Gordan Jennings Two-Stroke User Manual

Two-Stroke

TUNER’S HANDBOOK

By Gordon Jennings

Illustrations by the author

Copyright © 1973 by

Gordon Jennings

Compiled for reprint © 2007 by Ken

i

PREFACE

Many years have passed since Gordon Jennings first published this manual. Its
2007 and although there have been huge technological changes the basics are still the
basics. There is a huge interest in vintage snowmobiles and their “simple” two stroke
power plants of yesteryear. There is a wealth of knowledge contained in this manual.
Let’s journey back to 1973 and read the book that was the two stroke bible of that era.

Decades have passed since I hung around with John and Jim. John and I worked
for the same corporation and I found a 500 triple Kawasaki for him at a reasonable price.
He converted it into a drag bike, modified the engine completely and added mikuni carbs
and tuned pipes. John borrowed Jim’s copy of the ‘Two Stoke Tuner’s Handbook” and
used it and tips from “Fast by Gast” to create one fast bike. John kept his 500 until he
retired and moved to the coast in 2005. The whereabouts of Wild Jim, his 750 Kawasaki
drag bike and the only copy of ‘Two Stoke Tuner’s Handbook” that I have ever seen is a
complete mystery. I recently acquired a 1980 Polaris TXL and am digging into the inner
workings of the engine. I wanted a copy of this manual but wasn’t willing to wait for a
copy to show up on EBay. Happily, a search of the internet finally hit on a Word version
of the manual. Previous searches had unearthed usable but poor quality .pdf versions that
didn’t print very well.

I have proof read and edited this version and will print and bind a copy for
myself. Although it is not an authentic reproduction it will be a proud addition to my
library. I have to thank the anonymous creator of the file 2-stroke-guide_r2.zip.

I felt I had to share the background and history of the quest. Perhaps it will help
you enjoy this copy and in closing I would like to thank Gordon Jennings for the fine
material and all the effort he put into the original publication. Computers and all the fine
software have taken modifying and tuning two-strokes to a level higher that one could
have imagined in 1973. A lesser work would have surely passed into “Nostalgia

Heaven” long ago.

On page 51 Gordon uses the term “One-Percenter”, the term refers to an
American Motorcyclist association quote “that 99% of motorcyclists were law-abiding
citizens, and the last one percent were outlaws”.

ii

iii

FOREWARD

Only ten years ago the two-stroke engine was widely and quite understandably
thought to be a "reasonable alternative to the four-stroke only when minimum weight and
manufacturing cost were all-important considerations. The two-stroke was recognized as
having substantial theoretical promise, as it delivered a power stroke for each 360 degrees
of crankshaft rotation but the hard reality was that each individual power impulse was too
feeble to amount to much when totaled at the output end of the crankshaft. A very few
engines had begun to appear in which some of the theoretical promise was realized
however, and this encouraged engineers at MZ, Yamaha and Suzuki to persist in their
efforts to wring competitive power output from the racing two-stroke engine. To say that
they were ultimately successful would be gross understatement.

Those engineers were motivated by the need to demonstrate that the two-stroke
engine, per se, was worthwhile -as that would stimulate sales of their companies' ordinary
touring models. My own interest in the two-stroke, which had reached the level of an
obsession by 1963, was generated by comparative poverty. I like to tinker with engines,
and the complexities of the poppet-valve four-stroke make modifications very expensive.
One may think that a change in valve timing would do wonders for a four-stroke's power,
but getting a camshaft made to order costs hundreds of dollars. In contrast, a two-stroke
engine's valve timing may be altered simply by reshaping the holes in its cylinders, and
its power output markedly changed by utilizing inertia and resonant effects in its intake
and exhaust tracts. None of these modifications are costly.

On the other hand, while the two-stroke engine does not commonly require large
dollar inputs to raise its power output, it does require an in-depth under- standing on the
part of the man doing the modifications. In an attempt to acquire that understanding I
began a study of the high-speed, high-output two-stroke engine that has led to the
collection of a minor library of text books and SAE papers. And to an endless series of
experiments, some of them illuminating and many others raising more questions than
they have answered. At this stage I have arrived at more or less satisfactory explanations
for most of the gross phenomena, such as the general behavior of expansion chambers
and port time-area values, and I flatter myself to think that just that much is an acceptable
excuse for writing this book for the guidance of the layman experimenter. If it will not
supply all of the answers it will at least take care of the fundamental problems and
prevent the worst mistakes.

My special thanks to Mr. John Brooks, of McCulloch Engineering, who has done
much to dilute my once pure ignorance (but should not be held accountable for the
residue found herein). Also to the late Henry Koepke, who mistakenly assumed that I
knew something about two-stroke engines and supported my early research; to my old
friend Joe Parkhurst, who started me working on this book nearly ten years ago but never
got it; and finally to Tom Heininger, who wheedled, needled, pleaded, complained and
cajoled until I hammered my file of notes into publishable form.

iv

CONTENTS

Preface .....………………………………………………………………... ii

Foreword ….………………………………………………………………... iv

Contents …………………………………….…………………………….... v

FUNDAMENTALS …………………………………………………………….. 1

Predicting Power …………………………………………………..… 3
Piston Speed …………………………………………………………... 5
Piston Acceleration …………………………………………………… 7

CRANK TRAIN ………………………………………………………………. 17

The Piston …………………………………………………………… 17
Piston Rings ……………………………………………………………. 22
Wristpin/Crankpin Bearings …………………………………………… 25
Crank Assembly ……………………………………………………………. 30

CYLINDERHEADS …………………………………………………………….. 33

The Combustion Process ……………………………………………. 34
Squish Bands …………………………………………………………... 37
Plug Location …………………………………………………………….. 42
Head/Cylinder Sealing ……………………………………………………. 46

EXPANSION CHAMBERS …………………………………………………… 51

The Basic Process …………………………………………………… 52
Tuned Length .………………….……………………………………….. 55
Diffuser Proportions ……………………………………………………. 56

Baffle Cones ............................................................................................. 61

Outlet Pipes ............................................................................................. 64

Lead-in Pipes ............................................................................................. 66

Expansion Chamber Design Formula ......................................................... 73

PORT TIMING .................................................................................................... 75

Specific Time-Area ................................................................................. 76

