machines that can realistically load a motorcycle with completely repeatable test patterns, which mimic very closely a particular section of road or test track, to determine flexure and life expectancy characteristics.
Fig. 15.4 is an indication of the frequency of occurrence of different magnitude loads as found in practice. Note that the middle range of load occurs most frequently, while both low and high values occur much less often. This is to be expected as low loading is produced at low speeds on smooth roads, while high loading is a result of severe conditions — both extremes being encountered for only a relatively smal proportion of the machine’s life. Mathematical techniques are available for the treatment of such test data and, together with stress analysis, they enable us to investigate the frame’s fatigue behaviour. Fatigue analysis is based on summing the microscopic damage done during each load cycle, more damage is done at higher values of the load cycle. Such sophisticated testing methods are usually beyond the scope of the small chassis maker and the enthusiast building a one-off. Indeed, it is unlikely that even large manufacturers make full use of available techniques. However, good results can usually be achieved by rule-of-thumb and simpler calculations based on previous practice and experience. If the designer himself has the experience, so much the better. If not — and all designers must start somewhere — then a review of existing hardware is a good basis to build on. will result in a fragile structure, intolerant of hard use. Since metal fatigue is the most common cause of failure, in non-extreme circumstances, a study of the high-load conditions gives little indication of the life of a structure in normal service. (Fatigue, as explained in Chapter 10, is a result of pulsating stress.) Since conditions of use may vary widely from machine to machine, it is impossible to predict the exact future loading history of any particular motorcycle. The best we can do is to average out the expected life/load records. Such load cycle information can be obtained by use of a bike resembling the new design as closely as possible. This must be fitted with strain gauges and other transducers, then ridden under a wide variety of road conditions, with records kept of the time history of the loading. Given sufficient testing, we can thus build up a picture of the load cycles of a typical machine of that type over its expected lifetime. Additionally, data can be recorded of axle accelerations, these results can then be used to form the basis of the inputs to laboratory testing machines that can realistically load a motorcycle with completely repeatable test patterns, which mimic very closely a particular section of road or test track, to determine flexure and life expectancy characteristics.
Fig. 15.5 Called a “Two poster rig” this testing machine is designed to produce controlled vertical tyre loading on real motorcycles. The bike is supported by two powerful but fast acting servo-hydraulic rams. These are under the control of a computer programme and are vibrated to simulate real road surfaces, previously measured. Repeated tests can be done without further road testing. The side rods are to prevent the machine from falling over, without interfering with the vertical motions. The dummy rider is construction to mimic the dynamic behaviour of a real rider as closely as possible. (Dr Roahin Tithia NATS Gyvetame Carnnratinn)
Fig. 15.5 Called a “Two poster rig” this testing machine is designed to produce controlled vertical tyre loading on real motorcycles. The bike is supported by two powerful but fast acting servo-hydraulic rams. These are under the control of a computer programme and are vibrated to simulate real road surfaces, previously measured. Repeated tests can be done without further road testing. The side rods are to prevent the machine from falling over, without interfering with the vertical motions. The dummy rider is construction to mimic the dynamic behaviour of a real rider as closely as possible. (Dr Roahin Tithia NATS Gyvetame Carnnratinn)
The collection, processing and analysis of such data is a highly complex technical subject, and we can only just scratch the surface here. The new test systems would be largely impotent without the parallel development of advanced analysis techniques and software. The following is a very brief description of the principal steps involved to develop a system for testing and improving suspension settings and performance. Enormous strides have been made in these areas over the recent past, with large testing machines being built that can test, in repeatable laboratory conditions, whole bikes in conditions very close to real road situations. MTS Systems Corporation in the USA are important constructors of such simulators and software (to whom thanks are due for permission to use the following illustrations), which can, in the laboratory, re-create the same damaging chassis loads and motions that the motorcycle experiences on the road or proving ground. With these simulators, durability and other forms of repeatable testing can be carried out significantly faster.
The collection, processing and analysis of such data is a highly complex technical subject, and we can only just scratch the surface here. The new test systems would be largely impotent without the parallel development of advanced analysis techniques and software. The following is a very brief description of the principal steps involved to develop a system for testing and improving suspension settings and performance. Enormous strides have been made in these areas over the recent past, with large testing machines being built that can test, in repeatable laboratory conditions, whole bikes in conditions very close to real road situations. MTS Systems Corporation in the USA are important constructors of such simulators and software (to whom thanks are due for permission to use the following illustrations), which can, in the laboratory, re-create the same damaging chassis loads and motions that the motorcycle experiences on the road or proving ground. With these simulators, durability and other forms of repeatable testing can be carried out significantly faster.
Fig. 15.7 Showing the degree of accuracy with which the test rig in fig. 15.5 can reproduce real road inputs. There are two curves over-laid and it can be seen that there are only minor differences. One curve is of axle accelerometers durin a road test and the other curve shows the same event simulated on the rig. The principle event in the centre of the curves was hitting a 50 mm. step at 48 km/h. The vertical axis is vertical acceleration, and time is shown on the horizontal. Modern data collection is generally digital in nature, which means that data measurement is not continuous, but is sampled at short time intervals. This sample rate puts a limit on the maximum data frequency that can be resolved. This frequency is known as the “Nyquist frequency” and is half of the sampling frequency. To prevent errors (called aliasing) in the subsequent transformation into frequency response data, it is very important to filter out any road input frequencies higher than the Nyquist rate. This must be done by means of “anti-alias” filtering at the data collection source, it can’t be done by software later in the process, the information just will not in the data to do that.
Fig. 15.7 Showing the degree of accuracy with which the test rig in fig. 15.5 can reproduce real road inputs. There are two curves over-laid and it can be seen that there are only minor differences. One curve is of axle accelerometers durin a road test and the other curve shows the same event simulated on the rig. The principle event in the centre of the curves was hitting a 50 mm. step at 48 km/h. The vertical axis is vertical acceleration, and time is shown on the horizontal. Modern data collection is generally digital in nature, which means that data measurement is not continuous, but is sampled at short time intervals. This sample rate puts a limit on the maximum data frequency that can be resolved. This frequency is known as the “Nyquist frequency” and is half of the sampling frequency. To prevent errors (called aliasing) in the subsequent transformation into frequency response data, it is very important to filter out any road input frequencies higher than the Nyquist rate. This must be done by means of “anti-alias” filtering at the data collection source, it can’t be done by software later in the process, the information just will not in the data to do that.
Patented axle mounted wheel force transducers. These can measure three axle forces, up and down, fore and aft, and axial in addition to two axle bending moments. They are small enough to install with a wheel, and would be very useful when investigating some fundamental dynamic aspects of cornering etc. (Dr. Robin Tuluie, MTS Systems Corporation)
Patented axle mounted wheel force transducers. These can measure three axle forces, up and down, fore and aft, and axial in addition to two axle bending moments. They are small enough to install with a wheel, and would be very useful when investigating some fundamental dynamic aspects of cornering etc. (Dr. Robin Tuluie, MTS Systems Corporation)
Suspension testing. On the left an USD fork leg under going test. To the right is a rear damper unit. In this case the visible side connection is a cable to apply a lateral load to the damper. The cable passes over a pulley and is tensioned by a pneumatic cylinder, these are partially hidden by the vertical columns. Hydraulic testing machines like this can apply a wide range of input displacements to the dampers. For example, step inputs, sine waves, variable frequency sine wave sweeps and white noise shaped to match typical road spectra. The traditional mechanical test machines used a crank mechanism (like a crank from a single cylinder engine) and this produced a pseudo-sine wave only, similar to the piston movement in anengine. (Dr. Robin Tuluie, MTS Systems Corporation)
Suspension testing. On the left an USD fork leg under going test. To the right is a rear damper unit. In this case the visible side connection is a cable to apply a lateral load to the damper. The cable passes over a pulley and is tensioned by a pneumatic cylinder, these are partially hidden by the vertical columns. Hydraulic testing machines like this can apply a wide range of input displacements to the dampers. For example, step inputs, sine waves, variable frequency sine wave sweeps and white noise shaped to match typical road spectra. The traditional mechanical test machines used a crank mechanism (like a crank from a single cylinder engine) and this produced a pseudo-sine wave only, similar to the piston movement in anengine. (Dr. Robin Tuluie, MTS Systems Corporation)
Fig 16.1 When a weld cools it contracts. In this case the contraction will pull the vertical member to the right. The addition of the bracing member prevents distortion at the expense of residual stresses in the welds.
Fig 16.1 When a weld cools it contracts. In this case the contraction will pull the vertical member to the right. The addition of the bracing member prevents distortion at the expense of residual stresses in the welds.
The way in which a gusset stiffens a structure is not always understood. Imagine a tube attached to a rigid member and subjected to a bending force, as shown in figure 16.4. If there is no gusset to support it, the tube will bend over its entire length whereas, with a gusset, the bending is virtually restricted to the unsupported portion. The amount of flexure is proportional to the cube (third power) of the unsupported length, so reducing that by a half increases the stiffness by a factor of eight. Hence even small gussets can stiffen a multi-tube frame considerably. However, where a tube is stressed only in tension or compression a gusset can have no more than a minor effect. For that reason, they are seldom found on well-designed triangulated structures, except perhaps to provide mounting points. These are used at corner joints to stiffen the structure, spread loads and/or provide mounting points. Their design must be given careful consideration if excessive stress concentration and consequent weakening of the frame are to be avoided. We have already explained the benefit of attaching the gussets along the tube’s neutral axis but there are additional methods of improving the assembly. For example, it is better to taper the ends of the gusset rather than cut them off abruptly. Here, ease of welding and manufacture conflict with the requirements of minimizing stress concentration. Sometimes, where similar gussets are used on both sides of a tube, they are formed by one folded pressing, so that the two sides are joined by a web. In this case it is often better not to complete the weld round the tube (i.e. at the ends of the web) as this would introduce more stress concentration.
