HEAD / CYLINDER SEALING (2)

Nominal compression ratios, as I have said before, have little meaning in high-output two-stroke engines. However, you can work with trapped compression ratios almost as effectively as by measuring cranking pressures. An engine's trapped compression ratio is the ratio between the cylinder volume at the moment of the exhaust port's closing and the volume with the piston at the top of its stroke. To find this, you must first measure the combustion chamber volume, with the piston in position at top center. The job can be done with the engine assembled, using n graduated cylinder and pouring in oil until the level comes up to the spark plug hole. Or you can calculate the volume. When the combustion chamber has a simple shape (part-spherical, conical or cylindrical) I prefer to do the job by calculation, but more complex shapes send me scurrying for a can of oil and a graduated cylinder. In fact, the process of actual measurement may appeal to you as a regular thing, because you will need a graduated cylinder for more than this single task, and a slide-rule may not be a part of your basic equipment. In any case, remember when figuring the compression ratio, that it is not the ratio between piston displacement and combustion chamber volume, but between cylinder volumes from the point of exhaust port closing to top center, as in the following formula :

?????? At site
Where CR is compression ratio
V1 is cylinder volume at exhaust closing
V2 is combustion chamber volume
Traditionally, compression ratios have been measured "full stroke". That is to say, V, would represent the combustion chamber volume plus piston displacement from bottom center to tog center. Thus, a combustion chamber volume of 28cc and a piston displacement of 250cc, calculated full-stroke, would be
?????? At site
CR= 9.93:1

But a far more realistic figure is obtained when V1 represents the cylinder volume above the upper edge of the exhaust port, and if we assume that our hypothetical engine has an exhaust port height equal to 45-percent of stroke, then V1 becomes 55-percent of piston displacement plus V2, and calculation goes like this:
?????? At site
CR= 5.91:1

Coincidentally, that compression ratio (5.91:l) is very nearly all a non-squish combustion chamber will permit in an otherwise fully-developed two-stroke engine. With small-bore engines you may push the compression ratio up to perhaps 6.5:l without serious consequences, using a non-squish cylinder head, but that is very near the limit. Good squish-band cylinder heads, on the other hand, permit compression ratios up to as much ns 9.5:l in motocross engines with exhaust systems that provide a wide boost without any substantial peaks, but for road racing engines I cannot recommend anything above 8.5:l even when unit cylinder size is only 125cc. You will find that higher compression ratios than those suggested can produce marvelously impressive flash readings on a dynamometer; as soon as the engine has a chance to get up to full temperature, the output will drop well below that sustained by an otherwise identical engine with a lower compression ratio. Sustained, and not flash horsepower, is what wins races.

HEAD / CYLINDER SEALING

A major problem with cylinder heads on high-output engines that began life as low output engines is persistent leakage around the head/cylinder joint. The combined increases in temperature and pressure seem always to be too much for the joint, and you will find evidence of fire puffing past on the surfaces after disassembly even though you may not have observed anything out of the ordinary when the engine was running. This leaking will occur even if you have retained the engine's stock compression ratio, and it may become very serious if the head has been thinned to get a compression ratio increase. Many manufacturers, perhaps most, feel some awful compulsion to skimp on section thicknesses when they make a cylinder head, a habit that often stands revealed as a questionable economy when you test their handiwork on a dynamometer: first, the thin sections often do not have the cross-sectional area required to transfer heat away from the head's lower surface quickly enough to keep the spark plug temperatures stabilized; second, most of these cylinder heads are secured to their cylinders by only four widely-spaced bolts, which presumes heavily on their beam-strength to maintain a tight seal at the joint.
This last situation becomes especially marginal when metal has been machined away to raise the engine's compression ratio, and the stock head gasket (usually cut from light-gauge, soft aluminum) will in many cases not be strong enough to hold even the pressure increases involved in a simple switching of exhaust systems. Shave the head (which both weakens the head's beam strength and increases the forces acting upon it) and you'll very likely find that it becomes impossible to hold the head/cylinder seal- the gasket will fail after only minutes of running. Also, attempting to use the stock cylinder head, in either standard or modified form, often will increase the heat input around the spark plug to such extent that the engine becomes impossibly fussy about plug heat range. Use a plug cold enough to avoid trouble at maximum output, and it will foul at anything less than full-throttle operation. There is nothing like masses of metal to equalize the temperature gradients through the cylinder head, and - sad to say - those masses are not provided in many stock cylinder heads.
Cylinder head design also can strongly affect overall cylinder cooling. When the cylinder head's lower surface is cooler than the cylinder itself, heat will be drawn away from the latter; conversely, a cylinder head can also put heat into the cylinder if the situation is reversed. All things considered, the engine's best interests probably are served by isolating, to such extent as is possible, the cylinder and head- which means restricting the contact area at the cylinder/head joint to a narrow sealing band which bulges to encompass the hold-down bolts, or studs. In that way any cooling problems will be isolated, and can be dealt with separately. That, of course, assumes that it will be possible to improve cylinder cooling should such improvement become necessary. Actually, making a new cylinder head is fairly easy (it can be either cast or simply machined from a block of aluminum) while the cylinder itself presents a far more difficult problem in fabrication. So you may very well want to use an oversized, deeply-finned cylinder head to help cool a particular engine's stock, cast-iron cylinder. And if that should be the case, remember that you'll need a maximum contact area between head and barrel, and surfaces that will seal without any kind of gasket. There is a very sharp temperature gradient across any joint, and even a solid copper gasket presents one more pair of surfaces across which heat must flow.You may find that providing a seal between the head and barrel is one of the more difficult facets of the overall job. As I have said, stock aluminum gaskets are almost certain to fail, being a bit weak at ambient temperatures anyway -and impossibly frail at the temperatures to which they will be subjected. Copper is a better material, for while it is nearly as soft as aluminum at ambient, its hot-strength properties are better. Copper is soft enough to make a good gasket in the annealed state, but hardens in use, and must be re-annealed frequently to keep it soft and thus retain its properties as a gasket. Brass should never be used as a gasket material, but steel may be used if it is very thin and has one or more corrugations rolled, in rings, around the bore - in the manner of the head

?????? At site
gaskets used in some automobile engines. You can also get a good seal by machining a narrow groove in the cylinder's upper face and inserting in it a soft copper ring (made from wire) to bear against the head's lower surface. Other, even better seals may be had with gas-filled metal O-rings, piston rings (they'll work here, too) and one of the best sealing rings I've seen has a V-shaped section, laid on its side, with the V's point aimed away from the bore. Gas pressure tries to force the V open, bringing one arm to seal down against the cylinder while the other is pressed against the cylinder head. Another sealing ring that works in roughly the same fashion is a hollow metal O-ring with vent- holes drilled through from its inner diameter to admit gas pressure from the cylinder - which expands it outward and thus creates a seal even between somewhat uneven surfaces.