Angle-Area ............................................................................................. 80

Time-Area Combinations ..................................................................... 81

Emphasis on Area ................................................................................. 84

Timing Limits ............................................................................................ 85

Rotary-Valve Timing ................................................................................ 88

Porting …...................................................................................................... 90

v

CONTENTS

CRANKCASE PUMPING …............................................................................... 91

Resonance Effects ................................................................................. 92

Carburetor Location …............................................................................. 94

Crankcase Volume ................................................................................. 95

Reed Valves ............................................................................................. 98

The Rotary Valve ................................................................................. 101

Intake Port Shape ................................................................................. 102

CYLINDER SCAVENGING ............................................................................... 105

Exhaust Port ............................................................................................. 109

Port-Edge Chamfers ................................................................................ 112

Flow Patterns ............................................................................................ 115

Multiple Transfer Ports …............................................................................. 119

Subtleties ............................................................................................. 121

CARBURETION, IGNITION ……...................................................................... 129

The Basic Carburetor ................................................................................. 130

Adjusting Mixture ................................................................................. 134

Ignition Fundamentals …............................................................................. 143

Spark Plugs ............................................................................................. 152

Ignition Timing ............................................................................................. 157

TIPS FOR BIG BORING CYLINDERS ………………………………………. 158

Note: An original manual was scanned, and that .pdf was converted to a MS-Word document. I copied that
document from the site http://edj.net/2stroke/jennings/
of correcting typos, changing phases, adding some notes and a few quotes from Gorr.

and prior to printing for my own use took the liberty

vi

vii

FUNDAMENTALS

Throughout this book it will be assumed, inconvenient though that assumption
may occasionally be, that the reader has progressed to at least a superficial knowledge of
the manner in which a piston-type internal combustion engine - with particular reference
to those operating on the two-stroke cycle principle-converts quantities of fuel and air
into useful power delivered at the end of its crankshaft. People who need enlightenment
in that regard will find a wealth of explanatory literature collected on the shelves of any
public library; no real purpose would be served by lingering over the matter here.
Neither will I attempt to instruct you in the elementary mathematics and physics required
to grasp much of what follows, as again the public library is an entirely adequate source
of information. What will be provided is a kind of “state of the art” report about high­speed, high-output two-stroke engines for laymen-who in most cases do not have access
to the literature (SAE papers, etc.) available to engineers and thus must rely upon
hunches (often wrong) and folklore (almost invariably wrong) for guidance. Many have
learned, to their sorrow, that it is distinctly possible to lavish enormous amounts of time
and money on the two-stroke engine without realizing a return appropriate to the
investment. The information to be provided here will not make you a Kaaden, or Naito;
it will help you to avoid some of the more serious mistakes.

The first serious mistake a layman experimenter can make is to assume that those
who designed and manufactured his particular engine didn't know what they were doing.
In point of fact, the professional engineer knows very well, and if the engine in question
is something other than what the experimenter has in mind, there are excellent reasons:
all engines are compromised, from what you might consider an ideal, in the interest of
manufacturing economy and broad usefulness. For example, ports may derive their shape
as much from what the design engineer intended to be a low scrap-rate casting as from
consideration of flow characteristics. In other words, even something like ports-design
always will be influenced by the demands of mass-production manufacturing. Similarly,
designing for mass-market sales implies that an engine must be agreeable to many
different uses - even though that inevitably means that it will do no single thing
particularly well. In these areas will we find the latitude for “improving” an engine, and
one should always be mindful that the real task is simply to tailor a mass-use product to a
very specific application- and that in the tailoring process one inevitably will incur all the
various expenses the engine's designer has avoided. Hours of labor may be required to
finish rough-cast ports; dollars will be spent correcting other things that are the creatures
of manufacturing economies; power added at maximum revs will be power subtracted at
lower crankshaft speeds, while the increased speeds required to obtain large
improvements in power output will be paid for in terms of reliability.

1

Two Stroke TUNER’S HANDBOOK

Another mistake commonly made, sometimes even by those who have enjoyed
some success in modifying two-stroke engines, is to believe in a kind of mechanistic
magic. Bigger carburetors, higher compression ratios, altered port timings and expansion
chambers often do bring an improvement in power output, but more and bigger is not
magically, instantly better. All must work in concert with the basic engine, directed
toward the particular application, before they constitute a genuine improvement. You
cannot treat them as a voodoo incantation, hoping that if you mutter the right phrases and
stir the chicken entrails in the prescribed manner, your mild-mannered, all-purpose
chuffer will be transformed into a hyper-horsepower fire-breather. With a lot of luck,
you might get that result; the chances heavily are that you won't.

With all the mysticism filtered out, horsepower at any given displacement is
simply a function of average pressure in the cylinder during the power stroke and the rate
at which power strokes occur, minus work absorbed by friction and scavenging. Raise
pressure and/or the delivery rate of the power strokes, or reduce friction and pumping
losses, and the engine's net output will rise. Unfortunately, there are limitations on all
sides: Pressure must be limited because of thermal considerations (and is further limited
by an engine's restricted ability to recharge its cylinder with a fresh air/fuel mixture
between power strokes). The limit for power strokes per unit of time is established by
what is tolerable in terms of crankshaft rotational speeds, and tolerable here is what the
bearings, rod and piston will survive, in inertia loadings, for what you consider an
acceptable service life; the design engineer has already expressed his opinion in this
matter. Pumping losses can be reduced - relative to the mass flow through an engine ­with a properly designed exhaust system, but otherwise are an inevitable and almost
invariable consequence of pulling air from the atmosphere, moving it through the engine,
and out the exhaust port. Some improvement in output may be obtained with reductions
in friction, but the scope for such improvements is very small compared to what may be
accomplished with cylinder pressure and engine speed.

Obviously, pressure in a cylinder will vary continuously throughout an engine's
entire power stroke. Knowing what those pressures may be in a given engine is useful,
but more useful still is knowing what they should and are likely to be, as such knowledge
can keep you from that futile exercise commonly known as flogging a dead horse - and
from believing a lot of lies about how much power various people are getting from their
engines. Engineers have an overall efficiency rating called “brake mean effective
pressure” (bmep), which they calculate by working their way back through torque
readings observed on the dynamometer, the leverage provided by crankpin offset, and
piston-crown area. Thus, bmep says little about peak cylinder pressures (those
measurements being taken with a pressure transducer and oscilloscope) but it is an
excellent relative indicator of performance and highly useful in projecting power output
from a modified engine.

