When performing your own experiments with intake tract length, you will be tempted to mount the carburetor as close to the cylinder as possible, and make adjustments by varying the length of an intake stack added on the carburetor's mouth. Don't do it! That certainly is the most convenient method, but a carburetor placed too far back toward the intake port window is going to be subjected to radical pressure fluctuations, due to wave activity in the system, and those pressure fluctuations do terrible things to the carburetor's ability to meter fuel. Worse, the largest departure from the desired mixture strength will occur precisely at the engine speeds where intake ramming is strongest, which makes the task of selecting a main jet impossible. So the carburetor must be located out at the end of the intake tract, rather than close to the cylinder- which is unfortunate, as maximum air delivery into the crankcase is obtained when the reverse is true. Perhaps some form of fuel injection is the answer.
Mixture-strength problems can also occur due to the plumbing between the carburetor and air cleaner, and it should go without saying that such plumbing may also add to the intake tract's effective tuned length. The common practice of connecting carburetor mouth and air cleaner with a section of rubber hose may have much to recommend it from the standpoint of convenience; it also is likely to establish either an extension that becomes part of the effective tract length, or a secondary resonating system that heterodynes at some frequencies with the main tract and thus upsets its proper functioning. Therefore, it is good practice, if sometimes inconvenient, to make connecting plumbing both as short and as large in diameter as is possible. Air cleaners having plastic-foam elements may be mounted very close to the carburetor without penalty, but the paper-type filter may, if it is too close to the carburetor mouth, become saturated with oil and fuel - in which case it will refuse to pass any air at all.
At one time, it was every two-stroke tuner's habit to begin any serious attempt at extracting more-than-standard horsepower from a given engine by increasing the engine's primary compression ratio - that is to say, the ratio between crankcase volume with the piston at top center, and at bottom center, as in the following expression:
Where, CRp is primary compression ratio
V1 is crankcase volume at BDC
V2 is piston displacement
I suspect that the popularity of this practice, “stuffing” crankcases, was derived from the fact that the old T-crank Villiers was then the two-stroke engine most frequently being modified. This engine was built more with an eye toward manufacturing cost and long-term reliability than power output (of which it had only extremely modest amounts) and the configuration of its crankshaft and crankcase provided only a very low primary compression ratio. Too low, in fact, for anything even remotely approaching high speed operation, so that any efforts at increasing its primary compression ratio were immediately reflected in a power increase. But a lot of people simply assumed that if increasing the Villiers engine's primary compression ratio from, say, 1.2: to 1.4:1 was good, then raising it even further would be better -and that all two-stroke engines would benefit from being given the same treatment. There, they fell into error.
Referring again to the work of Fujio Nagao (and his results have been verified by other researchers) we find some very interesting conclusions with regard to primary compression ratios: First, Nagao tells us that for given port areas, there is an engine speed at which maximum air delivery to the cylinder occurs, and that this engine speed is inversely proportional to crankcase volume, but that the maximum value changes only slightly with changes in crankcase volume. To put it another way, the crankcase-pump's volumetric efficiency is nearly constant, but the engine speed at which it attains maximum efficiency rises as crankcase clearance volume is reduced. Significantly, too, Nagao goes on to say that any deficiency in air delivery due to a crankcase volume too great for a given engine speed is fairly well compensated by properly tuned intake and exhaust pipes. He says in conclusion that “little advantage is obtained by making the crankcase volume excessively small.”
Later work in the same field indicates that Nagao understated the case, and that there may well be positive disadvantages in excessive reductions of crankcase volume. Hiroshi Naito, who I believe was responsible for the fantastic Yamaha GP racing engines, has indicated that there is little improvement to be had even from ultra-high speed engines with primary (crankcase) compression ratios above 1.5:l. This apparent anomaly can be explained by studying the effects of the whole scavenging system, from the intake tract through the crankcase and scavenging ports and right on out into the exhaust system. And we must think in dynamic, rather than static terms. Starting with the exhaust system, we find that it is possible to evacuate the cylinder to well below atmospheric pressure by using the “extractor” effects of the expansion chamber's diffuser. If the lengths and proportions of the exhaust system are properly established, the fresh charge moving up through the transfer ports will not only be pushed through by the pressure below but will get additional aid from what is, in effect, an exhaust-produced vacuum in the cylinder. Further, this vacuum may well be communicated down into the crankcase, via the transfer passages, and crankcase pressure “trapped” at a below atmospheric value when the transfer ports close. This factor can be very important, as it produces an air-delivery volume greater than would be possible with crankcase pressure alone to impel the fresh charge into the cylinder. But that isn't the whole story, obviously, for the “pull” on the exhaust side of the cylinder is matched by a “push” where mixture from the carburetor enters the crankcase - with a force equal to about 1.5 atmospheres in advanced engines. Thus, we have evacuated the crankcase to something below atmospheric pressure before refilling it with a fresh charge, and the crankcase will have been charged to something above atmospheric pressure by the pulsations in the intake tract. All this has implications in terms of crankcase volume, for if we assume that the positive and negative pressures applied remain constant, then bulk flow through the cylinder will increase with increases in crankcase volume. Does that seem obscure? Then consider that there is more air in a 1000cc flask compressed to 1.5-atmospheres than in one of 500cc capacity. In light of this, you might assume that it is time to start reducing crankcase compression ratios -especially as pumping losses (horsepower absorbed in doing this work) rise as to the third power of compression ratio. Actually, this isn't quite true either, for reductions in crankcase compression ratio cause an equal reduction in the amplitude of wave activity in the intake tract, which in turn reduces the ramming pressures available to charge the crankcase. So, crankcase compression ratios must be established to balance the conflicting requirements of volume and wave strength, and if it is pure, peak horsepower that interests you, then you will find that a primary compression ratio of 1.5:l represents something very near the optimum. Of course, this will only be true if carburetor size also is selected with an eye toward maximum horsepower.
