Datsun camshafts & valve timing by Racer Brown

Datsun camshafts & valve timing by Racer Brown

Racer Brown was a licensed mechanical engineer and was once technical editor for hot rod magazine. Racer Brown cams are standard part numbers for Mopar, Nissan Competition Department and others.

Grab yourself a stubby, get comfortable and read into the fascinating world of Datsun and other engine valve train operation. Once read a couple of times you will likely understand more about camshafts, valve train geometry and the things affecting them than most of your mates and some automotive engineers as well.


Datsport has been extremely fortunate to be permitted to reproduce what is widely recognised at being the best article ever written on camshafts and their operation.

We thank Racer Brown Incorporated for this opportunity. Please be aware that this information is copyright and may not be reproduced in any format whatsoever. This article has not been reproduced in its original form since 1972.

Please respect the copyright as we do not wish to jeopardise the right to reproduce this technical article as it will deprive Datsun enthusiasts of the most important information affecting the performance of Datsun L series engines.

Chapter one



Piston at top centre (TC) compression stroke, both valves closed.

Explosion BANG! Piston forced to bottom centre (BC) during power stroke by explosion in cylinder. Piston at BC of power stroke. Exhaust valve opens. Piston pushes exhaust gases from cylinder by moving toward TC.

Top centre exhaust stroke. Exhaust valve closes. Intake valve opens. Piston moves toward BC on induction stroke.

Bottom centre of induction stroke, intake valve closes. Piston moves toward TC on compression stroke, compressing combustion fuel, both valves closed. Cycle complete.

Ready for another explosion BANG! Sound dull? It is dull. The mechanical-textbook world must necessarily and primarily be concerned with theoretical aspects of mechanical problems and their solutions. Practical considerations are secondary to the teaching and learning of theories. In the "real" world, the situation is usually reversed. And so it is, and has been with the four-stroke cycle internal (infernal, if you prefer) combustion piston engine. Practical applications of concepts, ideas and modifications have had such incredibly mind-boggling results since the inception of this type of power plant that the original theoretical approach, as it is still taught and learned, is as antiquated as if it had been carved by hand on tablets of stone. This doesn't suggest that the basic premise, as taught and learned, is necessarily wrong. It does suggest that this premise has been oversimplified to unattainable extremes. Perhaps this isn't all bad for the novice, but for more advanced students of the four-stroke cycle engine, this most basic approach falls flat on its nose, simply because it is inadequate. And DULL. The latter factor alone has probably been the direct cause of more dropouts in this area of study than such unrealistic teaching methods could ever hope to gain.
While the study of the four-stroke cycle engine may not be the most stimulating pursuit in the world, it is anything but dull. Moreover, the practitioners of practicalities - the imaginative designers, innovators, inventors, doers - have almost completely rewritten the original theoretical premise so the original tablets of stone, carved by hand, may be safely consigned to the gravel pit.

An explosion!

BANG! Good grief! But there was an explosion back about 1680 when, in one of the earliest recorded attempts, a gentleman named Christian Huygens tried to make a functional engine using gunpowder as fuel. But history doesn't tell us about the immediate after-effects. Probably a case of no survivors during a one-shot experiment. BANG! (FACT!)
Now bear with me through the dreary details of a most basic four-stroke cycle engine. The assembly consists of a crankcase or cylinder block, in which resides a crankshaft. The cylinder block is bored to accept a piston assembly that reciprocates in the bore. From a front or rear elevation, the cylinder bore centre (usually, not always) coincides with the crankshaft axis. Longitudinally, the bore centre is perpendicular to the crankshaft axis. The crankcase also contains main bearings, which support the crankshaft main bearing journals. Between main bearing journals, the crankshaft has a crankpin parallel to the crankshaft axis but offset from it by an amount equal to exactly one-half the piston stroke. The lower part of a connecting rod contains connecting rod bearings and this end of the rod is clamped to the crankpin. The top end of the connecting rod hooks up to a piston pin located in the piston assembly, the piston pin also being parallel to the crankshaft axis. These components represent the mechanical linkage necessary to transform the piston's reciprocating motion into rotary motion at the crankshaft.
A cylinder head assembly, usually removable from the cylinder block, is located at the top end of the cylinder bore, away from the crankshaft. Among other bits of ironmongery, the cylinder head assembly contains an intake port and (usually) poppet valve, an exhaust port and (usually) a poppet valve and some provision to accept a spark plug, or other means of igniting a combustible fuel under pressure. A cavity in the cylinder head, surrounding the valve and spark plug, forms the combustion chamber, although it can be formed in the top of the piston, or a combination of both.
There are sub-systems for providing the cylinder with the combustible fuel (induction system); for carrying the products of combustion from the cylinder to the atmosphere - very difficult to do legally, these days - (exhaust system); for opening and closing the valves at the proper times and for the proper intervals (valve train system); for delivery of lubricant under pressure to critical areas (lubrication system); for carrying off excess heat generated by normal engine operation and dissipating the heat to the atmosphere (cooling system); for supplying electrical energy to the spark plug for igniting the combustible fuel in the cylinder under pressure (ignition system).
This is the basic lump, the assembly from which we expect some useful work to be done for us because we are too lazy to do it ourselves. Of course, there are other accessories such as p/s, p/b, a/c, 8-track stereo, etc. But I'm talking about the standard item, the poverty-model.
A "combustible fuel" was mentioned, without which the engine couldn't fire its first power stroke. It really doesn't matter if the fuel is a mixture of hydrogen peroxide, diesel oil and oxygen (not for amateurs), the more common alcohols, to gasolines, to old socks as long as its behaviour under heat, compression and ignition is uniform and predictable. However, in this context, plain old pump-type gasoline is the indicated fuel. We will make the assumption that the induction system does an acceptable job of delivering a mixture of air and fuel in the correct proportions to the cylinder. Plain old atmospheric-type air is necessary to support combustion of the fuel in the cylinder, the oxygen of the air being the necessary compound.
As the term implies, the four-stroke cycle engine completes its cycle in four complete strokes of the piston within the cylinder. One stroke of the piston is defined as the motion of the piston from the top of the cylinder bore (top centre) to the bottom of the bore (bottom centre), or from BC to TC. With the direct mechanical link between the piston and the crankpin of the crankshaft formed by the connecting rod, the motion of the piston can be, and is, related to any number of degrees of crankshaft rotation from zero to 720. That's right - 720 degrees - not 719 or 721, but exactly 720 degrees of crankshaft rotation is required for the completion of one cycle of the four-stroke cycle piston engine. The four strokes of the piston are all named to minimise confusion because the piston positions are duplicated for each complete cycle. Starting in order, the correct sequence of strokes is: Power (or expansion): Exhaust: Induction (or intake): Compression. Cycle complete and ready to start the next four strokes to form the next cycle. You can start anywhere you like, but the correct sequence of strokes must be followed. Example: Induction: Compression: Power: Exhaust. Cycle complete. You cannot transpose the compression stoke for the exhaust stroke any more than you can transpose the induction stroke for the power stroke. The sequence of strokes is fixed and cannot be changed.
Ignoring for the moment Tablet I (that explosion BANG! Episode) and looking at Tablets II, III, and IV in the volume of hand-carved stone tables, we find that the theorists would have us believe that at BC of the power stroke, the exhaust valve opens, at TC of the exhaust stroke, the exhaust valve closes. And the intake valve opens, and at BC of the induction stroke, the intake valve closes. How do the valves get themselves open or closed? And why at these precise points of piston travel? They are not vague when they say "valve opens" or Valve closes," but we are left to assume that the valves are fully open or fully closed, with no lapse between the two valve positions, and we are left to further assume that the valve motion is instantaneous. Swell. If valve motion can be instantaneous, why can't piston motion be just as instantaneous?? NO WAY!!
This is where the original hypothesis, as writ by hand on tablets of stone, drops completely dead. The theorists conveniently neglected to include the on vital, absolutely essential element in the proceedings. Time. That's right. T-I-M-E. There is NO action, reaction, force, counterforce, or whatever, on the face of the earth or in the entire universe that is instantaneous. Each and every occurrence in the universe, no matter how infinitesimal or how vast, no matter what its nature, all have one common denominator: They all take time. Maybe light-years, maybe milliseconds, but time is the essential element. Even that BANG! Stupid explosion takes time. If it were mechanically and physically possible to bring about instantaneous events, and if they were applied to our laboriously-constructed engine, we'd better be able to duck instantaneously because all we'd ever have to show for our efforts would be a loud and widespread case of instant shrapnel. So while time allows our engine to function in a normal manner, it also allows us opportunities to correct our mistakes.
Because of the necessary time factor, valve motion overlaps the piston positions of top and bottom centres and during on period of the cycle the valve motions overlap themselves. Even the cycles overlap each other. Sometimes I wish it were possible to view a functional engine through some kind of simplified but cock-eyed glass prism so that all engine events could be seen directly and in their proper perspective, rather than relying upon mere words.




