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Back in the days when pistons were uniformly poor and two-stroke engines wouldn't be run very fast, wrist pin bearings were almost always a simple brass bushing. Such bushings work very well in four-stroke engines, but lubrication is much less lavish in the crankcase-scavenged two-stroke and added difficulties are created by the essentially uni-directional loads placed upon it, which prevent the piston pin from lifting away from the lower part of the bearing and admitting oil to the load-carrying surfaces. For those reasons, the plain bushing has now almost universally been replaced by “needle” roller bearings, which are more easily penetrated by such oil as is available and in any case need much less oil. This last is of very particular importance in high output engines, as the heat flowing down from the piston is certain to thin any oil present to a viscosity approaching that of water. But all these difficulties not withstanding, the needle-roller bearing is wonderfully trouble-free, and if you encounter problems at the hinge between the connecting rod and piston pin, those problems will almost invariably be with breakage of the bearing cage. Given the extremely low rotational speed of the bearing in question, no cage is really needed except to make engine-assembly easier: the cage holds all the needle-rollers in place while the piston is being fitted to the connecting rod. The arrangement certainly makes working on the engine less complicated, but as it happens, the cage becomes the bearing's weakest link. Piston acceleration at high speeds is also applied to the bearing cage, and it may shatter under the strain - which sends a shower of particles from the broken cage and loose needles down into the crankcase. The debris thus liberated invariably gets pumped up through the transfer ports, into the cylinder, and more often than not a roller will get trapped hanging half out of a port by the piston with dire consequences to both.

Yamaha's TD1 was particularly prone to small end bearing cage failures, and I learned the hard way to replace these bearings if I saw over 11,000 rpm on the tachometer even for a moment, for their cages required only a moment's battering before cracks would start to spread and outright disintegration soon followed even if I indulged in no more excursions past the red-line. This difficulty has been overcome with cages made of tougher material; it is possible to accomplish the same thing by using crowded needles and no cage at all, which does require that a washer be fitted on each side of the connecting rod, to take up clearance so that the rollers cannot escape. Getting the thing assembled (with the roller glued in place with grease) is enough to make strong men weep with frustration, but it absolutely insures reliability at this point in the engine and is a measure worth remembering if problems with broken wrist-pin bearing cages do occur.

McCulloch, the chain-saw people, have used an arrangement similar to the one just described for years, but they have reasons other than simply working around bearing cage failures at the wrist-pin end of the rod. It was discovered at McCulloch that failures at the crankpin bearing were traceable to the thrust washers most manufacturers of two-stroke engines use to center the rod on the crankpin. These washers usually are made of brass, or steel with a copper coating, and they do not find high rubbing speeds and scanty lubrication at all agreeable. At very high crankshaft speeds, they register their protest by overheating, and this causes a rise in temperature all around the connecting rod's big end, which thins the oil present enough to create yet more friction, more overheating, until at last the thrust washers, roller bearing and cage are hot enough to “flash” the oil. At that point, lubrication is nil and friction quickly melts the bearing cage and wears flats on the rollers. McCulloch's engineers reasoned that the point of failure could be pushed upward materially simply by removing the thrust washers, which is what they did. Of course, the connecting rod still had to be centered over the crank, but this task was given to a pair of thrust washers up inside the piston. The improvement in terms of elevating the McCulloch kart engine's maximum crank speed was in the order of 1500 rpm, and it is worth noting that Yamaha borrowed this idea for use in the 17,000 rpm GP engines the company raced in 1968. It is interesting that in those engines, the piston rings were only 0.6mm in thickness.

Crankpin bearing failures also stem from the use of excessively heavy bearing cages. Sheer rotational speed is not enough to burst a cage of such small diameter and mass, but the fact that the cage must accelerate and decelerate, relative to the crankpin as the connecting rod swings, will cause difficulties unless the bearing cage is very light. In effect, the rollers must push the cage up to speed and then slow it, and if the cage has enough inertia it will resist this pushing and pulling enough to skid the rollers - at which point they momentarily become a plain bearing- a job for which they are poorly constituted. The skidding rollers generate a lot of heat, through friction, and the heat leads the bearing into the same deteriorating cycle to outright failure as was outlined for the thrust washers. Most modern engines have steel crankpin bearing cages, copper- or tin-plated to provide a low-friction surface to bear against the rollers, crankpin and connecting rod eye. These replace the phosphor-bronze cages of the recent past - which replaced the inelegant aluminum and brass cages of a yet-earlier era. But the best current “big-end” bearing cages are made of titanium and silver-plated. Experimenters with near-unlimited funds may like to try titanium bearing cages, but when having them made they should know that the bearing retaining slots must be machined with edges parallel to within 1/200 with each other and with the crankpin (assuming a parallel condition between cage and crankpin axis). It is not a job for someone with a bench-vise and a file. On the other hand, if employing silver-plated titanium cages and moving the thrust washers from the crankpin to the piston will elevate your engine's red-line by 2000 rpm, then they clearly will pay dividends in horsepower - if port-timing, etc., is adjusted correspondingly.