Broad range performance is another matter entirely. For motocross, you will require much more flexibility from an engine than would be true of road racing or even flat-track. In the large-displacement classes particularly, where sheer engine size makes it altogether too easy to get more horsepower than a motorcycle's rear tire can apply to the ground, an engine's ability to pull strongly over a very wide speed range is much more important than any peak reading. For such applications, the best engine is likely to be one with a primary compression ratio well below 1.5:1, a long, slow-taper expansion chamber and a relatively long intake tract. With regard to this last item, I might add that it is possible with intake lengths of about 30-inches to boost an engine's crankcase delivery ratio (volume of air pumped, per cycle/piston displacement) to as much as 1.2:l. However, due to restrictions in available time, these ultra-long pipes will not work except at very low engine speeds and, moreover, tend to work only over a very narrow :peed range. Indeed, all tuned intake pipes effectively reduce an engine's operating range, though this is compensated by their providing a power boost at some engine speeds, and it might be that a near-zero length would provide the best spread of power. The difficulty here is that some sort of smooth passage must be provided for the transition from the substantially-square intake port window to the round carburetor throat, and if you add the length of that passage to the carburetor's length, then you have a resonating intake system even if you don't want one. The choice is thus reduced to selecting a length that provides the best results overall.
There is one means of halting resonant effects in the intake tract, and to accomplish that you have only to reduce the carburetor throat in diameter until its cross-sectional area represents 35-percent, or less, of the intake port area. Curiously, a carburetor of that small size placed at the port window, behind the intake pipe, will not stall wave activity in the system- but one located out at the end of the intake pipe acts in effect as a wave-damper. I mention this only for its value as a curiosity, as there is no point in fitting such a small carburetor - except, perhaps, on a trials motorcycle - and if fitted at the port window it would be subject to the same mixture-strength instability as any other carburetor preceded by a length of resonating pipe.
Another, indirect means of gaining the benefits of a resonating, ram intake tract (as well as certain others) is to interpose a check-valve between carburetor and cylinder - which will insure a one-way flow, and trap in the crankcase anything that passes through the valve. Now as it happens, the only check-valve currently capable of functioning fast enough to keep pace with a high-revving two-stroke engine is the one generally called a “reed-valve”. The name fairly accurately describes the valve, which has a thin, broad metal or phenolic reed seating against an aperture and clamped at one end so that it may bend away from said aperture. There usually are multiple reeds and apertures in any reed-valve assembly; also in most instances these will be set at an angle oblique to the direction of gas flow, so as to minimize flow losses past the reeds. Usually, too, there will be four separate reeds, clamped against the sides of either a pyramid or a wedge, with the carburetor feeding the interior and the point being aimed into the crankcase. At one time many attempts were made to use comparatively stiff reeds, to overcome a tendency for the reeds to fracture and drop off at high engine speeds, but now everyone utilizes thin reeds backed by stops. The stops, which are simply thick, curved strips of metal clamped over the reeds, control both the reeds' travel and the shape of reed- bending. You will appreciate that a reed held only at its end will bend quite abruptly right next to the clamp that holds it in place, and that it may well bend too far for its own well-being under some conditions. Reed-stops prevent both of those things from happening, forcing the reed to curve very evenly around the stop and allowing it to bend only as far as the fatigue-resisting properties of the reed material enable it to withstand.
Only rarely, today, do reeds fracture at their root and drop into an engine's crankcase - but it still happens, and for that reason many manufacturers who build reed-valve engines prefer to employ plastic (usually a fiber-reinforced phenolic sandwich) reeds. Steel reeds can do terrible damage to an engine's interior when they break-off and are aspirated up through its transfer ports; the phenolic reed simply gets gnawed into fragments and expelled out the exhaust port. This danger, with steel reeds, apparently is very real, for there can be little doubt but that engine performance is improved by the use of steel over phenolic in the reed material - simply because steel is the more flexible material.