Now follow the piston motions and the interrelated valve motions through four strokes of the piston for one complete four-stroke cycle. Let's start at exact TC of the power stroke, with both valves closed. Ignition of the air/fuel mixture occurred at an earlier point so that at exact top centre there is seething, flaming, furiously violent activity within the cylinder and combustion chamber, as normal combustion is in a process.

This process generates a very rapid but relatively uniform increase in cylinder pressure and temperature. However, at TC the point of maximum cylinder pressure has not been reached. If it had, the piston/connecting rod/crankshaft assembly would simply be shoved out the bottom of the crankcase. The piston, being the only moveable component within the combustion chamber at the moment, is forced toward BC by the still-expanding and pressurising gases in the combustion chamber.

At a point represented by about 15 crankshaft degrees of rotation past TC, pressure within the cylinder reaches its maximum value. This places enough of a mechanical leverage advantage from the piston, to the connecting rod, and to the crankpin to force the crankshaft to rotate, producing useful work at the crankshaft, in rotary motion we can capture and measure as torque and/or horsepower.

At a point before the piston reaches BC of the power stroke, the exhaust valve starts to open. At this point, most (not all) of the force of the expanding and still-aflame gases in the cylinder has been captured by the piston and transmitted to the crankshaft. Although the gases are still expanding, cylinder pressure is diminishing, and what pressure remains is relatively insignificant to the total process. The exhaust valve opening point before BC gives a period of cylinder "blow-down," during which most of the remaining cylinder pressure escapes from the cylinder past the exhaust valve, through the exhaust port and into the exhaust system. Opening the exhaust valve before BC permits the valve to be fairly well off its seat as the piston reaches BC, although the valve is not yet at full lift, which reduces some of the negative work or pumping losses applied to the piston during the exhaust stroke.

Bottom centre of the power stroke. End of piston stroke 1. Also end of cylinder "blow-down" period. From the piston positions of exact TC to exact BC of the power stroke, the crankshaft rotates exactly ½ revolution, or 180 degrees.


Exhaust Stroke


Next on the sequential piston-stroke agenda is the exhaust stroke. However, evacuation of exhaust gases from the cylinder began with the exhaust valve opening at a point before BC of the power stroke. This represents the first of four periods of the cycle during which valve motion overlaps piston position.

As the piston reaches BC of the power stroke, it changes direction and the exhaust stroke begins, the exhaust valve is still opening, and the piston starts to push the remaining exhaust gases out of the cylinder past the exhaust valve. Whatever pressure is applied to the top of the piston by the exhaust gases represents negative work or pumping loss; the higher the pressure the higher the pumping loss, which reduces the effectiveness of the positive work gained in the power stroke. With this in mind, we can readily see the advantage of the blow-down period in reducing residual cylinder pressure, thereby reducing pumping loss during the exhaust stroke. At a point well before TC of the exhaust stroke, the exhaust valve reaches maximum lift and starts to close.

As the piston approaches TC of the exhaust stroke, the intake valve starts to open. This is the second of four periods during the cycle that the valve motion overlaps piston position. This also signifies the start of the valve overlap period, a segment of piston motion and/or crankshaft rotation during which the exhaust valve (still closing) and the intake valve (starting to open) are both open at the same time. The piston is slowing down before it reaches TC, therefore it doesn't have the same amount of force that it did during the first part of the exhaust stroke. As a result, the departing exhaust gases are also slowing and it becomes a practical impossibility to remove all of the exhaust gases from the cylinder before the exhaust valve closes. In fact, some of the exhaust gases are diverted from their intended path toward the exhaust valve and instead point themselves toward the intake valve and the partially exposed intake port. There is no hesitation or indecision. This occurs simply because the pressure around the closing exhaust valve is momentarily higher than it is around the opening intake valve. The lazy exhaust gases take the path of least resistance ahead of the advancing piston as an escape route. This action is inevitable as long as both valves are open at the same time for some amount of piston travel and/or crankshaft rotation. Exhaust gases in the intake port represent a diluting agent to the air/fuel mixture in the immediate vicinity. And, the exhaust gases, having some velocity and therefore inertia, cannot conveniently do an about-face and march toward the exhaust port until the momentary pressure conditions in the combustion chamber and around each valve are favourable for them to do so.

At TC of the exhaust stroke, the piston has completed its second sequential stroke of four, and during the piston travel from BC to TC, the crankshaft has rotated a second ½ revolution for a total of exactly 360 degrees since the cycle began. Exhaust stroke completed. Cycle half-completed. Although the piston motion around top centre is relatively slow, the degree of activity in and around both intake and exhaust ports and valves, and in the combustion chamber, is fast and furious indeed.


Induction Stroke


As the piston reaches TC, changes direction and starts toward BC, the induction stroke begins. Prior to this, however, the air/fuel mixture within the induction system and intake port has been more-or-less "stacked" around the intake valve, awaiting the time when the intake valve has opened far enough to permit entry of part of the air/fuel mixture into the combustion chamber by weight of its own inertia. Therefore, the penetration of the escaping exhaust gases into the intake port usually doesn't get too far before they are damped out and their direction reversed by the advancing air/fuel charge. However, the exhaust gases do dilute the initial portion of the air/fuel charge.

Opening the intake valve before TC of the exhaust stroke so the valve is pretty well of its seat at TC reduces the pumping loss exerted on the piston in a manner similar to the "blow-down" period, although cylinder pressure conditions and direction of gas flow are both reversed.