2

FUNDAMENTALS

PREDICTING POWER

An average, well-developed stock engine intended for use in a sports/touring
motorcycle will have a bmep of about 70-psi. It is possible, and I must stress that word
"possible", to raise this to perhaps 115-psi - an improvement of some 64-percent - which
(if accomplished) will yield a 64-percent increase in power output without raising the
engine's operating speed. Similarly, a 64-percent increase in operating speed without a
change in bmep would have the same effect on output. You will see this in the following
formula for calculating horsepower:

PLAN

BHP =

33,000

Where BHP is brake horsepower

P is brake mean effective pressure, in psi
L is piston stroke, in feet
A is the area of one piston, in square inches
N is the number of power strokes per minute

Obviously, when the values of L and A are held constant, as would be the case
with an engine having a piston displacement at the limit established for a particular racing
class, then increases in power may only be obtained by increasing the values for P and N

- and you will find that in practice it is a lot easier to increase the latter than the former.

As already stated, bmep figures for stock, touring-type engines with flow­restricting air cleaners and mufflers, and with porting/carburetion compromised in favor
of smooth low-speed running, will be around 70-psi. Typical figures for engines with
porting and other plumbing arranged solely (and effectively) for maximum horsepower at
peak revs would be about 115-psi - with a few small, highly-developed two-stroke
engines operating up at 125-psi. The exact number will vary according to unit cylinder
displacement and the width of an engine's useful power band, but one may reasonably
expect that engines suitable for motocross will fall in the 85>95-psi range - with big
cylinders tending toward the lower figure and small cylinders vice versa. Road racing
engines, tuned to exert a maximum effort over a very narrow speed range, will usually
show a bmep of 100-115-psi, and of course the same remarks regarding the influence of
cylinder size apply. Bimotion © 2000 suggested a BMEP of 11 bar (160 psi) for road

race, 9 bar (130 psi) for motocross and 8 bar (116 psi) for enduro.

These numbers have usefulness beyond the mere satisfaction of vulgar curiosity:
they may be used very profitably to determine an engine's suitability for some particular
application. For example, they shed light on the future prospects of those who are trying
to transform Kawasaki's F-5 “Bighorn” engine, a 350cc single, into a prime-mover

3

Two Stroke TUNER’S HANDBOOK

capable of ending the Yamaha TD-2's absolute domination in road racing. Much has
been made, by the Kawasaki's supporters, of the usefulness of a broader power range
inherent with the F-5's disc-valve induction and the l00cc advantage it gets, over the TD­2, by having only a single cylinder (this, under the present American Motorcycle
Association rules). Now while it is true that a racing motorcycle having a wide power
band is easier for its rider to manage, and may offer an absolute if very slight advantage
on short, extraordinarily twisty circuits, one must not overlook the fact that the TD-2 has
been blessed with an excellent close-ratio transmission and a number of riders quite
capable of coping with any problems introduced by the need for frequent gear changes.
Viewed realistically, the situation facing any serious challenger to Yamaha's supremacy
is one in which horsepower must be met with horsepower. And what are the Kawasaki's
prospects of developing that kind of horsepower? Let's have a look at the numbers:

Assuming that the man who modifies the Kawasaki F-5 knows his business, but
doesn't have all the development time in the world, (probability favors the latter far more
than the former) then he very likely will arrive at a combination of porting, etc., good for
a bmep of about 105 psi-which is about all that can be expected with a single cylinder of
350cc displacement. To expect more would be to ignore the considerable difficulties in
scavenging efficiently the F-5's large-bore (3.17-inch) cylinder. Further assuming (and as
we shall see later, this assumption is far from safe) that the F-5 engine will remain in one,
working piece for the duration of a longish race with its rider observing a red-line of 9000
rpm, with a power peak at 8500 rpm, then,

BHP =

000,33

BHP = 47.6

So, a well developed F-5 would deliver 47.6 brake horsepower. How does that
compare with the Yamaha TD-2? With all the years that have gone into the TD-2's
development, and giving due thought to Yamaha's proven expertise in these matters, it
seems safe to assume that this engine would be operating with a bmep of 115 psi at its
power peak- which seems to be at 11,000 rpm. Thus, working from those numbers and
the 250cc Yamaha twin's bore/stroke dimensions of 56mm and 50mm, respectively,

BHP =

000,33

BHP = 48.0

8500789.7223.0105 ×××

000,2281.3164.0115 ×××

4

FUNDAMENTALS

Clearly then, those who would try to beat the Yamaha with a Kawasaki F-5 have
taken upon themselves a task of considerable magnitude. The only bright spot in the
picture, for them, is that while they are 0.4 bhp down on the Yamaha (assuming near­optimum work on their part) they probably will have the advantage in terms of average
horsepower, figured from the moment a gear is engaged - when revs fall somewhat below
those for peak horsepower -until the red-line is reached and it is time for a change to the
next higher gear. There will be no advantage in frontal area, for although the F-5 engine
is narrower than that of the TD-2, the fairing must be wide enough to shroud the rider,
and the minimum width that requires is sufficient to encompass either engine. Moreover,
moving from the theoretical to the practical for a moment, it is highly unlikely that the
Kawasaki could be made as reliable at 8500 rpm as is the Yamaha at 11,000 rpm, and not
because the F-5 engine is badly designed or shoddily constructed. The simple truth is
that any single-cylinder 350cc engine with the F-5's bore/stroke dimensions and red-lined
at 9000 rpm is going to be stressed very near its absolute limit - a limit imposed by the
properties of available materials.