Actually, the failure of a complete steel reed is an extremely rare event. Much more frequently the failure will be confined to a piece breaking away at the reed's tip - which is caused by the repeated impacts of the reed against the reed block. For many years this problem seemed to defy solution, as very thin reeds were not strong enough to resist these fretting fracture, and thicker reeds were shattered by their own greater inertia. Yamaha found the solution, which is beautiful in its simplicity. The 1972 Yamaha reed-valve (they call it "Torque Induction") engines have reed blocks coated with neoprene, and while the coating is very thin, it has enough resilience to cushion the impact of the reeds and prevent fretting failures. I understand that the Yamaha reeds are of stainless steel, with thicknesses of 0.008- and 0.006-inch for the 250cc/360cc and 100cc/125cc/175cc engines, respectively. The use of a stainless steel as a reed material will be explained if you consider the devastating effect of even a very little rust on such thin strips of metal.
I am much inclined to doubt that reed-valves will become popular in road racing engines, as they do introduce a flow restriction into the intake system that must have a depressing effect on peak horsepower at ultra-high engine speeds. But there is every reason to assume that “Torque Induction” will come into very wide use for every kind of off-road motorcycle: There is, for example, the reed-valve's ability to make the most out of positive intake-resonance effects while stalling out-of-phase resonance. Also, the reed-valve has, by its very nature, the ability to effectively adjust intake timing to suit all engine speeds. Finally, it does seem that reed-valving eases problems with broad-range carburetion, ending the oft-noted tendency for carburetors to produce wildly-varied mixture strengths at different engine speeds.
Another benefit that accrues with reed-valving is that with all possibility of back-flowing out the intake removed, it becomes possible to use the rear cylinder wall for something other than a support for the piston. Yamaha, long an advocate of multiple transfer ports, has added another port, opposite the exhaust port, in its reed-valve engines. This port is as high as the other (four) transfer ports, but is very narrow. Significantly, this extra transfer port is not fed mixture from the crankcase; it relies entirely upon the extractor effect of the exhaust system to pull mixture from the carburetor through the reeds and reed cavity, and up through the port into the cylinder. My friends at McCulloch - who have an enormous accumulated experience with reed-valve engines, tell me that power characteristics can be altered very radically by widening and narrowing this boost-port, and/or by arranging for mixture compressed in the crankcase to flow up through the boost port instead of relying purely on the depression created in the cylinder by the exhaust system to do the job. But they also seem to think that the boost-port's width is something one determines almost entirely through experimentation, so it would appear to be wise to start with a very narrow port window and widen it gradually until the desired power characteristics are obtained.
THE ROTARY VALVE
Better than reed valves in function, if not in mechanical convenience, is the rotary valve. Specifically, the disc-type rotary valve, which is the only variety currently in use in motorcycle engines. There is little point in dwelling on this kind of intake valving over-much, because it really is only practical on a GP road racing engine. Touring-type engines, if they have more than a single cylinder (and the trend obviously is toward multis) become inconveniently wide when a disc-valve assembly is added at each end of the crankshaft, and this type of intake valving is all but impossible to apply to an in-line engine with more than two cylinders. The rotary-valve's extra width is, of course, an embarrassment even on single-cylinder engine in off-road applications.
If you ignore the mechanical disadvantages of the rotary-valve, it becomes highly attractive from the standpoint of not only peak power, but broad-range power. The reason for this is that it does not leave the engine to rely upon pressure waves marching back and forth in a pipe to prevent charge loss back through the carburetor. It is possible to extract just as much horsepower, at the peak, from a given engine with either rotary- valve or piston-port induction, but the latter will lose power very rapidly either side of the bhp-peak while the rotary valve continues to function and crankcase pumping remains effective.
INTAKE PORT SHAPE
Intake port timing is covered elsewhere in this book, but this certainly is the place for a discussion of intake port shape - which has a definite influence on crankcase pumping efficiency. In general, the best flow coefficient for any given timing-area value will be obtained with the widest-possible port. That is to say, a port that wraps around the cylinder as far as is permitted by mechanical considerations (stud placement, transfer port location, etc.). Of course, with a very wide port there is a tendency for the rear edge of the piston skirt to snag at the bottom of the port window, which means that it may be necessary to use a window shape more nearly round than square to prevent rapid wear at the bottom of the piston skirt. Which is often good practice in any case. A rounded port window, or one with a V-shape to its lower edge, provides what effectively is a slower rate of port opening, which is very useful in reducing intake roar - a point that must be considered in a time when statutory noise limits are popping up all over the place. Also, the more-gradual opening of the port tends to extend the duration of the sonic wave that is used, on its return trip, to “supercharge” the crankcase, and that has the effect of broadening an engine's power band. Finally, a port with rounded corners has a much better flow-coefficient than one that is square. The same may not be said for rounding back the lower edge of the piston skirt, as that extends the intake timing- to permit backflow as the piston descends -without producing any measurable improvement in flow coefficient. It is, however, possible to improve flow with a down-turned lip at the top edge of the intake port window. But the primary thing one must remember when carving away at an intake port is that ripples in the port walls, or any sudden change in cross section, have a far more damaging effect on flow-coefficient than a slightly rough finish in the port. Therefore, it is vastly more important to smooth the port than to give it a mirror-finish. And it should be obvious that the port face, the gaskets and heat-block (if any) and carburetor should all align very neatly, without any steps between parts - even if that means doweling everything in place to assure alignment.