The question of overlap breathing is always good for an argument and for good reason: It's never the same between engines of different types, and it can be minimal or substantial, depending upon application, engine speed, load, and a number of other influential factors. It isn't a question of "exchanging breaths" between exhaust gases and air/fuel mixture. It seems to be more a question of how much fuel can be forced out the exhaust port before the exhaust valve closes. There is no doubt at all that this does occur in all engines to some extent. A large amount of fuel forced out the exhaust port in a race engine seems to signify that the remaining cylinder/combustion chamber space not covered by the piston has been forcibly "scavenged" - swept clean of all traces of residual exhaust gases. I think not. Evidence points to air/fuel separation during the valve-overlap period - a divergence of directional paths - whereby the heavier fuel molecules tend to travel in straighter lines - whether or not the straighter lines lead them to the exhaust port - while the air molecules, being very much lighter in molecular weights, are much more easily deflected, turned, bent, whatever, from a straight-line path. The term overlap breathing is perhaps an unfortunate choice of words, because it does refer to a one-directional path from intake port to exhaust port.


One more factor: In spite of the high cylinder pressure, high cylinder temperature, and the intensity of the combustion flame front, some particles of air and fuel within the cylinder escape combustion during the power stroke. These lurk about in relatively isolated and inaccessible pockets of the combustion chamber, around the top piston ting and between the piston and the cylinder bore, etc. This occurs because the turbulence of the advancing flame front is dampened and cooled by the proximity of two or more surfaces to the point where ignition of these particles cannot take place. Being heavier than the exhaust gases and being in a less-active state, these particles are among the last guests to depart from the combustion chamber party and then usually after the piston has passed top centre of the exhaust stroke and has started toward bottom centre of the induction stroke, when they are exposed to the general activity within the cylinder and are more-or-less free to find their way out past the closing exhaust valve. Even then, some of these particles don't make it past the exhaust valve.

In any case, and every case, some residual exhaust gases and unburned air and fuel are trapped within the cylinder after the valve overlap period is one of extreme and intense activity, even though piston motion is relatively slow. Perhaps this will explain part of what happens during this period and why.

After the piston passes the exhaust stroke TC and heads toward the induction stroke BC, the first significant event is the exhaust valve closing. This is the third of four periods that the valve motion overlaps piston position. Meanwhile as the piston descend, the intake valve continues to open until at some point well before BC, it reaches maximum lift and starts to close.

The piston's descent during the induction stroke generates a pressure-reversal within the cylinder to a point less than ambient atmospheric pressure - a negative pressure, partial vacuum or pressure differential. The amount of pressure differential within the cylinder is more-or-less dependent upon how far the intake valve is open as the piston starts its descent toward BC: The higher the valve lift at this point, the less the pressure differential (closer to ambient atmospheric pressure) and vice versa. The pressure differential again represents a pumping loss exerted on the piston but there must be some pressure differential, otherwise the air/fuel mixture would have no incentive for moving into the cylinder to fill the void left by the descending piston. Under favourable conditions, the air/fuel mixture could actually start to enter the combustion chamber before the piston reaches the exhaust stroke's TC.

The force behind the motion of the air/fuel mixture is ambient atmospheric.


Compression Stroke


Now the piston changes direction and heads for TC again on the compression stroke. Meanwhile, the intake valve is closing but is not yet seated. Piston motion before and after BC is at its "laziest" during the cycle, which means the crankshaft swings through a fair-sized arc before the piston moves significantly toward TC. The cylinder-filling action of the air/fuel mixture is continuing because the inertia of the mixture charge outweighs the effect of the rising piston. At least for the moment.

You know what inertia is. This is what hurts when you whack you thumb with a hammer. Use a lighter hammer, or swing it at a lower speed, or both, and it won't hurt as much nor for as long. Swing the hammer at a faster speed, or use a heavier hammer, or both, and it will hurt more and for longer. More scientifically and in this context, inertia is a property of the air/fuel mixture which causes the mixture to resist any change in its motion. Your thumb represents resistance to the motion of the hammer. Similarly, the air/fuel mixture continues to fill the cylinder until the rising piston becomes enough of a resistance to slow the mixture, stop it, or reverse it.

The time to close the intake valve is before the flow of the air/fuel mixture into the cylinder is stopped, at some point well before the compression stroke TC. Closing the intake valve represents the fourth and last period during the cycle that valve motion overlaps piston position. From this point until the exhaust valve opens during the latter part of the power stroke. Both valves remain closed.

Whatever volume of air/fuel mixture in the cylinder as the intake valve closed is trapped and is the source of energy for the power stroke and to carry the piston and crankshaft through subsequent strokes, with enough left over to measure as usable power. Being a compressible gas, the air/fuel mixture is compressed by the rising piston and as a function of compression, the mixture is also heated, which further increases the cylinder pressure. At this stage, the mixture is a highly-agitated mass caused by the motion of the piston and its compressive action, during which time the air particles are heated further and attempting to expand in a volume that is progressively decreasing as the piston approaches top centre. Meanwhile, the fuel particles are being forced into a state of separation and vaporisation and into more intimate contact with the air particles by the same heat of compression. And there is motion. Oh, boy! Is there motion! As the piston approaches top centre, this violently turbulent mass is just ripe for…

No. I won't say explosion BANG! But it is time to light the fire. At a point arbitrarily chosen to be 30 crankshaft degrees before top centre of the compression stroke, several thousand volts of jolt (electrical energy) are delivered to the centre spark plug electrode, causing a spark to jump the gap between the centre and ground electrodes in the spark plug. The spark intensity and duration ignites the air/fuel mixture closest to the plug electrodes, setting of a chain reaction that will envelope the entire combustion chamber area, except those few relatively tiny isolated inaccessible pockets that resist flame penetration. The flame front moves away from the point of ignition and expands more-or-less uniformly while the unburned portion of the mixture is forced into even higher degrees of turbulence and compression by the advancing flame front, all of which causes a relatively abrupt but uniform increase in cylinder temperature and pressure.

All of this is going on as the piston reaches the compression stroke TC. The piston has completed the last stroke of the four-act-stroke cycle comedy. This is accompanied by an additional 1/2 revolution (180 degrees) of the crankshaft, which brings the total number of crankshaft degrees of rotation to 720 for the complete cycle. Now any dumbhead knows there are just 360 degrees in exactly one revolution of anything that turns, so it should be obvious by now that it takes exactly two revolutions (720 degrees) of the crankshaft to complete the four piston strokes that form one cycle of the four-stroke cycle engine. Compression stroke complete.

Cycle complete.

But is it? Earlier, it was stated that the cycles overlap each other, so we just can't walk off and leave it. Besides, hell hath no fury like a chamberful of ignited violent vapors and an immovable piston. From the arbitrary 30-degree before top centre ignition point until maximum cylinder pressure is reached at about 15 degrees after TC, the crankshaft has rotated about another 15 degrees after the theoretical end of cycle. But now we can see how the finishing phase of one cycle is the beginning phase of the following cycle, and also how on cycle overlaps another.


