## Induction System Volumetric Efficiency

here are two real world effects that determine how much fuel/air charge can get into the cylinder. The first is that air is compressible, the second is the dynamics (acceleration/deceleration) of the air. The compressibility of the air becomes a factor when the air enters the intake port around the intake valve. The intake port/valve forms a constriction, like the throat of a nozzle. Because air is compressible, it can only be pushed through a constriction so fast. Regardless of how much pressure you apply, the maximum velocity possible through the throat of a nozzle is a velocity equal to the speed of sound .

The same effect happens at the intake valve. The ratio of the typical velocity to the intake sonic velocity is called the inlet Mach index. From the science of fluid mechanics the controlling velocity in a compressible flow system is usually the intake valve opening. For a given cylinder and valve design, the inlet Mach index is proportional to the piston speed, and that the fuel/air charge flows in faster when the piston moves down faster. Of course, at some point the constriction of the valve opening starts to limit this. When the inlet Mach index exceeds 0.5 (intake velocity equal to half the speed of sound), the volumetric efficiency falls rapidly with increasing speed. Therefore, enginest are typically designed so that the inlet Mach index does not exceed 0.5 at the highest rated speed.

The effect of this constriction shows up as a pressure drop through the intake valve. Why don’t we just open the intake valve further? Because when the valve is lifted a distance equal to 1/4 of its diameter, the area of a cylinder around the valve (that the fuel/air charge passes through, not the engine cylinder) is equal to the area of the valve face and intake port, ignoring the valve stem. Mathematically, the area of the cylinder is (2 r)(d/4). Since d = 2r, this evaluates to r2, which is the area of the intake port, the amount of additional flow through the intake port increases very slowly as the lift of the valve increases beyond 1/4 of the valve diameter.

Because of the dynamics of the fuel/air charge, the intake valve normally closes at some time after the piston passes bottom dead center. As the piston moves down, it draws the fuel/air charge into the cylinder. This movement builds up momentum in the intake manifold. When the piston reaches bottom dead center, the fuel/air charge is still flowing into the cylinder as a result of this residual momentum. Thus, at the speed desired for maximum torque, the intake valve closing is timed to correspond with the velocity of the fuel/air charge through the intake port dropping to zero. This closing will occur at some time after the piston has started the compression stroke, and will result in the maximum amount of fuel/air charge being drawn into the cylinder. This maximizes the volumetric efficiency, and maximizes the torque delivered to the crankshaft, ignoring friction effects. The angle of the crankshaft at the time the intake valve closes is called the intake valve closing angle.

So what effects does this later valve closing have at other speeds? At low speeds, the momentum built up in the intake manifold will be small, such that part of the fuel/air charge will be pushed back into the intake manifold as the piston starts up prior to the intake valve closing. At speeds above the speed for maximum torque, the constriction of intake valve opening will cause a pressure loss which will reduce the amount of fuel/air charge entering the cylinder. In either case, the amount of fuel/air charge in the cylinder is reduced, and thus the torque is reduced.

The design of the intake manifold also affects the amount of momentum built up in the flow of the fuel/air charge. The momentum of the fuel/air charge is the sum of the effect of standing waves built up from previous intake strokes keep in mind that any tube will have a resonant frequency and effect the transient wave caused by the current intake stroke. While the standing waves contribute to the overall effect, there are no sudden changes in the volumetric efficiency when the RPM of the engine is an even multiple of the natural frequency of the intake manifold.

Long, skinny intake manifold pipes give high volumetric efficiencies at low piston speeds because high momentum lots of velocity is built up in the pipe during the intake stroke. At high piston speeds, the small diameter of the intake pipe causes a constriction and the volumetric efficiency falls. Fat intake pipes show a maximum volumetric efficiency at intermediate piston speeds. However, at high piston speeds, the larger mass of the fuel/air charge in the fat intake pipe is slow to accelerate, and thus the volumetric efficiency falls off.

As the manifold pipes get shorter, the maximum gain in volumetric efficiency over having no intake manifold at all decreases. However, the gain you do get with shorter intake pipe happens over a greater range of piston speeds. Basically, it comes down to the intake manifold pipe should be designed according to the engine requirements. If you need high torque at slow piston speeds, use long skinny intake pipes. For high torque at intermediate piston speeds, use long fat intake pipes. For high torque over a wide range of piston speeds (i.e. a flat torque curve), use shorter intake pipes.

## Ignition components

Inductive Discharge Coils – Ignition spark for motorcycle is accomplished by the iginiton coil, coils have 2 sets of windings, a primary and a secondary. A typical coil will have around 250 turns of wire on the primary and about 25,000 on the secondary for a ratio of 100 to 1. The secondary section often uses an iron core to increase its inductance. Coil resistance on the primary will be from .3 to .5 ohms usually and on the secondary, between 5000 and 12,000 ohms. The inductance and resistance of the coil will determine how quickly a coil can be charged and discharged.

A transistor is used to switch the current flow off and on in the primary coil. When the transistor is switched on, current rapidly builds from 0 to a maximum value determined by the coil inductance and resistance. This current flow induces a magnetic field within the primary. When the current is turned off, this magnetic field collapses which cuts the windings of the secondary coil and induces a high voltage surge.

Output voltage is determined by the rate of field collapse and the windings ratio between primary and secondary. Because the path to ground for the current involves the spark gap, initial resistance is extremely high. This allows the voltage to build to a high value until it gets high enough to jump the plug gap. The difference must be high enough to first ionize the gas between the electrodes. The ionized gas creates a conductive path for the current to flow, at this point the arc jumps and current flow is established.

If only 10,000 volts are required to jump a plug gap under a given condition, that will be the maximum delivered. It is also important to note that the spark duration is determined by coil inductance and total resistance of the circuit, plus spark plug gap. Most inductive discharge systems have a spark duration of between 1 and 2 milliseconds.