PISTON SPEED

All this asks the question, “How does one determine the limit, with regard to
engine speed?” Unfortunately, establishing this limit with any precision is not only
extremely difficult in terms of the mathematics involved, but also requires data
concerning metallurgy, etc., seldom available outside the record-rooms of the factories
from which the engines originate. Still, there are guide-lines which, if lacking in absolute
precision, do at least have the virtue of simplicity, and will provide an indicator to keep
us away from certain trouble. It is almost impossible to establish the point, in engine
speed, between zero trouble and the possibility of trouble; there is much less difficulty in
determining a red-line between

A quick and easy method of establishing a limit for crankshaft speed is by
working with piston speed. Actually, with "mean" piston speed: pistons do not travel at
uniform velocity; they move from a dead stop at each end of their stroke, accelerate up to
a maximum speed that often is in excess of 120 mph, and then brake to another complete
stop. For convenience, we use just the mean piston speed and the safe limit for that, for
engines having bore-stroke dimensions within the range considered normal for
motorcycles, is about 4000 feet per minute. And mean piston speed may be calculated
very easily by applying the following formula:

some trouble and nothing but trouble.

5

Two Stroke TUNER’S HANDBOOK

= 0.166 x L x N

C

m

Where: C

Thus, using again the Kawasaki F-5 engine as an example, with L being 2.68-inches and
N given as 9000, we find that

Here we have a theoretically-predicted limit that seems to agree quite closely with
observable reality in the field: Reports from those actually racing modified F-5
Kawasaki’s indicate that the engine does in fact retain acceptable (within the framework
of that word's meaning in racing) reliability when red-lined at 9000 rpm, and ravels with
horrifying abruptness if pressed further. Of course, it must be
engines, the F-5 not excepted, retain more than marginal reliability at mean piston speeds
of 4000 ft/min, and even this presupposes frequent replacement of the piston and the
crank/rod bearings.

You will be on far more solid ground if your engine is not asked to endure
piston speeds above 3500 ft/min. Anything above that takes an engine into the twilight
zone of reliability, and the ground between 3500 ft/min and the near absolute limit of
4000 ft/min is covered with unpleasant possibilities, but these often may be minimized
with the proper selection of materials and lubrication. I should note here that there are
exceptions to this rule among some of the old-fashioned, long-stroke engines, which tend
to have very light (and strong) reciprocating parts relative to their absolute stroke. An
example that comes to mind is the Bultaco 125cc TSS, which had a stroke of no less

2.36-inches (decidedly long for a 125) but which would, in “factory”
up to 11,500 rpm, just like the Yamaha TD-2 (with a much shorter, 1.97-inch stroke), and
that represents a mean piston speed of 4500 rpm. Obviously, Bultaco held the opinion
that the resulting thin-ish margin of reliability was acceptable, but their TSS never was as
predictably trouble-free as Yamaha's TD-2, which at the same crankshaft speed (11,500)
has a mean piston speed of only 3775 ft/min.

While on the subject of bore/stroke dimensions, I would like to say that there is
much in favor of long stroke two-stroke cycle engines in many applications. They are not
superior (as many people seem to think) compared to the present day short-stroke designs
in terms of low-speed torque, as torque is entirely a function of displacement and bmep,
and wholely unrelated to bore/stroke ratios. With a long stroke, there is (at any given
displacement) a reduction in bore, and with it a loss of piston area against which gas

is mean piston speed, in feet per minute

m

L is stroke, in inches
N is crankshaft speed, in revolutions per minute

Cm = 0.166 x 2.68 x 9000
Cm = 4000 ft/min

stressed here that few

mean

road racing trim run

6

FUNDAMENTALS

pressure can exert its force, that exactly balances the loss of leverage in a short-stroke
engine (which is, in turn, compensated by a gain in piston area). The only thing wrong
with the long-stroke engine is that its crankshaft speed is limited by inertia loadings, and
that in turn limits its absolute power potential as compared with the “modern” short­stroker. On the other hand, it is compensated by having a much more compact
combustion chamber, which makes for more efficient burning, and by lower thermal
loadings on the piston as a result of the smaller crown area into which heat from the
combustion process may soak. Finally, there is an advantage in port area for the long­stroke design resulting from its relatively large cylinder wall area. This area increases in
the long-stroke engine because displacement rises only in direct proportion to stroke, but
is increased by a factor of 3.1416 (the constant, π) with enlargements in bore. These are
very real advantages, but they are not enough, usually, to prevail against the short-stroke
engine's sheer ability to rev. Crankshaft speed is the only thing subject to much juggling
in the horsepower equation- and is a far more potent factor in determining power output
than the relatively slight improvements in bmep obtainable with the marginally better
combustion chamber and porting in the long-stroke engine. A 10-percent improvement
in our Kawasaki F-5 engine's bmep (a large improvement indeed) would raise its output
to 52.3 bhp; leave the bmep unchanged, but shorten the stroke and spin it 11,000 rpm and
you would have 61.3

Sadly, while there is no substitute for revs, there are plenty of barriers: piston
speed is one, as was already noted. But that is a rather indirect limit, as it ignores the fact
that it is not speed so much as all the starting and stopping of pistons that does the
damage, or at least the worst of any damage. The acceleration forces generated by the
starting and stopping are felt even in an engine's main bearings, but they are at a peak in
the connecting rod and piston and have a particularly disastrous effect on the latter, as
any attempt to make a piston stronger is apt also to make it heavier-which aggravates the
very situation the strengthening of the piston should improve. Even so, an engine's true
Achilles heel, the problem that may most strongly resist solution, often is the disastrous
effects piston acceleration may have on the piston's rings.

It often is thought, and quite wrongly, that rings maintain a seal between the
piston and the cylinder's walls simply through their properties as springs. A
should convince you that such cannot be the case, for most rings, compressed in the
process of installation, press outward against the cylinder with a force amounting to about
30-psi. Gas pressure in that cylinder may easily exceed 750-psi, and it should be obvious
that a 30-psi force will not hold back one
rings do form an effective seal. How? Because they get a lot of help from the cylinder

bhp. There is indeed no substitute for revs.

PISTON ACCELERATION

little thought

circa 750-psi. Still, equally obviously, piston

7

Two Stroke TUNER’S HANDBOOK

pressure itself: gas pressure above the ring forces it down against the bottom of its groove
in the piston, and also (acting behind the ring, in the back of the groove) shoves it out
hard against the cylinder wall. Thus, in the normal course of events, sealing pressure at
the interface between cylinder wall and ring always is comfortably higher than the
pressure it must hold back.