Chapter two

Changing Valve Events Around


Of course, there's a great deal more to it than this simplified version, some of which comes later. Now let's start at the commencing again and see how by changing the valve opening and closing points, the entire character, personality and performance level of an engine can be changed, for better or worse.


Exhaust Valve Opening


What happens if the exhaust valve is opened earlier than normal; that is, at a piston position further away from BC? Obviously, the "blow-down" period is extended, which is some times beneficial in an engine that consistently operates at higher engine speeds. Most of the useful work applied to the piston during the power stroke is used up by the time the piston reaches about 80 to 90 degrees before BC, so very early exhaust valve opening can rob the engine of some power…most noticeable at lower engine speeds. This practice also releases additional heat to the exhaust system, which makes control of oxides of nitrogen emissions very much more difficult. Early exhaust valve opening points are usually associated with longer valve open periods (longer duration's) in race engines where control of exhaust emissions is not a factor (yet). If kept within reasonable limits, the exhaust valve opening point has less effect upon general engine performance than the exhaust valve closing point and the intake valve opening and closing points.

Later exhaust valve opening with the piston closer to BC of the power stroke helps engine performance at lower engine speeds by capturing more energy applied to the piston. Of course, the "blow-down" period is reduced, which may (not always) reduce maximum power output at the top end of the engine speed range. Very late exhaust valve opening points, when combined correctly with other valve events, can help exhaust emissions by containing cylinder heat for a longer period, allowing more time for the heat to dissipate within the cylinder before releasing the exhaust gases. Particularly, this reduces oxides of nitrogen (NOx) emissions, but carbon monoxide (CO) and unburned hydrocarbons (HC) are reduced as well. Properly done, this will improve engine performance as well through most of the engine speed range, if not all of it.


Intake Valve Opening


Next in sequence, the intake valve opening point. Ah! The romance of high valve overlap periods! Early intake valve opening accomplishes half of that. When the intake valve opens early (with the piston further away from TC on the exhaust stroke than normal), the engine will usually immediately respond by being rough and balky at low engine speeds. This occurs because of the greater dilution effect the exhaust gases have on the air/fuel mixture charge as the mixture attempts to enter the cylinder. As engine speed increases, the velocity and inertia of the mixture charge overcomes most (not all) exhaust gas dilution and helps power output at higher engine speeds. Very early intake valve opening points will kill performance in the low and mid-range speeds, making engine power and response acceptable only in the highest engine speed ranges. Early intake opening points are associated with relatively long intake valve durations in race engine applications. Later opening of the intake valve (with the piston closer to TC on the exhaust stroke) smooths out engine operation during idle and off-idle conditions, and in the low and mid-range engine speeds. Engine vacuum, necessary for vacuum booster operations of power brake systems, etc., doesn't go down the drain with later intake valve opening points, assuming the other three valve opening and closing points are within reason. Some power may be lost at the top end of the engine speed range, but it may be worth this penalty to get better low and mid-range performance, better idle and off-idle characteristics, perhaps even an improvement in fuel economy if the throttle is treated with some respect. Very late intake valve opening points have shown to be beneficial in reducing exhaust emissions, which can be accomplished with a general improvement in performance level.


Exhaust Valve Closing


The exhaust valve closing point constitutes the other half of the long valve overlap romance. Late exhaust valve closing (piston further away from TC on the induction stroke) contributes its share to evil engine operation at lower engine speeds. This occurs because the air/fuel mixture charge is exposed to two paths: One, it can enter the cylinder, become trapped there and perform some useful work; or, it can enter the combustion chamber only long enough to march right out the exhaust pipe, thereby escaping unused, except for some slight cooling of the exhaust valve in passing. High valve overlap periods are on reason why race engines will tolerate high compression ratios and full, or nearly full spark advance at relatively low engine speeds. Under these conditions, maximum cylinder pressures are quite low so there isn't much chance for an abnormal combustion condition to develop. As engine speed increases, late exhaust valve closing does allow a higher percentage of exhaust gases to be evacuated from the cylinder because of the direction of exhaust gas flow and the inertia of the gases. However, there is still some "bleed-off" of the air/fuel mixture, which could conceivably limit maximum power output. Late exhaust valve closing by itself, or in combination with early intake valve opening, detracts from acceptable low and mid-range performance, driveability, fuel economy, etc. Very late exhaust valve closing could easily contribute to a loss of maximum power, as well as all the other disadvantages. Earlier exhaust valve closing brings with it smoothness and docility, although not necessarily dullness in performance level. It makes engine operation much more civilised at lower engine speeds and could help top end power as well, and contributes to overall engine flexibility, particularly if it is combined with a later intake valve opening point. Very early exhaust valve closing can reduce exhaust emissions and again, if all other factors fall in place, this can be accompanied with a higher performance level.


Intake Valve Closing


Now we get to the intake valve closing point. AHA! This is the one that can make it or break it. Intake valve closing point probably has more influence on engine operating characteristics than the other three valve opening and closing points combined. Late intake closing (with the piston further away from BC on the compression stroke), if optimised for a race engine, can beneficially capture a larger volume of air/fuel mixture within the cylinder before a pressure-reversal condition could cause reverse-pumping of some of the mixture past the closing valve, generated by the rising piston. Optimising the intake valve closing point in a highly sensitive engine can be very itchy indeed. Later intake closing points are acceptable and necessary at consistently high average engine speeds for the above reason. But at low engine speeds, this practice is purpose defeating because reverse-pumping action, if not killed outright, is at least seriously wounded, again by the direction and inertia of the tag end of the air/fuel charge. Very late intake valve closing makes the effects of reverse-pumping worse and through a broader engine speed range, and can make engine response so lazy it may never be able to pull itself to its most favourable engine speed. Late intake closing usually relates to longer intake valve durations for race engines. Believe me, late intake closing does nuthin' at all for exhaust emissions except to increase them. A fine balancing act goes on here, particularly if we try to avoid the trap of a "peaky" engine; that is, one that operates at its best at high engine speeds, and then only through a relatively narrow speed range. The bike boys call this "comin' up on the cam." But if the intake closing point is delayed too long, you will undoubtedly find some satisfying profanity to go with it and you'd probably get better results by pushing the whole soggy mess off a cliff.

Earlier intake valve closing is what brings and engine to life, gives it flexibility, range and muscle at just about any reasonable point in the speed range. It helps idle, off-idle and part-throttle steady-state cruise conditions possibly more than you could imagine. This occurs because no appreciable reverse-pumping action takes place, allowing a larger volume of air/fuel mixture to be trapped in the cylinder and put to work at lower engine speeds. Very early intake valve closing is beneficial in reducing exhaust emissions, improving fuel economy, etc., and is usually accompanied by a performance increase.

There is on dreaded word that rears its ugly head in the area of valve timing for any specific application: Compromise. It's a compromise between what you want and what you can get, or are willing to accept. It's a trade-off of something, someplace, for something else, someplace else. If there is a direction indicated, it is certainly toward conservatism. The sorriest, soggiest engine in the world is one that is "overcammed." Unfortunately, this is a trap that is all too easy to fall into, and the best words of advice are to avoid it in the first place.


Chapter Three


Where is the piston and what is it doing?