When cylinder pressure increases, the voltage required to jump the plug gap increases. The second problem on high performance engines with high rev limits, is that there is less time to charge the coil with increasing rpm, high rpm and high output puts greater demands on the ignition system.

Coil Charge Time and Saturation – The time it takes to charge the coil or bring the current to maximum in the primary windings is called charge time. Input voltage and coil resistance are the main parameters relating to charge time, when the current has reached its maximum value in the primary, it is said to be fully saturated.

If current is applied longer than the time needed to fully saturate the primary, energy is wasted and there is nothing to be gained. If the current is cut off before saturation is achieved, the maximum spark energy available will be reduced.

Coils require charge times of between 2.1 and 6 milliseconds. Obviously, a coil requiring 6 milliseconds to saturate would be unsuitable on a high revving engine as there is not 6 milliseconds available to charge it between discharges at high rpm. For this reason, most racing coils have low primary resistances between .3 and .7 ohms and are fully saturated in less than 3 milliseconds permiting full coil output at very high rpms.

Capacitive Discharge Ignition – On very high output engines, an inductive discharge coil is inadequate to supply spark at high rpm and high cylinder pressures. A CD ignition or CDI is used to reduce charge times. The MSD system is very popular worldwide.

In normal inductive discharge coils, only 12-14 volts is available from the battery to charge the primary. The CDI charges capacitors to store a high voltage kick to fire to the primary side, putting between 30 and 500 volts onto the primary windings which reduces the charge time substantially. A coil that would take 3 milliseconds to become fully saturated with 12 volts is now fully saturated in less than 1 with a CDI. The same engine now will be able to turn twice the RPM and experience a major increase in cylinder pressure before encountering misfire.

Some CDIs also include a multispark function where more than 1 spark is generated after the first spark. This improves ignition probability besides the high rpm coil saturation advantages and a greater resistance to plug fouling.

Ignition Wires (Spark Leads) – The purpose of the ignition wires is to conduct the maximum coil output energy to the spark plugs with a minimum amount of radiated electromagnetic interference (EMI) and radio frequency interference (RFI). There are 3 basic types of conductors used in racing applications: carbon string, solid and spiral wound. Most production engines come equipped with carbon string. The solid core types are used exclusively for racing, mainly with carbureted or non-computer controlled engines because they offer no EMI or RFI suppression. They generally have a low resistance stainless steel conductor. These types are rapidly losing favor, even in racing circles.

The carbon string type is the most common and work just fine in racing applications. The conductor is usually a carbon impregnated fiberglass multistrand. Suppression qualities are fine with resistances in the 5K to 10K ohms per foot. They are cheap and reliable for 2 to 5 years usually, then they may start to break down and should be replaced. High voltage racing ignitions will likely hasten their demise. Dynatech makes low priced wire set which works well in performance applications.

Ignition Wires (Spark Leads) – The spiral wound type is probably the best type for any application. The better brands offer excellent suppression, relatively low resistance and don’t really wear out. Construction quality and choice of material vary widely between brands. NGK makes low priced wire sets which work well in performance applications.

Some amount of resistance is required along with proper construction to achieve high suppression levels. Resistance is also important to avoid damaging some types of coils and amplifiers due to flyback and coil harmonics. Beware of wires claiming to have very low resistance. These CANNOT have good suppression qualities.

Beware of any ignition wires claiming to increase hp. Ignition wires CANNOT increase hp. As long as the wires that you have are allowing the spark to jump the gap properly, installing a set of \$200 wires is strictly a waste of money.

Lately, some truly “magic” wires have come onto the market claiming to not only increase power but also to shorten the spark duration from milliseconds to nanoseconds. As previously mentioned, spark duration is determined primarily by coil inductance and coil resistance so these wires CANNOT shorten the spark duration by the amount claimed. The wire resistance has a minimal effect on discharge time because of the high voltage involved. A very short duration spark is in fact detrimental to ignition because of lower probability.

These same wires claim to increase flame front propagation rates and the ability to ignite over-rich fuel mixtures for more power. Once ignited, the mixture undergoes a flagregation process and that the progression rate of the flame front is totally independent of the spark. As previously mentioned most gasolines will not ignite nor burn at air fuel ratios richer than 10 to 1, period, and that maximum power is actually achieved at around 12 / 13 to 1 AFR so the claim also has no basis in fact.

These wires use a braided metal shield over the main conductor which is grounded to the chassis. This arrangement offers poor suppression because it does not cover the entire conductor. Any energy leaking out of the main conductor by induction is actually wasted to ground and will not make it to the spark plug. These wires also have very low resistance which as mentioned can have a detrimental effect on coils and ignition amplifiers due to severe flyback effects which are normally damped by circuit resistance.

Other claims for these wires include current flows of up to 1000 amps. The current flow in the ignition circuit is determined by the coil construction and drive circuits, not by the ignition wires. Most ignition systems are current limited to between 5 and 15 amps. The most powerful race systems rarely exceed 30 amps. To flow current at 1000 amps, you would require #0 welding cable for the ignition system!

Spark Plugs – The final part in the ignition system is the spark plug itself. The average plug consists of steel shell which threads into the cylinder head, a ceramic insulator, an iron or copper core leading to a nickel or platinum center electrode and a ground electrode of similar material. The spark jumps between the center and ground electrode. Certain special application plugs may have multiple ground electrodes. Different heat ranges are available depending on application. For constant high power applications, a colder than stock plug is usually selected to keep internal temperatures within limits.

Again, many “trick” plugs come onto the market from time to time expounding the virtues of their incredible new design, usually offering more hp of course. Split electrode plugs are a waste of money because the spark will only jump to one of the electrodes at a time in any case.

