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 we have not straightforward "organpipe" resonance, but a resonating flask consisting of the crankcase and intake track. As it happens, there is a fairly simple formula for determining the resonant frequency of flasks, which is
Where Vs is sonic speed
A is the cross-sectional area of the inlet.
L is the inlet pipe length.
Vc is the flask (crankcase) volume.
From the formula we 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 the intake pipe cross-sectional area; inversely proportional to the square root of intake pipe length. It has been established empirically that the intake tract length for maximum air delivery should be that which provides 0.75 pressure fluctuations "coincident with the period of inlet port opening". That is to say, if an engine's intake period in degrees is 120 degrees and its torque peak is 6000 rpm, then its intake period in time will be 0.003 seconds 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 if 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, ant the intake tract is in most engines far from being a simple tube (not only in 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.
Mathematical deduction of the intake track length is an arduous and lengthy task. 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, we have to fit out 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 3 ½ inches (measured from port window to the end of the exhaust stub) should be satisfactory. Quite obviously the sheer noise level produced by the stub -exhaust will make some form of muffling a necessity in many areas, and hence we 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 engine and the attached muffler should have sufficient internal diameter to prevent any back pressure from developing in the system. And that effect will be unmistakable, for the engine will pull very strongly when it comes "on the pipe". We will also find the intake pipe length can move the stub exhaust-equipped engine's power peak over a very wide speed range. Pipe lengths vary with any changes in intake timing, crankcase volume and intake tract diameter. It is quite possible to make a given intake timing work over a wide speed range by varying intake tract length and 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.
Mixture-strength problems can also occur due to plumbing between the carburetor and air cleaner, and it should go without saying that such a 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 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.
A common practice for extracting more-than-standard horsepower from a given engine was to increase 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:
Primary compression ratio =
Or CRp =
Where CRp is primary compression ratio
V1 is crankcase volume at BDC
V2 is piston displacement
For given port areas there is an engine speed at which maximum air delivery to cylinder occurs, and this engine speed is inversely proportional to crankcase volume, but 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 any deficiency in air delivery due to a crankcase volume too great for a given engine speed is fairly compensated by properly tuned intake and exhaust pipes. It is evident that little advantage is obtained by making the crankcase volume excessively small.
Starting with the exhaust system, we find that it is possible to evacuate the cylinder to well below the 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 the crankcase pressure " trapped" at a below-atmospheric value than 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 the atmospheric pressure by the pulsation 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. One might get the impression then, that it is time to start reducing crankcase compression rations - especially pumping loses (horsepower absorbed in doing the work) rise as 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 is desired, then we will find that primary compression ration of 1.5:1 represents something near optimum.
Broad range matter and the subject we are truly concerned in our micro car engine is another matter altogether. For this, we require much more flexibility from an engine. In the large-displacement classes particularly, where sheer engine size makes altogether too easy to get more horsepower, an engines ability to pull strongly over a very wide range is much more important than any peak reading. For such applications, the best engine is likely to be the 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, it is possible with intake lengths of about 30 inches to boost an engines' crankcase delivery ratio (volume of air pumped, per cycle/piston displacement) to as much as 1.2:1. 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 speed 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 winder to the round carburetor throat, and if we add the length of that passage to carburetor's length, then we will have a resonating intake system even if we don't want one. This choice is thus reduced to selecting a length that provides the best result overall.
Intake port shape has a definite influence on crankcase pumping efficiency. In general, the best flow coefficient for any given timing-area value will be obtained with 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 locations 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 he bottom of 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 that effectively is a slower rate of port opening, which is very useful in reducing intake roar. Also, more gradual opening of port tends to extend the duration of the sonic wave that is used, on is return trip, to "supercharge" the crankcase 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 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 however is 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 wall, 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 the carburetor should all align very neatly, without any steps between parts.
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 introduces 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 engine, it occupies a minimum a full 360-degrees of crankshaft rotation, with one piston stroke begin devoted to pushing exhaust products from the cylinder, past a valve in the cylinder head. Two-stroke cycle engines deliver power strokes twice as often at 360-degree intervals. But in such engines one finds an even greater mechanical complexity than in 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 the entire cylinder clearing and recharging for which 360-degress if crank rotations are reserved in the four-stroke engine must occur while the piston is halted at the bottom of the stroke. And lacking time for a more leisurely exchange of gases, the process must be helped along by extremely large port areas and high multiple exhaust valves in their cylinder heads. And a ring of windows around the cylinder's base through which scavenging air is forced by engine driven pump.