PLUG LOCATION (4)

There is another solution to the problem that has nothing whatever to do with the cylinder head: you simply add metal to the piston crown, and that, too, will tend to equalize skirt temperatures - but it also makes the piston heavier. Even so, it is a solution much-loved by manufacturers, as adding thickness in the piston costs virtually nothing, while any departure from symmetry in combustion chamber configuration entails multiple machining operations (it being extremely difficult to cast, with sufficient accuracy, the combustion chamber's small volume) and machining-time is expensive.
There may be another reason for employing an asymmetrical combustion chamber, and/or relocating the spark plug from its normal position over the bore axis. In loop-scavenged two-stroke engines, the fresh charge is directed upward, and at, the rear cylinder wall, as it emerges from the transfer ports. Ideally, the mixture streams converge and sweep up and over at the top of the cylinder to clear away exhaust products and push them out the exhaust port, following the rear cylinder wall upward, and then curling back smoothly under the cylinder head. In practice, the scavenging stream tends to be much less ordered in its habits, and the general turbulence can make it leap and dodge all over the place, impinging strongly at one point and only eddying at others. This leads, in some engines, to a reshaping and repositioning of the combustion pocket - the purpose of such changes being to aid scavenging by using the combustion chamber's form to give the scavenging stream direction.
In such cases, the spark plug may also be moved to a position where it will be washed by the mixture stream, which tends to cool the plug between firings, and thus make the engine somewhat less sensitive to plug heat range. Also, as noted before, the plug may be moved away from the combustion chamber center to create a slightly longer path for flame travel, which lowers the rate at which pressure in the cylinder rises during the combustion process and, in some instances, makes for smoother running. To a lesser extent, the same treatment may be used to combat a tendency toward detonation, as the lower pressure-rise rate gives all the pockets of end-gases time to lose their heat into the surrounding metal. This last effect is, of course, better obtained with a conical combustion chamber, rather than by offsetting the plug. Incidentally, moving the spark plug over too close to any edge of the bore is usually poor practice: At times, particularly when starting from cold, the piston ring will scrape oil off the cylinder walls and pitch it up at the cylinder head, and if you place the spark plug in the line of fire, it definitely will show a weakness for oil-fouling.

PLUG LOCATION (3)

Not all cylinder heads have their spark plugs and combustion chamber pockets centered over the cylinder bore, and there are good reasons for most of the variations in form one sees in the products of the major manufacturers: For instance, piston crown temperatures seldom are even, and while the overall temperature distribution pattern is understandably inclined toward maximums in the center of the crown, circumstance can also lend a bias toward the exhaust port. That bias comes not from any heat-input pattern, but rather from the manner in which the piston crown is cooled- by heat transference into the air/fuel mixture below, and into the piston skirt, from whence it is transferred out into the cylinder walls. Cooling provided by the turbulent crankcase charge is more or less even; the same cannot be said of heat losses into the cylinder, for the temperature gradients around the cylinder's walls are most uneven. The area around the exhaust port is hotter than that back at the intake port, even though the exhaust-side of the cylinder is in most instances the recipient of the direct cooling-air blast. Moreover, the exhaust-port side of the piston skirt is bathed in fire every time the port opens at the end of a power stroke. The overall result is to move the maximum temperature point on the piston crown toward the exhaust port.
Now, when that maximum temperature bias begins to seriously overheat the side of the piston, you are likely to see some severe piston ring problems develop: Too-high temperatures will eventually be a disaster for the ring itself, but more often it will not have a chance to show its displeasure because another disastrous situation will already have developed, with the lubricating oil. Sometimes, if a relatively high ash-content or inadequately de-gummed oil is used, the ring will be glued solidly in its groove by varnish and carbonized oils. More often, the temperatures prevailing in that section of the piston skirt adjacent to the exhaust port will cause a breakdown of the oil film in that area and the piston will seize. And this can happen even though a generous margin of safety still exists all around the rest of the piston skirt. A common, and highly sensible solution to this problem is to move the combustion chamber pocket away from the bore axis, toward the back (inlet) side of the cylinder. This measure shrouds more of the piston crown's exhaust side under the squish band - which becomes crescent-shaped, instead of being a symmetrical ring - and reduces heat input there from combustion (the skirt will still be getting plenty of heat when the exhaust port opens) enough to provide a more even distribution of heat around the piston skirt. Then, with piston-skirt temperatures evened-out, a slightly higher compression ratio may be used without incurring seizure, or localized overheating of the piston ring.

PLUG LOCATION (2)

Fortunately, most engines usually are relatively insensitive to plug location as long as the gap isn't moved too close to the piston. Which raises an interesting point: The common practice of shaving material from the cylinder head's lower surface not only raises the compression ratio, and thus the thermal load on the piston, but it brings the plug gap close to the piston crown -compounding the problem. A better approach to obtaining

?????? At site
?????? At site
increases in compression ratio is to purchase a cylinder head developed to do the job properly. Yamaha's GYT-kit heads, for example, provide the right compression boost, have their spark plugs properly located, etc. Other made-for-the-job cylinder heads offer the same fundamental advantage, which is that you get to buy a lot of other people's engineering at a very low cost.

PLUG LOCATION

Tests have shown that the best location for the spark plug is, by and large, squarely in the center of the combustion chamber, and with its gap as close to the center of the volume of trapped mixture as possible - which is logical, as that position provides the shortest flame travel in all directions. However, a number of other considerations do intrude. First, the plug gap will necessarily be at the periphery of any partly-spherical chamber, and not at its center, and trying to form a knob in the chamber roof - to move the plug deeper into the mixture volume -will upset the chamber's surface/volume ratio. Secondly, moving the plug too close to the piston seems to cause a local overheating of the piston crown, which can impose an unnecessarily low ceiling on compression ratio.
This last consideration has, in many instances, led development engineers to use combustion chambers with forms that allow the plug to be positioned well away from the piston: modified spheroids; conical sections, etc. Also, chambers with higher roofs (like those shaped as cones) with their spark plugs up at the top and the broader base down at the piston, provide a slightly slower pressure rise as combustion progresses, and are in consequence a bit more kind to bearings. Other switches in plug location may be made in the interest of easing the job of plug replacement: it is difficult to change a plug centered in the cylinder head when the bottom of a fuel tank, or frame tube, is directly overhead.