The way in which a gusset stiffens a structure is not always understood. Imagine a tube attached to a rigid member and subjected to a bending force, as shown in figure 16.4. If there is no gusset to support it, the tube will bend over its entire length whereas, with a gusset, the bending is virtually restricted to the unsupported portion. The amount of flexure is proportional to the cube (third power) of the unsupported length, so reducing that by a half increases the stiffness by a factor of eight. Hence even small gussets can stiffen a multi-tube frame considerably. However, where a tube is stressed only in tension or compression a gusset can have no more than a minor effect. For that reason, they are seldom found on well-designed triangulated structures, except perhaps to provide mounting points. These are used at corner joints to stiffen the structure, spread loads and/or provide mounting points. Their design must be given careful consideration if excessive stress concentration and consequent weakening of the frame are to be avoided. We have already explained the benefit of attaching the gussets along the tube’s neutral axis but there are additional methods of improving the assembly. For example, it is better to taper the ends of the gusset rather than cut them off abruptly. Here, ease of welding and manufacture conflict with the requirements of minimizing stress concentration. Sometimes, where similar gussets are used on both sides of a tube, they are formed by one folded pressing, so that the two sides are joined by a web. In this case it is often better not to complete the weld round the tube (i.e. at the ends of the web) as this would introduce more stress concentration.
Fig 16.5 Showing how an old flat lathe bed can provide a solid and accurate basis for a frame jig suitable for one-offs. It is also easy to see that if the frame were to be made in this position, then an additional tall jig structure would need to be made to support the head stock. This extra structure will reduce accuracy and rigidity. A jig is simply a structure to hold the component parts of our frame in their correct relationship during welding. It may be simple or complex, depending on the quantity and quality of the frames to be produced. For one-off jobs or experimental development work, the most versatile system is to use a rigid flat surface as a base to support easily made sub-jigs. An old flat lathe bed or table from a milling machine is ideal for this purpose, as squaring off the sides is easy and the centre gap or T-slots are ideal for bolting on fixtures.
Fig 16.5 Showing how an old flat lathe bed can provide a solid and accurate basis for a frame jig suitable for one-offs. It is also easy to see that if the frame were to be made in this position, then an additional tall jig structure would need to be made to support the head stock. This extra structure will reduce accuracy and rigidity. A jig is simply a structure to hold the component parts of our frame in their correct relationship during welding. It may be simple or complex, depending on the quantity and quality of the frames to be produced. For one-off jobs or experimental development work, the most versatile system is to use a rigid flat surface as a base to support easily made sub-jigs. An old flat lathe bed or table from a milling machine is ideal for this purpose, as squaring off the sides is easy and the centre gap or T-slots are ideal for bolting on fixtures.
Fig 16.6 Compare this with fig. 16.5. The head stock is simply bolted at right angles to the lathe bed with only a parallel! spacer block between. This is both accurate and rigid. The down side is that it is more difficult to visualize the frame as you are fabricating it, and a taller support jig would be necessary to hold the seat tubes, but required accuracy here is of a lower order than between the head stock and the swing-arm pivot. A final tip on jigging. It is natural to think of building a frame in the same orientation that it will adopt when finished. However, sometimes it is easier and more accurate to jig a frame ina different attitude to that which it will adopt on the road, even though you may find it a bit harder to visualize the placing of the various components. To jig a frame with its normal positioning may require more complex jigging, including a long, high steering-head fixture, particularly when a flat bed is used as a base. If the frame is properly drawn first (say to ‘4 or % scale) then simpler jigging may be possible, with the steering head clamped directly to the bed, as shown below.
Fig 16.6 Compare this with fig. 16.5. The head stock is simply bolted at right angles to the lathe bed with only a parallel! spacer block between. This is both accurate and rigid. The down side is that it is more difficult to visualize the frame as you are fabricating it, and a taller support jig would be necessary to hold the seat tubes, but required accuracy here is of a lower order than between the head stock and the swing-arm pivot. A final tip on jigging. It is natural to think of building a frame in the same orientation that it will adopt when finished. However, sometimes it is easier and more accurate to jig a frame ina different attitude to that which it will adopt on the road, even though you may find it a bit harder to visualize the placing of the various components. To jig a frame with its normal positioning may require more complex jigging, including a long, high steering-head fixture, particularly when a flat bed is used as a base. If the frame is properly drawn first (say to ‘4 or % scale) then simpler jigging may be possible, with the steering head clamped directly to the bed, as shown below.
Practical examples of tube profiling by two straight saw cuts. On the left we see symmetrical cuts to produce a right angle joint. The other example shows the two cuts at different angles to produce a joint of 50°. To obtain a perfectly fitted joint this example needed a small amount of adjustment with a half round file at points denoted by (A). Except in the case of square or rectangular tubing, it is necessary to shape the tube ends to match the mating structure. This is often done with a milling cutter of appropriate diameter set at the required angle. While this method presents no problem for a well equipped workshop, it is seldom that a specials builder has access to the necessary facilities. There are available, tube notching saws. Basically just a normal hole saw with a supporting fixture for the tubing. Personally I’ve not been too impressed with these tools. In most cases, however, two straight saw cuts at the end of a round tube will produce a nicely fitting joint. It is not very high tech. but when hand fitting tubes to a one-off frame I’ve found this method to be about the quickest. If the cut profile does not match up exactly with the mating tube then a quick adjustment with a file or grinder is all that’s necessary.
Practical examples of tube profiling by two straight saw cuts. On the left we see symmetrical cuts to produce a right angle joint. The other example shows the two cuts at different angles to produce a joint of 50°. To obtain a perfectly fitted joint this example needed a small amount of adjustment with a half round file at points denoted by (A). Except in the case of square or rectangular tubing, it is necessary to shape the tube ends to match the mating structure. This is often done with a milling cutter of appropriate diameter set at the required angle. While this method presents no problem for a well equipped workshop, it is seldom that a specials builder has access to the necessary facilities. There are available, tube notching saws. Basically just a normal hole saw with a supporting fixture for the tubing. Personally I’ve not been too impressed with these tools. In most cases, however, two straight saw cuts at the end of a round tube will produce a nicely fitting joint. It is not very high tech. but when hand fitting tubes to a one-off frame I’ve found this method to be about the quickest. If the cut profile does not match up exactly with the mating tube then a quick adjustment with a file or grinder is all that’s necessary.
So, if we know that the bending loads to which a frame member is subjected are mainly in one direction, we can tailor the tubing shape to achieve the most efficient structure. A typical example is a simple rear swing-arm, where any sideways load produces different bending moments about the neutral axes of the fork arms. See figure16.9. In this fairly typical layout - 325 mm. wheel radius and 250 mm. between fork arms — the vertical bending forces due to a sideways load are, as shown in the diagram, 2.6 times the horizontal forces. This depends on the tube shape and, except for a round section, differs about the neutral axis according to direction. In each case of fig. 16.8 the second moment of area is greater about the XOX, axis than about the YOY, axis, hence the resistance to bending is also greater about the XOX; axis.
So, if we know that the bending loads to which a frame member is subjected are mainly in one direction, we can tailor the tubing shape to achieve the most efficient structure. A typical example is a simple rear swing-arm, where any sideways load produces different bending moments about the neutral axes of the fork arms. See figure16.9. In this fairly typical layout - 325 mm. wheel radius and 250 mm. between fork arms — the vertical bending forces due to a sideways load are, as shown in the diagram, 2.6 times the horizontal forces. This depends on the tube shape and, except for a round section, differs about the neutral axis according to direction. In each case of fig. 16.8 the second moment of area is greater about the XOX, axis than about the YOY, axis, hence the resistance to bending is also greater about the XOX; axis.
Tube diameter and wall thickness are determined by the size, weight and power of the machine, also by the design family of the structure — e.g. triangulated, backbone, etc. Consequently the following is intended only as a general guide. (approximately 1.25 in.), and the second moment of area of that is 15900 mm.* —i.e., 1.22 times that of the square tube. So, even on its most favourable axes, the square tube is less stiff in bending besides being less capable of handling torsional and compressive loads.
Tube diameter and wall thickness are determined by the size, weight and power of the machine, also by the design family of the structure — e.g. triangulated, backbone, etc. Consequently the following is intended only as a general guide. (approximately 1.25 in.), and the second moment of area of that is 15900 mm.* —i.e., 1.22 times that of the square tube. So, even on its most favourable axes, the square tube is less stiff in bending besides being less capable of handling torsional and compressive loads.
