Blow-back during the period between top center and intake closing is the limiting factor in establishing time-area values for all two-stroke engines except those fitted with automatic intake valves (i.e. reed-valves). But the problem is much less severe and more easily managed when an engine has rotary-disc intake valving. Because of the considerable mechanical complication attending this method of crankcase filling, rotary-valves have not been used much in mass-produced engines, and the inconvenient width they add now seems unacceptable as quite good results can be obtained with either reed or piston-port induction. Be that as it may, the rotary valve still is best in terms of sheer engine performance, whether arranged for maximum power or for an ultra-broad power range. The rotary valve is free of the really serious blow-back problem afflicting piston-controlled valving, and it offers much less resistance to flow than reeds. People who are currently so infatuated with the reed-valve concept should consider that in the world of karting, where there is much more experience with both reeds and rotary-valves than motorcyclists have accumulated, the two types of engines have been separated into different classes. Why? Because while the reed-valve engines are inexpensive, they cannot match the performance of those with rotary-valves. So, even though the rotary-disc valve is for the moment out of fashion, the pressure of competition may eventually return it to the fore despite its handicap in cost and bulk.
Principally because it is so free of blow-back problems, the proper time-area of a rotary intake valve is much higher than for the piston-controlled port. This does not mean that one may indiscriminately whack away at the valve disc without getting into trouble, especially on the port-closing side of the disc cutaway. The valve's opening point has an influence on power output, but in general it is relatively insignificant as compared with port-closing. Curiously, the best port-closing timing for a very wide variety of disc-valve engines is about 65-degrees after top center, but before deciding to apply that timing you should consider that any appreciable delay in closing the port, after the piston has started down from top center, will cause some low - speed blowback. At higher engine speeds inertia effects in the intake tract will overcome the slight blowback caused by the delayed intake closing, but there will be a loss of low-speed power. All things taken into account, the best approach here is to increase the valve-closing delay in very small stages, not more than 2-degrees at a time, until the desired result is obtained. Remember that retarding the intake-closing point moves the engine's power peak higher, while reducing power at the lower end of the range. Remember, too, that changes in the length or diameter of the overall intake tract, such as would occur in substituting a carburetor of some different size, will alter the point at which port-closing delay reaches its optimum.
It is fortunate that rotary-valve engines are relatively insensitive to the point of intake-opening, because there is no clear pattern in existing examples to lend our efforts direction. If there is a rule, it is that the point of opening for rotary valves is best established right at the point of transfer-closing if you want a very broad range of power. On the other hand, maximum power is obtained by opening the intake port somewhat earlier: from 130- to 145- degrees before top center, which means that the intake and transfer ports have overlapping open periods. This presupposes that the engine will have been fitted with a proper expansion chamber. The diffuser section in such chambers is capable of returning a negative-pressure wave to the cylinder having a below-atmospheric value of something like minus-7.0 psi, and as this partial vacuum is communicated to the crankcase via the transfer ports, the pressure inside the crankcase itself will drop well below atmospheric, At lower-than-peak engine speeds, particularly, there is a tendency for this below-atmospheric crankcase pressure to equalize itself by drawing part of the charge back down through the transfer ports, which neatly cancels an equal and important part of the work done by the exhaust system. There is little that may be done to counter this in a piston-port engine, beyond working with an exhaust system proportioned to give a long-duration scavenging pulse that will maintain the negative pressure in the cylinder until after transfer-closing. But when the engine in question has arotary valve, it is possible to open the intake side just as transfer passage back-flowing is about to occur, and balance the pressures with mixture drawn in through the carburetor instead of robbing from the cylinder. As you might guess, this neat trick doesn't work properly unless the intake-open period is delayed until after pressure in the crankcase has been pulled down to atmospheric or below - which means that intake-open timing is very closely tied to the exhaust system's pulsing and the overall flow characteristics of the transfer ports. A few hours work with an oscilloscope and pressure-transducers would get you right on target, because you would then know with a high degree of certainty the precise point at which crankcase pressure did in fact fall to atmospheric, but few people have that very expensive equipment and most will have to achieve the same result through a laborious process of cut-and-try. The same kind of cut-and-try in fact, as is needed to locate the optimum point for intake-closing.