SQUISH BANDS (7)

The clearance space between piston and cylinder head must be enough to avoid contact at high engine speeds, yet close enough to keep the mixture held there cooled during the combustion process. This vertical clearance between squish band and piston should not be greater than 0.060-inch, and it is my opinion that the minimum should be only barely enough to prevent contact -usually about 0.015-inch in small engines (with tight bearings and cylinder/rod combinations that do not grow, with heat, disproportionately) and up to about 0.045-inch in big engines.
Some disagreement exists as to the validity of claims that the squish band aids combustion by causing turbulence in the combustion chamber as a result of the piston "squishing" part of the charge between itself and the head. I don't know about that, but I do know that holding squish band clearance to a minimum means that there will be the smallest volume of end-gases escaping the combustion process, and that can be more important than you might think. For example, a 250cc cylinder with a full-stroke compression ratio of 10:1 will pack its entire air/fuel charge into a volume of only 28cc by the time its piston reaches top center. Assuming that it has a 3-inch bore, and a 50-percent squish band with a piston/head clearance of .045-inch, then the volume of the charge hiding in the squish area will be in the order of 2.6cc, or almost 10-percent of the total. That can be reduced to 5-percent merely by closing the squish band's clearance to 0.020-inch - and you'll never find an easier 5-percent horsepower difference. True, the difference measured at the crankshaft might prove to be more like 2-1/2-percent, but the addition of those small percentages can make a very large final difference.

SQUISH BANDS (6)

But if you want to use a true (measured from exhaust-closing) compression ratio much over 6.5:1, on a high-output engine, combustion control beyond that afforded by a non-squish cylinder head will be necessary. Considerable variation is possible, but a

?????? At site
good rule to follow is to make the cylinder head's squish band about 50-percent of the cylinder bore area. For example, in a 3-inch bore -which has a total area of 7.07-inches2 the squish band would be wide enough to represent an area of just about 3.5 in2. Assuming that you have centered the combustion chamber proper on the bore axis, then your squish band would be a ring having the same outer diameter as the bore, and an inner diameter of just over 2-inches. The combustion chamber itself, to meet the previously-stated minimum surface/volume requirement, would again be a spherical segment - with a radius that provides the total volume, added with that from the clearance space between piston and squish band, to give the desired compression ratio.

SQUISH BANDS (5)

No matter what the compression ratio you ultimately use, it will have been influenced much more than you probably suspect by the combustion chamber configuration, and by certain gross characteristics of the head itself. Over the years, I have seen the fashion in combustion chamber forms swing back and forth, hither and yon, with first hat-section chambers in favor and then trench-type chambers, and torus-type chambers and so on and so forth ad infinitum. I was not, and am not, impressed. Combustion chamber form should be established with an eye toward only a very few special considerations, and these cannot account for even half the chamber shapes I have seen. Listed, though not really in order of importance, these are: surface / volume ratio; spark plug location; thermal loadings; and combustion control. We will consider each of these in turn.Surface to volume ratio is important because even in the part of the combustion chamber fully exposed to the advancing flame front, there will be a mixture layer adhering to the metal surfaces that does not burn. These layers, like that trapped within the squish band, are cooled by their proximity with the cylinder head, or piston, and simply never will reach ignition temperature. And, like the end-gases from the squish band, they eventually find their way out the exhaust port, having taken no part in the conversion of fuel and air into horsepower. Thus, the best combustion chamber shape - taken strictly from the standpoint of surface/volume ratio - would be a simple spherical segment sweeping in a continuous arc from one side of the cylinder bore to the opposite side. No tricky changes in section, no squish bands, no nothing. And that is, in point of fact, precisely the shape employed in nearly all non-squish cylinder heads

SQUISH BANDS (4)

In an engine intended purely for road racing, with a torque peak virtually coincidental with its power peak and driving through a very close-ratio transmission (enabling the rider to hold engine-speed within narrow limits), making this beggar's choice is a fairly straight-forward proposition: you play with jetting until the motorcycle runs fast. However, road racing conditions allow you to stay right on the mixture-requirement hump; you don't have to worry about what happens two-thousand revs below the power peak, because that's below what you'll use in a race. Motocross racing is another matter entirely, and an engine with a mixture-curve hump will drive you absolutely mad. Jet a motocross engine so that it doesn't melt a piston every time it pulls hard at its torque peak, and (if its mixture-curve is humped) it will be huffing soot and losing power above and below that speed.
The answer to this problem is to iron out that mixture-requirement hump, because no matter how much work you do with the carburetor, it never will be able to cope with the engine's needs. All the carburetor knows, really, is how much air is moving through its throat, and it adds fuel to the air in proportion to the rate of air-flow; don't expect it to know when the piston is getting hot and respond by heaving in some more fuel. How do you get rid of the hump? You do it mostly by substituting a somewhat less effective expansion chamber: one that gives more nearly the same boost all the way through the speed range you are obliged to use by racing conditions, without any big surges. That will result in a drop in peak power, obviously, but you can compensate for it to a considerable extent with the higher compression ratio you previously were forced to forego in the interest of keeping the piston crown intact when the expansion chamber did its big-boost routine. Again, it is all a matter of finding the balance.

SQUISH BANDS (3)

One of the most undesirable side-effects that comes with too-high compression ratios is an enormous difficulty in getting an engine to "carburet" cleanly. When the compression ratio is too high, you'll find that an engine's mixture-strength requirement has a sharp hump right at its torque peak that no motorcycle carburetor can accommodate. You'll realize, after working with high-output two-stroke engines, that all of them are to some degree liquid-cooled - and that the cooling liquid is gasoline. It is

?????? At site
true that an over-rich mixture tends to dampen the combustion process, and reduce power, but here again we find ourselves faced with the necessity for finding a balance between evils: We have overheating to rob power on one side, and we can cool the engine with gasoline, but too much fuel also robs power. The solution is a beggar's choice, in which we try to find the cross-over point between overheating and over-rich mixtures.