With the relationship of valve events and the piston positions of top and bottom centre pretty well established, let's take a look at the piston between these two positions to get a clearer picture of what going on. At any given engine speed, the angular rotation of the crankshaft, and therefore of the crankpin, is constant. However, the motion of the piston is not constant, simply because it stops twice and changes direction twice during each revolution; once at TC and again at BC, even though the crankshaft continues to rotate at a constant speed. An erroneous belief would have us swallow the idea that at points of crankshaft rotation exactly 90 degrees before and after TC the piston has completed exactly ½ its total stroke. No way. This could only happen if the stroke length were infinitely short and the centre-to-centre length of the connecting rod were infinitely long. But we are dealing with finite numbers here, so this idea goes in the trashcan. The piston actually travels more than ½ the total stroke from exact TC to a crank angle position of 90 degrees after TC. This being true (and it is, there being a number of complex equations to prove the point), piston speed is faster during the first 90 degrees of crank rotation than it is during the second 90 degrees of crank rotation because the piston can only move from exact TC to exact BC in exactly 180 degrees of crank rotation, no more, no less. From this, it can be seen that maximum piston velocity is reached at some point before 90 degrees of crank rotation from TC. This occurs when the axes of the crankshaft, crankpin and piston pin form an angle of exactly 90 degrees. This shows the relationship of the centre-to-centre length of the connecting rod and the radius of the crankpin from the crankshaft axis (exactly ½ the total piston stroke). The equation is: Rod length divided by crankpin radius is equal to the tangent of the angle at which maximum piston velocity is reached.

A simplified diagram is shown for the two basic engine configurations under discussion. Piston pin offset has been omitted for clarity. In the Datsun L-16/L-24 versions, maximum piston velocity occurs at 74 degrees, 31 minutes of crank rotation after TC. From TC until this point is reached, the piston is accelerating. From this point until BC is reached, which requires 105 degrees, 29 minutes of crankshaft rotation, the piston is decelerating. This holds true for both the power and induction strokes. From BC until maximum piston velocity is reached in 105 degrees, 29 minutes of crank rotation on both exhaust and compression strokes. The piston is accelerating again. From the point of maximum piston velocity, the crank rotates 74 degrees, 31 minutes for the piston to reach TC, during which time the piston is decelerating. Corresponding crank angles for the Datsun L-18 are 73 degrees, 21 minutes and 106 degrees, 39 minutes. Not much difference between the two, but enough to make a considerable difference in functional characteristics, as will be shown later. The alternate but unequal periods of piston acceleration/deceleration exert some powerful influences in optimum valve timing events for specific applications.


EXAMPLE 1: The exhaust valve reaches full lift at about the same time the piston reaches it maximum rate of acceleration on the exhaust stroke to minimise pumping losses.

EXAMPLE 2: The intake valve should open at a point about where the piston is at is maximum rate of deceleration during the exhaust stroke to prevent serious air/fuel charge dilution as the forefront of the mixture charge attempts to enter the cylinder.

EXAMPLE 3: The intake valve should be closed before the piston reaches its maximum rate of acceleration on the compression stroke to minimise pressure-reversal disturbances generated by the piston. (There are enough disturbances going on in the engine at sonic velocities, so we don't need another one, but more about that later.)


The point here is that the piston behaves in exactly the same manner as it moves from BC to TC during the exhaust stroke and the compression stroke, in relation to maximum acceleration, maximum rate of acceleration, maximum velocity, maximum deceleration, maximum rate of deceleration, etc. Only the strokes have been changed. Therefore, the three valve events that occur during the upstroke of the piston must be pretty carefully tailored, keeping in mind where the piston is and what it is doing, for optimum results. In order of importance, intake valve opening point second, with maximum lift point of the exhaust valve a poor third because of inherently more latitude.

On the down-stroke of the piston, from TC to BC of both the power and induction strokes, the piston again behaves in exactly the same manner related to acceleration/velocity/deceleration, but it is NOT - repeat NOT - the same as on the upstroke. Look at the diagram. Three valve events also occur during two piston down-strokes.

EXAMPLE 4: Exhaust valve opens during the period of piston deceleration on the power stroke, when most of the energy of the power stroke has been converted to useful work, but when the cylinder pressure is still considerably higher than atmospheric. This allows the remaining cylinder pressure to unload itself past the exhaust valve so that when the piston reaches BC, residual cylinder pressure is only slightly above atmospheric, which of course, reduces the pumping loss when the piston begins the exhaust stroke.

EXAMPLE 5: Exhaust valve closes during the period of piston acceleration on the induction stroke, before the piston reaches its maximum rate of acceleration to prevent the escape of excess air/fuel mixture past the exhaust valve and also to prevent air/fuel charge dilution by drawing the tag end of the exhaust gases back into the cylinder as the cylinder pressure condition changes from a pressure vessel to a vacuum vessel.

EXAMPLE 6: Intake valve reaches full lift during the period of piston deceleration but before it reaches its maximum rate of deceleration. Thus the intake valve "lags" behind the piston somewhat to generate as much of a pressure differential between the cylinder and the atmosphere, for as long a period of crank rotation as practicable to impart high rates of velocity and inertia to the entering air/fuel charge, so the velocity and inertia can be utilised as a cylinder-filling expedient well after the piston has started the compression stroke. Maximum lift point of the intake valve is therefore somewhat more critical than maximum lift point of the exhaust valve.

The six valve events within four strokes of the piston, or two revolutions of the crank, which ever occurs first (You dummy! They both occur at exactly the same time!), listed in order of importance and significance to engine operation are:

1. Intake valve closing point, by far;

2. Intake valve opening point (probable);

3. Exhaust valve closing point (probable; could be reversed with (2)) under certain conditions);

4. Exhaust valve opening point;

5. Maximum lift point of intake valve;

6. Maximum lift point of exhaust valve.

The mechanical, physical, chemical actions and reactions, forces and counterforces at work within a four-stroke cycle engine are known and recognised to be exceedingly complex and inextricably interrelated. Today some glimmer of knowledge and understanding can be seen emerging from some of the engine's innermost secrets which, even a few years ago, defied reasonable explanation. It is ironic, even amusing, to dwell on the thought that since its inception in 1862, 111 years ago, until the first functional four-stroke cycle engine appeared in 1878, 95 years ago (by different people, incidentally) until now, literally billions of these engines have been built and used satisfactorily, but some of its deeper mysteries may never be solved satisfactorily.

So get your feet wet and your hands dirty: You may be the one to find a secret or two lurking under that blob of cast iron. The four-stroke cycle engine may be a lot of things, but it is NOT dull. NEVER!


Chapter Four




The Datsun L-16, L-18 and L-24 camshaft and valve train system could be described as "contemporary single-over-head camshaft with rocker arms," but with some variations. The basic camshaft and valve train layout is not new. It's been used by such venerable makes as Mercedes-Benz, Porsche 911 series, the old OSCA 750's, various Ferraris, Maserati, BMW, Rover, etc.

Later-comers to the same scene include small Peugeots, some Mazda and Toyota models, Dodge Colt, Ford Pinto 2000 and others not as well known in this country.