You will find that most reputable engine builders for racing use standard NGK, Bosch or Champion plugs with a standard electrode setup. A properly selected, standard plug will easily last 25,000 miles of hard use in most engines. A platinum tipped plug will easily last twice as long on most engines. There is no rocket science here, modern spark plugs coupled to modern ignition systems in a modern engine are extremely cheap and reliable, even on race engines, a \$2, off the shelf, NGK plug will work just fine.

## Ignition / combustion criteria

Some people think that when a spark plug fires, the fuel/air mixture explodes instantaneously, driving the piston down. If this really happened, engines would last only a few minutes before they literally explode.

Looking at the dynamics involved from the moment that the intake valve is fully open. With the piston moving down the bore, cylinder volume increases, cylinder pressure decreases, allowing the higher pressure in the intake tract to push the fuel/air mixture into the cylinder. As the piston starts back up and the intake valve closes, cylinder volume decreases and cylinder pressure increases.

When the crankshaft reaches about 30 degrees before top dead center (TDC), the spark jumps the gap between the spark plug electrode. The purpose of the spark is to raise the temperature of a very small portion of the fuel/air mixture above its ignition temperature. This is the point where true combustion begins. As the reaction starts, the mixture directly adjacent to the spark plug is also ignited and the process progresses out from the spark plug in a roughly spherical shape.

At about 20 degrees before top dead center (BTDC), the rate of heat release causes the cylinder pressure to rise above the compression line which is what the cylinder pressure would be at a given piston position without ignition. Notice that it has taken 10 degrees of crank rotation to generate this pressure level. This is known as the ignition-delay period.

The rate of pressure rise is a function of the rate of energy release vs. the rate of change of combustion space. The rate of energy release is directly related to the flame propagation rate and the area of reacting surface. The flame speed is dependant on fuel/air ratio, charge density, charge homogeny, fuel characteristics, charge turbulence and reaction with inert gasses and the combustion chamber, cylinder walls and piston.

No two combustion cycles progress at the same rate or uniform rate. Some start slow and end slow, some start slow and end fast, some start fast and slow down. Generally, only the ones that end too fast will lead to detonation / knocking / pinging as the rapid pressure rise may happen too soon with the cylinder volume still decreasing or not increasing fast enough. Usually, not all cylinders will detonate / knock / ping at the same time or on the same cycle because of this.

By the time the crank is at about 5 degrees after top dead center (ATDC), the cylinder pressure is about double that of the compression line. From this point to roughly 15 degrees after top dead center (ATDC) the combustion process is fast due to the increasing area of inflamed mixture and the high rate of energy release. The peak cylinder pressure (PCP) occurs between 10 and 20 degrees after top dead center (ATDC) on most engines and the combustion process is complete by 20 to 25 degrees after top dead center (ATDC). The peak temperature within the combustion gasses will reach somewhere around 5000 degrees Fahrenheit and pressures may be anywhere from 300 to 2500psi depending on the engine.

Obviously it is very important to have the crankpin at an advantageous angle before maximum cylinder pressure is achieved in order that maximum force is applied through the piston and rod to the crankshaft. If the mixture was ignited too early, much of the force would simply try to compress the piston, rod and crank without performing any useful work. In a worst case scenario, the cylinder pressure would be rapidly rising before the piston reached top dead center (TDC) which would have the cylinder volume decreasing at the same time. This will often result in detonation/knock/ ping which is counterproductive to maximum power and engine life.

Detonation, knock or pinging is defined as a form of combustion which involves too rapid a rate of energy release producing excessive temperatures and pressures, adversely affecting the conversion of chemical energy into useful work. Detonation usually involves ignition and literal explosion of the end gases, these are the gases not in contact with the initial spark or the progressing flame front.

If peak cylinder pressure (PCP) is achieved too late, again, less work would be performed. Most of the useful work is done in the first 100 degrees of crank rotation. Most combustion must be done with the piston in close proximity to the chamber so that the minimum amount of heat (energy) is lost and the maximum amount of energy is delivered to the crankshaft.

## Frankenstein’s guide to oil cooled engines

Before anything, I would like to have it said that I wrote this in my best knowledge and do
not want to be held responsible for any mistakes. I’m confident about what I’ve seen and done,
but since I’m not the only one messing around with gixxers, I can hardly ever be sure that
the engine I find in a 89 1100R is really an 89 1100R. I’ve left the types before 88 out,
since I have not much experience with them.

Frankenstein@robbynitroz.nl

There are mainly 2 types of 750’s, the 88-89 short stroke, and the pre-88 and 90-91 long stroke.
(The 750F is basicly the same motor as the 88-89 short stroke, the B6 and GSXF600 are basicly
the same as 90 long stroke with a smaller bore).
1100R motors from 88-92 are similar to the 1100F and B12 motors. The 1100G is also similar,
but has an axle drive. They all have the same stroke, and only the B12 has a 1mm bigger bore.

Apart from the color, all the GSXR, GSXF, GSXG and Bandit ignition covers are the same (except
the 750RK).

The clutch covers are depending on the clutch operation, there are 3 possibilities:
(The dry clutch is left out, to avoid making it more confusing).
1.The GSXF600, GSXF750 and B6 have the clutch cable connected to a mechanism on the sprocket
cover, and the clutch is operated by a push pin through the primary gear box shaft.
2.The 750R has the clutch mechanism in the clutch cover (on the right side). The 88-89 clutch
cover is recognizable by a smooth clutch cover, the 90-91 has a bubble in the center. They are
very similar, but since the engines have a different clutch, I don’t think these covers can be
swapped, I haven’t tried though.
3.The 1127’s and B12 all have the clutch mechanism on the sprocket cover, like the 600’s and
the 750F, but then hydraulically operated. The mechanisms on the sprocket cover can all be
swapped, so it’s possible to put a cable operation from a B6, F6 or F750 on an 1127 (and v.v.),
although it might need some adjustment of the length of the pushrod.
This also means that, since the clutch covers on the 600’s, 750F and 1127’s are nothing but
covers, they can be swapped.