SQUISH BANDS (2)

Our application here, of course, is strongly biased toward maximum horsepower, and that points toward a squish-band head - which is what you will have in most motorcycles in any case. I will warn you, now, that it may be unwise to follow the old-time tuner's habit of increasing an engine's compression ratio as an opening gambit in the quest for better performance. Indeed, before your work is done you may find it necessary to reduce your engine's compression ratio below the stock specification. You see, in the final analysis it is not so much compression ratio as combustion chamber pressure that determines the limit - and these are not at all the same things. Your stock engine, with a carburetor size and porting chosen to lend it a smooth idle and easy starting, does a much less effective job of cylinder-filling than will be the case after it has been modified. More important, it will probably have an exhaust system that has more to recommend it as a silencer than as a booster of horsepower. These factors, in combination, make a very great difference between the cylinder pressures at the time of ignition in the stock and modified engine. Even given a certain willingness on your part to use a fairly cold spark plug - changing it frequently - and a further willingness to replace pistons and bearings more often in payment for added power, it may still be necessary to stay with the stock specification for compression ratio. Or, as I have said, to lower the engine's compression ratio from the stock condition. This last will be particularly true if you succeed in creating a much better than stock exhaust system.
By and large, you would be well-advised to ignore the whole business of compression ratios in favor of cranking pressures. There is, after all, a big difference between the kinds of numbers you get by performing the traditional calculations to find compression ratio, and what is happening as the engine turns. My experience has been that you can use cranking pressures of 120 psi without worrying much about overheating anything. Maximum power will be obtained at cranking pressures somewhere between 135 and 165 psi. Going higher with compression, in a conventional motorcycle engine, can give a neat boost in low speed torque, but the thermal repercussions of higher cranking pressures will surely limit maximum output. On the other hand, fan-cooled kart engines perform very well at cranking pressures up at 200 psi, and water cooled engines behave much the same.

SQUISH BANDS

Ricardo solved the problem, once he had determined its nature, by lowering the underside of the cylinder head in that part of the chamber over the piston. Thus, most of the mixture was concentrated right at the ignition source, and would be more likely to burn without detonating. The small part of the mixture caught between the cylinder head's squish band and the piston was still subject to compression heating, but was fairly effectively shielded from radiation and was, moreover, spread in such a thin layer that it would resist ignition from any cause - as it would lose heat into the relatively cool piston and cylinder head too fast to ignite.
That still is the secret of the squish-type cylinder head: It concentrates the main charge in a tight pocket under the spark plug, and spreads the mixture at the cylinder-bore's edges too thinly to be heated to the point of ignition. These “end gases” do not burn with the main charge, and are only partly consumed as the piston moves away from top center and releases them from their cooling contact with the surrounding metal. And right there is the disadvantage that comes with the squish-band cylinder head, for mixture that does not burn is mixture that contributes nothing to power output. Of lesser importance, though only in this context, is that those end-gases contribute heavily to the release of unburned hydrocarbons out the exhaust pipe and into the atmosphere, and for that reason automobile manufacturers are now relying much less heavily on squish-band chambers for combustion control. You may be interested to know, too, that in many cases a non-squish combustion chamber, with its complete utilization of the mixture to offset the power-limiting effects of a necessarily-lower compression ratio, has proven to be best in absolute terms of power and economy. McCulloch, for example, make engines with both squish and non-squish cylinder head configurations - having found that both have their applications.

THE COMBUSTION PROCESS (3)

We have England's Harry Ricardo to thank for this type combustion chamber, which he created to cope with conditions that ceased to exist long before most of us were born. During the conflict that wracked Europe just after the turn of this century, there were not only shortages of internal combustion engine fuels, but the fuels available were of very poor quality – and would detonate severely in the side-valve engines of that

?????? At site
period unless the engines were operated with a much-retarded spark, or their compression ratios lowered to about 4:1, or both. These measures had a terrible effect on fuel economy, naturally, and the problem led Ricardo to do serious research into the nature of detonation. We now know that the side-valve engine is particularly prone to detonation, as it of necessity has a very long combustion chamber. Ignite a fire at one end, and it will be a long while reaching the far corners of the chamber. In the interval between ignition and the completion of burning there is ample opportunity for the unburned part of the charge to overheat and ignite.

THE COMBUSTION PROCESS (2)

Should this detonation continue, it will overheat the engine's upper end to the point where ignition occurs before there is a spark: compression heats the mixture in any case, and when a lot more heat is added from the piston crown, etc., the mixture will be

?????? At site
brought to “pre-ignite”. Detonation has a very bad effect on power output; pre-ignition (thought by some to be the same phenomena) is even worse in that regard, but will not long continue unnoticed as it will very rapidly overload the piston - in both the thermal and mechanical sense - beyond the point of failure. Knowing that, you will appreciate that detonation is to be avoided if at all possible. One way to avoid detonation would be to simply hold the compression ratio to some very low number, as they would reduce the pre-combustion temperatures and thereby make detonation unlikely if not impossible. But that method is mostly (the exception I will deal with shortly) too expensive in terms of power-output efficiency. A better method is one employed in most engines today: use of a "squish type combustion chamber, in which the mixture is trapped in a small pocket under the spark plug, and the rest of the cylinder head surface over the bore is made to fit closely against the piston crown when the piston is at top center.

THE COMBUSTION PROCESS

Not too surprisingly, the equilibrium described is influenced by combustion chamber design-as is the point at which smooth burning gives way to the outright explosions we call detonation. This aspect, too, is widely appreciated, but not widely understood. In truth, most people have very little understanding of the events that follow ignition; events that are highly complex if studied with regard to their chemistry but really quite straightforward taken in less narrow terms. Much of the misunderstanding that exists has been created by the popular press, which insists upon saying that a piston is driven downward on its power stroke by a burning mixture. In reality, the burning of fuel in the cylinder is simply a means of raising the temperature of the working gas (air; actually a mixture of gases) and thereby raising its pressure. This relationship was formulated long ago by Boyle as:

?????? At site
Where, of course, P is pressure and T is temperature. The whole business gets complicated in the internal combustion engine by the changes in the cylinder's contents due to the combination of elements in the working gas with fuel, but it still basically is a case of raising the working gases' temperature and thus raising their pressure, and it is that which pushes the piston down and makes the horsepower. In fact, burning will have been all but completed by the time the piston starts downward on its power stroke.
Here, for anyone who cares, is what happens from the moment of ignition: Several thousandths of an inch of travel before the piston reaches the top of its compression stroke, representing somewhere between 20- and 45-degrees of crank rotation, the trapped air/fuel charge is ignited by the spark plug and burning commences. At first, the process proceeds quite slowly (relative to subsequent crank rotation before TDC). A small bubble of fire expands gently away from the point of ignition between the spark plug's electrode and ground wire, and if all combustion were to continue at this pace it would hardly be completed in time for the following compression stroke. However, this small flame quickly heats the remaining mixture enough to enormously increase the rate at which burning occurs, and after the initial delay, the flame-front accelerates outward from its point of origin with ever-increasing rapidity - sweeping throughout the combustion chamber. And if the engine has been given the proper amount of spark advance, the piston will have just moved up to the top of its stroke as the rapid phase of combustion begins, so that the bulk of the burning is done while the piston is virtually stopped and the mixture compressed to minimum volume. By the time the crankshaft has rotated a few more degrees, and the piston is once again moving downward, the combustion process will have been almost entirely completed.
The preceding is what happens in the normal course of events; combustion does not always occur that neatly. The most common, regrettable combustion irregularity is detonation, the harsh knocking you hear just before an engine seizes, or melts a piston - and the noise you would hear, when running an engine on a dynamometer, as the needle on the scale begins an ominous retreat. Unhappily, the very process by which the mixture in the combustion chamber is re-heated before its actual contact with the flame-front advancing from the spark plug, and rapid combustion thus made possible, is the process that may also lead to the sudden explosion of the combustion chamber's contents that we call detonation. Here's how it happens: It has already been noted that as the flame-front advances, the combustion chamber's remaining unburned mixture is heated, and this heating is caused not only by direct contact with the flame, but also by radiation and the overall pressure rise within the chamber. If the temperature of this remaining mixture is raised to its ignition point, all of it is consumed at the same instant in a single explosion. This explosion creates a shock, due to a fantastically rapid pressure rise, that strikes out against all its surroundings hard enough to make detonation’s characteristic knock - and it is a shock with a force often sufficient to break the spark plug insulator's tip and damage both the piston and bearings. Even so, its worst effect is to force a lot of heat out into the piston, cylinder head and the cylinder walls. These are thus brought to abnormally high temperature, which tends to overheat the next air/fuel charge and make it detonate even more quickly and severely.