Scavenging, in the context of piston-type internal combustion engines, is the process in which the products of combustion are cleared from a cylinder at the end of the power stroke and a fresh air/fuel charge is introduced in preparation for the compression and power strokes to follow. This process is common to all Otto-cycle engines, but it can be accomplished in two entirely different ways: In the four-stroke cycle engine, it occupies at minimum a full 360-degrees of crankshaft rotation, with one piston stroke being devoted to pushing exhaust products from the cylinder, past a valve in the cylinder head; the return stroke aspirates a fresh charge in through another port and past another valve. Thus, there is a fairly complete mechanical separation of the gases involved, and while valve timing will commonly be arranged so that there is some overlapping of the exhaust and intake phases of this operation, little dilution of the fresh charge by exhaust gases is possible, and any short-circuiting of this air / fuel mixture out the exhaust port during the early stages of the intake-open period has no adverse effect on the weight of the charge ultimately trapped in the cylinder at intake-closing. Obviously, the mechanism required to operate a four-stroke engine's valves adds considerable complication to the basic crank-rod-piston assembly, but the very efficient scavenging obtained is, for most applications, considered to be worth the valve-gear complexities. And the very high brake mean effective pressures resulting from this high scavenging efficiency - a bmep in the order of 200 psi for highly-tuned examples of the type - offset the scheme's single disadvantage, which is that power strokes occur at 720-degree intervals.
Two-stroke cycle engines deliver power strokes twice as often, at 360-degree intervals, and in those intended for industrial (trucks, electrical generators, etc.) or marine applications essentially the same bmep as are obtained from four-stroke engines are to be expected. But in such engines one finds an even greater mechanical complexity than in the four-stroke design, for in all two-stroke engines the scavenging process occurs in time borrowed from the compression and power strokes. In effect, this means that all of the cylinder clearing and recharging for which 360-degrees of crank rotation are reserved in the four-stroke engine must occur while the piston is halted at the bottom of its stroke. And, lacking time for a more leisurely exchange of gases, the process must be helped along by extremely large port areas and high scavenging pressures. Usually, engines of the type being discussed will have multiple exhaust valves in their cylinder heads, and a ring of windows around the cylinder's base through which scavenging air is forced by an engine-driven pump. Such engines almost invariably are diesels, in which fuel is injected only after all the valves and ports are closed (injection beginning at TDC and continuing for perhaps 70-degrees of crank angle) and the scavenging pump delivers air in excess of what is required to fill the cylinder, so there is no loss of charge nor any dilution to cause a loss of power. Unfortunately, not only is this type engine very complicated and expensive, it cannot be scaled-down to a size useful in motorcycle terms - simply because the exhaust valves will not open and close fast enough to keep up with the kind of crankshaft speeds needed for the power outputs we have come to expect from our small-displacement engines.
In the end, only the familiar “piston-port” transfer and exhaust valving arrangement is suitable for motorcycle engines, and that is -with an exception, of sorts, existing in the use of disc- or reed-type intake valving –precisely what has come to be universally applied. Present motorcycle engines are all scavenged through windows in their cylinder walls, with scavenging air being supplied from their crankcases. This system is beautiful in its simplicity, but it does have serious short-comings: First, there is the relative incapacity of the crankcase as a scavenging-air pump, which prevents even the hope of having excess air to use in clearing the cylinder. Secondly, the use of the piston's motions to open and close (actually, to uncover and cover) the exhaust and transfer ports creates enormous difficulties in a number of areas related to clearing and recharging the cylinder. The low scavenging pressure available makes it absolutely essential that pressure in the cylinder be no more than slightly higher than atmospheric when the transfer ports open, which means that the exhaust phase must begin well in advance of the uncovering of transfer ports. And, because the piston-controlled exhaust timing is necessarily symmetrical, the exhaust port will remain open long after the transfer ports close - leaving an unobstructed opportunity for the fresh charge to escape the cylinder. Indeed, the charge injected into the cylinder has every reason to escape, as the upward motion of the piston, moving to close the exhaust port and begin the effective compression stroke, is displacing the gases above its crown. Gas pressures always try to equalize, and those in the cylinder can only do that by moving back into the transfer ports, while these are still open, and out the exhaust ports. Thus, it is virtually inevitable that some portion of the fresh charge will be lost into the exhaust system, and that the upward stroke of the piston will also tend to aspirate some of the charge back down in the crankcase.
Difficulties inherent in the piston-port scavenging system are not confined to charge loss, or backflow into the crankcase. One of the great problems is created by the lack of mechanical separation of the exhaust gases and the incoming fresh charge. We expect that the engine's exhaust gases will choose to escape from the exhaust port, and that the charge coming in through the transfer ports will push the residual exhaust products ahead of it to completely clear the cylinder, but the actual process is by no means that tidy. The cylinder pressure may drop very neatly to atmospheric, or even below, but it still will be filled with exhaust gases, and these will not necessarily be swept out the exhaust port merely because other gases have entered the cylinder. In point of fact, it is possible to short-circuit the scavenging flow straight from the transfer ports to the exhaust port and leave the exhaust residuals in the upper cylinder entirely undisturbed. This possibility has haunted the design engineer throughout the two-stroke engine's long history, and many an elaborate system of ports and piston-crown convolutions has been created to confound this worst of all demons.