Depending on detail design, 0.75 to 1.0 mm. steel sheeting may be used here, with possible thickening in regions of high stress or load application. Thicker material would be needed when made with aluminium, probably 1.2 to 1.5 mm. This type of frame derives its stiffness from its large cross-section, overall shape is not usually critical. Once more, however, the need to avoid local buckling cannot be over-emphasized and considerable thought must be given to all connection points if a satisfactory life is to be expected. Even on smaller machines, the main member is unlikely to be stiff enough if much under 50 mm diameter x 1.2 to 1.5 mm. wall. On larger machines, though, there is little to be gained by going above about 75 mm. diameter x 1.2 to 2.0 mm. wall. With this diameter it is better to avoid excessively thir walls, otherwise cracking may result at joints as a result of local buckling. When designing this type o frame, great care must be taken to feed in the loads in such a way as to minimize the risk of such loca buckling. This can be done by spreading the loads over a wide area, welding gussets and plates alonc the neutral axis wherever possible, and sometimes incorporating local reinforcement as shown in figure 16.10.
Depending on detail design, 0.75 to 1.0 mm. steel sheeting may be used here, with possible thickening in regions of high stress or load application. Thicker material would be needed when made with aluminium, probably 1.2 to 1.5 mm. This type of frame derives its stiffness from its large cross-section, overall shape is not usually critical. Once more, however, the need to avoid local buckling cannot be over-emphasized and considerable thought must be given to all connection points if a satisfactory life is to be expected. Even on smaller machines, the main member is unlikely to be stiff enough if much under 50 mm diameter x 1.2 to 1.5 mm. wall. On larger machines, though, there is little to be gained by going above about 75 mm. diameter x 1.2 to 2.0 mm. wall. With this diameter it is better to avoid excessively thir walls, otherwise cracking may result at joints as a result of local buckling. When designing this type o frame, great care must be taken to feed in the loads in such a way as to minimize the risk of such loca buckling. This can be done by spreading the loads over a wide area, welding gussets and plates alonc the neutral axis wherever possible, and sometimes incorporating local reinforcement as shown in figure 16.10.
are many people that | know that claim that high end CAD software is as essential a tool as good fabrication equipment, welders and lathes etc. but generally these are people that use CAD in their daily employment and hence are skilled in the method. So if you are serious about building your own frame then use whichever design method is best for you, if you are familiar with CAD then it can be an extremely powerful tool and design aid but if you have never used it and your main concern is building and riding your own machine in a reasonable time period, then be aware that the time spent learning to use this type of software may well be longer than the time spent in the workshop actually building your dream. On the other hand if you plan on getting some parts machined elsewhere then having previously designed these parts on computer and produced compatible files, you might find that having them made at a CNC machining centre may well save you some money.
are many people that | know that claim that high end CAD software is as essential a tool as good fabrication equipment, welders and lathes etc. but generally these are people that use CAD in their daily employment and hence are skilled in the method. So if you are serious about building your own frame then use whichever design method is best for you, if you are familiar with CAD then it can be an extremely powerful tool and design aid but if you have never used it and your main concern is building and riding your own machine in a reasonable time period, then be aware that the time spent learning to use this type of software may well be longer than the time spent in the workshop actually building your dream. On the other hand if you plan on getting some parts machined elsewhere then having previously designed these parts on computer and produced compatible files, you might find that having them made at a CNC machining centre may well save you some money.
| was asked to stiffen two frames for a 750 Kawasaki, the current model in the late 1980s. Despite favourable comments regarding the handling of this machine in many road tests, deficiencies were soon brought to light when subject to the much higher rigours of the race track. These particular frames were to be raced in a class where the rules forbade any modifications which entailed removing any parts, but it was permitted to add. As weight was obviously a priority, brute force was not the way to go, any added material had to earn its keep.
| was asked to stiffen two frames for a 750 Kawasaki, the current model in the late 1980s. Despite favourable comments regarding the handling of this machine in many road tests, deficiencies were soon brought to light when subject to the much higher rigours of the race track. These particular frames were to be raced in a class where the rules forbade any modifications which entailed removing any parts, but it was permitted to add. As weight was obviously a priority, brute force was not the way to go, any added material had to earn its keep.
The photographs show the finished modifications that were done on this Kawasaki 750., anc incorporated are examples of many of the techniques mentioned in other chapters.
The photographs show the finished modifications that were done on this Kawasaki 750., anc incorporated are examples of many of the techniques mentioned in other chapters.
The completed modified frame. Note the triangulation of the front down tubes, the asymmetric layout was determined by location of exhaust pipes and crankcase shape. The pyramid above the swing arm pivot provides in-plane torsional stiffness as well as lateral bracing.
The completed modified frame. Note the triangulation of the front down tubes, the asymmetric layout was determined by location of exhaust pipes and crankcase shape. The pyramid above the swing arm pivot provides in-plane torsional stiffness as well as lateral bracing.
The large open four sided area under the engine could be seen to be lozenging when loaded on | stiffness testing rig, but the engine itself prevented any cross bracing and so the two rear corners we gusseted as much as room. allowed. The similar area above the engine was approximating to a Ic thin triangle anyway, and so needed no similar treatment.
The large open four sided area under the engine could be seen to be lozenging when loaded on | stiffness testing rig, but the engine itself prevented any cross bracing and so the two rear corners we gusseted as much as room. allowed. The similar area above the engine was approximating to a Ic thin triangle anyway, and so needed no similar treatment.
As the wheel moves relative to the body the transducers sensing load, wheel position, acceleration ar velocity, send their signals to the computer together with the output of pitch and roll accelerometer mounted in the body. These signals are processed to determine whether the wheel movement is due f bumps, cornering roll or to acceleration/braking induced pitch. Signals are then passed to the contr valves to leave, increase or decrease the amount of fluid in each actuator. The result is a system whic can eliminate pitch, roll and ride height variations while at the same time giving a very comfortable rid Often the requirements for a smooth ride are in conflict with those other needs, and so as usual compromise must be struck. Active suspension can greatly reduce this need for compromise. Since th first Lotus designs were unveiled much progress has been made, particularly in the electronics ar control algorithms and even though these developments have been mainly aimed at cars, it can b argued that motorcycles actually have more to gain from improved suspension response. The most significant milestone came from Lotus Cars with the application of electronics to control hydraulic actuators that provided total suspension support without springs nor dampers. The drawing shows a simplified layout of this design.
As the wheel moves relative to the body the transducers sensing load, wheel position, acceleration ar velocity, send their signals to the computer together with the output of pitch and roll accelerometer mounted in the body. These signals are processed to determine whether the wheel movement is due f bumps, cornering roll or to acceleration/braking induced pitch. Signals are then passed to the contr valves to leave, increase or decrease the amount of fluid in each actuator. The result is a system whic can eliminate pitch, roll and ride height variations while at the same time giving a very comfortable rid Often the requirements for a smooth ride are in conflict with those other needs, and so as usual compromise must be struck. Active suspension can greatly reduce this need for compromise. Since th first Lotus designs were unveiled much progress has been made, particularly in the electronics ar control algorithms and even though these developments have been mainly aimed at cars, it can b argued that motorcycles actually have more to gain from improved suspension response. The most significant milestone came from Lotus Cars with the application of electronics to control hydraulic actuators that provided total suspension support without springs nor dampers. The drawing shows a simplified layout of this design.
The 2WD Rokon makes riding over logs and through deep mud look easy. The model on the right has sealed aluminium disc wheels which with the balloon tyres allows the machine to be floated across deep streams. (Rokon) the front wheel to spin slightly slower than the rear, and have a free-wheel system, so that in norma conditions the drive is only through the rear wheel. When the rear loses enough grip and spins then the front wheel becomes driven also. The Rokon is the only 2WD machine that made it into production, anc that uses a fixed drive ratio of 1:1 with a free-wheel system.
The 2WD Rokon makes riding over logs and through deep mud look easy. The model on the right has sealed aluminium disc wheels which with the balloon tyres allows the machine to be floated across deep streams. (Rokon) the front wheel to spin slightly slower than the rear, and have a free-wheel system, so that in norma conditions the drive is only through the rear wheel. When the rear loses enough grip and spins then the front wheel becomes driven also. The Rokon is the only 2WD machine that made it into production, anc that uses a fixed drive ratio of 1:1 with a free-wheel system.
The Drysdale 2WS and 2WD motorcycle. The five piston radial hydraulic motor can be clearly seen in the hub of the front wheel shown on the left. Note also the cross-wise cylinder, which is for steering. The photo to the right shows the rear wheel in a steered position. This machine now resides in the Donington Collection Museum. (Drysdale) Australian lan Drysdale completed such a 2WD off-road machine, which also featured hydraulic two wheel steer. To allow for steering and suspension movement it is necessary to use some form of flexible connection between the wheel mounted motors and the main chassis but flexible hoses become extremely stiff when pressurized. The Drysdale surmounted this problem by using swing-arm suspension at each end and using oil tight rotating joints mounted coaxially with the pivots for the arms. This machine reportedly worked extremely well in slippery conditions.
The Drysdale 2WS and 2WD motorcycle. The five piston radial hydraulic motor can be clearly seen in the hub of the front wheel shown on the left. Note also the cross-wise cylinder, which is for steering. The photo to the right shows the rear wheel in a steered position. This machine now resides in the Donington Collection Museum. (Drysdale) Australian lan Drysdale completed such a 2WD off-road machine, which also featured hydraulic two wheel steer. To allow for steering and suspension movement it is necessary to use some form of flexible connection between the wheel mounted motors and the main chassis but flexible hoses become extremely stiff when pressurized. The Drysdale surmounted this problem by using swing-arm suspension at each end and using oil tight rotating joints mounted coaxially with the pivots for the arms. This machine reportedly worked extremely well in slippery conditions.