CYLINDERHEADS

For the Otto-cycle engine, of which the two-stroke is an example, there is a theoretical level of efficiency, in terms of converting heat into work, referred to in basic engineering texts as “air standard efficiency”. In this, it is assumed that the cylinder is filled only with dry air, and heat then added, which ignores the fact that in practice the air contains some moisture and a considerable percentage of hydrocarbon fuel. Even so, this theoretical level of efficiency, calculated against compression ratio, provides a useful yardstick against which actual efficiency can be measured - and it tells us a lot about the effects, on power output, of compression ratio. For example, at a compression ratio of 5:1, air standard efficiency is 47.5-percent, while at 10:1, it is 60.2-percent. That is, of course, a very great gain, and the consequences - measured at an engine's output shaft-are the reason for many experimenters' fixation on “raising the compression”. Certainly, increases in compression ratio, which may be accomplished simply by trimming a few thousandths of an inch from the cylinder head's lower surface, can work minor miracles with an engine's performance.
But higher compression ratios can also bring about a mechanical disaster: improvements in power gained in this manner are purchased at a disproportionate cost in peak cylinder pressure, leading to reduced bearing life and sometimes to an outright failure of a connecting rod or crankpin. Moreover, because the higher pressures are reflected in a proportionately greater side thrust at the piston, frictional losses are such that net power gains are always less than the improvement one would expect from the calculated air standard efficiency. Finally, heat flow from the combustion gases into the surrounding vessel (piston crown, cylinder head, and cylinder walls) rises increasingly sharply with compression ratio, so that a number of thermal-related problems intrude into the already complicated relationship between compression ratio and power.
The worst of these problems is the overheating of the piston crown. A too high compression ratio will raise piston crown temperatures to the point where heating of the mixture below the piston, in the crankcase, reduces the weight of the charge ultimately trapped in the cylinder during the compression stroke to such extent that net power suffers -no matter what Mr. Otto's air standard efficiency formula may say. And if the compression ratio is high enough, heat input into the piston may raise the crown temperature to the point where detonation and then pre-ignition occur. These phenomena will, in turn, very quickly further raise piston crown temperature to such extent that the piston material loses enough of its strength to yield to the gas pressure above – the piston crown then becoming either concave (which drops the compression ratio to a tolerable level) or develops a large hole (and that reduces the compression ratio to zero:zero).
Many people have encountered this last effect, and the tuner's one-time favorite ploy of “milling the head” has fallen into disrepute. But it also is possible to encounter trouble without recognizing it: There is a delicate balance between gains from increased compression ratios and losses due to increased temperatures -which appear not only at the piston's interior, but also throughout the crankcase, crankshaft, rod and all the rest of the engine's interior contacted by the air/fuel mixture. When these parts are hotter, the mixture's temperature is also raised, along with its free volume. Thus, the mixture's temperature-induced efforts to expand inevitably force part of it out the exhaust port, and as power is related very closely to the weight of the charge captured in the cylinder, this heating shows up as a power loss. The trick is to balance crankcase heating and compression ratio. There is an optimum combination for every set of conditions, but finding that optimum without heat-sensing equipment and a dynamometer is exceedingly difficult.

CRANK ASSEMBLY

There are gains in power and reliability to be had from carefully aligning your crankshaft and main bearing bores, and in getting the cylinder axis precisely perpendicular to the crankshaft. As it happens, there is more variation in production tolerances when the various parts of a crankshaft are made than can comfortably be tolerated in a racing engine. Crankpin holes in flywheels are not all precisely the same distance from the main shaft axis; factories "select-fit" these parts, and you can be fairly certain that a new crankshaft is true, but if you manage to ruin any of its flywheels, do not assume that a replacement flywheel, selected at random from the nearest parts bin, will be a satisfactory replacement. Crankpin holes, in facing flywheels, should be matched to within 0.0002-inch with regard to their offset from the main shaft. If your local source cannot supply a single replacement wheel within that tolerance limit, I strongly urge that you purchase a complete, new crankshaft - with flywheels matched at the factory. And when rebuilding a crankshaft, with new crankpins and bearings, be certain that it is aligned to at least the tolerances suggested by the manufacturer's workshop manual. Also, check your crankcases for main bearing-bore alignment - and, more important yet, that the cylinder is exactly perpendicular with the crank axis, for any tilting will be reflected in added friction in the hearings (especially at the thrust washers) and in the piston itself.
Do not attempt to second-guess the manufacturer with regard to crankshaft and crankpin bearings unless you have very specialized knowledge in this field or can obtain the advice of someone who is an expert. Main bearings, particularly, should not be replaced with just anything that will fit, as n very special kind of bearing is employed in these applications, with clearances to accommodate the expansion and contraction of aluminum bearing housings. And the same cautionary note must be added with regard to crankshaft seals, which in the high-speed, two-stroke engine must survive extremes in temperatures and rotational speeds with very scanty lubrication. Not so very long ago, seal failures were common, but now that means have been found to Teflon-coat seal's lips, trouble is usually encountered only when the seals have been damaged in the course of installation. So handle the seals carefully, and pre-coat them with a good high-temperature grease before assembling your engine. You can also improve their reliability somewhat by polishing the area on the main shafts against which they hear to a glassy finish. The seals themselves will polish the shaft eventually, but at considerable expense to their working life.
By and large, problems with piston, connecting rod bearings, crankshaft and seals can be avoided simply by following the recommendations made in the manufacturer's shop manual. The single exception to this is in the fit between piston and wristpin, for the very high temperatures in a modified engine tend to cause a breakdown in the lubrication between pin and piston. Trouble can be avoided in the racing engine if the wristpin is a light, sliding fit through the piston; it should slip through of its own weight, without forcing, for if it is tight enough so that you have to tap it through with a mallet, you eventually may have to remove it with a hydraulic press. Too-light fits may be corrected by using an old wristpin as a lap, and a dash of some fine, non-imbedding lapping compound to polish out the piston's pin-bore to size.