Even a couple of domestic engines were involved: The short-lived Ford 427 SOHC V8 and Pontiac's in-line six-cylinder that ran from the 1966 through the 1970 model years. Oh yes, there was a Willys engine, too! With only three known exceptions, all of these engines suffered (or still suffer) from the same disease - usually in the earliest production runs. ALL had (or have) severe camshaft lobe and/or rocker arm wear... Three exceptions were Fords SOHC V8 and some Ferraris because these used rollers that bore on the cam lobes instead of the more usual radiused pad machined on, or inserted into the rocker arms. Ferraris were kind of shaky for a couple of reasons but the Ford was virtually indestructible. Datsun's L-24 is the third exception…not to imply that the L-24 has been totally free of camshaft and rocker arm failures. But with this engine, incidence of this type failure has been vastly reduced.

Basically, the sever cam-loge/rocker arm wear problem revolves around two significant and interrelated factors: (1) Incorrect metallurgy between the two rubbing surfaces of the cam lobe and the rocker arm; or, (2) Improper lubrication between the cam lobe and rocker arm interface; and (30 a combination of the first two. Datsun has had their share of the same problem, first with the discontinued U-20 four-cylinder two-litre cammer and with earlier L-16's. It isn't my intent to bad-mouth or be hypercritical of this or any other type of camshaft-valve train layout. Where inherent problems do exist, I believe they should be examined for cause, and hopefully, to show how such grief can be cured or at least avoided. So an overhead camshaft system by itself does not necessarily solve all valve train problems - no matter how attractive it may seem initially.




Now specifically to the current Datsun single-overhead-camshaft types.




Over head camshafts in L-16/18/24 are driven directly from the front of the crankshaft by a long single-stage double roller chain and sprocket drive assembly, with no intermediate sprockets. An automatic chain tensioner on the "slack" side of the chain (left side of engine viewed from front) is just above the crankshaft sprocket. Engine oil pressure and a compression spring combine to apply the necessary load to a piston, at the end of which is a curved "shoe" bearing directly on the chain. A curved guide extends from just above the shoe to just below the camshaft sprocket to control chain whip and vibration. A similar but straight guide does the same job on the "tension" side of the chain (right side of engine viewed from front). These two guides and the shoe are faced with wear-resistant plastic where they contact the chain. Twenty teeth on the crankshaft sprocket and 40 on the camshaft sprocket give the necessary two-to-one reduction between crank and cam. In this type cam drive system, the crank and cam rotate in the same direction; clockwise when viewed from the front.

The cam sprocket is secured to the camshaft nose with a 16mm bolt which also holds the separate steel fuel pump eccentric in place. A 6mm-dowel pin pressed into the cam nose is offset from the camshaft centre. It matches up with any one of the three holes in the camshaft sprocket. The pin locates the sprocket relative to the cam; the three holes and three timing marks allow adjusting valve timing, which will be explained later. A cam thrust plate bolts to the front side of the forward cam tower. Fore-and-aft movement of the camshaft is controlled by the thickness of this plate and the depth of the counterbore in the back of the camshaft sprocket. The thrust plate is available in three different thicknesses to allow minimising axial camshaft movement.


Cam Towers


The camshaft is supported in the aluminium alloy cylinder head by aluminium ally cam towers - four in the L-16/L-18, five in the L-24. There are no separate camshaft bearings as such; the camshaft journals ride directly in the cam tower bores.

Cam towers are bolted to the cylinder head located with large diameter hollow dowels, then align-bored after assembly at the factory. It is an ABSOLUTE NO-NO to remove the cam towers from the cylinder head because it is nearly impossible to restore correct cam bearing bore alignment after they have been removed.




The camshaft is a one-piece iron casting with induction-hardened cam lobe and bearing journal surfaces. There's no fancy metallurgy here: Analysis shows the material to be a close relative of plain grey cast iron. However, the casting technique is excellent and the castings show uniform density and are generally very good. Cam lobes are coated with a non-metallic phosphate compound to minimise scuffing with their mating rocker arm pads. Camshaft journals are left clean and bright because the phosphate coating material is not at all compatible with the aluminium tower bearings.


Rocker Arms


Rocker arm pivots thread into the steel bushings and are secured by locknuts bearing against the bushing tops. The pivot tops have spherical segments, the geometrical centres of which form the rocker pivot points. Just below the spherical segment and integral with the rocker pivot is a hex section that permits using an open-end wrench to raise or lower the pivot in the bushing, thereby adjusting valve lash. The rocker pivot locknut fits beneath the hex section in the rocker pivot and locks against the top of the threaded bushing.

Rocker arms are steel forgings spanning the distance between the rocker arm pivots and indirectly to the valve stem tips. Here, there are no questions or doubts about the use of hydraulic, mechanical or roll-type camshafts; all are mechanical with a positive and predetermined amount of lash or clearance between cam lobe and rocker arm pads. A hemispherical socket machined into the under side of the rocker arm corresponds to the spherical segment of the rocker pivot. Each rocker arm fits over its pivot and is thus located by the pivot and oscillates about the pivot's geometrical centre. The rocker arm tip, which normally bears against the valve stem tip but doesn't in this case, is also on the underside of the rocker and has a single-plane radius machined into the rocker arm tip pad. An oil hole drilled in the top of the rocker socket permits splash lubrication of the socket and pivot, adequate because of the small linear distance the rocker arm socket travels in relation to the pivot.

Topside on the rocker arm between pivot socket and rocker arm tip, is a pad with a single-plane radius in the same plane as the valve stem tip radius, that is, coaxial with the camshaft. This pad bears against the cam lobe. At this point there is some controversy about early and late rocker arm types, though I can see no valid reason for argument at all. Earlier rocker arms were made of one piece and the rocker pad that contacted the cam lobe had a slightly larger radius, which gave a slightly flatter arc to this pad. Some claim that these are the rocker arms to use in modified L-series engines because the larger pad radius results in a slightly higher effective rocker arm ratio and also slightly higher valve velocities and valve lifts.

As far as it goes, this is correct. However, what these proponents fail to recognise or realise is that the forged steel rocker pad contacting a relatively low-alloy iron cam lobe is metallurgically one of the worst possible combinations with commonly-used materials. This is true even with very light valve spring loadings and lubrication that, by sheer volume, makes up for what it lacks in direction and placement. Simply reversing the materials would have helped some but would not put a complete fix on it. Datsun learned about the same lesson from its failures that nearly all previous designers of this general type of camshaft/valve train layout had learned: Ya gotta have metallurgically compatible materials at the interface of the cam lobe and rocker arm rubbing pad. In a conventional pushrod engine the cam lobes have a slight taper, are offset from the lifter bore centres, and the lifter faces are slightly crowned. All of these combine to force the lifters to rotate as they rise and fall in the bores in accordance with cam motion. This action produces a very smooth and even ear pattern on both lighter faces and cam lobes. In this day and age this is completely predictable, if there is metallurgical compatibility in the first place.

In the Datsun and similar overhead camshaft layouts neither the cam lobe nor the rocker arm rubbing pad are protected in this way from each other, so the material for each surface must not only get along with the other - each must also be scuff and abrasion-resistant, even if loads between the two are high enough to squeeze out or vaporise the interface oil film. This problem was very effective in finally killing production of the Pontiac OHC-6.