The startermotor covers from the 1127’s are all the same (The startermotor covers from the 1052
engines are not the same) The 1127 covers can be recognized by a kind of bubble, to accomodate
the bigger starter motor. The 600′ and 750’s have a smaller starter motor, and the top line of
these covers is straight. (I believe the 1052 motors also have this smaller starter motor and
cover). Covers can be swapped among the 600’s and 750’s, but an 1127 cover only fits an 1127.

The oil pan on all 1127’s are the same, but the B12 is different. The 750F and 750R 88-89 have
the same oil pan as the 1127’s. The 91-750R and B6 have a similar or same oil pan as the B12,
I’m not sure. However, it is possible to swap these oil pans, as long is you change the oil
pickup as well. Oil hoses on the 1127 pans connect at the front, the others at the bottom.

The valve covers are different depending on the cam chain type, and the cylinder head size.
The B6 cover only fits the B6, the B12 cover only fits the B12. The 750R-90 and 91 covers
are the same. All the 1127 and the 750R-88/89 cam covers are the same.

There a 3 main items which make the difference in crankshafts.
1. Stroke
2. Clutch gear
3. Camchain type

1. The 1127’s and B12 all have the same stroke. The 600’s and 90-91 750R’s have the same
stroke. The 88-89 750’s and the 750F have the same stroke.
The stroke is important because this directly reflects on the number on teeth on the
clutch gear (ie. the gear diameter).
2. All GSXR1127 crankshafts are the same. The GSXF and G have a helical
cut gear, so when using a GSXF1127 crank You will have to use a GSXF1127 clutch basket as well.
3. All GSXR’s (both 750 and 1127) have the same type camchain, but the B6 and B12 are
different. Since the cam chain is driven from the crankshaft, this means these crankshafts
are not interchangeable with GSXR crankshafts, unless you also change the cam chain, tensioner,
guides, cam sprockets, cam covers, cam guiding between cam shafts.

All the 3 items above have to match. Swapping a crankshaft with a type that has the same
stroke, clutch gear and cam chain is no problem. If you start mixing, you have to match
clutch to the crankshaft (and in some cases gearbox), or cam chain stuff to the crankshaft.

Connecting rods from B6, 750R-90 and 750R-91 can be swapped. 1127 rods are all the same.
I have used B12 rods in 1127’s; I found there was a minor weight difference, but they could
easily be matched. This difference might have been incidental.

I left out the dry clutches on purpose, since I have no experience with them.

The GSXR1127 89-on and B12 have a diaphragm spring, the GSXF/G have normal springs.
The GSXR and B12 have a straight cut gear, the GSXF/G have a helical cut gear.
Because of the different gear on the clutch basket, the clutch basket is not swappable.
Since the types with a diaphragm spring have a longer shaft to accommodate the bolt for the
central spring, these parts are also not swappable. It is possible to use the internal clutch
parts from a ‘normal spring type’ in the basket (or actually on the gear box shaft) from a
‘diaphragm spring type’, but you need to fill the space on the longer shaft. It is not
possible to use the diaphragm style clutch on a GSXF gear box shaft, since the shaft is to short.

The 88-89 750R have a large (actually the largest) diameter but relatively flat clutch.
Although the gear box shaft is the same, the 88-89 clutch can not be swapped with the 91
clutch because the crankshaft diameter (and consequently tooth count) is different.
Although the B6 clutch is the same diameter as the 91 750R clutch (since they have the
same stroke), there is not a lot to swap there since the plates are different and the
gear box shaft are differently machined.

Cylinders block with pistons from 1127’s can all be swapped. B12 block+pistons fit the 1127
as well, or only pistons+have your 1127 block bored.
88-89 750’s is same as GSXF750.
B6 and 90/91 750R have 18mm wrist pins, whereas 88-89 750R, GSXF750, 1127’s and B12 have
20mm pins.
Since the B6 and later 750R 90-91 have the same stroke, cylinder block dimension, and wrist pin
diameter, the 90/91 block+pistons can be swapped with the B6 stuff (although you’ll have to
check that the pistons don’t hit the head/valves).

The long stroke engines (ie. B6, GSXF600, 90/91 750R) have the same dimensions, just the
combustion chamber and valves in the 750’s is bigger. So somebody who want less power could
fit a B6 top on a 90 750R. Camshaft type on the B6, GSXF600 and 90 750R is forked rocker,
meaning 1 cam for each pair of valves. 91 750R has shim type with 1 cam for each valve.
If swapping the camshafts as well, the 90 and 91 heads can be interchanged.
Both the 90 and 91 750R top ends can be used on a B6, but since the B6 has another type of
camchain, it is needed to maintain the B6 cam chain tensioner, guides, cam sprockets, valve
cover etc.

The 750 short stroke engines 88/89 heads have the same outside dimensions as the 1127/B12,
but the combustion chamber is smaller (although the valves are the same diameter).
The 1127R-91/92 has the same style head as the 750R-91, but
not much to swap; 1100 valve spacing differs (so camshafts can not be swapped), 1100 valves
are bigger, outside head dimensions differ.
As mentioned, 750R-88/89 valves are the same as 1127/B12, exception are the 1127R-91/92 valves.
These heads have shim type adjustment, and therefore different cams and longer valves.

It is possible to modify a 1127 shim head to a forked rocker head. It’s quite some work, and
you’ll need the valves from the forked rocker head, the rockers, cam shafts. You’ll need to
make all the spacers yourself, or in fact I believe there is a company that has or used to
have a modification kit.