Ohlins/Yamaha 2WD system. The photo on the left shows how the hydraulic motor is fitted to the front wheel hub. The size of the motor and the feed hoses indicates that only a small proportion of the engine power is directed to the front wheel. On the right is the detail of how the pump is driven at an increased speed from the rear drive sprocket. and feed hoses indicate that this system is only intended to pass a relatively small amount of power t the front wheel. Even a small amount of drive from the front would probably be of great assistance i terrain where the front can get bogged down, but when we see that motoX riders keep the front wheel i the air for large periods of time it is only natural to query whether there is much benefit to be gained fron this development at the top level of competition. It is probably the less skilled rider that would benefit th most in off-road terrain.
Ohlins/Yamaha 2WD system. The photo on the left shows how the hydraulic motor is fitted to the front wheel hub. The size of the motor and the feed hoses indicates that only a small proportion of the engine power is directed to the front wheel. On the right is the detail of how the pump is driven at an increased speed from the rear drive sprocket. and feed hoses indicate that this system is only intended to pass a relatively small amount of power t the front wheel. Even a small amount of drive from the front would probably be of great assistance i terrain where the front can get bogged down, but when we see that motoX riders keep the front wheel i the air for large periods of time it is only natural to query whether there is much benefit to be gained fron this development at the top level of competition. It is probably the less skilled rider that would benefit th most in off-road terrain.
Fig. 18.3 This plot shows the actual kinematic turn radii of the front wheel for different ratios between the steering angle of the rear and front wheels. —1.0 being opposite sense steering with equal steering angles, 0 is when the rear wheel is fixed and doesn’t steer. The individual curves are for different front wheel steering angles as marked in degrees. The vertical scale has been changed in the inset to highlight the singularity that occurs with equal steer angles at each end. Fig. 18.2 Two wheel steering. Left: The small turning radius of “opposite sense” 2WS is demonstrated, compare this with a normal front wheel only steer system, middie. “same sense” 2WS, right, with a 1:1 ratio between front and rear steering angles, clearly has the turn centre at infinity giving a straight line motion parallel to the wheel angles. The un- steered part of the bike would stay aligned with its original orientation. The machine would not negotiate a curve but could instead, adopt a directionally stable straight line attitude over a range of steer angles, resulting in the machine moving crab-wise as shown. The tendency to keep the wheels aligned with the chassis would be quite weak. The normal self aligning effects of trail would be satisfied as long as the two wheels remained parallel to the direction of travel.
Fig. 18.3 This plot shows the actual kinematic turn radii of the front wheel for different ratios between the steering angle of the rear and front wheels. —1.0 being opposite sense steering with equal steering angles, 0 is when the rear wheel is fixed and doesn’t steer. The individual curves are for different front wheel steering angles as marked in degrees. The vertical scale has been changed in the inset to highlight the singularity that occurs with equal steer angles at each end. Fig. 18.2 Two wheel steering. Left: The small turning radius of “opposite sense” 2WS is demonstrated, compare this with a normal front wheel only steer system, middie. “same sense” 2WS, right, with a 1:1 ratio between front and rear steering angles, clearly has the turn centre at infinity giving a straight line motion parallel to the wheel angles. The un- steered part of the bike would stay aligned with its original orientation. The machine would not negotiate a curve but could instead, adopt a directionally stable straight line attitude over a range of steer angles, resulting in the machine moving crab-wise as shown. The tendency to keep the wheels aligned with the chassis would be quite weak. The normal self aligning effects of trail would be satisfied as long as the two wheels remained parallel to the direction of travel.
Lateral movement - metres. Fig. 18.4 The kinematic (low speed) paths, of the mid-point of the wheelbase, travelled by the two opposite classes of 2WS and a normal front wheel only steer. All three curves are for a wheelbase of 1.5 metres and assume an instantaneously applied steering angle of 2°, and the two 2WS cases have front to rear steering ratios of 1:1. At the low values of steering angle used in normal riding (less than about 10°) the paths cross at one bike length.
Lateral movement - metres. Fig. 18.4 The kinematic (low speed) paths, of the mid-point of the wheelbase, travelled by the two opposite classes of 2WS and a normal front wheel only steer. All three curves are for a wheelbase of 1.5 metres and assume an instantaneously applied steering angle of 2°, and the two 2WS cases have front to rear steering ratios of 1:1. At the low values of steering angle used in normal riding (less than about 10°) the paths cross at one bike length.
The author constructed this 2WS coasting bicycle to test the basic balance and rideability characteristics. The steering link could be mounted on either side at the rear so providing same sense and opposite sense steering as required. The steering ratio was also adjustable between 1:1 and 2:1 (rear wheel steered through half the angle of the front). The riding shot shows same sense steering at a 1:1 ratio. Note the crab like motion similar to that shown on the right in fig. 18.2. As shown by the Drysdale example it doesn’t have to be done that way. The Drysdale machine very cleverly circumvents these problems by allowing the front to have approximately 5° of steering angle before the rear begins to steer. Balance and counter-steering movements normally require less than 5° and so gentle riding is accomplished in a completely normal fashion. It is only during hard off-road riding that steering angles are large and that is when the Drysdale brings the rear steering into effect. It was reported that this machine was very agile and had a strong resistance against the rear sliding out, but it must be remembered that it was 2WD as well as 2WS.
The author constructed this 2WS coasting bicycle to test the basic balance and rideability characteristics. The steering link could be mounted on either side at the rear so providing same sense and opposite sense steering as required. The steering ratio was also adjustable between 1:1 and 2:1 (rear wheel steered through half the angle of the front). The riding shot shows same sense steering at a 1:1 ratio. Note the crab like motion similar to that shown on the right in fig. 18.2. As shown by the Drysdale example it doesn’t have to be done that way. The Drysdale machine very cleverly circumvents these problems by allowing the front to have approximately 5° of steering angle before the rear begins to steer. Balance and counter-steering movements normally require less than 5° and so gentle riding is accomplished in a completely normal fashion. It is only during hard off-road riding that steering angles are large and that is when the Drysdale brings the rear steering into effect. It was reported that this machine was very agile and had a strong resistance against the rear sliding out, but it must be remembered that it was 2WD as well as 2WS.
quite a long time and important balance correcting manoeuvres need to be made within that time. II would seem therefore that very low speed balance might be enhanced with same sense 2WS. On the other hand the opposite sense system would initially suffer a delay (fig. 18.5) before the mid-point of the wheelbase would begin to move significantly, leading to a sluggish response.
quite a long time and important balance correcting manoeuvres need to be made within that time. II would seem therefore that very low speed balance might be enhanced with same sense 2WS. On the other hand the opposite sense system would initially suffer a delay (fig. 18.5) before the mid-point of the wheelbase would begin to move significantly, leading to a sluggish response.
The Quasar in motion. Despite its weight and long wheel-base, 1800 mm., it had an agile cornering performance. Even with a relatively underpowered engine, the superior aerodynamics enabled a high top speed. Built to heavy construction standards, this machine has been shown to withstand frontal impacts better than conventional motorcycles. (Clive Dixon) Even though there had been many historic designs with a seating position more akin to that of a car thar a conventional motorcycle, without doubt any resurgence of modern interest was given a boost by the Quasar design of the late Malcolm Newell.
The Quasar in motion. Despite its weight and long wheel-base, 1800 mm., it had an agile cornering performance. Even with a relatively underpowered engine, the superior aerodynamics enabled a high top speed. Built to heavy construction standards, this machine has been shown to withstand frontal impacts better than conventional motorcycles. (Clive Dixon) Even though there had been many historic designs with a seating position more akin to that of a car thar a conventional motorcycle, without doubt any resurgence of modern interest was given a boost by the Quasar design of the late Malcolm Newell.
Above, the Suzuki Burgman. The long wheelbase and small scooter wheels give a low reclining riding position, but the rider’s back rest is insufficient to give full back support. The pillion passenger is seated higher and benefits from a slightly bigger back rest. For many years this type of machine has been the preserve of a small band of enthusiasts but over the past decade there has been increasing interest from the major manufacturers, with several producing what are commonly referred to as “super scooters”, the Suzuki Burgman for example. These machines have quite low seating positions with a more reclining posture than traditional scooters, but are built considerably longer to achieve this. However, the lack of a full backrest and enclosing bodywork denies the rider the extra comfort otherwise possible from this seating layout.
Above, the Suzuki Burgman. The long wheelbase and small scooter wheels give a low reclining riding position, but the rider’s back rest is insufficient to give full back support. The pillion passenger is seated higher and benefits from a slightly bigger back rest. For many years this type of machine has been the preserve of a small band of enthusiasts but over the past decade there has been increasing interest from the major manufacturers, with several producing what are commonly referred to as “super scooters”, the Suzuki Burgman for example. These machines have quite low seating positions with a more reclining posture than traditional scooters, but are built considerably longer to achieve this. However, the lack of a full backrest and enclosing bodywork denies the rider the extra comfort otherwise possible from this seating layout.