For many years cross-flow scavenging was preeminent, principally because it makes maximum use of cylinder wall area. In the cross-flow engine, ports ring virtually the entire lower cylinder, with half being exhaust and those opposite being transfer. Given the direction of the flow emerging from the transfer ports, the charge would shoot straight across and out the exhaust side but for the shape of the piston crown - which lips up on the transfer side to form a deflector that redirects the transfer flow upward. Cross-flow scavenging is still employed in outboard-marine and model airplane engines, in the latter because it is a manufacturing convenience and in the former because its insensitivity to scavenging pressures and volume of flow provide superior low-speed, part-throttle running characteristics. The cross-flow system is, however, handicapped in terms of maximum power by the large surface area created in all that lumpiness in the piston's crown, which very greatly increases the heat flow into the piston and lowers the compression ratio a given cylinder will tolerate as compared to a piston with a flat or slightly domed piston crown. Apart from this thermal problem, there is much to recommend cross-flow scavenging, but the thermal problem is of sufficient magnitude-even in water-cooled engines - to remove it from serious consideration for any high-output, two-stroke engine.
Numerous scavenging systems not requiring a deflector-type piston have been tried: The pre-WW2 Villiers had exhaust ports on opposite sides of its cylinder and four transfer ports, in pairs, between them. A Barnes and Reinecke design had a ring of exhaust ports located above a ring of transfer ports and a part-conical piston crown, all of which sent the scavenging flow in a narrow column up the middle of the cylinder, and forced the exhaust outflow to follow a path down the cylinder walls. Curtiss employed multiple transfer and exhaust ports on opposite sides of a cylinder, and biased the direction of the transfers upward and to one side, so that the fresh charge spiraled up into the cylinder. But the best of the scavenging systems was one devised by a Dr. Schneurle, of Germany, in which a pair of mirror-image transfer ports flanked a single exhaust port, directing the scavenging flow toward the cylinder wall opposite the exhaust, and upward, to loop over and thus clear the cylinder. Schneurle's loop-scavenging method was patented by him, in 1925, and this had the effect of simultaneously elevating German industry's fortunes in the two-stroke engine field while forcing practically everyone else to seek alternative and less-efficient systems. Of course, now that the Schneurle patents have expired everyone employs some form of his scavenging method, although best results are being obtained with more than Schneurle's original pair of transfer ports.
Only a decade past, East Germany's MZ was considered to be the repository of really advanced research in high-speed two-stroke engine design, and one Walter Kaaden could be said to have the best grasp of the intricacies of scavenging systems of anyone working in the field. Today, no discussion of two-stroke engine scavenging is possible without concentrating almost exclusively on development in Japan. Japanese engineers did not invent the two-stroke engine, nor have they employed any system of scavenging ports that has not seen earlier service elsewhere. But they have done an enormous amount of basic research directed at quantifying what previously has been known only in terms of generalities; they have established very firm design criteria for the management of factors that once were decided almost purely through cut-and-try experimentation. Of course, none of this would be of more than incidental interest but for the fact that some of the Japanese firms have abandoned their once-absolute policy of secrecy and are sharing what they have learned with the rest of the world. Yamaha, particularly, has made a vast contribution to the overall state of the art by publishing fairly specific criteria for the port timings and areas required for engines of any given cylinder volume and operating speed. Like many others, I knew that port timing and area were interrelated factors, but the job of obtaining and sorting through data on a wide range of engines to establish a pattern, and then experimentally verifying conclusions was too time-consuming and expensive to even contemplate, as an individual. Yamaha has done that work for us, and published enough information on the subject to complete at least my understanding (a detailed discussion is presented elsewhere in this book as a chapter, headed, “Port Timing”). From a number of SAE papers from Japan - as well as examples from Germany and the United States -and my own experience, I have also accumulated much incidental information related to the shapes, number and disposition of ports. These factors profoundly influence scavenging flow, which influences horsepower very greatly, and we will for the moment concentrate on them alone.
The difference between success and failure with a modified engine can be the treatment of the exhaust port. Even assuming that no change is made in exhaust-port timing, simply widening the port window will result in a power increase; it also can result in drastically-shortened ring life, amounting in extreme examples to outright and nearly-instant breakage of the rings, and/or severe overheating of the piston crown. There are reasons for these problems: A two-stroke engine's piston rings always bulge out into any port window they pass, and while transfer port windows seldom are wide enough to permit this to an extent sufficient to cause difficulties, the same certainly may not be said of the exhaust port. A relatively mildly-tuned engine will have an exhaust port width equal to at least 50-percent of its bore diameter (which is to say, a 3-inch cylinder bore would have a 1.5-inch exhaust port width) and that is enough to allow the ring to spring out into the port window very perceptibly. Make that port square, with sharp comers and sharp edges, and the ring will be destroyed very quickly. And if you enlarge the port so that its width represents 70-percent of cylinder-bore diameter, ring failure would almost certainly occur during the first revolution of the crankshaft. Yet, racing engines have been run quite successfully with 70-percent port widths, and while I cannot recommend that kind of extremity for modified production-type engines, the mere fact that it has been done tells us that techniques exist to make it possible.