Cam shafts from the 1127F, 750F, 90-750R, B6, B12, 88/89 750R are theoretically all swappable,
but of course the profiles are different. The long stroke 750’s have a different tooth count
on the cam sprockets so they can not be mixed. B12 sprockets can only be used in the B12.
B6 sprockets can only be used in the B6. 1127F and 1127R sprockets are the same, 88/89-750R
sprockets are similar, but the timing marks are different. (Meaning they can only be used if
slotted and timed)

1127: Depending on the clutch type there are long and short shafts. Also the gears themsleves
from these boxes are different. It may be possible to swap a few gears between these boxes,
but the gearchanges might not be very smooth.
Apart from the clutch type, the 91-92 1127R has a double row bearing on the output shaft, and
therefore a slightly different crankcase (around the bearing area).

Gear boxes from all 750’s are swapable. I have no experience with swapping gears seperately.

The B6 has a different shaft, so it can only be used with it’s own clutch.

Although it might seem there are so many differences, a lot can be mixed, as long as the right
parts are choosen, a few examples.
(There are some basic guidelines to assemble an engine, like check compression, cam timing, valve
clearance etc., no matter what combo you’re making).

1. A 1052 crank fits in 88-89 750R and 750F cases, but a 1127 crank doesn’t (but the cases can
be modified to take the 1127 crank as well)

2. A 750R-90 or 750R-91 top end on a B6.
It’s actually very easy, and I think all the info you need is above. Both engines have the same
stroke, same wrist pin diameter. Theoretically, it would be possible to put only 750 cylinders
and pistons on a B6. However, the pistons are designed to fit the 750 head and since that also
fits, why not install a 750 head as well (with bigger valves). Since the B6 has another cam chain
the B6 cam chain tensioner, cam sprockets, cam chain guides and B6 valve cover need to be
used. Then there are 2 options: either go for a 750R-90 top end, which uses forked rockers
like the B6 does (so it’s possible to use either the B6 cams or the 750R cams), or go for a
750R-91 top end, which uses another type of rockers so it is not possible to keep the B6 camshafts.

3. A 750R-88/89 top end on a 750R 90/91 bottom end (or 86-87 bottom).
This is a bit more difficult, since it needs some more work and imagination then the plain
assembling of a B6/750.
The 750R-88/89 have a bigger bore, so the idea of this combo is to increase the capacity of the
engine. (You could also take this combo the ‘other way around’, and fit a 90/91 crankshaft + clutch
in a 88/89 engine.)
Since the dimensions of the heads are not the same, it is not possible to only put the 88/89 pistons
+cylinders on the 90/91; the head of the 90/91 would not fit the cylinder block. So the complete
88/89 top has to be installed on the 90/91. The wrist pins on the 90/91 are 18mm, on the 88/89 20mm,
so the small end of the 90/91 rods have to be bored to 20mm. Now the whole thing could mechanically be
assembled, but since the stroke of the 88/89 is smaller, the height of the cylinder block is smaller.
This has to be compensated by putting a spacer under the cylinderblock. (This spacer would very
roughly have to be 1/2 x the difference in stroke, but the only right way is to measure/calculate the
compression.

4. 750R 6 box in a 1127 motor
The only hard thing here is to have a hole drilled through the gear box shaft, for the pushrod.
The 750 6 boxes have a single row bearing on the output shaft, and the clutch does
not have a diaphragm spring. So the easiest 1127 engines to put a 6 box in are the ones with a
single row bearing on the output shaft, and no diaphragm clutch, ie. only the GSXF1127 engines.
In these engines the 6 box drops straight in, only the shaft has to be drilled.
Second easy would be an 1127R engine with a diaphragm clutch, but no double row bearing (88-90).
In this case the box would still drop in, but for the clutch one would have to use the inner
clutch parts from a GSXF1127 (with normal springs) and the outer clutch basket from the 1127R
(with a straight cut gear, not helical).
Most work is in a 91/92 1127R where one would have to match the clutch as above + find a
solution for the double row bearing (the solution is actually to turn the double row bearing
inside out, and make a little hole for the small pin).
Of course the shift drum and forks from the 6 box have to be used as well, but they drop in
any 1127 without problems.

5. 88-89 750R head or 750F head on a 1127 or B12
These heads fit as they are, and give higher compression, better ports, larger squish.
In the case of the 91-92 1127R you’ll need to use the 750 camshafts as well, since the
91-92 1127R uses shim type camshafts and the 750 head is forked rocker type. If you use the
91-92 1127R cam shaft sprockets they can be timed as in the manual.
In the B12’s case you could use the B12 or the 750 cams (although they have different profiles)
but will have to use the B12 cam shaft sprockets because of the different cam chain.
In the case of the 88-90 1127R you can use the 750 or 1127 cams, and use the 1127 cam sprockets
(timing ‘by the book’) or use 750 cam sprockets (timing to be done by yourself)

## Setting up CV carbs for a turbo

See below all I know about blowthrough CV carbs. This is worth several hundred hours of sorting out and at least €1000,- worth of dyno time, so feel free to thank me! (and/or make a paypal donation)

Bowl pressure
If pressure in venturi of carbs is raising due to boost, so should the pressure in bowl raise. Because if pressure in bowls is lower as it is in venturi, no fuel can be taken into the engine.
In my experience dynamic boost is absolutely nessecary, and it’s best to give each connection on the carbs it’s own spot on the tube going to the plenum: not in the plenum itself as there is to much pulsing. Make sure they are “angelcut” and in the middle of the airstream. (also known as pitot tubes)

On the GSXR1100 model 92 with the 40mm carbs you need to fasten the rubber T’s for instance with steel wire to prevent leakage and pressure drop.

Membrane pressure (CV carbs only)
Pressure above slidemembrane is not needed: it got it’s signal/pressure through the hole in the slide. Don’t enlarge the holes because the slides go up too fast and cause stuttering. (beware of dynojet kit modifications: larger holes and softer springs are a real pain in the butt for the midrange!)