BMW have adopted a somewhat different approach, weather protection and safety being the obvious motives behind their C1, this has a fairly upright seating position similar to traditional scooters but the rider is provided with a fully supporting seat including head rest. The bodywork and windscreen add
BMW have adopted a somewhat different approach, weather protection and safety being the obvious motives behind their C1, this has a fairly upright seating position similar to traditional scooters but the rider is provided with a fully supporting seat including head rest. The bodywork and windscreen add
The Calleja system with two rear wheels. Each wheel can freely move relative to the other to allow normal leaning in a corner, and a single suspension unit handles the bumps from either. This clever design does not increase overall width and would be ideal for enabling stationary balance for an enclosed machine. The Swiss built Eco-mobile, in production since 1987, provides a well streamlined shape which gives total occupant enclosure like a car. It is of necessity long and heavy by motorcycle standards, 3.7 m. and 450 kg.. The full enclosure prevents the rider from extending his leg to maintain stationary balance and this machine is provided with retractable out-rigger wheels mounted amidships for this purpose. These detract somewhat from the neat appearance and streamlining of the body. Like the BMW C1, seat belts are fitted and helmet law exemption has been granted.
The Calleja system with two rear wheels. Each wheel can freely move relative to the other to allow normal leaning in a corner, and a single suspension unit handles the bumps from either. This clever design does not increase overall width and would be ideal for enabling stationary balance for an enclosed machine. The Swiss built Eco-mobile, in production since 1987, provides a well streamlined shape which gives total occupant enclosure like a car. It is of necessity long and heavy by motorcycle standards, 3.7 m. and 450 kg.. The full enclosure prevents the rider from extending his leg to maintain stationary balance and this machine is provided with retractable out-rigger wheels mounted amidships for this purpose. These detract somewhat from the neat appearance and streamlining of the body. Like the BMW C1, seat belts are fitted and helmet law exemption has been granted.
Eco-mobile. The extremely long wheelbase can be clearly seen, as can the retractable out-rigger balance wheels. Powered by the flat 4 cylinder BMW engine, approximately 100 of these machines have been built in single (/eft) and dual (right) seat versions with fully supporting seats as in a car. On the right, designer Arnold Wagner demonstrates the capabilities of this interesting machine. Current models are capable of 300 km/h., largely due to the body shape. (Peraves AG.) considerably more weather protection than normally available on a motorcycle or scooter. This machine is fitted with seat belts and has been granted exemption from the helmet laws in some countries.
Eco-mobile. The extremely long wheelbase can be clearly seen, as can the retractable out-rigger balance wheels. Powered by the flat 4 cylinder BMW engine, approximately 100 of these machines have been built in single (/eft) and dual (right) seat versions with fully supporting seats as in a car. On the right, designer Arnold Wagner demonstrates the capabilities of this interesting machine. Current models are capable of 300 km/h., largely due to the body shape. (Peraves AG.) considerably more weather protection than normally available on a motorcycle or scooter. This machine is fitted with seat belts and has been granted exemption from the helmet laws in some countries.
The Spanish designed Calleja has two independently sprung rear wheels, which allow full leaning for cornering. Unlikely to find much application on traditional motorcycles this design is an excellent solutior to the problem of stationary balance for an enclosed motorcycle of either FF or traditional layout. The twe rear swing-arms are connected by a pivoted balance bar to allow for the differential movement of the twc arms, and this balance bar is fitted with a brake. The brake locks the two swing-arms together, thus holding the machine upright when stationary. Application of this brake can be rider controlled or unde the command of a computerized system. Fitment of this design to a vehicle like the C1 for example, would allow a higher degree of side enclosure and weather protection, because riders would not need to put their feet down until it was time to dismount.
The Spanish designed Calleja has two independently sprung rear wheels, which allow full leaning for cornering. Unlikely to find much application on traditional motorcycles this design is an excellent solutior to the problem of stationary balance for an enclosed motorcycle of either FF or traditional layout. The twe rear swing-arms are connected by a pivoted balance bar to allow for the differential movement of the twc arms, and this balance bar is fitted with a brake. The brake locks the two swing-arms together, thus holding the machine upright when stationary. Application of this brake can be rider controlled or unde the command of a computerized system. Fitment of this design to a vehicle like the C1 for example, would allow a higher degree of side enclosure and weather protection, because riders would not need to put their feet down until it was time to dismount.
To keep other variables to a minimum, the original frame and suspension were retained and the wheelbase remained unaltered. Two non-standard rake angles were tried. In each case the ground trail was kept to approximately the same as the standard value (i.e. 89 mm.). The first alternative set-up tried was with a rake of approximately 15 degrees and almost nil offset. This was achieved by bolting a superstructure to the frame to support the new headstock (see photo) and reversing the yokes, since their offset is very close to that of the wheel spindle, the overall offset was reduced virtually to zero. For the second setup, the rake was close to zero (i.e., vertical steering axis). This was achieved by reversing the complete front-fork assembly, thus giving the negative offset necessary to maintain the standard trail. The new headstock was supported by an extension of the
To keep other variables to a minimum, the original frame and suspension were retained and the wheelbase remained unaltered. Two non-standard rake angles were tried. In each case the ground trail was kept to approximately the same as the standard value (i.e. 89 mm.). The first alternative set-up tried was with a rake of approximately 15 degrees and almost nil offset. This was achieved by bolting a superstructure to the frame to support the new headstock (see photo) and reversing the yokes, since their offset is very close to that of the wheel spindle, the overall offset was reduced virtually to zero. For the second setup, the rake was close to zero (i.e., vertical steering axis). This was achieved by reversing the complete front-fork assembly, thus giving the negative offset necessary to maintain the standard trail. The new headstock was supported by an extension of the
The standard BMW R75/5 used as a basis for the experiments in rake and trail, maker’s figures were 27 degrees rake and 89 mm. ground trail. original super-structure. In both cases the handlebar was pivoted in the usual place and connected to the fork by a ball-jointed link, a side effect of this was an adjustable steering ratio — i.e., for a given steering angle at the fork the angle needed at the handle-bar could be varied. With the 15-degree rake the bike had full road equipment, including lighting, so that it could be ridden under everyday conditions; indeed, five riders covered over 3000 kms. between them, including wet and dry going, bumpy country lanes, London traffic and motorway trips. Throughout this period, no steering damper was fitted.
The standard BMW R75/5 used as a basis for the experiments in rake and trail, maker’s figures were 27 degrees rake and 89 mm. ground trail. original super-structure. In both cases the handlebar was pivoted in the usual place and connected to the fork by a ball-jointed link, a side effect of this was an adjustable steering ratio — i.e., for a given steering angle at the fork the angle needed at the handle-bar could be varied. With the 15-degree rake the bike had full road equipment, including lighting, so that it could be ridden under everyday conditions; indeed, five riders covered over 3000 kms. between them, including wet and dry going, bumpy country lanes, London traffic and motorway trips. Throughout this period, no steering damper was fitted.
the effect on directional stability is considerably smaller and the rider is less aware of the rut. Thus the steeper the steering axis the smaller the disturbing effect. It was previously suggested that balance might be enhanced, particularly at low speeds, by steepening the head angle. To verify this, much riding was done at very low speeds and balance was indeed improved by the modifications. The machine could easily be ridden more slowly than when in standard trim before the rider had to put a foot down. (Of course, champion trials riders can balance a stationary machine indefinitely, but that is exceptional and most of us need to be moving slightly to maintain balance.) In heavy traffic, it was noticeably easier to trickle along slowly on the modified BMW, making it less tiring to ride from one side of London to the other.
the effect on directional stability is considerably smaller and the rider is less aware of the rut. Thus the steeper the steering axis the smaller the disturbing effect. It was previously suggested that balance might be enhanced, particularly at low speeds, by steepening the head angle. To verify this, much riding was done at very low speeds and balance was indeed improved by the modifications. The machine could easily be ridden more slowly than when in standard trim before the rider had to put a foot down. (Of course, champion trials riders can balance a stationary machine indefinitely, but that is exceptional and most of us need to be moving slightly to maintain balance.) In heavy traffic, it was noticeably easier to trickle along slowly on the modified BMW, making it less tiring to ride from one side of London to the other.
The top photo shows the superstructure which gave an unaltered trail with zero offset and rake angle of 15 deg. The second shows the extended structure, with the reversed forks, giving equal trail with near zero rake angle.
The top photo shows the superstructure which gave an unaltered trail with zero offset and rake angle of 15 deg. The second shows the extended structure, with the reversed forks, giving equal trail with near zero rake angle.
This effect apart, one of the most interesting results (mentioned by all the riders) was the surprisingly normal feel of the modified machine, with the steering pleasantly light at low speeds but always totally stable. No special riding technique was required and cornering was accomplished normally. The variable steering ratio mentioned earlier was tried from 1:1 (equivalent to conventional direct steering) to 1:2 (steering angle doubled from handlebar to fork). In normal riding (dry roads) it was difficult to detect the difference, only when manoeuvring at a standstill, using large steering angles, was the heavier feel of the 1:2 ratio noticeable. It was under braking, however, that a disadvantage was noticed, in the form of severe shuddering in the fork as the braking force tried to bend it backward. Naturally, this was more severe with the steering axis upright. Such juddering was entirely absent with the standard rake, though quite bad at 15 degrees. However, since the steeper rakes made the steering lighter anyway, even the effort required with the 1:2 ratio felt similar to that with the standard machine. Indeed, the reduced handlebar swing could be a bonus when designing a non-steerable fairing, as handlebar clearance usually results in a bulky shape if steering lock is not to be restricted.