Basically, ring-life is improved - at any given port width -by A)rounding the shape of the port window, and B) breaking the window's sharp edges. Both of these measures are employed in all engines, but they reach special and somewhat exaggerated form in racing applications. The traditional port window shape is square, or rectangular, with its corners rounded to help prevent ring snagging. Assuming that the port width does not exceed 60-percent of bore diameter, the radii at the port window's corners should be about 15- or 20-percent of the port width, and that is just what you will find in most engines. But as the port is widened, those corner radii have to be made larger -to about 2s-percent of port width when the latter approaches being 70-percent of cylinder diameter. Actually, even these very large radii will not completely prevent ring snagging if they are not joined by straight-line edges. The upper and lower edges of the port window should be arched, on a radius equal to about twice the port width, in ports having a width that is 60-percent of bore diameter or less. Unfortunately, these simple rules-of-thumb are not adequate at port widths above the 60-percent level- and such widths are - becoming very common. With the rings we had a decade past, which were relatively thick, axially, and quite narrow, radially, (not to mention being made of ordinary and rather brittle cast-iron) the upper limit for port width was 62-percent of cylinder bore diameter; now, with our vastly improved rings it has become possible to widen the port out to 70-percent of bore, or slightly more, if we are very careful in shaping the port window.
Just as there is no means of predicting, with any great accuracy, what kind of “cam” and taper a piston will require to fit closely in a cylinder when both are at operating temperature, neither is there any firm rule for shaping ultra-wide exhaust port windows. Both are established, initially, on the basis of past experience, and then modified according to test results. It has been demonstrated, in practice, that a modified ellipse is the basic shape of port windows in the 62- to 70-percent (of bore) range. Thus, while the ring may actually bulge out into the port window enough to cause its instant destruction in a square port, or in one with straight-line edges are joined with simple radii, the contours of an elliptical port window will sweep the ring gently back into its groove. Then, the only problem that will be encountered is that the ring may bulge out, and be pushed back, unevenly -which may drive one end of the ring sharply against its locating pin and eventually cause the pin to loosen and come adrift. It should be obvious that this last difficulty will be most pronounced when the port window is not perfectly symmetrical, as any departure from symmetry will result in the ring being displaced to one side as it is pushed back into its groove.
Careful craftsmanship will prevent this asymmetrical displacement of the ring; it will not, of itself, forestall other problems associated with very wide exhaust port windows. In my opinion, one should never simply and arbitrarily widen a port out to the 70-percent limit. Instead, the safe and sensible approach is to begin at 62-percent, with a shape that is as nearly an ellipse as is possible. Quite obviously, sharp limitations are going to be imposed by the shape of the existing port window; the idea is to provide the most generous radii permitted by the basic shape with which one must begin. Obviously, too, this reshaping of the exhaust port window will be easier if you have opted for increasing the exhaust timing, as that will give you room to work above the existing port. Then, having established the initial shape, you will have to inspect the rings and the edges of the port window for evidence of scuffing or snagging. Seldom will there be any problem around the lower edge of the port, as the piston slows considerably near the bottom of its stroke. Most of any scuffing that appears will be around the comers of the port; outright snagging will make its presence known in the appearance of scratches leading upward from the center of the port window.
I have already indicated that the kind of reshaping possible is largely a function of the stock port window's shape, but alterations in shape are not the only cure for scuffing and snagging available to us. Practically everyone knows enough to round-off the port window's sharp edges; what most do not know is that a simple round-off is not what is needed. The purpose in breaking those edges at all is to prevent ring snagging by easing the ring back into its groove, and this job is done best not by a simple radius, but by surrounding the window with a very slight bevel, about 0.080-inch in width, and tapering in toward the window to a depth of perhaps 0.015-inch. It is of course necessary to work a slight radius where the bevel reaches the port window, just to be safe, but the real job of tucking the ring safely away in its groove is performed by the bevel. You will appreciate that the same kind of bevel is needed at all the port windows; those at the transfer ports need not be quite as large.
As regards the exhaust port, a secondary function is served by providing a bevel, and radiused edges, around the port window. There is a very considerably contraction of flow through any sharp-edged orifice, and such orifices may be made effectively larger by providing them with a rounded entry. Improvements in flow in the order of 30-percent could be had were it possible to give the port window edges a radius of, say, ¼ inch. Unfortunately, to do this would mean advancing the point of exhaust-opening a like amount, which in most engines would result in a very radical exhaust timing indeed. It is, on the other hand, often possible to carve just such a radius at the sides of an exhaust port - although it is questionable that this radius would be as effective as simply widening the port to the same extent. The radius approach does have the advantage of leaving intact much of the metal around the port, which can be important: Thick sections of metal tend to equalize cylinder temperatures and prevent the kind of local distortion that is such a potent cause of piston seizure. Also, in engines having exhaust ports closely flanked by cylinder hold-down studs, there may not be room enough to widen the port as much as would otherwise be desirable, and in that event the side-radiused ports become a necessity.