In any case, late L-16 engines and all L-18 and L-24 engines are equipped with two-piece rocker arms. The second piece is an insert furnace-brazed in place, and this insert contacts the cam lobe. When the two-piece rockers were introduced, the single-plane radius of the insert was, decreased slightly, slightly decreasing the effective rocker arm ratio and the valve velocity. This resulted in lower dynamic stresses at the cam lobe/rocker pad interface.

Analysis shows the current rocker arm insert pad material to be the approximate American equivalent of chilled cast iron. Radius of the current species of rocker arm pad is 50mm (approximately 2.00 inch).


Combustion Chamber/Valve Layout


Combustion chamber configuration of the L-16/L-18 is more =-or-less "conventional" wedge; the L-24 chamber is more like an "open" wedge. Valve are in-line in all engines and are tilted the relatively small angle of 12 degrees from the cylinder bore centralise. Valve stems are tilted to the right when viewed from the front. All intake and exhaust ports are on the same right side. On the opposite side, the aluminium head casting is drilled and tapped to accept hexagonal head steel bushings, which are threaded internally and externally. Above the hex section, the bushings are grooved to accept "butterfly" type spring clips. Each of these bushings, one for each valve, is screwed tightly into the head casting.

Valve guide inserts are a dense cast iron alloy similar to Meehanite. These are an interference fit in the head with the head heated and the guides frozen. Nominal guide bore diameter is 8mm and nominal valve stem diameter is close to 5/16 inch, a point of occasional convenience, which will be discussed later. Valve stems are hard-chrome plated for wear resistance.

Aluminium-bronze valve seat inserts are used for all intake valves. Exhaust seats are cast iron alloy inserts. Valve seat inserts are fitted to the cylinder head in the same manner as the guides.


Chapter Five


Valve Springs


All late L-series engines have inner and outer valve springs without flat damper coils. Valve open/closed spring loads vary with the type of engine and the valve lift.


Spring Retainers & Lash Pads


Valve spring retainers are upset steel stamping's with separate flanges for inner and outer valve springs. Tapered holes near the bottom accept single-groove split valve locks, mating with single grooves in the valve stems to secure valves to retainers in the normal manner. Nothing remarkable here. Spring retainer tops are counterbored to accept what is variously called a valve lash pad or a rocker arm guide. Whatever, it performs both functions. This pad is cylindrical with a slot machined through the centre of one side. This leaves two ears - one on each side of the slot extending upwards, away from the valve stem tip, and above the top of the springs retainer. The pad slip fits the counterbore in the retainer and the pad's flat bottom surface contacts the valve stem tip; they merely locate the rocker on the pivot. In addition, some of the side thrust transmitted from the rocker arm tip to the valve stem tip is eliminated and absorbed by the slip fit of the lash pad in the retainer, resulting in reduced wear of valve guide bores and valve stems.

In all of the Datsun cammer engines, valve lash is measured between the centre of the heel of the cam lobes and the rocker arm pads. Thus the amount of valve lash between rocker arm tip and the lash pad becomes a function of rocker arm ratio.

A "mousetrap" spring, with coils that straddle the rocker arm and which fits in a small groove behind the rocker pivot socket in the rocker, has free ends that hook into the butterfly clip in the rocker pivot bushing mentioned earlier. This spring exerts a fairly light load on the rocker arm to keep the rocker arm pad in (presumably) constant contact with the cam lobe as further insurance for correct rocker arm alignment.

Datsun's design involves a few more pieces than in more conventional single overhead camshaft layouts - but the number of pieces in any mechanism does not signify efficiency or inefficiency. Functionalism is the key here. In this respect the Datsun design is both original and clever. And it works!




Pressurised lubrication is delivered to the camshaft bearings by oil holes in each cam tower that line up with oil holes in the cylinder head, which in turn intersect with the main oil gallery in the cylinder head. In L-16/L-18 engines, the camshaft itself forms two additional oil galleries. These are drilled axially and on centre from each end for something less than half the total camshaft length, leaving a wall between the two galleries near the centre. The second and third camshaft bearing journals are grooved and drilled with an oil entry hole in each groove to admit oil into each of the two oil galleries in the cam. Each lobe is drilled to release oil under pressure to lubricate the cam lobe/rocker pad interfaces. Cam lobes for No. 1 cylinder exhaust, No. 2 intake, No. 3 intake, and No. 4 exhaust have oil holes drilled in the centre of the lobe heel. Exhausts 2 and 3 have the oil holes on the lobes opening flanks. For some mysterious and inexplicable (and unexplained reason, intakes 1 and 4 have the oil holes located on the lobes closing flanks, which makes no rhyme, reason or sense at all. Historically, and as one might expect, No. 1 and 4 intake lobes are the ones most frequently damaged or distressed, along with their mating rocker pads - even in completely stock and unabused engines. It hardly seems likely that these two misplaced oil holes represent an oversight. However, it eliminates the need for a fourth indexing station, plus related tooling and machinery for drilling the am lobe oil holes at the factory. In any case, one of two possible fixes can make it right. Pressurised oil is contained within the camshaft by a press-in plug at the back of the camshaft and by the cam sprocket retaining bolt at the front. L-24 cam bearings are lubricated in the same way as in the L-16/L-18 engines, but the cam lobe/rocker pad interfaces are lubricated in a different and much more satisfactory manner. An external steel tubing oil gallery bolts to Nos. 1,3 and 5 cam towers on the rocker. Pivot side. Oil to this gallery is supplied under pressure by drillings in the No. 3 cam tower. Short transverse pipes aim oil streams directly at each of the cam lobe/rocker pad interfaces.

This system works admirably; so well in fact, that it has been more-or-less duplicated on a good number of modified L-16/L18 engines with equal success. When this has been correctly done on the four-cylinder engines, there is no longer any need for interface lubrication from the cam lobe oil holes. Oil sources to the hollow camshaft should be blocked off because they create a relatively large and unnecessary oil leak, which robs oil from the main and connecting rod bearings.

Oil entry holes in the grooves of numbers 2 and 3 camshaft journals must be plugged by drilling and tapping the holes for small socket-head setscrews. Use a bottoming tap to about 3/8-inch depth so the setscrews will bottom out tightly in the oil holes. The heads of the setscrews must not extend beyond the bearing journal surfaces. Loctite the setscrews to secure them. Also, remove the bolt at the front and the plug at the back of the camshaft so all chips can be scrubbed out. Otherwise the chips will inevitably find their way to the cam lobe oil holes with catastrophic results when they become trapped between the cam lobes and the rocker pads. Upon completion, there is no need to replace the rear camshaft plug (which you had to drill out to remove).

Enter SCCA. This August group occasionally comes up with some seemingly odd and misguided rules and regulations - and that's being charitable. According to their 1973 rules, SCCA hath decreed that the L-24 type of external oil gallery is not legal in L-16/L-18 engines. However, some clever lad who shall remain nameless circumvented this ridiculous rule by soldering copper "cooling fins" to his modified L-24 bolt-on external oil gallery. SCCA makes no distinction about the number or location of oil coolers for any on vehicle, so the cam lobe/rocker pad interface lubrication system magically becomes and "oil cooler" - entirely legal and acceptable by SCCA.