This effect apart, one of the most interesting results (mentioned by all the riders) was the surprisingly normal feel of the modified machine, with the steering pleasantly light at low speeds but always totally stable. No special riding technique was required and cornering was accomplished normally. The variable steering ratio mentioned earlier was tried from 1:1 (equivalent to conventional direct steering) to 1:2 (steering angle doubled from handlebar to fork). In normal riding (dry roads) it was difficult to detect the difference, only when manoeuvring at a standstill, using large steering angles, was the heavier feel of the 1:2 ratio noticeable. It was under braking, however, that a disadvantage was noticed, in the form of severe shuddering in the fork as the braking force tried to bend it backward. Naturally, this was more severe with the steering axis upright. Such juddering was entirely absent with the standard rake, though quite bad at 15 degrees. However, since the steeper rakes made the steering lighter anyway, even the effort required with the 1:2 ratio felt similar to that with the standard machine. Indeed, the reduced handlebar swing could be a bonus when designing a non-steerable fairing, as handlebar clearance usually results in a bulky shape if steering lock is not to be restricted.
With the modified BMW, the author demonstrates the stability achieved with the 15 degree rake angle. In both experiments the handlebar was connected to the fork by a rose-jointed link providing for steering ratios anywhere between 1:1 (direct) and 2:1 (geared up). This was accomplished by sliding the bar atop the forks toward greater or lesser radii.
With the modified BMW, the author demonstrates the stability achieved with the 15 degree rake angle. In both experiments the handlebar was connected to the fork by a rose-jointed link providing for steering ratios anywhere between 1:1 (direct) and 2:1 (geared up). This was accomplished by sliding the bar atop the forks toward greater or lesser radii.
Trail To test this, a double-wishbone front-suspension system was fitted to the BMW (see photo), making i possible to alter the trail by adjusting the wishbone lengths to give variations in rake. There was nc offset, and trail values from 50 to 100 mm. were tried varying the rake angle either side of 15 degrees Although the machine was perfectly ridable over the full range of settings, the front end proved livelier and the steering more sensitive to bumps as the trail was shortened (steeper rake), albeit never so muct as with the standard set-up. In the upper range of trail settings the bike was very steady but could stil be manoeuvred quickly. At about 76 mm. trail there was a tendency for the machine to lift itself out of é bend when cornered at moderate banking angles (say, 15 to 20 degrees), though no such effect was detectable at higher cornering speeds requiring, say, 35 to 40 degrees of lean. When a machine is banked into a bend, trail gives rise to two opposing effects: (1) directional stability, tending to make the machine run straight and (2) the self-steering effect (also dependent on rake and wheel diameter} mentioned in Chapter 3, tending to steer the machine into the bend. To achieve neutral handling, these effects have to be properly balanced (in concert with several other parameters), and the problem jus' mentioned, whilst not properly understood, is thought to have been caused by an unsuitable combinatior of these effects for that particular machine at a critical rake or trail. It is thought that at the greater bank angles, the self-steering effect would have outweighed the straight-ahead tendency.
Trail To test this, a double-wishbone front-suspension system was fitted to the BMW (see photo), making i possible to alter the trail by adjusting the wishbone lengths to give variations in rake. There was nc offset, and trail values from 50 to 100 mm. were tried varying the rake angle either side of 15 degrees Although the machine was perfectly ridable over the full range of settings, the front end proved livelier and the steering more sensitive to bumps as the trail was shortened (steeper rake), albeit never so muct as with the standard set-up. In the upper range of trail settings the bike was very steady but could stil be manoeuvred quickly. At about 76 mm. trail there was a tendency for the machine to lift itself out of é bend when cornered at moderate banking angles (say, 15 to 20 degrees), though no such effect was detectable at higher cornering speeds requiring, say, 35 to 40 degrees of lean. When a machine is banked into a bend, trail gives rise to two opposing effects: (1) directional stability, tending to make the machine run straight and (2) the self-steering effect (also dependent on rake and wheel diameter} mentioned in Chapter 3, tending to steer the machine into the bend. To achieve neutral handling, these effects have to be properly balanced (in concert with several other parameters), and the problem jus' mentioned, whilst not properly understood, is thought to have been caused by an unsuitable combinatior of these effects for that particular machine at a critical rake or trail. It is thought that at the greater bank angles, the self-steering effect would have outweighed the straight-ahead tendency.
Normally the equations describing gyroscopic motion are derived in physics books using the idea of momentum acting along the axis of rotation, but the mathematics of this are beyond this book and the explanations don’t provide an intuitive feel for why precession works as it does. Presented here is an explanation, based on the instantaneous linear momentum of two particles contained within the rotating wheel, which clearly demonstrates the basic mechanism and direction of gyroscopic reactions. Rotating objects such as motorcycle wheels and crankshafts are subject to the demands of the conservation of momentum but exactly how is somewhat less obvious than in the cases of linear o curvilinear motion as explained more fully in the appendix on basic mechanics. Each particle of the object has its own instantaneous linear momentum and could be considered individually and the overa effect determined by summing the separate changes in momentum. In this case the momentum of eac! particle is in a different changing direction and so it is rather difficult to visualize and discuss the situatiot on that basis. In order to provide an easy way of representing the momentum of rotating objects (calle: angular momentum) we adopt the convention that angular momentum acts along a straight line coaxie with the axis of rotation, as in fig. A4.1. This is a universal way of representing angular momentum bu we must always remember that it is only a man-made convenience, the “real” momentum of individue particles is really still in the plane of the disc’s rotation. As with the linear case above we have to appl some outside force to change the momentum as shown, in the case of angular momentum this force is i the form of torque, a twisting force.
Normally the equations describing gyroscopic motion are derived in physics books using the idea of momentum acting along the axis of rotation, but the mathematics of this are beyond this book and the explanations don’t provide an intuitive feel for why precession works as it does. Presented here is an explanation, based on the instantaneous linear momentum of two particles contained within the rotating wheel, which clearly demonstrates the basic mechanism and direction of gyroscopic reactions. Rotating objects such as motorcycle wheels and crankshafts are subject to the demands of the conservation of momentum but exactly how is somewhat less obvious than in the cases of linear o curvilinear motion as explained more fully in the appendix on basic mechanics. Each particle of the object has its own instantaneous linear momentum and could be considered individually and the overa effect determined by summing the separate changes in momentum. In this case the momentum of eac! particle is in a different changing direction and so it is rather difficult to visualize and discuss the situatiot on that basis. In order to provide an easy way of representing the momentum of rotating objects (calle: angular momentum) we adopt the convention that angular momentum acts along a straight line coaxie with the axis of rotation, as in fig. A4.1. This is a universal way of representing angular momentum bu we must always remember that it is only a man-made convenience, the “real” momentum of individue particles is really still in the plane of the disc’s rotation. As with the linear case above we have to appl some outside force to change the momentum as shown, in the case of angular momentum this force is i the form of torque, a twisting force.
Gyroscopic reaction and precession are all about this relationship between a torque and an angular velocity each acting on a pair of mutually perpendicular axis, which are themselves both perpendicular to the axis of rotation of the object. If a torque is applied to one of this pair of axis then the object will generate an angular velocity about the other, this is known as precession. Conversely, if a disc is precessing about one of these axis then we know that a torque is being applied about the other. The torque about a horizontal axis has produced a velocity about a vertical axis, no wonder the reactions sometimes seem magical and counter-intuitive.
Gyroscopic reaction and precession are all about this relationship between a torque and an angular velocity each acting on a pair of mutually perpendicular axis, which are themselves both perpendicular to the axis of rotation of the object. If a torque is applied to one of this pair of axis then the object will generate an angular velocity about the other, this is known as precession. Conversely, if a disc is precessing about one of these axis then we know that a torque is being applied about the other. The torque about a horizontal axis has produced a velocity about a vertical axis, no wonder the reactions sometimes seem magical and counter-intuitive.
There is another little talked about aspect of gyroscopic reaction that affects the final balanced lean angle of a bike. As we negotiate a turn the whole bike is subject to a yaw velocity equal to the angular velocity of the bike about the turn centre. Depending on the lean angle of the bike there will be a component of this yaw velocity about the local “vertical” or z axis of the bike, this must be supported by a roll torque acting away from the turn, which in turn must be balanced by an increased lean angle of the bike. In practice this only amounts to about 1 degree at maximum cornering speed. We know that we only turn the steering by a very small amount at normal riding speeds and so it follows that the steering angular velocity must be very low, and so the gyroscopic coupling between roll torque and steering velocity is relatively unimportant. However, during the lean-in process the roll velocity is significant and so the coupling between roll velocity and steering torque is of prime importance. This is further elaborated upon in chapter 4.