WRISTPIN/CRANKPIN BEARINGS (4)

Connecting rods should not be lightened, or even polished, unless you intend going all the way in this direction and will finish the job by having the part shot-peened. Forgings acquire a tough skin in the process of being pounded into shape, and I know of instances where connecting rods that were entirely satisfactory in standard condition promptly broke after having been polished. I do think, on the other hand, that there is a margin of safety to be gained by smoothing off the rough edges where the flash has been sheared away from the forgings. Notches are, in the engineer's language, “stress raisers” and you can do the connecting rod no harm in removing them. Lightening the connecting rod is, however, a poor choice of ways to use one's time, because a rod intended for the loads at, say, 8000 rpm is going to be overstressed at 10,000 rpm and if anything, material should be added to the rod, not removed. On the other hand, one sometimes can improve bearing reliability by opening slightly the oil channels at the ends of the connecting rod. I do not recommend that you actually cut into the bearing surface, but oil delivery to the bearing will be improved by tapering the entry. Do not extend the taper all the way to the bearing surface, as the sharp edges thus formed will flake away as the engine runs and cause a bearing failure.Crankshaft main bearings seldom are troublesome, except in engines that have

?????? At site
been in storage for a long time and have had corrosion at work in these bearings -or unless the bearings have been mishandled. Bearing steels are very tough, but you definitely can pound small pits in the races by injudicious use of a hammer, and pits also can be formed by rusting. Bearings damaged in either fashion should be replaced, as the pits will soon spread and become minor trenches, as a result of an activity called “Brinelling”, which actually is a form of work-hardening. The bearing's rollers and races have casehardened surfaces, but the metal under this thin case is relatively soft, and it is compressed and released (at any given point) as the bearing turns under a load. If the load is high enough, or the bearing in service long enough, the repeated compressions will literally fatigue the metal, and tiny particles of the surface will start flaking away - which becomes visible as the “tracking” seen in the races of a worn-out bearing. Any bearing will start flaking at some point in its life; bearings with races damages by rust, etc. will begin such flaking almost immediately. Incidentally, in very highly loaded bearings the flaking may be started by the sharp edges around any interruption in the bearing's surface, if the rollers pass over those edges. Oiling slots in the rod's big-end are prone to develop this kind of failure, and the same sort of flaking is sometimes observed around the oil feed holes in the crankpins of engines equipped with “direct-injection” oiling systems, like the Suzuki’s and Kawasaki’s. Remove the sharp edges, and you remove the problem - if any. There is sufficient margin of strength in stock production engines so that the problem does not occur; you may find it in the course of reaching for crank speeds substantially above the stock specification.
Somebody is always telling me about having an engine “balanced”, and I always smile nastily when the engine in question has fewer than four cylinders. In point of fact, the single-cylinder motorcycle engine cannot be brought into dynamic balance, for if you counterweight the crankshaft to compensate for the full weight of the piston and rod, you will simply have moved the shaking force from being in-plane with the cylinder axis 90-degrees. "Balancing" one of these engines consists of finding a balance factor, in percentage of reciprocating mass, which is kind to the engine's main bearings and does not excite resonance in the motorcycle's frame. In-line twin- and three-cylinder engines always have a rocking couple. By and large, the stock crankshaft counter-weighting will be correct for most applications, and unless you want to get into a really lengthy experimental program there is nothing to be gained in making changes.

WRISTPIN/CRANKPIN BEARINGS (3)

Crankpin bearing failures also stem from the use of excessively heavy bearing cages. Sheer rotational speed is not enough to burst a cage of such small diameter and mass, but the fact that the cage must accelerate and decelerate, relative to the crankpin as the connecting rod swings, will cause difficulties unless the bearing cage is very light. In effect, the rollers must push the cage up to speed and then slow it, and if the cage has enough inertia it will resist this pushing and pulling enough to skid the rollers - at which

?????? At site
point they momentarily become a plain bearing- a job for which they are poorly constituted. The skidding rollers generate a lot of heat, through friction, and the heat leads the bearing into the same deteriorating cycle to outright failure as was outlined for the thrust washers. Most modern engines have steel crankpin bearing cages, copper- or tin-plated to provide a low-friction surface to bear against the rollers, crankpin and connecting rod eye. These replace the phosphor-bronze cages of the recent past - which replaced the inelegant aluminum and brass cages of a yet-earlier era. But the best current “big-end” bearing cages are made of titanium and silver-plated. Experimenters with near-unlimited funds may like to try titanium bearing cages, but when having them made they should know that the bearing retaining slots must be machined with edges parallel to within 1/200 with each other and with the crankpin (assuming a parallel condition between cage and crankpin axis). It is not a job for someone with a bench-vise and a file. On the other hand, if employing silver-plated titanium cages and moving the thrust washers from the crankpin to the piston will elevate your engine's red-line by 2000 rpm, then they clearly will pay dividends in horsepower - if port-timing, etc., is adjusted correspondingly.

WRISTPIN/CRANKPIN BEARINGS (2)

McCulloch, the chain-saw people, have used an arrangement similar to the one just described for years, but they have reasons other than simply working around bearing cage failures at the wrist-pin end of the rod. It was discovered at McCulloch that failures at the crankpin bearing were traceable to the thrust washers most manufacturers of two-stroke engines use to center the rod on the crankpin. These washers usually are made of brass, or steel with a copper coating, and they do not find high rubbing speeds and scanty lubrication at all agreeable. At very high crankshaft speeds, they register their protest by overheating, and this causes a rise in temperature all around the connecting rod's big end,

?????? At site
which thins the oil present enough to create yet more friction, more overheating, until at last the thrust washers, roller bearing and cage are hot enough to “flash” the oil. At that point, lubrication is nil and friction quickly melts the bearing cage and wears flats on the rollers. McCulloch's engineers reasoned that the point of failure could be pushed upward materially simply by removing the thrust washers, which is what they did. Of course, the connecting rod still had to be centered over the crank, but this task was given to a pair of thrust washers up inside the piston. The improvement in terms of elevating the McCulloch kart engine's maximum crank speed was in the order of 1500 rpm, and it is worth noting that Yamaha borrowed this idea for use in the 17,000 rpm GP engines the company raced in 1968. It is interesting that in those engines, the piston rings were only 0.6mm in thickness.