In the vast majority of engines there will be a fairly large increase in cross-sectional area between the stock exhaust window and the actual exhaust outlet. Indeed, this increase often is too large to give best results with expansion chamber exhaust-systems: What may seem to be no more than a flow-improving enlargement in area leading into the exhaust pipe during the outflow phase of scavenging becomes a sudden constriction for waves returning to the cylinder from the expansion chamber. In fact, if the difference in the areas at the port window and the outer end of the port becomes as great as 1:3, virtually all of the expansion chamber's resonant effects will be lost. What happens, in such cases, is that the waves returning to the cylinder are reflected back into the chamber by the abrupt constriction of the port. Maximum transmission of these waves into the cylinder will, of course, be obtained with a 1:l port-window / port-outlet ratio, but that kind of straight-through passage represents something less than the optimum in minimized flow resistance during the blow-down phase of scavenging (the period beginning when the exhaust port cracks open and ending with the opening of the transfer ports). Thus, the walls of the exhaust port should diverge somewhat, giving a progressively-increasing cross-sectional area out to the exhaust flange. The most important thing to remember, here, is that sudden changes in section should definitely be avoided. Neither gas-flow nor the effects of sonic waves in the exhaust tract are served by a bunch of lumps and jogs - this being far more important than a mirror finish on the port walls. Given an absolutely free hand with an exhaust port (which seldom is possible, when you're modifying an existing engine) I would be inclined to make the exhaust passage's areas equal to those in an 8-degree cone. However, unless you have a dynamometer available for the verification of exhaust passage experiments, the best approach to this whole matter is to leave unaltered the exhaust-flange end of the port, and carefully blend the enlarged exhaust window into the rest of the port.
There is one more item to be covered here: the “bridged” exhaust port. The bridge to which I refer is a vertical rib that divides the exhaust port and creates two separate exhaust windows. Few engines have bridged exhaust ports, but if yours does, it should neither be removed nor narrowed – despite the obvious advantage in gas-flow to be gained thereby. One, and perhaps the most important, reason why we do not see more bridged exhaust ports is that the bridge is inevitably poorly cooled. In consequence, it tends to distort back into the cylinder slightly, and that occurrence becomes the more unfortunate because the piston/cylinder pressure point thus created is very poorly lubricated and localized seizing along the bridge is common. Narrowing the bridge reduces the heat-path into the metal surrounding the port, making the bridge's temperature just that much higher, while weakening it and making serious distortion all the more probable. Despite these considerable shortcomings, the bridged exhaust has been employed and will continue to be in certain applications, for it offers an opportunity to make the exhaust port window area extremely large without the usual attending difficulties with ring-snagging. On the other hand, it makes worse one of the real problems inherent in oversize exhaust windows -which is piston-crown overheating. Atthe moment of exhaust port opening, a great blast of fire goes jetting down the side of the piston, and this puts a lot of heat into the piston skirt. Clearly, any enlargement of the window exposes more of the piston to this kind of heating, while depriving the skirt of part of its contact with the relatively cool cylinder wall. Raising the exhaust port height has the same effect, and either of these power-enhancing modifications may, as they are carried progressively forward, require a compensating reduction in compression ratio to protect the piston.
Some engines do not respond nearly as well to increased exhaust port width as they should, which brings us to the overall problem of flow in the cylinder during the scavenging operation, and the transfer ports. As was stated earlier, we may imagine the mixture flowing from the transfers and neatly sweeping away residual exhaust products, but it does not really happen in quite that way. For one thing, there simply will not be a volume of gases delivered up from the crankcase sufficient to clear all the exhaust products from the cylinder. Seldom will the delivery ratio (the ratio of piston displacement to air actually pumped into the cylinder) be much better than 1:0.8, or 80-percent. Therefore, a cylinder having a piston displacement of 250cc will only have 200cc of air/fuel mixture coming into it through the transfer port - leaving, by implication, at least 50cc of exhaust gases trapped in the cylinder even if we assume a near-perfect separation of exhaust products and the incoming charge. Actually, there will be some mixing of the two due to turbulence, with the result that some part of the charge is lost out the exhaust port and there is a greater dilution of the fresh charge, with exhaust products, than would be assumed from the delivery ratio alone.
Delivery ratio is almost entirely a function of crankcase pumping efficiency, and the transfer ports' time-area factor - which is to say, the volume of the charge delivered into the cylinder is entirely independent of the number and disposition of the transfer ports. Those things are dealt with elsewhere in this book; we are concerned here with making the most of the mixture actually delivered, and in that regard the importance of the transfers' shapes and placement cannot be exaggerated. Often, the most subtle changes yield very large differences not only in peak power, but in the shape of the entire power curve, and it is all too easy to deal an engine a considerable injury while performing some minor alteration with a steady hand and the best of intentions. In this respect, I think it most unfortunate that the two-stroke engine cannot be driven below a minimum level of operating efficiency by even the most awful butchery of its transfer ports, as an engine thus served will continue to run, and run fairly cleanly, after that kind of surgery. How much better it would be if a serious departure from the optimum would produce a great fit of misfiring, or some other obvious evidence of distress. Unhappily, it will not, which means that an engine's transfer ports should be left strictly alone unless you have both the knowledge and the tools to make any modifications properly. The tools you will have to acquire on your own; knowledge is what I hope to provide with this book.