Present low-end two-stroke 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 shortcomings: First, there is the relatively incapacity of the crankcase as a scavenging-air pump, which prevents even the hope of having excess air to help in clearing the cylinder. Secondly, the use of 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 transfer ports open, which means that the exhaust phase must begin well in advance of 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 movement of the piston, moving to close the exhaust port and begin the effective compression stroke, is displacing the gases above the crown. Gas pressures always 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 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 greatest 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 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 of 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.
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 the 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 over-heating 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 and that is enough to allow the ring to spring out into the port window very perceptibly. And if the port is enlarged so that its width represents 70 percent of cylinder bore diameter ring, failure would almost certainly occur during the first revolution of the crankshaft.
Basically, the ring life is improved - at any given port width - by
Both of these measures are employed in all engines. The traditional port window shape is rectangular or square, with corners rounded to help prevent ring snagging. Assuming that port width does not exceed 60 percent of bore diameter, the radii at the port width should be about 15 to 20 percent of the port width. But as the port is widened those corner radii have to be made larger- to about 28 percent of port width when the latter approaches being 70 percent of the cylinder diameter. Actually, even these large radii will not 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 twice the port width, in ports having a width that is 60 percent of the bore diameter or less.
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 port windows 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 to loosen and some adrift. It should be obvious that this last difficulty will be most pronounced when the port window is not perfectly symmetrical 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. The safe and sensible approach is to begin at 62 percent, with a shape that is 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 we have opted for increasing the exhaust timing, as that will give us room to work above the existing port. Then, having established the initial shape, we 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 corners of the port; outright snagging will make its presence known in the appearances of scratches leading upward from the center of the port window.
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. The purpose of rounding-off the port window's sharp edges 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. It is of course necessary to work 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.
As regards to the exhaust port, a secondary function is served by providing a bevel, and radiused edges, around port window. There is 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 out just such a radius at the sides of an exhaust port. 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 hold-down studes, there is not enough room to widen the port as mush as would otherwise be desirable, and in that event the side-radiused port becomes a necessity.
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 be a flow increasing enlargement in area leading into the exhaust pipe during the outflow phase of the scavenging becomes a sudden constriction for waves returning to the cylinder from the expansion chamber. In fact, if the difference in the area at the port window and the outer end of the port becomes as great as 1:2, virtually all of the expansion chamber's resonant effects will be lost. What happens in such a case is that 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 1:1 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 of the transfer port). 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, her 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.
Sometimes the engine does not respond to the increased exhaust port width as it should, which brings us to the overall problem of flow in the cylinder during the scavenging operation, and the transfer ports. We may imagine the mixture flowing from the transfers and neatly sweeping away residual exhaust products, but it does not really quite happen 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 be much better than 80%. Therefore, a cylinder having a piston displacement will have air/fuel mixture coming into it through the transfer port-leaving by implication, at least 25% 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 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. We are concerned 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 to 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, we think it is most unfortunate that the engine cannot be driven below a minimum level of operating efficiency by even the most awful butchery of its transfer ports.
Perhaps the most valuable bit of information that we can supply is that unless we plan to alter fairly radically the speed at which the 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 has to be done with some caution 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 it may have improved the delivery ratio slightly, the scavenging pattern will have been upset and power output will drop. The upper reaches of the transfer passages should be left entirely alone, unless to remove conspicuous casting effect. 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 centre. This method shortens the intake timing, so we 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. 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 a thickness equal to that of the spacer from the top of the cylinder to return to the original compression ratio. When either of these 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 we think is what really we need.
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 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. 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 bulk flow capabilities while maintaining a high degree of flow control.
The proper direction of the scavenging streams, is important for reasons beyond the reduction of turbulence and fresh charge/exhaust products mixing. Cylinders 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 an high delivery ratio avails the 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. This danger increases as the transfer ports are crowded closer to the exhaust port, too, a degree of crowding is almost inevitable. It would seem that the point at which short-circuiting becomes a problem is when the separation between the exhaust ports side wall and the forward edge of the transfer port is decreased below .350-inch. It should be understood that this proximity is acceptable only when determined effort has been made to direct the scavenging streams sharply towards the rear cylinder wall.