WRISTPIN/CRANKPIN BEARINGS

Back in the days when pistons were uniformly poor and two-stroke engines wouldn't be run very fast, wrist pin bearings were almost always a simple brass bushing. Such bushings work very well in four-stroke engines, but lubrication is much less lavish in the crankcase-scavenged two-stroke and added difficulties are created by the essentially uni-directional loads placed upon it, which prevent the piston pin from lifting away from the lower part of the bearing and admitting oil to the load-carrying surfaces. For those reasons, the plain bushing has now almost universally been replaced by “needle” roller bearings, which are more easily penetrated by such oil as is available and in any case need much less oil. This last is of very particular importance in high output engines, as the heat flowing down from the piston is certain to thin any oil present to a viscosity approaching that of water. But all these difficulties not withstanding, the needle-roller bearing is wonderfully trouble-free, and if you encounter problems at the hinge between the connecting rod and piston pin, those problems will almost invariably be with breakage of the bearing cage. Given the extremely low rotational speed of the bearing in question, no cage is really needed except to make engine-assembly easier: the cage holds all the needle-rollers in place while the piston is being fitted to the connecting rod. The arrangement certainly makes working on the engine less complicated, but as it happens, the cage becomes the bearing's weakest link. Piston acceleration at high speeds is also applied to the bearing cage, and it may shatter under the strain - which sends a shower of particles from the broken cage and loose needles down into the crankcase. The debris thus liberated invariably gets pumped up through the transfer ports, into the cylinder, and more often than not a roller will get trapped hanging half out of a port by the piston with dire consequences to both.
Yamaha's TD1 was particularly prone to small end bearing cage failures, and I learned the hard way to replace these bearings if I saw over 11,000 rpm on the tachometer even for a moment, for their cages required only a moment's battering before cracks would start to spread and outright disintegration soon followed even if I indulged in no more excursions past the red-line. This difficulty has been overcome with cages made of tougher material; it is possible to accomplish the same thing by using crowded needles and no cage at all, which does require that a washer be fitted on each side of the connecting rod, to take up clearance so that the rollers cannot escape. Getting the thing assembled (with the roller glued in place with grease) is enough to make strong men weep with frustration, but it absolutely insures reliability at this point in the engine and is a measure worth remembering if problems with broken wrist-pin bearing cages do occur.

Piston Rings (3)

Another point of trouble can be the ring's locating pin, and if you encounter difficulties with locating pins working loose, the source of the trouble nearly always will be in the exhaust port. The racing engine's very wide exhaust port (width representing, in extreme instances, up to 70-percent of cylinder bore diameter) leaves a lot of the ring's diameter unsupported when the piston is down in the lower half of the cylinder, which allows the ring to bulge out into the port. Making the port opening oval and chamfering its edges will prevent the ring from snagging, as these things ease the ring back into its groove as the piston sweeps back upward. However, while the ring may not snag on the

?????? At site

port, it does get stuffed back into its groove fairly rudely, and that may have a very bad effect on the locating pin: On most two-ring pistons, the locating pins are positioned adjacent to the areas of blind cylinder wall between the intake and transfer ports -placed about 90-degrees apart - to provide a long path for gas leakage. Thus, when the ring bulges out into the exhaust port and then is stuffed back, the end of the ring is pushed into hard contact with the pin, and after a sufficient number of hard blows (and these accumulate rapidly at, say, 10,000 rpm) the pin begins to loosen and it will gradually enlarge the hole in which it is inserted enough to work completely loose. Then the ring is free to rotate, and it quickly works its way around to catch the end in a port. At risk of seeming immodest, I will admit to having isolated this problem for Yamaha several years ago and today that firm's racing engines have pistons with locating pins positioned 180-degrees from the exhaust port. Touring engines, which have much narrower exhaust port windows and thus treat their rings more gently, usually benefit from having their two rings' end-gaps placed more nearly on opposite sides of the piston, as described before.
In some racing applications, the standard rings are adequate to the engine speeds anticipated, but overall performance may dictate a much wider-than stock exhaust port. Then, the “offset” ring-locating pin may prove prone to precisely the sort of loosening and subsequent failure described in the preceding paragraph, which will lead you into a piston modification that can be very tricky: installing a new locating pin in the back of the ring groove. This gets tricky because in many cases the pin will be half-in, half above, the ring groove and it is impossible to drill the hole for a new location after the groove is machined. Impossible, unless you cut a small piece of aluminum to exactly fit the ring groove, filling it flush, in which case you drill your hole half in the piston and half in the filler piece. Then you remove the filler and your hole is ready for the pin - which introduces yet another problem: what to use for a pin? Steel wire is a good choice on grounds of strength, but is likely to work loose simply because the aluminum in which it is pressed grows and contracts so much with changes in temperature. A small-diameter “split pin” (which is like a rolled tube) is a better choice, but if suitable sizes are not available, then n pin made of hard brass is at least as good.

Piston Rings (2)

Ring sticking is a problem to be faced with all high-output two-stroke engines. Carburized oil may lock the ring in its groove after a remarkably short period of running if the ring lacks sufficient vertical clearance (usually, from 0.0015- to 0.0040-inch) or if the ring is located too near the piston crown. More frequently, the problem stems from the oil being used for lubrication, and it is most unfortunate that the very oils providing the best lubrication are the ones most likely to cause ring sticking. Castor-based oils, particularly, will build up thick layers of varnish inside the ring groove, unless the oil contains a considerable percentage of detergent chemicals.
Apart from the L-section Dykes ring, most piston rings have a basically rectangular cross-section, but you will find many minor variations on this arrangement. Currently very popular is the “keystone” ring, which has a tapered section, with either the upper or lower surface, or both, sloping away from the ring's outer face. The reason for this primarily is to keep the ring and its groove scrubbed free of carbon and varnish. In four-stroke engines the rings are free to rotate, and do, and their rotation performs this scrubbing. Two-stroke engines nearly always have their rings pinned, to prevent them from rotating and the ring's ends from springing out and becoming trapped in a port. Hence, the need for some other form of scrubbing action. Seldom is the taper in a keystone-type ring more than 7-degrees, and it is all too easy to attempt installing one of them upside-down, so you should give particular attention to the ring's markings. Such markings vary in kind, but without exception they will be on the ring's upper surface.