All of the major influences on time-area requirements have been covered. There are others that could be major, but for practical reasons are not. Crank case compression ratio is one. Both intake and transfer time-area requirements do vary in inverse proportion with crankcase compression ratios, simply because as crankcase pressures are reduced so are the pressure differentials that cause gases to move. You are spared having to worry overmuch about this factor because virtually all modern engines have primary (crankcase) compression ratios very near 1.5: 1 - this having proven to be best for almost every application, and is mechanically easy (it is what you get with flywheels and crankcase of normal proportions). Exaggerated angles of entry into the cylinder could also create ports with window areas misleadingly larger than the passages themselves, but again this condition- while always present in some degree - is seldom serious enough to warrant one's worry while working time-area calculations. Finally, extraordinarily poor casting techniques and/or badly misshapen port cavities could lend ports such low coefficients of flow as to upset one's calculations, but one almost never sees outstandingly bad work from any of the major manufacturers. All of their casting work is quite good, and will not be much improved by even the most painstaking carving and polishing on your part -which may be discouraging, but does tend to preserve the validity of the time-area values I have presented here.
Back before the unpleasantness of the early 1940s, a number of two-stroke motorcycle engines were built with external scavenging-air pumps – much in the fashion of today's GMC truck engines -despite the resulting penalty in bulk and manufacturing cost. It was then thought that no engine relying on ordinary crankcase pumping for scavenging-air delivery could hope to compete with the better four-strokes in terms of specific power output, and there was every reason for that pessimism: For one thing, no one in his right mind would design a piston-type air pump with as much clearance volume as inevitably exists in an engine's crankcase. Neither could it be considered desirable to pre-heat the scavenging air before it is delivered to the cylinder, which is precisely what happens in the crankcase-scavenged two-stroke engine. Finally, using the two-stroke engine's crankcase as a scavenging pump condemns the connecting rod bearings to a diet of too much raw fuel and too little oil. Only the fact that resorting to a separate, external scavenging pump roughly doubles the cost of a single- or twin-cylinder engine, making it more expensive than a four-stroke engine of equivalent power output, has kept us at work on the crankcase-scavenged two-stroke. And over the years, out of necessity, we have learned to make the crankcase function as a pump with an effectiveness that would astonish the engineers of fifty years ago.
Much of the improvement in the air-delivery capabilities of crankcase pumping can be traced back a half-century to a two-stroke stationary engine (used to drive an electrical generator) that neither employed an external scavenging pump nor relied upon its crankcase to do the job. Instead, the work of moving air through the cylinder was performed by the effects of sonic waves and inertia in the engine's intake and exhaust pipes. The former was connected directly to the cylinder's scavenging ports; the latter was a conventional, if lengthy, straight pipe. This engine had to be motored up to its operating speed and a blast of compressed air directed into its intake pipe to start it firing. But then a combination of wave and inertia activity in its intake and exhaust pipes would take over, to scavenge and recharge the cylinder, and the engine would thud-thud merrily along at that speed until it ran out of fuel or broke. Being utterly dependent on the resonant frequencies of the attached plumbing, it would of course run at only one speed-but that is a virtue, rather than a disadvantage, in an engine used to turn a generator. And while the engine described was neither very powerful for its size nor particularly efficient, it was simple and trouble-free… and pointed the way for the hyper-powerful racing engines of another, later time.
Obviously, engines capable of running at only one speed, and which must be cranked up to that speed before they will start, are of limited usefulness in the context of motorcycling (although a couple of road racing engines approaching that condition come to mind). Even so, few of the current crop of high-performance engines would perform at all well without help from resonant pipes. Rotary-valve engines function very nicely without much assistance from intake-tract resonance, but those using piston-controlled intake ports -which necessarily have their intake-open period spread symmetrically before and after the piston reaches top center - tend to regurgitate a lot of the mixture drawn into the crankcase by the ascending piston unless this backflow is in some manner prevented. Combined inertia and wave activity, which may attain pressure values in excess of 1.5 atmospheres, are enough to balance any pressures created by the descending piston - even if port closing is delayed until as much as 100-degrees after top center, as is the case in highly-developed road racing engines. Of course, these effects are extremely transitory, and will prevent backflow only if their arrival is properly timed.
Timing wave arrivals on the two-stroke engine's exhaust side is a relatively simple task, as the exhaust system's resonate frequency is almost exclusively a function of its length. A far more complex situation exists on the engine's intake side, for there you have not straightforward “organ pipe” resonance, but a resonating flask consisting of the crankcase and intake tract. As it happens, there is a fairly simple formula for determining the resonant frequency of flasks, which is
Where Vs is sonic speed (usually about 1100 ft/sec)
A is the cross-sectional area of the inlet
L is the inlet pipe length
Vc is the flask ( crankcase ) volume
From the formula, you will see that resonance in a flask, or in an engine's intake system, ( which includes the crankcase ) , frequency is directly proportional to the square root of intake pipe cross-sectional area; inversely proportional to the square root of crankcase volume; and also inversely proportional to the square root of intake pipe length. Researcher Fujio Nagao, of Kyoto University, who has done much of the basic investigation in this field, has established that intake tract length for maximum air delivery should be that which provides 75 pressure fluctuations “coincident with the period of inlet port opening”. That is so say, if an engine's intake period in degrees is 120 degrees and its torque peak is at 6000 rpm, then its intake period in time will be 0.003-second and the intake system's frequency coincident with 75-percent of that, or 125 cycles per second.