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

Crankshaft main bearings seldom are troublesome, except in engines that have been in storage for a long time and have had corrosion at work in these bearings -or unless the bearings have been mishandled. Bearing steels are very tough, but you definitely can pound small pits in the races by injudicious use of a hammer, and pits also can be formed by rusting. Bearings damaged in either fashion should be replaced, as the pits will soon spread and become minor trenches, as a result of an activity called “Brinelling”, which actually is a form of work-hardening. The bearing's rollers and races have casehardened surfaces, but the metal under this thin case is relatively soft, and it is compressed and released (at any given point) as the bearing turns under a load. If the load is high enough, or the bearing in service long enough, the repeated compressions will literally fatigue the metal, and tiny particles of the surface will start flaking away - which becomes visible as the “tracking” seen in the races of a worn-out bearing. Any bearing will start flaking at some point in its life; bearings with races damages by rust, etc. will begin such flaking almost immediately. Incidentally, in very highly loaded bearings the flaking may be started by the sharp edges around any interruption in the bearing's surface, if the rollers pass over those edges. Oiling slots in the rod's big-end are prone to develop this kind of failure, and the same sort of flaking is sometimes observed around the oil feed holes in the crankpins of engines equipped with “direct-injection” oiling systems, like the Suzuki’s and Kawasaki’s. Remove the sharp edges, and you remove the problem - if any. There is sufficient margin of strength in stock production engines so that the problem does not occur; you may find it in the course of reaching for crank speeds substantially above the stock specification.

Somebody is always telling me about having an engine “balanced”, and I always smile nastily when the engine in question has fewer than four cylinders. In point of fact, the single-cylinder motorcycle engine cannot be brought into dynamic balance, for if you counterweight the crankshaft to compensate for the full weight of the piston and rod, you will simply have moved the shaking force from being in-plane with the cylinder axis 90-degrees. "Balancing" one of these engines consists of finding a balance factor, in percentage of reciprocating mass, which is kind to the engine's main bearings and does not excite resonance in the motorcycle's frame. In-line twin- and three-cylinder engines always have a rocking couple. By and large, the stock crankshaft counter-weighting will be correct for most applications, and unless you want to get into a really lengthy experimental program there is nothing to be gained in making changes.




There are gains in power and reliability to be had from carefully aligning your crankshaft and main bearing bores, and in getting the cylinder axis precisely perpendicular to the crankshaft. As it happens, there is more variation in production tolerances when the various parts of a crankshaft are made than can comfortably be tolerated in a racing engine. Crankpin holes in flywheels are not all precisely the same distance from the main shaft axis; factories "select-fit" these parts, and you can be fairly certain that a new crankshaft is true, but if you manage to ruin any of its flywheels, do not assume that a replacement flywheel, selected at random from the nearest parts bin, will be a satisfactory replacement. Crankpin holes, in facing flywheels, should be matched to within 0.0002-inch with regard to their offset from the main shaft. If your local source cannot supply a single replacement wheel within that tolerance limit, I strongly urge that you purchase a complete, new crankshaft - with flywheels matched at the factory. And when rebuilding a crankshaft, with new crankpins and bearings, be certain that it is aligned to at least the tolerances suggested by the manufacturer's workshop manual. Also, check your crankcases for main bearing-bore alignment - and, more important yet, that the cylinder is exactly perpendicular with the crank axis, for any tilting will be reflected in added friction in the hearings (especially at the thrust washers) and in the piston itself.