This very desirable situation will be maintained unless something happens to
upset things, and most-insistent among the several “something’s” that may intrude is
excessive piston acceleration. When piston acceleration exceeds the sum total of gas
pressures holding the ring in place, the ring will lift upward (as the piston nears the top of
its stroke, and is being braked to a halt). Instantly, as the ring lifts, the gas pressure
previously applied above and behind is also applied underneath the ring, at which point
its inertia takes over completely and the ring slams up hard against the top of its groove.
This last action releases all pressure from behind the ring, leaving it entirely to its own
feeble devices in holding back the fire above, and as its 30-psi outward pressure is no
match for the 750-psi pressure in the upper cylinder, it is blown violently back into its
groove. The ring's radial collapse opens a direct path down the cylinder wall for the high
temperature and pressure combustion gases - but only for a microsecond - for the action
just described instantly applies gas pressure once again behind the ring and that sends its

8

FUNDAMENTALS

snapping back into place against the cylinder wall. Unhappily, it cannot remain there, as
gas pressure immediately bangs it back into its groove again - to repeat the process over
and over until the piston is virtually stopped and the ring's inertia is no longer enough to
counter gas pressure.

The net result of all this activity is that over the span of several degrees of crank
rotation, immediately preceding the piston's reaching top center, the ring will be
repeatedly collapsed radially and at the same time hammered hard against the top of its
groove. Understandably, the ring is distressed by this, as it not only receives a fearful
battering but also is bathed in fire while being deprived of the close contact with piston
and cylinder that would otherwise serve to draw off heat. Equally damaging is that the
piston is having much the same problem, with high-temperature gases blowing down past
its skirt to cause overheating, to burn away the film of oil between itself and the cylinder
wall, and with its ring, or rings, all the while trying to pound their way up through the
piston crown. A mild case of what is quite accurately termed “ring flutter” eventually
results in the destruction of the ring and sometimes the dimensional integrity of its
groove; a more serious case is certain to lead rapidly into lubrication failure, overheating,
and piston seizure. Fortunately, this drastic problem can be avoided, thanks to the work
of the researcher Paul de K. Dykes, whose investigation of the ring flutter phenomenon
yielded most of what we know about it - and who invented the flutter-resistant ring that
bears his name. Dykes showed us the cause of ring flutter, and engineers' understanding
of the cause is reflected in their designs of the modern piston ring, which is very thin,
axially, with a very considerable width, radially. Thus, gas pressure bears down on a
large surface, providing an equally large total down-force, but is opposed by a relatively
small upward load as the ring, being thin, is light and in consequence has little inertia.
Still, even with very thin rings, flutter will occur if inertia loadings are high enough. To
settle the question, with regard to any given engine, apply the following formula for
determining maximum piston acceleration:

2

G

Where G

N is crankshaft speed, in revolutions per minute

L is stroke, in inches

A is the ratio of connecting rod length, between centers, to stroke

×

=

max

is maximum piston acceleration, in feet per second squared

max

LN

2189

1



1






+



A2



9

Two Stroke TUNER’S HANDBOOK

To illustrate how high these forces may sometimes be, let's use as an example the
Yamaha TD-2, using 11,000 rpm for N. The formula tells us that at that speed, maximum
piston acceleration will be (with the answer rounded off by my slide rule; I'm too lazy to
do it all with paper and pencil) no less than 135,000 ft/sec2. Now if you will recall for a
moment that the acceleration of gravity is only 32 ft/sec2, it will be clear that the load on
the Yamaha's pistons - and thus on its rings - is very high indeed. But is the loading high
enough to make the Yamaha's rings flutter? Obviously, it is not, as the engine remains
not only reliable but crisp in comparatively long races. The limit, for the TD-2 engine, is
slightly higher than 135,000 ft/sec2 - but not much higher, as
following table listing ring thicknesses and the accelerations at which they begin to
flutter.

2

For rings having a 0.125-inch thickness, 40,000 ft/sec

0.094 “ “ 53,000 ft/sec

0.063 “ “ 80,000 ft/sec

0.047 “ “ 106,000 ft/sec

0.039 “ “ 138,000 ft/sec

2

2

2

2

The Yamaha, with rings having a thickness of 1mm, or 0.039-inch, and a
maximum piston acceleration of 135,000 ft/sec2 at 11,000 rpm,
operating very near the limit - as indeed it is. But it probably is not quite as near the limit
as the numbers suggest, for a racing ring (with its exaggerated thickness/width cross­sectional aspect) is somewhat less subject to flutter than a ring made for application in a
touring engine. Still, the numbers given are fairly close for rings with normal-range
proportions, and if you have an engine with rings for which flutter is predicted at 80,000

2

and intend using crankshaft speeds that would raise maximum piston acceleration

ft/sec

2

to something more like 100,000 ft/sec

, then I strongly urge you to fit new pistons with
thinner rings. You may interpolate between the figures given to find the safe acceleration
levels for ring thicknesses not listed.

There are piston rings that resist very strongly piston acceleration's efforts toward
making them flutter. The best known of these is the Dykes-pattern ring, which has an L­shaped cross-section and fits into a similarly-shaped groove in the piston. The Dykes
ring is made flutter-resistant by the fact that its horizontal leg fits quite closely in its
groove, as compared to clearances around the vertical leg, and therefore even if
acceleration lifts the ring it cannot lift high enough to close off the pressure behind the
ring's vertical leg. In consequence, the ring's sealing abilities are maintained at
accelerations that would be the undoing of rings in the conventional rectangular-section
pattern. However, the Dykes ring's ability to maintain a seal does not free it of all the
unpleasantness attending too-high piston acceleration: while it may seal under those

you will see in the

would seem to be

10

FUNDAMENTALS

conditions, it is still being rattled about vigorously and if the rattling continues long
enough, the Dykes ring, and the groove trying to restrain it, both become badly battered.
At that point, its ability to seal vanishes and mechanical failure of the ring, piston, or
both, follows very closely. Bultaco has long used Dykes-pattern rings, as have certain
others, but most manufacturers prefer rings that do not require such careful and intricate
machining. There are other flutter-resistant rings, and many excellent reasons for using
rings of conventional configuration, but these details are discussed elsewhere in this book
and in greater depth than would be appropriate here.