Unfortunately, this straightforward picture begins to get very complicated as one tries to apply any of the preceding to the concrete example. We can, for instance, only guess at sonic speed, as it is dependent on temperature and one can only guess at the temperature of a stream of gases simultaneously being cooled by evaporating gasoline and heated by surrounding metal. But that is not the greatest difficulty one faces in calculating the proper length: The flask in question here has a volume that varies continuously with the piston's motions, and the intake tract is in most engines far from being a simple tube (not only area but shape will vary from the carburetor intake bell to the port window). Moreover, the inertia of the fast-moving column of gases in the intake tract must also have an effect. In short, calculating intake tract length is a highly complex problem, and while a trained engineer with a heavy background in higher mathematics could probably do the job given sufficient time, it really is a task for the computer - and even then it will be necessary to use abbreviated formulae to avoid accumulating too many expensive computer hours on the problem.
In all probability, those who will read this book will have neither a computer nor the spare hours for a theoretical determination of intake tract length. Thus, it becomes necessary to arrive at this length experimentally, which fortunately is a much shorter (and more accurate) means of accomplishing the same thing. The first step in that direction is to isolate the influence of intake tract length on the engine, which means removing any effect the exhaust system may have on the results obtained. To do this, you will have to fit your engine with a stub pipe too short to be effective within the engine's projected operating range, yet long enough to prevent the “back-breathing” of air in through the exhaust port to upset mixture strength and thus mask intake-length influence. For small-displacement cylinders, a total exhaust length of 31/2-inches (measured from port window to the end of the exhaust stub) should be satisfactory. Middle-displacement engines should have a 4-inch stub-exhaust, and those with a 350-400cc cylinder displacement a length of 4 1/2-inches. Quite obviously, the sheer noise level produced by stub-exhausts will make some form of muffling a necessity in many areas, and where such is the case you can add muffling without upsetting the experiment by introducing a large-volume chamber into which the exhaust-stub empties. This chamber should be as large as can be fitted on the motorcycle without making it impossible to ride, and the attached muffler should have sufficient internal diameter to prevent any backpressure from developing in the system. The hardware that makes up this no-effect exhaust system need not be beautiful, nor tucked in on the motorcycle neatly enough to permit anything more than straight-line riding, but it should be easy to install because you will have to re-tune the intake length with every change in port timing and/or carburetor diameter. There will, of course, be a change in the system's frequency with every change in crankcase volume as well, but such changes will be too small to worry about in most instances.
For intake-length experiments, you will need not only the stub-exhaust system, but a tachometer on your motorcycle and either a long, straight road (or trail) or a hill. These last are required because you must load the engine heavily enough to allow you a good, seat-of-the-pants reading of the engine speed where the pipe has its effect on power. And that effect will be unmistakable, for the engine will pull very strongly when it comes “on the pipe”. You will also find that intake pipe length can move the stub exhaust-equipped engine's power peak over a very wide speed range. My friends at McCulloch, who acquainted me with the stub-exhaust testing technique, tell of tests they have performed with an engine having an intake period of 120-degrees, and with pipes ranging from 5- to 9 1/2-inches in length they were able to move the power peak anywhere from 3000 rpm to 8000 rpm. The 9 1/2-inch length gave substantially the same power at 3000 rpm and 4000 rpm; at 5000 rpm, an 8 1/2-inch was best; at 6000 rpm and 7000 rpm the same power could be had with either a 9 1/2- or 6 1/2-inch length; and the same was true at 8000 rpm, but with lengths of 5 1/2 and 8-inches. From these tests, it was determined that the best compromise length for the entire range 3000-8000 rpm was 8-inches; a 5 1/2inch length was slightly better for the range 6000-8000 rpm. It should be understood that these lengths only apply to this specific example, and would vary with any changes in intake timing, crankcase volume and intake tract diameter. They are given here only to illustrate that: A) it is possible to make a given intake timing work over a wide speed range by varying intake tract length; and B) that the same length may produce more than one peak, or conversely that more than one length may be effective at any given power peak.