Perhaps the most valuable bit of information I can supply is that unless you plan to alter fairly radically the speed at which your engine makes its maximum output, there is no need to do anything beyond smoothing the casting flaws out of the transfer ports -and even that should be approached with some caution. Why? Because in scavenging efficiency so very much depends on symmetry of flow. Get one transfer port flowing conspicuously better than its mate on the opposite cylinder wall, and while you may have improved the delivery ratio slightly, the scavenging pattern will have been upset and power output will drop. As a matter of fact, I am inclined to think that the upper reaches of the transfer passages should be left entirely alone, unless to remove some conspicuous casting defect. It is in any case extremely difficult to do precise work up near the port window, and almost impossible to raise the port's roof without altering its angle - which is the wrong thing to be doing. All things considered, it is probably easiest to raise the transfer ports, when you want to increase the transfer timing, by raising the entire cylinder. A spacer under the cylinder will accomplish this, and it is usually a simple matter to trim the lower edges of the transfers and exhaust port to align with the edge of the piston crown at bottom center. Of course, this method shortens the intake timing, and so you will have to do a bit of trimming there as well, but anything is easy compared with trying to carve higher transfer ports with the port roofs held to their original configuration. Unless you discover that lifting the cylinder raises the intake port to the point where the piston-ring ends spring out into it when the piston moves down to the bottom of its stroke, or unless, for some reason, it is not possible to machine athickness equal to that of your spacer from the top of the cylinder to return to the original compression ratio. When either of those things present a problem, changes in transfer timing should be effected by cutting shallow troughs in the piston crown - which is a measure that can be used on the exhaust side, too, and should be used as a preliminary experiment to see whether the port timing you think you want is what you really need.
All present loop-scavenged motorcycle engines follow Dr. Schneurle's original pattern fairly closely, and many -perhaps most - are more or less exactly the same: Twin streams of incoming charge emerge from twin transfer ports flanking the exhaust port, and angle back across the piston crown and slightly upward, joining into a single stream at a point approximately two-thirds of the way back from the exhaust port. This stream is deflected upward by the rear cylinder wall, and then it sweeps up to the top of the cylinder to be directed back down the forward cylinder wall - moving the residual exhaust gases out of the exhaust port as it advances in that direction. There is, as you might expect, some turbulence generated by this activity, which is unfortunate because turbulence promotes the very kind of churning and mixing that should be avoided. But the turbulence is minimized when the flow is symmetrical, and there will therefore be less dilution of the fresh charge trapped in the cylinder at exhaust-port closing. Skewing either transfer port to one side, or lifting the upper edge of one slightly higher than the other will badly upset the scavenging pattern
Curiously, the ill-effects of such misalignments are more strongly felt at engine speeds below that where maximum horsepower is produced. Only a slight weakening of an engine's peak output is noted when modifications to its transfer ports leave the scavenging flow badly aimed, but there appears a marked instability at the point of maximum torque and below. This phenomenon would bear investigation, but I suspect that it reflects the fact that there is an almost explosive entry of the scavenging streams into the cylinder at high speeds. Also, it is a matter of record that the modern, short-stroke engine is generally less sensitive to scavenging patterns than the long-stroke engines of years past - this because sheer bulk flow, even if badly directed, does a much better job of scavenging the kind of low, large-diameter cylinder presented by the short-stroke layout than a taller, more slender cylinder. Still, and despite the fact that high horsepower numbers make good conversation, power range is going to be an extremely important consideration until such time as we have transmissions providing infinitely-variable ratios. So the best scavenging system is one that has good bulk-flow capabilities while maintaining a high degree of flow control.
This last item, the proper direction of the scavenging streams, is important for reasons beyond the reduction of turbulence and fresh charge / exhaust products mixing: Cylinders for high-performance engines need very wide ports to avoid excessive timing durations, which means that the ports must be crowded together too closely to entirely avoid the dangers of “short-circuiting” the charge. Having a high delivery ratio avails an engine nothing if the mixture streams emerging from its transfer ports are allowed to divert from their intended path and disappear out the exhaust port. Obviously, this danger increases as the transfer ports are crowded closer to the exhaust port; obviously, too, a degree of crowding is almost inevitable. On the available evidence, it would seem that the point at which short-circuiting becomes a problem is when the separation between the exhaust port's side-wall and the forward edge of the transfer port is decreased below 0.350-inch - that minimum dimension holding true for cylinders of 125cc and up. But it should be understood that this proximity is acceptable only when determined effort has been made to direct the scavenging streams sharply toward the rear cylinder wall.