After establishing all these mechanical limits, with regard to piston speed and
acceleration, and after deciding how much power you are likely to get from a particular
engine, you should subject the engine to a complete survey. This would include the
measuring of port heights and widths, combustion chamber and crankcase volumes, and
charting piston travel against crank rotation. This last effort may at first seem rather
pointless, but as your work progresses you will find that the chart, which will show
almost but not quite a sine curve, provides an instant readout between degrees at the
crankshaft and the position of the piston from top center that is most useful. It will tell
you, for example, how much to raise the top edge of an exhaust port to make a given
change in timing, and how much to trim from the piston skirt (in a piston-port engine) to

11

Two Stroke TUNER’S HANDBOOK

get the intake period you want - or think you want. The chart also will provide you with
all the mean port-open points, and it will provide an exceedingly useful relationship
between ignition timing expressed in degrees and in piston travel from top center. You
may devise your own methods for deriving all this information according to your
preference and resources; I have explained my own techniques elsewhere in this text, in
the appropriate chapters.

An item that must be included in any discussion of the two-stroke cycle engine's
basics is general gas dynamics. You can get information on the subject at your local
library, but the applicable particulars are likely to be widely scattered there, so I will
cover the subject in brief here. The manner in which what follows applies at specific
points throughout the engine and its related plumbing will be covered later, but you
should know a few of the fundamentals now and thus save me from becoming
unnecessarily repetitious later.

One thing you must know, for example, is that the air moving through the engine,
a mixture of gases, has many of the properties of a fluid. It even has the ability to “wet” a
surface, and has viscosity, which means that air will cling to all surfaces within an engine
in a layer that moves hardly at all no matter what the midstream velocity may be. This
boundary layer's depth is influenced by gas temperature, and by the temperature of the
surface on which it forms, as well as by the shape of the surface. Please understand that
the layer is not solid; it is “shearing” with general flow throughout its depth – which may
be as much as 0.100-inch - with movement increasing as to distance from the surface on
which it is formed. And as close as 0.020-inch from the surface, flow may still be in the
order of 80-percent of that in midstream, which means that the restriction formed by the
boundary layer is not very great. Nonetheless, it is there, and it accounts for such things
as round ports having less resistance to flow than square ports, area for area, and for the
ability of a single port to match the flow of a pair of ports of somewhat larger area. It
also accounts for the fact that flow resistance increases in direct proportion with the
length of a port, and much of the resistance resulting from the shape of a particular port is
due to that shape's creating a thick boundary layer, which becomes literally a plug inside
the port.

Generally speaking, boundary layers will be held to minimum depth on surfaces
that “rise” (relative to the direction of flow) and gain in thickness on any surface that falls
away. Thus, an intake trumpet (
from the inlet end to the carburetor-by perhaps 2-3 degrees - in the interest of holding
boundary layer thickness to a minimum. In that configuration, it will have appreciably
less resistance to flow than a straight, parallel-wall tube. Similarly, transfer ports should
diminish in cross-sectional area from their entrance in the crankcase toward their outlet in
the cylinder.

velocity stack), for example, should be tapered in slightly

12

FUNDAMENTALS

These gases also have inertia: once set in motion they tend to remain in motion;
when at rest they resist all efforts to get them moving. In practice, this means that there
always is a lag between the intake port's opening and the movement of air in the intake
tract. Fortunately, this lag can be amply compensated toward the end of the intake
period, when the pressure inside the crankcase has risen to a level that should push part of
the charge back out the port-but cannot because of the effect of inertia on the incoming
gases. Inertia also has its effect on the flow of gases through the transfer ports and out
the exhaust system, but I will deal with that while treating those subjects separately.

These inertia effects are useful, but difficult to manage as something apart from
other processes occurring as the engine runs. For example, intake tract length usually is
established more with an eye toward resonances than inertia, and its diameter set by the
flow rate required by the carburetor to meter properly - balanced against the resistance
that attends high gas velocities. Therefore, virtually the only thing we can do about
inert ia effects is to attempt to find the intake timing that will make maximum use of those
provided by an intake system proportioned mostly to suit other requirements.

Resonances are another matter. Sound waves will travel through any elastic
medium, such as air, and in their passage they pull together or force apart molecules, just
as the similar energy waves traveling through the ocean pull the water into peaks and
troughs on its surface. And, as in the ocean, the waves move steadily onward away from
their source but the transmitting medium does not. Take, for example, the activity
surrounding a single condensation, or positive-pressure wave, as it moves through the air.
In its center, molecules have been pulled together, condensed, but as it travels it releases
those molecules and compresses others as it reaches them. In the same manner, a
rarefaction, or negative-pressure wave, pushes molecules apart. Both waves behave in a
curious, but useful way when confined in a tube and the effects of inertia are mixed with
them. For one thing, they will be reflected back when reaching the end of the tube ­whether that end is open or closed. But at the tube's open end, the wave changes in sign:
a condensation is inverted and becomes a rarefaction, and v
the wave will be reflected, but retains its sign.

How is all that useful? For example, in the intake system the opening of the
intake port exposes the crankcase end of the tract to a partial vacuum, and that in turn
sends a rarefaction shooting off toward the opposite, atmospheric, end of the tract. It
travels out to the intake bell, inverts in sign to become a condensation, and instantly
moves back toward the crankcase - to arrive there as a clump of compressed molecules,
which surge into the crankcase to be trapped, if the piston then closes the intake port, as
part of the scavenging charge. That effect, over-layed with inertia in the inrushing gases,
makes all the difference in getting the job of charging done in two-stroke engines - which
provide only an absurdly short time for such chores.

ice versa; at the closed end,

13

Two Stroke TUNER’S HANDBOOK

θ

θ

θ

How short a time? That is at the same time one of the least complicated and most
depressing calculations you can perform. Let us consider the Yamaha DT-1, which in
fully developed configuration had an intake duration of 160-degrees, a transfer duration
of 123-degrees, and an exhaust duration of 172-degrees. Yamaha claims a power peak at
7000 rpm. Let's have a look at the actual time, in fractions of a second, available for the
completion of these functions. To arrive at these times, use the following formula:

60

T

Where T is time, in seconds

N is crankshaft speed, in revolutions per minute

×=

360N

is port open duration, in degrees

T

(This formula can be abbreviated to

Thus, to find T for the 160-degree intake duration,

160

60

T −=×=

7000

With application of the same formula to the transfer and exhaust periods, we find
that the former is open 0.0029-second, and the latter open 0.0041-second. Even the
longest of these, the exhaust-open duration, is only 41/10,000-second, and that is not very
much time in which to empty exhaust gases out of the cylinder. Actually, that particular
process is substantially finished in the 29-degrees, or 0.0007-second, between exhaust­and transfer-opening. In that short period, pressure in the cylinder must fall to something
very near atmospheric, or the exhaust gases would force their way down into the
crankcase through the transfer ports. Of course, the exhaust gases are provided quite a
large aperture by means of which they may make their escape, and that they do so,
successfully, is less remarkable than the fact that the fresh charge compressed in a two­stroke engine's crankcase is able to make its way through the far more restricted transfer
ports, propelled by a far lower pressure, to refill the cylinder in the extremely brief
moment available. It seems nothing short of astonishing that this recharging operation is
accomplished in the 0.0027-sec provided by the Yamaha DT-1's 114-degree transfer
period; that the same process takes place in a Yamaha TD-2 engine in only 0.0017-sec
appears a minor miracle. Obviously, divine intervention is not really a factor in the
functioning of two-stroke engines, and cylinder recharging is possible simply because the
process gets a lot of help from the activities of the exhaust system, gas velocities through
the transfer ports have a mean value in the order of 300-ft/sec, and the cross-sectional

360

sec0038.0

.

=

)

6N

×

14

FUNDAMENTALS

areas of the ports involved are relatively large as compared with the volume of gases to
be transferred.

As it happens, it is possible to calculate correct combinations of port-open times
and port areas for any motorcycle engine, at any engine speed. The maximum safe speed
for any engine is also calculable, as explained earlier in this chapter, along with
expansion chamber dimensions, carburetor size and many other factors influencing both
maximum power output and overall power characteristics. It should be noted here that
none of the values derived purely from calculations are necessarily optima, and fine
adjustments must always be made experimentally, but it is far better to employ the simple
formulae presented in the chapters to follow than to attempt a purely-experimental
approach. The mathematics involved are not terribly complicated, though sometimes the
arithmetic is laborious, and you can use paper and pencil to arrive at a basic engine/pipe
combination that will be very near the optimum. Much nearer, in fact, than would be
obtained by even the most experienced tuner's unsupported guesswork, and near enough
to a fully developed configuration to minimize the outlay of time and money entailed in
the building of a racing engine. You start by determining, mathematically, an upper limit
for engine speed, then use more math in establishing a maximum for piston-ring
thickness, in establishing all the port dimensions to suit the projected engine speed, in
selecting a carburetor, and in designing an expansion chamber. Suitable values for
compression ratios, both primary and secondary, are provided in the chapters dealing
with crankcase pumping and cylinder heads, respectively, and with the rest of the
material included in this book it all adds up to being a fairly complete engine redesign
manual for the two-stroke engine-fixated “tuner”. My own experience indicates that
engines built along the lines suggested here never fail to deliver high specific horsepower
(which is more than may be said for any cut-and-try system) even without the benefit of
experiment-indicated adjustments. I dislike guesswork, have made a serious effort to
eliminate it from my own projects, and am hopeful that the lessons learned - and outlined
in this text - will reduce the generally high level of guesswork among most
experimenters. If I have forgotten to cover anything, the omission is inadvertent, because
my distaste for

Secret” in the game: “To know what you are doing, and to do it thoroughly.”

“

Speed Secrets is even greater than for guesswork. There is only one

15

Two Stroke TUNER’S HANDBOOK

16

THE CRANK TRAIN

As was noted in the chapter of this book dealing with basics, power output from
an engine of any given displacement is a function of gas pressure in the cylinder during
the power stroke, and the number of power strokes per unit time. Implicit therein is the
suggestion that the horsepower ultimately to be had from an engine has little to do with
port shapes and port timings, exhaust systems, carburetion or indeed any of the things on
which our attention usually is fixed. Why? For one thing, increases in gas pressure bring
corresponding increases in heat flow into the piston - and no high-output two-stroke
engine can operate beyond its thermal limit. Similarly, you cannot increase the rate at
which power strokes occur without increasing crankshaft speeds, with increases in this
direction sooner or later taking you beyond the engine's mechanical limit. The
horsepower you ultimately will extract from any given engine depends therefore very
directly upon your ability to expand those thermal and mechanical limits to the greatest
extent possible, and only then to make the most of the territory thus gained.

THE PISTON

For a very long time subsequent to Dugald Clerk's creation of the two-stroke
engine, the thermal limit was the only limit, but it was enough to hold power output from
such engines to extremely modest levels. Then, as now, it was primarily a limit imposed
by available piston materials. Cast-iron has its advantages in terms of wear resistance,
hot-strength and low thermal expansion rates, and it was used quite frequently in the low
speed engines of years past. Unfortunately, iron is heavy, and heavy is the last thing you
want in a piston - which in modern engines is subjected to accelerations well in excess of

2

100,000-ft/sec
today, is conveniently light, but disagreeably insists on melting at much lower
temperatures than that of the fire to which it is directly exposed. Moreover, it loses
strength very rapidly with increases in temperature above ambient, so that piston failures
do occur at crown temperatures well below the material's melting point. Finally,
aluminum is a high expansion-rate metal, which makes a piston made of it a variable­clearance fit in any cylinder. But aluminum is a very light metal, and that alone was
enough to recommend it for use in pistons, even though the drawbacks listed were
enough to severely limit the specific power outputs attainable with two-stroke engines for
a long time.

Aluminum-based piston alloys improved slowly over the years, with the addition
of small percentages of, say, copper, to improve their hot-strength, but it was not until
means were found to add considerable amounts of silicon that large improvements were
made. Today, the best piston alloys contain between 15- and 25-percent silicon, and this
addition has all but transformed the “aluminum” piston. Add mixtures of silicon in
excess of 15-percent not only drastically reduce aluminum's expansion rate, they also

. Aluminum, used as the primary constituent in virtually all piston alloys

17

Two Stroke TUNER’S HANDBOOK

affect a proportionate increase in hot-strength and improve the piston's wear-resistant
properties. In all of these respects the improvement is large enough to almost exactly
equal the percentage gains in horsepower during the years in which aluminum-silicon
alloys have been in use. I am inclined to think that most of what we consider to be
“modern” improvements in two-stroke engine design - with particular reference to
expansion-chamber type exhaust systems -might have been applied as much as fifty years
ago had good pistons been available. There was little point in such development work
without the aluminum-silicon piston; aluminum or aluminum-copper pistons would melt
at specific power outputs well below what we now consider only average.