Pressure below the membrane is needed. On the GSXR1100 model 92 with the 40mm carbs you have a seperate “venting” system with external hoses. Those are not suitable to pressurize. I removed that system, plugged the holes and drilled holes from the bellmouth towards under the diafram. When you drill you cross a not used hole. You have to plug that up also to prevent leakage and pressure drop. So it is made like the earlier models: the 36 and 38 mm CV carbs do not have this system.

Fuel pressure
The pressure of the fuel going to the bowls should be higher than the pressure in the bowl. If not, not any fuel is flowing into the carbs causing starvation as soon as boost starts to build.
Therefore you need a pump capable of making enough pressure to overcome boost + 2-3 psi at sufficient amount of fuel. The standard membrane pump on some carb bikes is definitely not up to the job: don’t even try it. An automotive EFI pump coming out of a car with the same amount of horsepower you are aiming for should work allright.

If you don’t use a regulator: you will have the maximum pressure on the bowls the pump can handle, and that will be around 90 psi and the end of your carbs.
A regulator is used (I use a malpassi and highly recommend it) to give the carbs a 2-3 psi above boost, so difference in pressure between pump and carbs is always the same independent of boost. Make sure you use a bypass type, not a deadend type as it is less accurate. If you can get away with 5 psi on the carbs without leaking, you can use static pressure and don’t need to modify the malpassi. It you have trouble with it, as I have, you need to shorten the spring in the malpassi a bit to achieve 2-3 psi. I also needed a bigger 8mm return line to be less restrictive. Than you definately need the dynamic pressure to prevent starvation under boost!

Jetting+tuning
Don’t make holes in the slides bigger, or use soft dynojet springs! They make things worse in the midrange area.
If you have your carbs and fuelsupply correctly setup: you don’t need to make big adjustments to needles and jets. Mine is actually on 125 mainjets (stock!) functioning just fine at 11.5:1 A/F
Don’t get an A/F any leaner then 12:1 A/F when you use some serious boost. You will burn up some pistons/valves. 11.5:1 is safe imho.
Use an A/F meter to see what’s going on. Looking at plugs is not saying too much, because if their getting hotter then normal on a turbocharged motorcycle they are white anyway.
Symptoms of lean and rich while driving can sometimes be very similar so in doubt always consult the A/F on a dyno!
Because of the lowered fuellevel it is nessecary to give more fuel for the idle, in my case 8 complete turns out for the mixture screws.

Extra tips:
Place a fuelfilter between pump and regulator. Not before pump because you restrict too much, and not after regulator because you messing with pressure which is critical on a carb/turbo setup. Make sure the fuelfilter can hold the pressure!
The return line of the regulator should be as less restrictive as possible: inside minimum 6mm returning into TOP of tank, not below fuellevel in tank. Otherwise fuel pressure cannot go as low as 2-3 psi. Don’t ever (I mean ever!) bend this line because fuel pressure can reach scary levels damaging your carbs seriously!!!!
Mount a fuelpressure gauge direcly next to the boost gauge so you can easy troubleshoot. Remove later if you want to.
Remove filter in your petcock, or better mount a less restrictive petcock without filter. Pingel makes these, but I machined a custom one with build in fuelreturn to the top of the tank so you don’t have to weld in a separate fitting in the tank
If it’s a vacuum operated petcock make sure you use the bypass mode: otherwise on boost the petcock will shut! This is a nice one, as you easily overlook it. On some models the petcock is really restricted in bypass mode so you have to modify that in some way.
If carbs overflow: put in new needles and valves. Make sure fuellevel is not too high: pressurized bowls have a slightly higher fuellevel so you maybe have to adjust it a few milimeters.
On very high horsepowerlevels there can’t flow enough fuel past the needles. Put in thinner/sharper dynojetneedles to solve this. Typical symptom is a sharp leaning out around peak torque, were you need the most fuel every stroke. This makes it harder to get the midrange ok.
At higher boost you can press out the choke plungers! This causes unwanted rich situations and stalling at closing the throttle. Put in stronger springs to solve.

Disclaimer
This info is true for my bike and my application. Some of this info is also true on other bikes/carbs but you have to check yourself. If it doesn’t work or you burn your engine up: I don’t accept any claims. If you crash due to an exploding engine: I’m really sorry and I will send flowers to your family but I am not responsible!

Good luck!

>> Check out the hints section of www.turbo-bike.com, site has an illustrated guide for “converting” CV’s for turbo use.
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The small 34mm gsxr750 carbs with alloy throttles and 4 screws on the caps and floats are actually the best you can use when going turbo. A friend (at is cranking out 420 rwhp at 2.2 Bar of boost.
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I’m also using 34mm carbs on my turbo GSX with alu caps. One more tip: Replace the o-rings in the T-s that go to the floatbowls; 9 times out of 10 they are old and brittle and will leak your precious dynamic boost so the motor runs like a pig at the transit from vaccuum to boost.
One important thing that I noticed during dyno-testing is that you (usually) start with std. mains, then lean it out a bit. If you go down-or upsize the mains, they have a bigger effect on fuelling than it would on an NA engine. So where you would normally say take it 3 sizes down at a time, you need to take one size at at time with this setup.

Marco.

for 38-40 mm carbs

http://www.gsg-mototechnik.de/

## Extra shift detent spring

EXTRA SHIFT DETENT SPRING: “There Is Enough Tension In Drag Racing”

A lot of my hands on mechanical knowledge has been gained thru many decades of working as a R&D Engineer for various Automobile and Motoccyle Maufactuers such as G.M. Toyota, Isuzu, Suzuki, Yamaha, Kawasaki and H-nda. An advantage of working for these Companies is that I would spend a lot of time with a large variety of data acquisition instruments. Fortunately I was able to use and apply many of these instruments to my EFE Drag Bike.

Many of our OSS members are well aware of the fact that Drag Races are won or lossed by a thousand of a second. I have had the opportunity to do A -B -A testing methods using test instruments that are capable of taking measurements in miliseconds, comparing the use of a single or double detent springs with a MRE Pro-Airshifter.