There is another little talked about aspect of gyroscopic reaction that affects the final balanced lean angle of a bike. As we negotiate a turn the whole bike is subject to a yaw velocity equal to the angular velocity of the bike about the turn centre. Depending on the lean angle of the bike there will be a component of this yaw velocity about the local “vertical” or z axis of the bike, this must be supported by a roll torque acting away from the turn, which in turn must be balanced by an increased lean angle of the bike. In practice this only amounts to about 1 degree at maximum cornering speed. We know that we only turn the steering by a very small amount at normal riding speeds and so it follows that the steering angular velocity must be very low, and so the gyroscopic coupling between roll torque and steering velocity is relatively unimportant. However, during the lean-in process the roll velocity is significant and so the coupling between roll velocity and steering torque is of prime importance. This is further elaborated upon in chapter 4.
Most people are familiar with degrees and there are 360° in a complete revolution. Radians are sometimes referred to as a natural unit and do often occur naturally when deriving basic behaviour of many handling characteristics. It is therefore worth becoming familiar with their use. A radian is approximately equal to 57.3 degrees and is the angle formed by a segment of a circle where the length around the part circumference is equal to the radius. There are not an integral number of radians in a full circle, unlike degrees, instead there are 2x or 6.283.
Most people are familiar with degrees and there are 360° in a complete revolution. Radians are sometimes referred to as a natural unit and do often occur naturally when deriving basic behaviour of many handling characteristics. It is therefore worth becoming familiar with their use. A radian is approximately equal to 57.3 degrees and is the angle formed by a segment of a circle where the length around the part circumference is equal to the radius. There are not an integral number of radians in a full circle, unlike degrees, instead there are 2x or 6.283.
The behaviour of most physical systems is controlled by a set of rules known as Newton’s three laws, which were published in 1687. These are not some arbitrary man-made laws which decree how things must behave, rather they are statements which formalize how things have been observed, over a long period of time, to behave. In fact they cease to properly describe physical behaviour when objects travel very fast, that is when the velocity is a significant proportion of the speed of light. Even a GP 500 machine comes nowhere near this and so for all practical purposes we must consider these rules as unbreakable laws.
The behaviour of most physical systems is controlled by a set of rules known as Newton’s three laws, which were published in 1687. These are not some arbitrary man-made laws which decree how things must behave, rather they are statements which formalize how things have been observed, over a long period of time, to behave. In fact they cease to properly describe physical behaviour when objects travel very fast, that is when the velocity is a significant proportion of the speed of light. Even a GP 500 machine comes nowhere near this and so for all practical purposes we must consider these rules as unbreakable laws.
The units of moment are newton.metres Nm. and is represented by the letter “M” although “T” is also used for torque.
The units of moment are newton.metres Nm. and is represented by the letter “M” although “T” is also used for torque.
In vehicle terms there is no "force" trying to make the vehicle move outwards from a turn. What appear: to be such a force is just its momentum wanting to continue as before, that is; continue in a straight line tangential to its instantaneous velocity. On the other hand we need a "real force" to push the machine around a bend, i.e. change the direction of its forward momentum. This force is _ suppliec by the tyres and is known as centripetal force. Centripetal means "inward force" but as we know ever force has to be balanced, in this case it is balanced by the changing momentum, however, it is ofter much more convenient to think in terms of an actual balancing "force", and so we have adopted the outwardly acting conceptual force called "centrifugal force". In other aspects of dynamics some people use (although its now largely regarded as an unnecessary technique) what's called the "d'Alembert force which is just a way of expressing the reaction to a force needed to accelerate an object "Centrifugal force" is just the "d'Alembert force" for the special case of a curved motion, it is the conceptual reaction force needed to balance the real Centripetal force.
In vehicle terms there is no "force" trying to make the vehicle move outwards from a turn. What appear: to be such a force is just its momentum wanting to continue as before, that is; continue in a straight line tangential to its instantaneous velocity. On the other hand we need a "real force" to push the machine around a bend, i.e. change the direction of its forward momentum. This force is _ suppliec by the tyres and is known as centripetal force. Centripetal means "inward force" but as we know ever force has to be balanced, in this case it is balanced by the changing momentum, however, it is ofter much more convenient to think in terms of an actual balancing "force", and so we have adopted the outwardly acting conceptual force called "centrifugal force". In other aspects of dynamics some people use (although its now largely regarded as an unnecessary technique) what's called the "d'Alembert force which is just a way of expressing the reaction to a force needed to accelerate an object "Centrifugal force" is just the "d'Alembert force" for the special case of a curved motion, it is the conceptual reaction force needed to balance the real Centripetal force.
Work, energy and power Work in a technical sense is identical to energy. Work involves movement and is the product of the force required to move an object and the distance that it moves. Fig. A5.8 shows a moving motorcycle, this has a total drag force of Fd, including rolling resistance etc., to keep it moving the engine must supply a final force at the tyre / road interface that is equal and opposite to Fd, call it Ft. To move over a distance of L the engine must do Ft.L amount of work. The amount of energy expended or work done does not depend on how long it takes. We have to talk about power when time is a factor.
Work, energy and power Work in a technical sense is identical to energy. Work involves movement and is the product of the force required to move an object and the distance that it moves. Fig. A5.8 shows a moving motorcycle, this has a total drag force of Fd, including rolling resistance etc., to keep it moving the engine must supply a final force at the tyre / road interface that is equal and opposite to Fd, call it Ft. To move over a distance of L the engine must do Ft.L amount of work. The amount of energy expended or work done does not depend on how long it takes. We have to talk about power when time is a factor.
Work, energy and power are scalar quantities, they have magnitude only but no direction. This might seem strange because both force and distance are vectors, but there are two methods of vector multiplication and one yields scalars and the other vectors. Imagine pushing a motorcycle on a level road, you will have done some work to cover say 100 metres, but if you turn around and push back to the start the work will not be returned, in fact you'll have to do as much work again to get back, therefore work is not affected by direction and so neither is power. Work and energy have units of newton.metres, Nm which are also called joules, J. Power has the units of joules/sec. which is expressed as Watts, W.
Work, energy and power are scalar quantities, they have magnitude only but no direction. This might seem strange because both force and distance are vectors, but there are two methods of vector multiplication and one yields scalars and the other vectors. Imagine pushing a motorcycle on a level road, you will have done some work to cover say 100 metres, but if you turn around and push back to the start the work will not be returned, in fact you'll have to do as much work again to get back, therefore work is not affected by direction and so neither is power. Work and energy have units of newton.metres, Nm which are also called joules, J. Power has the units of joules/sec. which is expressed as Watts, W.
Fig. A6.1 Motion of a 4 bar mechanism shown top left. Top right: isolates link A, if the pivot is moved anywhere along a co-linear line as shown for example at 1a or 1b, the instantaneous motion at 2 will be unchanged. The lower diagram shows how a virtual pivot is defined and also demonstrates clearly how it changes position as the mechanism is moved about its real pivots. Consider what’s called a four bar linkage as in fig. A6.1, top left.
Fig. A6.1 Motion of a 4 bar mechanism shown top left. Top right: isolates link A, if the pivot is moved anywhere along a co-linear line as shown for example at 1a or 1b, the instantaneous motion at 2 will be unchanged. The lower diagram shows how a virtual pivot is defined and also demonstrates clearly how it changes position as the mechanism is moved about its real pivots. Consider what’s called a four bar linkage as in fig. A6.1, top left.
This appendix is just a brief guide to calculating the CoG position of a motorcycle with and without rider. The CoG is an imaginary point through which we consider all the weight of the bike to act. That is to say, for many calculation purposes we can consider the machine to be composed of a point weight equal to the machine weight but all acting at the CoG. For all practical purposes this is the same as the mass centre and the terms are often used interchangeably, even though strictly speaking they have different meanings. Viz: Two wheels, engine / transmission and a rider. Other component parts are added in the same way.
This appendix is just a brief guide to calculating the CoG position of a motorcycle with and without rider. The CoG is an imaginary point through which we consider all the weight of the bike to act. That is to say, for many calculation purposes we can consider the machine to be composed of a point weight equal to the machine weight but all acting at the CoG. For all practical purposes this is the same as the mass centre and the terms are often used interchangeably, even though strictly speaking they have different meanings. Viz: Two wheels, engine / transmission and a rider. Other component parts are added in the same way.
Basically we must equate the moment about this point of the total weight acting at the CoG with the sum of all the moments of all the individual components. Hence: We add all the other components of the machine in like manner. Just add the weight of a part multipliec by is distance from the reference point. Note that we didn’t include the front wheel, that is because we choose the front wheel as our reference point. We calculate the vertical CoG location in the same wa‘ but just substitute the vertical distance of the individual parts. In this case we must add the front whee because our reference was at ground level not at axle height. In practice the easiest way to do this is tc setup a spreadsheet on a computer.
Basically we must equate the moment about this point of the total weight acting at the CoG with the sum of all the moments of all the individual components. Hence: We add all the other components of the machine in like manner. Just add the weight of a part multipliec by is distance from the reference point. Note that we didn’t include the front wheel, that is because we choose the front wheel as our reference point. We calculate the vertical CoG location in the same wa‘ but just substitute the vertical distance of the individual parts. In this case we must add the front whee because our reference was at ground level not at axle height. In practice the easiest way to do this is tc setup a spreadsheet on a computer.