Piston Rings

Of all the problems that can be experienced with a modified engine, those connected with the pistons' rings are the most insidious. Borderline sealing failures can send fire shooting down along the pistons' sides to cause seizures and/or holing of the piston crown that appear to be the result of lean mixture, excessive ignition advance or too-high compression, but are not. These failures are, I suspect, much more frequent than is commonly supposed, for the 2.0mm rings that have become almost standard will begin to flutter when piston acceleration rises above about 60,000 ft/sec2 and it is entirely too easy to exceed that limit with a modified touring engine. Therefore, I would again urge you to do your homework before starting a development program with any engine. A formula for predicting the onset of ring flutter is provided in the chapter headed, “Fundamentals”, and you may save yourself a lot of grief by determining your engine's red-line with paper and pencil instead of through experimentation. At the same time, I must caution you against simply assuming that very narrow rings are an advantage in all engines. In fact, there is no detectable power difference between the standard 2.0mm ring and the “racing” 1.0mm ring below 7000 rpm, and the wider ring has the advantage of better durability right up to the point where piston acceleration starts it fluttering. Neither is there any advantage, below 7000 rpm, in the use of single-ring pistons. Above that level the lower friction of the single-ring piston begins to make a difference, but in the lower speed ranges you may as well use two-ring pistons and take advantage of their “second line of defense” capability.
Selection of ring-type will usually have been made for you by the piston manufacturer, and my advice is that you do not try to improve upon his judgment, which will be almost impossible in any case. You cannot, obviously, re-machine a piston made for 2.0mm rings to take 1.0mm rings -unless you cut a new ring groove above the existing grooves, and that would position your ring perilously close to the piston crown and almost certainly lead to immediate ring failure. The only way around this is to fit a Dykes-pattern ring, right up at the piston crown - as was noted previously. Such modifications can be very successful, if you have the right ring for the application and cut the groove correctly for the ring, but I cannot recommend the procedure simply because there is so much room for error. In general, I think it is far better to replace the stock piston with one fitted with thinner rings - even if the replacement piston is cast of somewhat inferior material, as is often the case. After all, the best of pistons will fail if its rings are not suited to the job it is being asked to perform. On the other hand, rings of less-than-desirable material will perform very well in racing applications if replaced frequently, and if they have not been crudely finished. Much of the ring's ability to function is related to this latter aspect. The ordinary cast-iron ring is fragile, and will shatter very quickly if allowed to flutter, but it will perform entirely satisfactorily if its lower surface is smooth and true, and seals against the bottom of the ring groove. Rings made of nodular cast-iron have the same wear-resistant properties, and are vastly stronger, for which reasons this material is almost universally used. Surface coatings, ranging from chromium to Teflon, are often applied to the piston's ring's face, to improve service life and /or prevent scuffing during break-in.

The piston (6)

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 and closing of the ports; with the ring in its usual position, about 0.200-inch below the piston crown, there is a tendency for gases to leak down the side of the piston, and the port-opening process thus becomes more gradual than is desirable. The effect is slight, but it is there, and for that reason ring location always is a matter of juggling the conflicting requirements of keeping the ring cool, and obtaining sharp, clean port-opening characteristics. And in most instances, the balance of this compromise will be in favor of the former, for an overheated ring quickly fails. The cause of this failure is twofold: first, excessively high temperatures effectively anneal the ring, and it loses its radial tension; second, an overheated ring warps like a potato chip, and no longer maintains close contact with the bottom of its groove. In both of these cases, the ring's ability to seal is reduced, which allows fire to start leaking down past the ring, and that further raises its temperature -starting a cycle that soon results in outright ring failure.
The single exception to the unpleasantness just described is the L-shaped “Dykes” ring, which also is excepted from the immediate effects of ring-flutter (described elsewhere). A number of engines have been fitted very successfully with Dykes rings located right at the tops of their pistons, and the dire effects of excessive heating are avoided because the Dykes ring's vertical leg has enough area in contact with the cooler cylinder wall to draw away heat faster than it can be added by the ring's contact with high-temperature gases. At least, that's the way the situation can be, if everything is right. On the other hand, it is worth remembering that many users of Dykes-pattern rings have been forced to fabricate them from stainless alloys to overcome temperature related troubles, and even then have experienced problems with oil carbonizing in the ring grooves. Probably the best thing to be said for Dykes-pattern rings from the experimenter's viewpoint is that they can be used to overcome the problem of using stock pistons at very much higher than stock crankshaft speeds. If, for example, you would like to use the stock piston, but cannot because it has been grooved for rings 2.0mm thick and you must use 1.5mm rings to avoid ring flutter, you can simply cut a new groove at the top of the piston for a Dykes ring and the problem is solved -unless you encounter some of the other difficulties just discussed.

The piston (5)

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

?????? At site
clearance-banded piston will absorb a lot of punishment before it is damaged sufficiently to cause retirement from a race.

The piston (4)

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 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 over-simplification 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.

The piston (3)

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

?????? At site
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 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.

The piston (2)

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.

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 100,000 ft/sec2. Aluminum, used as the primary constituent in virtually all piston alloys 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 variable-clearance 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. Admixtures of silicon in excess of 15-percent not only drastically reduce aluminum's expansion rate, they also 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.

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.

Piston Accelerator (3)

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:

?????? At site

Where T is time, in seconds
N is crankshaft speed, in revolutions per minute
is port open duration, in degrees
(This formula can be abbreviated to
?????? At site)

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

?????? At site

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 exhaust- and 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 two-stroke 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 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 Speed Secrets is even greater than for guesswork. There is only one “Secret” in the game: to know what you are doing, and to do it thoroughly.

Piston Accelerator (2)

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 you will see in the following table listing ring thicknesses and the accelerations at which they begin to flutter.

For rings having a
0.125 -inch thickness, 40,000 ft/sec2
0.094 -inch thickness, 53,000 ft/sec2
0.063 -inch thickness, 80,000 ft/sec2
0.047 -inch thickness,106,000 ft/sec2
0.039 -inch thickness,138,000 ft/sec2

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, would seem to be 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 cross-sectional 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 Butter is predicted at 80,000 ft/sec2 and intend using crankshaft speeds that would raise maximum piston acceleration to something more like 100,000 ft/sec2, 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 L-shaped 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

?????? At site

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 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, for example, should be tapered in slightly 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.
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 inertia 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 vice versa; at the closed end, 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.

Piston Accelerator

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 little thought 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

?????? At site

that a 30 psi force will not hold back one circa 750 psi. Still, equally obviously, piston rings do form an effective seal. How? Because they get a lot of help from the cylinder 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 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:

G max = ?????? At site

Where G max is maximum piston acceleration, in feet per second squared
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