The results of the tests (confirmed repeatability) is that a bike utilizing an MRE Pro-Airshifter will engauge a gear faster with a single spring as compared to using a double spring.

You can consider the above information as another “SSR Race Trick” donated to the OSS site. I still have several more when it comes to Suzuki transmissions that will remain propriety information.

“May The Shift Be With You”

## Exhaust System Efficiency

Part of getting a large fuel/air charge into the cylinder (volumetric efficiency) has to do with getting the combustion products of the previous cycle out of the cylinder. At first thought, it would seem that simply making the exhaust valve bigger would help get the combustion products out. As it turns out, the exhaust valve can be as small as 50% the size of the intake valve without affecting the volumetric efficiency over the usual range of inlet Mach speed. Normally the exhaust valves are at least 60% the size of the intake valve. This effect may arise because the combustion products are “pushed” out of the exhaust port by the piston, while the fuel/air charge is “sucked” in the intake port, pushed only by the manifold pressure.

To enhance the removal of the combustion gasses, the intake valve is opened prior to the end of the exhaust stroke. Since both valves are open at this point, this is referred to as valve overlap. If the pressure in the intake manifold is greater than the pressure in the exhaust manifold, the in rushing fuel/air charge will help scavenge the remaining combustion products in the cylinder as the piston reaches top dead center by pushing them out the exhaust port. While some of the fuel/air charge may go out the exhaust port, an engine tuner tries to design the timing such that the exhaust valve closes just as the last of the combustion gasses leave the exhaust port. An additional benefit of valve overlap is that the intake valve is essentially fully open at the start of the intake stroke, thus reducing the pressure loss through the intake port during the intake stroke. The angle that the crankshaft turns between the intake valve opening and the exhaust valve closing is called the valve overlap angle.

Of course, scavenging does not occur at all speeds. At low speeds, the throttle valve reduces the pressure in the intake manifold, such that the intake manifold pressure is less than the exhaust manifold pressure. In this case, a small portion of the combustion products enter the intake manifold, to be pulled back into the cylinder on the intake stroke. Additionally, the combustion gasess in the space above the piston at top dead center are not scavenged. Even so, at low power settings this is not a problem.

CONCLUSION- In general, we have seen that the torque, and thus the horsepower produced by an engine depends on the amount of air that can be pumped through the engine. The more fuel/air charge drawn into the cylinder, the higher the volumetric efficiency. The higher the volumetric efficiency, the higher the torque. The biggest factor affecting the volumetric efficiency is the valve timing, specifically the valve overlap angle and the intake valve closing angle. Volumetric efficiency can also be improved by the intake manifold design. Since the camshaft used determines the valve timing, changing the camshaft will change the shape of the torque curve, and thus the horsepower curve.

## Exhaust Reversion

Exhaust Reversion

There is a myth that an earlier opening of the intake valve even by 2 or 3 degrees causes the phenomenon known as reversion. This misconception is false from which other incorrect conclusions are made. When you focus on overlap you are on the wrong end of the cam-timing event.

Reversion or the effect of the backing up of the intake fuel air mixture is normally associated with longer duration high-performance camshafts, is actually caused by the intake valve closing later. The answer is in the basic principles of physics. just as with trigonometry and geometry the truth does not change because a person chooses to ignore it.

When the intake valve opens some 40 or more degrees before T.D.C. at the end of the exhaust stroke, very little exhaust gases remain in the cylinder. The piston is close T.D.C and no threat is posed to the incoming intake charge.

A false reversion theory when taken to an extreme would lead to a false conclusion that any overlapping of the intake and exhaust valves is totally undesirable. Engineers of the late 1800’s and early 1900’s used to think this way and they feared of overlap so much so they actually employed negative overlap of – 5 or -10 degrees to be sure none would occur.

The results were that these engines were severely limited to low speeds and marginal output. Engineers in the early 1920’s performed experiments with longer duration cams and proved that camshaft overlap fears are false, as both power, RPM and performance were actually improved. These engineers demonstrated that overlap did not cause engines to lock or backfire.

To further prove that reversion is not caused by earlier intake opening and the resulting extension of valve overlap, look what happens when you advance any camshaft, the intake and the exhaust valves open earlier, this advancement of the cam does not cause more reversion, yet throttle response and torque are improved.

If this myth were correct an engine would run poorly especially at lower RPM. By investigating what is occurring on the other end of the valve timing event will give you the explanation.

When a camshaft is advanced, not only do both valves open earlier but they also close earlier and there is the answer to reducing intake reversion. Closing the intake valve earlier and the reversion of the intake charge as the piston rises on the compression stroke will be reduced. It is not a mystery it is just the truth.

## Exhaust Performance Criteria

Exhaust Performance Criteria

When the piston approaches top dead center the spark plug fires a spark kernel igniting the fuel mixture into a fireball just as the piston rocks over into the power stroke. The piston transfers the energy of the expanding gases to the crankshaft as the exhaust valve starts to open in the last part of the power stroke.

The gas pressure is still high (70 to 90 p.s.i.) causing a rapid escape of the gases. A pressure wave is now generated as the valve continues to open. Gases can flow at an average speed of over 350 ft/sec, but the pressure wave travels at the speed of sound (Mach 1) and is dependent on the gas temperature. The expanding exhaust gases now rush into the port and down into the primary header pipe and then the gases and waves converge at the collector. In the collector, the gases expand quickly as the waves enter into all of the available orifices including the other primary tubes. The gases and some of the wave energy flow into the collector outlet and out the exhaust pipe.