With all that, high silicon-content piston alloys still are not universally employed.
As it happens, such alloys do have their disadvantage, which is that they are difficult to
manufacture. Just casting pistons of aluminum-silicon alloy is a task for specialists using
specialized equipment; machining the raw castings into finished pistons is an even more
formidable task. You may encounter this last difficulty if you have occasion to modify a
cylinder cast from the material in question - and you will find that it blunts cutting tools
of any kind with remarkable rapidity. For you, that will be an inconvenience; for the
mass-producer of pistons it is a disaster, as the need for frequent re-sharpening of tool
bits entails losing output from his machinery while such repairs are made, and it means
the expense of the man-hours required for the repairs. Thus, the manufacturer has every
reason to restrict the silicon content of the piston alloys he uses to the minimum required
by the use to which his engines will be put, which is the reason why Yamaha, for
example, uses different alloys for touring and racing pistons.

In point of fact, the Japanese seem to manage high silicon-content pistons better
than anyone else, which may well account for their notable superiority in coaxing power
from two-stroke motorcycle engines. All of the major Japanese manufacturers employ
piston alloys in their touring engines having percentages of silicon high enough to be
considered “racing only” in much of the rest of the world. And, sad to say, many of the
“racing” pistons being offered by speed equipment manufacturers are inferior in this
regard to the ordinary off-the-shelf parts you'll find at your local dealer in Japanese
motorcycles. For that reason, I am inclined to use either stock or “GYT-kit” pistons
when I am working with engines carrying a “made in Japan” label, rather than waste my
money on a specialty replacement. There are, of course, exceptions to this rule, which
evolve principally around ring widths, and I will deal with that in due course.

Unless you happen to be a piston manufacturer, there isn't much you can do about
piston alloys, beyond seeking out pistons having a high silicon content. Neither is there
anything you can do about piston shape - which is most unfortunate, because a piston is
not, as it first appears, simply cylindrical. Even with the use of aluminum-silicon alloys,
pistons do expand as they are heated, and they do not expand at all evenly. The greatest
increase in diameter will occur up at the crown, because that is both the area of maximum

18

THE CRANK TRAIN

mass and highest temperature. So there must be more clearance, measured cold, up at the
piston's crown than is required down around the lower skirt. In fact, clearances vary
continuously from the piston's crown to the bottom of its skirt - and from side to side, as
the piston is elliptical rather than round. Someday, someone may be able, with the help
of a computer, to actually calculate all the clearances and ellipse ratios involved; for the
present they are decided in a process of trial-and-error by even the most experienced of
manufacturers.

Presumably, you will not have the facilities to alter whatever shape your engine's
piston(s) may have, but you can vary running clearances by changing cylinder bore
diameter. The problem here is one of “How much?” and I regret to say that it is a
problem for which there is no convenient solution. Clearances, measured at the piston's
maximum diameter, across its thrust faces, may vary from about 0.002 to as much as

0.007-inch, depending on: the shape and composition of the piston itself; the absolute
cylinder bore diameter; the material from which the cylinder is made, as well as its
configuration; and the thermal loadings to which the piston will be subjected - which will
themselves vary according to gas pressure, fuel mixture, cylinder configuration and the
vehicle's rate of motion. Many people have expressed great faith in rules relating

19

Two Stroke TUNER’S HANDBOOK

clearance to cylinder bore diameter; I have not found the choice to be that simple. If
there is a rule, it would be that you can add perhaps 0.0005- to 0.001-inch to the
clearance recommended by your engine's maker, but even this is a gross over­simplification and I mention it only because it is somewhat better to have too much
clearance than too little. In the former, the excessive clearance adversely influences heat
transfer from the piston to the relatively cooler cylinder walls and may lead to any of the
several unpleasantries associated with overheating the piston, which range from a
tendency for oil to become carbonized in the ring grooves, to the appearance of a large
hole in the piston crown. Too little clearance will reveal itself in the form of scuffing, or
outright seizure - unless the piston is only marginally too tight, in which case the only
symptom of distress will be a power loss in the order of 2- to 3-percent.

Often, in modified engines, you will find that the straightforward increase in
overall piston clearance by slightly enlarging the cylinder bore is not a complete answer.
If the manufacturer has done his work properly, his pistons will, as they expand with
temperature, assume a round shape when the engine is hot. Your problem will be that
with the modifications you have made, more heat will be forced into the piston's crown,
raising its temperature above the level anticipated by the manufacturer, which results in a
completely different set of temperature gradients down the length of the piston.
Specifically, while the whole piston will assume a diameter slightly larger than that
planned for by its maker, the area around the crown will “grow” more than the rest. It
will thus be impossible to correct for the altered conditions simply by honing the cylinder
bore larger, for if you enlarge the bore enough to provide running clearance for the top of
the piston, its skirt will be given too much clearance (leading to rocking, and trouble with
the rings). In such cases, which are not the exception, but the rule, the solution is to
machine what is called a “clearance band” around the top of the piston. Usually, this
band will extend down from the crown to a point about 0.125-inch below the ring groove,
or grooves, and the piston's diameter reduced by perhaps 0.002-inch over the entire
band's width. Although the clearance band is not a particularly clean solution to the
piston-expansion problem, it is one that can be applied by anyone with access to a lathe,
and it has one advantage over the generally more desirable “pure” contouring of the
piston: if a piston with a clearance band seizes partially, aluminum will not be smeared
above and below the ring groove - an event which will lock the ring in its groove and
upset its ability to seal against gas pressure, In practical terms, this means that the
clearance-banded piston will absorb a lot of punishment before it is damaged sufficiently
to cause retirement from a race.

Excessive deep clearance bands must be avoided, for they expose the sealing ring
to too much heat, and heat has a devastating effect on the service life of a piston ring.
But for these effects, there would be every reason to locate the ring as close to the piston
crown as is mechanically possible, because we would then obtain the cleanest opening

20