Do not attempt to second-guess the manufacturer with regard to crankshaft and crankpin bearings unless you have very specialized knowledge in this field or can obtain the advice of someone who is an expert. Main bearings, particularly, should not be replaced with just anything that will fit, as n very special kind of bearing is employed in these applications, with clearances to accommodate the expansion and contraction of aluminum bearing housings. And the same cautionary note must be added with regard to crankshaft seals, which in the high-speed, two-stroke engine must survive extremes in temperatures and rotational speeds with very scanty lubrication. Not so very long ago, seal failures were common, but now that means have been found to Teflon-coat seal's lips, trouble is usually encountered only when the seals have been damaged in the course of installation. So handle the seals carefully, and pre-coat them with a good high-temperature grease before assembling your engine. You can also improve their reliability somewhat by polishing the area on the main shafts against which they hear to a glassy finish. The seals themselves will polish the shaft eventually, but at considerable expense to their working life.

By and large, problems with piston, connecting rod bearings, crankshaft and seals can be avoided simply by following the recommendations made in the manufacturer's shop manual. The single exception to this is in the fit between piston and wristpin, for the very high temperatures in a modified engine tend to cause a breakdown in the lubrication between pin and piston. Trouble can be avoided in the racing engine if the wristpin is a light, sliding fit through the piston; it should slip through of its own weight, without forcing, for if it is tight enough so that you have to tap it through with a mallet, you eventually may have to remove it with a hydraulic press. Too-light fits may be corrected by using an old wristpin as a lap, and a dash of some fine, non-imbedding lapping compound to polish out the piston's pin-bore to size.



For the Otto-cycle engine, of which the two-stroke is an example, there is a theoretical level of efficiency, in terms of converting heat into work, referred to in basic engineering texts as “air standard efficiency”. In this, it is assumed that the cylinder is filled only with dry air, and heat then added, which ignores the fact that in practice the air contains some moisture and a considerable percentage of hydrocarbon fuel. Even so, this theoretical level of efficiency, calculated against compression ratio, provides a useful yardstick against which actual efficiency can be measured - and it tells us a lot about the effects, on power output, of compression ratio. For example, at a compression ratio of 5:1, air standard efficiency is 47.5-percent, while at 10:1, it is 60.2-percent. That is, of course, a very great gain, and the consequences - measured at an engine's output shaft-are the reason for many experimenters' fixation on “raising the compression”. Certainly, increases in compression ratio, which may be accomplished simply by trimming a few thousandths of an inch from the cylinder head's lower surface, can work minor miracles with an engine's performance.

But higher compression ratios can also bring about a mechanical disaster: improvements in power gained in this manner are purchased at a disproportionate cost in peak cylinder pressure, leading to reduced bearing life and sometimes to an outright failure of a connecting rod or crankpin. Moreover, because the higher pressures are reflected in a proportionately greater side thrust at the piston, frictional losses are such that net power gains are always less than the improvement one would expect from the calculated air standard efficiency. Finally, heat flow from the combustion gases into the surrounding vessel (piston crown, cylinder head, and cylinder walls) rises increasingly sharply with compression ratio, so that a number of thermal-related problems intrude into the already complicated relationship between compression ratio and power.

The worst of these problems is the overheating of the piston crown. A too high compression ratio will raise piston crown temperatures to the point where heating of the mixture below the piston, in the crankcase, reduces the weight of the charge ultimately trapped in the cylinder during the compression stroke to such extent that net power suffers -no matter what Mr. Otto's air standard efficiency formula may say. And if the compression ratio is high enough, heat input into the piston may raise the crown temperature to the point where detonation and then pre-ignition occur. These phenomena will, in turn, very quickly further raise piston crown temperature to such extent that the piston material loses enough of its strength to yield to the gas pressure above – the piston crown then becoming either concave (which drops the compression ratio to a tolerable level) or develops a large hole (and that reduces the compression ratio to zero:zero).

Many people have encountered this last effect, and the tuner's one-time favorite ploy of “milling the head” has fallen into disrepute. But it also is possible to encounter trouble without recognizing it: There is a delicate balance between gains from increased compression ratios and losses due to increased temperatures -which appear not only at the piston's interior, but also throughout the crankcase, crankshaft, rod and all the rest of the engine's interior contacted by the air/fuel mixture. When these parts are hotter, the mixture's temperature is also raised, along with its free volume. Thus, the mixture's temperature-induced efforts to expand inevitably force part of it out the exhaust port, and as power is related very closely to the weight of the charge captured in the cylinder, this heating shows up as a power loss. The trick is to balance crankcase heating and compression ratio. There is an optimum combination for every set of conditions, but finding that optimum without heat-sensing equipment and a dynamometer is exceedingly difficult.


Date: 2015-12-17; view: 306

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