Due to the above there are two basic phenomenon that are created in the exhaust system, gas particle movement and pressure wave activity. The absolute pressure difference between the cylinder and the atmosphere determines gas particle speed. When the gases travel down the pipe and expand their speed decreases. The pressure waves, base their speed on the speed of sound (Mach 1). The wave speed also decreases as they travel down the pipe due to gas cooling, the speed will increase again as the wave is reflected back up the pipe towards the cylinder. All the time the speed of the wave action is much greater than the speed of the gas particles.

Waves behave much differently than gas particles when a junction is encountered in the pipe. When two or more pipes come together such as in a collector, the waves travel into all of the available pipes backwards as well as forwards. Waves are also reflected back up the original pipe, but with a negative pressure. The strength of the wave reflection is based on the area change compared to the area of the originating pipe.

The reflecting negative pulse energy is the basis of wave action tuning. The concept is to time the negative wave pulse reflection to coincide with the period of overlap this low pressure will pull in a fresh intake charge as the intake valve is opening and helps to remove the residual exhaust gases before the exhaust valve closes. This phenomenon is controlled by the length of the primary header pipe. Due to the critical timing aspect of this tuning technique, there may be areas of the power curve that may be harmed.

The gas speed characteristics is a double edged sword. Too much gas speed indicates that that the system may be too restrictive hurting top end power and too little gas speed tends to make the power curve very peaky hurting low end torque. Larger diameter tubes allow the gases to expand and this will cool the gases by slowing down both the gases and the waves.

Exhaust system design is a balance all of these events and their timing. Even with the best compromise of exhaust pipe diameter and length, the collector outlet sizing can optimize or minimize the best design.

The bottom line on any racing exhaust system is to develop the most useful power curve. the final design is how the engine responds to the exhaust tuning on both the dyno and on the race track.

The following components must be considered, Header primary pipe diameter whether constant size or stepped pipes, the primary pipe overall length, the collector design including the number of pipes per collector and the outlet sizing and the megaphone design.

The header pipe sizing and the primary pipe sizing is related to exhaust valve and port size. A header pipe length is dependent on wave tuning. Usually longer pipes tune for lower r.p.m. power and the shorter pipes favor high r.p.m. power. The collector package is dependent on the number of cylinders, and their configuration firing order and their design objectives and the collector outlet size is determined by primary pipe size and exhaust cam timing.

## Porting (general)

There are two ways to port cylinder heads: The right way and the wrong way.

The right way is to refine the flow characteristics of the head and intake manifold so as much air as possible enters the cylinders at the engine’s peak power curve. Every engine is different so there’s no ‘standard’ port configuration that is guaranteed to deliver maximum air flow on every application. The port profile that works best will be limited by the physical dimensions of the cylinder head.

Limiting factors include the size, position and angle of the stock ports, the size configuration and angle of the valves, the thickness of the casting around the ports.

But other factors must be taken into account, too, such as engine displacement, the engine’s bore and stroke, the shape of the combustion chambers, compression ratio, the depth and angles on the valve seats, total valve lift, camshaft profile (duration, overlap,), and type of intake manifold and induction system.

One of the basic goals of head porting is to minimize obstructions so air can flow relatively unimpeded from the throttle plate to the valves.Two things that get in the way are the valve guides and valve guide bosses. Using valves that are necked down just above the valve head improve the air flow.

Transition areas in the port also need to be reworked so air will flow more easily around corners with a sharp radius and into the seat throat just above the valves. Any sudden changes in the cross-section of the port can disrupt this effect and restrict air flow.

The point where the intake manifold and cylinder head intake port meet also is a critical area. If the runners in the rubber intake manifolds are not perfectly aligned with the ports in the head, sharp edges can interrupt normal air flow and impair performance. The same goes for exhaust ports. The head ports must be aligned with the header openings so the exhaust gases can pass freely out of the engine without encountering any sharp edges or obstacles.

The right way to improve air flow is to locate the best places to remove metal. This takes experience, knowing what changes work and what ones don’t and using the right tools for reworking the various portions of the ports, valve pockets and intake manifold

The wrong way to go at it is to grab a die grinder and start hogging out the intake and exhaust ports with no idea of where you’re going or what you’re trying to accomplish other than to open up the ports.

Bigger is not always better. Grind away too much metal and you may end up ruining the casting. But even if you don’t grind all the way through, removing metal in the wrong places can actually end up hurting air flow more than it helps.

“THE SECRET TO MAXIMIZING AIR FLOW AND ENGINE PERFORMANCE IS TO MAXIMIZE VOLUMETRIC EFFICIENCY AND AIR FLOW VELOCITY (SSR)”.

Big ports with lots of volume will obviously flow more air than a smaller port with less volume, but only at higher rpm. A lot of people don’t know that. At lower rpm and mid-range, a smaller port actually flows more efficiently and delivers better torque and performance because the air moves through the port at higher speed. This helps push more air and fuel into the cylinder every time the valve opens. At higher rpm, the momentum of the air helps ram in more air, so a larger port can flow more air when the engine needs it.

The bottom line is this, to realize the most power and performance out of an engine, air flow has to match the breathing requirements of the engine within the engine’s rpm range where it is designed to make the most power.

As a rule, the roof of an intake or exhaust port has much more influence on air flow than the floor or sides of the port. The greatest gains in air flow can often be realized by removing metal from the top of the port only and leaving the sides and floor relatively untouched. The shape of the port is far more critical than the overall size of the port. The largest gains in horsepower are found on the intake side by raising the roof of the port. On exhaust ports, if you tried to match the port to a header gasket you’d probably destroy the port. The secret of exhaust porting today is not how big the port is, but the shape of the port and the velocity of the exhaust flowing through it. Any time you start making the ports bigger on the exhaust side, you usually end up killing air flow in the head.

As for polishing, a smooth finish is great for exhaust ports, but a rougher finish flows better on the intake side. A slightly rough surface texture in the intake ports creates a boundary layer of air that keeps the rest of the air column flowing smoothly and quickly through the port.