Measuring the vertical CoG position on an existing bike can be done in a variety ways but is more difficult than obtaining the horizontal value, especially with the rider on board. Probably the easiest way is to lift the rear wheel onto a raised block and then weigh each end again. Depending on the height of the block there will be a reduction in rear wheel load. In this case we calculate the CoG height according to:
Measuring the vertical CoG position on an existing bike can be done in a variety ways but is more difficult than obtaining the horizontal value, especially with the rider on board. Probably the easiest way is to lift the rear wheel onto a raised block and then weigh each end again. Depending on the height of the block there will be a reduction in rear wheel load. In this case we calculate the CoG height according to:
Fig A7.3 Articulated rider model made from helmet visor. Fig A7.4 Typical mass distribution of rider. To assist with the estimation of the CoG of the rider | find it useful to use an articulated model of a rider in combination with some averaged values of the distribution of masses between the limbs. The following figures show such a model made from sheet plastic (actually an old helmet visor), and typical mass distribution. Remember that humans vary in shape and so these figures are only a guide.
Fig A7.3 Articulated rider model made from helmet visor. Fig A7.4 Typical mass distribution of rider. To assist with the estimation of the CoG of the rider | find it useful to use an articulated model of a rider in combination with some averaged values of the distribution of masses between the limbs. The following figures show such a model made from sheet plastic (actually an old helmet visor), and typical mass distribution. Remember that humans vary in shape and so these figures are only a guide.
Wet weights of a range of motorcycles. Note that a large number of examples range between 180 and 250 kgf. The range over 250 kgf. is inhabited by large touring machines and cruisers, and below 180 kgf. is generally the preserve of the smaller capacity bikes. The perception of weight is affected by the ease of stationary balance and this depends on CoG height as well. The “perceived weight index” chart, at the end, probably gives a better idea of this. This appendix lists various data from a range of typical motorcycles. Most are models from 1998 onwards, although a small number date back to the 1970s, and are generally of 650 cc. or greater capacity but with a small number of lower capacity machines. The data has been compiled from various sources (the majority from the American “Motorcycle Consumer News”). Some data is as supplied by the manufacturers who like to cheat where possible, for example the wheelbase of chain driven machines can vary by the amount of their chain adjustment, and some quote the most forward position in order to enhance a sporty image. The CoG height and dependent data is probably subject to greater measurement error than the other parameters. Nevertheless, the following is an interesting collection of parameters as a guide to typical values. The data is presented in a series of ordered bar and cluster charts to provide an easy visual appreciation of the data ranges.
Wet weights of a range of motorcycles. Note that a large number of examples range between 180 and 250 kgf. The range over 250 kgf. is inhabited by large touring machines and cruisers, and below 180 kgf. is generally the preserve of the smaller capacity bikes. The perception of weight is affected by the ease of stationary balance and this depends on CoG height as well. The “perceived weight index” chart, at the end, probably gives a better idea of this. This appendix lists various data from a range of typical motorcycles. Most are models from 1998 onwards, although a small number date back to the 1970s, and are generally of 650 cc. or greater capacity but with a small number of lower capacity machines. The data has been compiled from various sources (the majority from the American “Motorcycle Consumer News”). Some data is as supplied by the manufacturers who like to cheat where possible, for example the wheelbase of chain driven machines can vary by the amount of their chain adjustment, and some quote the most forward position in order to enhance a sporty image. The CoG height and dependent data is probably subject to greater measurement error than the other parameters. Nevertheless, the following is an interesting collection of parameters as a guide to typical values. The data is presented in a series of ordered bar and cluster charts to provide an easy visual appreciation of the data ranges.
The light coloured bars indicate the ground trail and the dark bars show the real trail. The majority of values are in the range or 90 — 110 mm. for ground trail and 80 -- 100 mm. real trail. Generally it is the cruiser machines that have much larger trail values, these machines also tend to have greater rake angles for styling reasons.
The light coloured bars indicate the ground trail and the dark bars show the real trail. The majority of values are in the range or 90 — 110 mm. for ground trail and 80 -- 100 mm. real trail. Generally it is the cruiser machines that have much larger trail values, these machines also tend to have greater rake angles for styling reasons.
Cluster chart of trail against rake angle, the circular symbols are ground trail and the others are real trail. Trend lines have been added, the solid one is for real trail. Note that both real trail and ground trail tend to increase with rake angle.
Cluster chart of trail against rake angle, the circular symbols are ground trail and the others are real trail. Trend lines have been added, the solid one is for real trail. Note that both real trail and ground trail tend to increase with rake angle.
Using the same symbols as the previous chart, this shows the trail compared to weight. Although there is more data scatter the trend lines indicate that in practice more trail is generally used on heavier bikes.
Using the same symbols as the previous chart, this shows the trail compared to weight. Although there is more data scatter the trend lines indicate that in practice more trail is generally used on heavier bikes.
Trail against wheelbase, as with the trail relationship to weight there is a degree of scatter but the trend lines show a general increase in trail with greater wheelbase.
Trail against wheelbase, as with the trail relationship to weight there is a degree of scatter but the trend lines show a general increase in trail with greater wheelbase.
The rake angles of various motorcycles. The vast majority fit into the range of 23-27 degrees, with modern sports bikes at the lower values. Generally heavy-weight tourers and cruisers go for angles above about 29 degrees.
The rake angles of various motorcycles. The vast majority fit into the range of 23-27 degrees, with modern sports bikes at the lower values. Generally heavy-weight tourers and cruisers go for angles above about 29 degrees.
Modern sports bikes have short wheelbases starting from about 1380 mm. to help turn quickly. Average general machines tend to be around 1450 — 1500 mm., and some large tourers go a bit over 1700 mm.
Modern sports bikes have short wheelbases starting from about 1380 mm. to help turn quickly. Average general machines tend to be around 1450 — 1500 mm., and some large tourers go a bit over 1700 mm.
Weight distribution - % on front This shows the percentage of the weight carried by the front wheel. As with all the characteristics show here, this is of the bike alone without the rider aboard. Mostly 40-50% are the most common values. Recently there has been a trend to increase this value, especially on high performance sports machines for stability and steering response. Cruisers have a more rearward weight bias and hence give a lower percentage on the front.
Weight distribution - % on front This shows the percentage of the weight carried by the front wheel. As with all the characteristics show here, this is of the bike alone without the rider aboard. Mostly 40-50% are the most common values. Recently there has been a trend to increase this value, especially on high performance sports machines for stability and steering response. Cruisers have a more rearward weight bias and hence give a lower percentage on the front.
CoG height for various machines. There is a reasonably even distribution between 400 — 600 mm.
CoG height for various machines. There is a reasonably even distribution between 400 — 600 mm.
Ratio of CoG height to wheelbase This uses the same sample of bikes as in the previous chart, but relates the ratio of the CoG height to the wheelbase. This parameter varies by a two to one ratio over the set of sample bikes. The CoG height will be increased significantly with the rider on-board. This parameter is important because it affects load transfer under braking and acceleration.
Ratio of CoG height to wheelbase This uses the same sample of bikes as in the previous chart, but relates the ratio of the CoG height to the wheelbase. This parameter varies by a two to one ratio over the set of sample bikes. The CoG height will be increased significantly with the rider on-board. This parameter is important because it affects load transfer under braking and acceleration.
CoG height - mm. Cluster chart showing the relationship between CoG height and wheelbase. As we might expect the longer bikes tend to have lower CoG heights, a trend that will continue with a rider seated. Longer bikes tend to have lower seat heights.
CoG height - mm. Cluster chart showing the relationship between CoG height and wheelbase. As we might expect the longer bikes tend to have lower CoG heights, a trend that will continue with a rider seated. Longer bikes tend to have lower seat heights.
60 mph to zero (96.6 km/h) stopping distances. These have been measured from a computer controlled radar gun and the braking was started at 65 -— 70 mph to avoid reaction times etc. from affecting the results. Aerodynamic drag will have a minimal effect on stopping distance at these speeds.
60 mph to zero (96.6 km/h) stopping distances. These have been measured from a computer controlled radar gun and the braking was started at 65 -— 70 mph to avoid reaction times etc. from affecting the results. Aerodynamic drag will have a minimal effect on stopping distance at these speeds.
Stopping distance from 60 mph. - metres Stopping distances against the ratio of CoG height to wheelbase. This chart would be more meaningful if the CoG height included the rider but that information was not available. There is a lot of scatter but the trend seems to be that the bikes with a higher CoG/wheelbase ratio tend to stop in a shorter distance. These bikes are more likely to be sports machines. (Data from Motorcycle Consumer News)
Stopping distance from 60 mph. - metres Stopping distances against the ratio of CoG height to wheelbase. This chart would be more meaningful if the CoG height included the rider but that information was not available. There is a lot of scatter but the trend seems to be that the bikes with a higher CoG/wheelbase ratio tend to stop in a shorter distance. These bikes are more likely to be sports machines. (Data from Motorcycle Consumer News)
The perceived weight of a bike depends very much on the CoG height as well as the true weight itself, and inertia effects vary as the square of the CoG height. The author’s “Perceived weight index” combines these factors into a guide number to give a rough idea of how the bike will feel at rest, balanced by the rider. This is only a rough guide because seat height and the angle at which the rider’s leg meets with the ground also have an effect.
The perceived weight of a bike depends very much on the CoG height as well as the true weight itself, and inertia effects vary as the square of the CoG height. The author’s “Perceived weight index” combines these factors into a guide number to give a rough idea of how the bike will feel at rest, balanced by the rider. This is only a rough guide because seat height and the angle at which the rider’s leg meets with the ground also have an effect.

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