If one wishes to retain the stock L-16/L-18 cam lobe oiling system, it is a simple matter to drill intake lobes 1 and 4 with an additional oil hole each, this time locating the new oil holes on the opening flanks of the cam lobes. The cam lobes aren't all that hard, so tungsten-carbide drills are not required. A good, sharp high-speed steel twist drill will do the job. Use a number 47 drill (0.078-inch diameter) or a 2mm drill. Again, thoroughly clean both camshaft galleries before reassembly. This time, replacement of the rear camshaft plug is essential.

With either oiling system. It isn't difficult to imagine the volume of oil being sloshed about under the camshaft cover. Looks like Signal Hill all over again. This condition demands an effective oil drainback system from the cylinder head to the crankcase. In the front, a large hole empties into the crankcase through the timing chain cover. At the back, a 9/16-inch diameter hole empties into a matching hole through the block. This seems small but the top of the head is nicely channelled to direct drainback oil in either direction, effectively preventing an increasing oil level in the head…the wrong place to be of any value.


Valve Stem Seals


All L-16/18/24 engines have valve stem oil seals which they need with all the oil splashing about and which should be retained in any stock or modified engine.



Chapter Six




With the camshaft lobe between the valve stem tip and rocker arm pivot point, as it is in the Datsun Design, it will be seen that obtaining a valve lift curve basically symmetrical on either side of the maximum valve lift point requires using an asymmetrical cam lobe profile. And so it is with the L-16/18/24 engines. A quick look at a stock camshaft lobe for any of these engines shows the lobe opening flank is almost a straight line. It isn't really, it is slightly convex. The closing flank has much more of a convex bulge to it. This is because the initial contact point of the opening flank with the rocker pad is as close as it gets to the valve stem tip and as far away from the rocker pivot, therefor the effective rocker arm ratio at this point is at its lowest. As the camshaft continues to rotate, the cam lobe/rocker pad contact point moves progressively further away from the valve stem tip and closer to the rocker pivot, increasing the effective rocker arm ratio until maximum valve lift is reached. The cam lobe/rocker pad contact point is as far away from the valve stem tip as it is going to get at maximum valve lift, so the effective rocker arm ratio is at its highest. Beyond maximum valve lift on the closing flank of the cam lobe, the condition reverses itself because the contact point is moving progressively closer to the valve stem tip, calling for a somewhat "fatter" closing flank to accommodate the progressively-decreasing effective rocker arm ratio. Thus the asymmetrical cam lobe profile produces a valve lift curve essentially the same on both sides of the maximum valve lift point. From the darkest and earliest days of the four-stroke cycle engine, right up until today - and maybe tomorrow - a symmetrical valve lift curve has seemed desirable. So what happens to the valve lift curve when the cam lobe is basically symmetrical? Right, the valve lift curve is asymmetrical. With a symmetrical cam lobe profile in the Datsun cammer engines, valves open more slowly than they close due to the system's physical layout. So who says the valve lift curve must be symmetrical? Not me! In fact, there are a couple of valid points to the contrary. First, the engines produce better torque at lower engine speeds, which contributes to the second; a broader engine operating speed range. This means the engines aren't as "peaky" because the torque curve and the power curve are both flatter. This is beneficial in nearly any application because the Datsun's, like any other relatively small piston displacement engines, need all the torque they can get, particularly at lower engine speeds Datsun engines under discussion respond in an extremely encouraging manner to valve timing changes from ridiculously (perhaps deceptively) mild where exhaust emissions are a factor to some real giants in purely race engines. No one in their right mind would initially think that and engine with from 398.75 cubic centimetres per cylinder (24.32 cu. In.) to 500cc's per cylinder (30.5 cu. In.) - as represented by bored and stroked L-18's - could possibly use camshafts with effective duration well over 300 degrees with valve lifts from 0.600 to about 0.640-inch. Properly modified Datsun cammer engines can and do, and they love it. In fact, the intake port of the so-called FIA-cylinder head (not available on any vehicle imported into the US) continues to increase in airflow until nearly 0.750-inch valve lift. But enough of such exotica for the moment.




The biggest problem with a camshaft change in any Datsun cammer engine lies in restoring the rocker arm geometry as closely as possible to the stock condition. Because special semi-finished camshaft blanks are next to impossible to obtain for these engines, stock camshafts must be reground and processed. Any material removed from the base circle radius of the cam lobes (1/2 the base circle diameter) must be replaced at the valve stems as a function of rocker arm ratio. The stock cam base circle diameter is 1.300 inches 0 a base circle radius of 0.650-inch. Now assume that the base circle diameter of the new cam is 1.200 inches (0.600-inch base circle radius), 0.050 inches smaller on the radius than stock. The book gives the stock rocker arm ration as 1.5 to 1, but I have found that 1.48 to 1 is a more reliable number, although we may be quibbling. Multiply the 0.050-inch difference in base circle radius by the rocker arm ratio --say 1.5 to 1, and you come up with a number 0.075-inch. This is the amount the calve lash pad must be increased in thickness over and above the stock lash pad thickness for the rocker arm geometry to come out someplace close to stock, assuming nothing else has been changed. It isn't all that difficult to make the Datsun cammer rocker geometry come out correctly, but it does take some time, some basic knowledge of rocker geometry theory and practice, plus some very close observation of the related pieces.

Theoretically the rocker arm tip radius should provide a purely rolling action across the valve lash pad face - with zero sliding action. This rolling action is desirable to minimise any frictional loss between the rocker arm tip and the valve lash pad, and also to minimise any side thrust that can be transmitted from the rocker arm tip to the lash pad to the spring retainer and finally to the valve stem. This all minimises any frictional loss between the valve stem and the valve guide bore, which can easily contribute to rapid valve guide bore wear.

In practice, some sliding action is inevitable but its effects can be minimised by following the rules and regulations of the game. With the Datsun cammer engines, everything you touch related to the valve seats in the cylinder head, valve faces, valve stem lengths and the camshaft affects rocker geometry - for better or worse!

Rocker arm geometry can be considered correct when the centre of the rocker arm tip radius contacts the valve lash pad on the valve stem centreline axis at exactly one-half total valve lift. Some say coincidence of the centre of the rocker arm tip radius and the valve stem centreline should occur at higher or lower percentages of valve lift. We have found that the one-half valve lift figure makes life easier on the entire valve train.

One highly important item may contribute some conflict of interest at this point: The actual area, and location of the area, of the rocker arm pad contacted by the cam lobe, hereafter referred to as "rocker pad contact patch." Some cam lobe profiles do not allow much if any latitude here. With these engines it is AB-SO-LUTE-LY ESSENTIAL that the cam lobe does NOT extend beyond either end of the rocker pad. For the most part, this occurs on strictly race engines, but if the cam lobe profile is sufficiently desirable, the rocker arm position must be adjusted at the pivot end so that the cam lobe does not override either end of the rocker pad. If, by mischance, miscalculation, or simply the wrong cam lobe profile, an override condition exists at either or both ends of the rocker pad, it spells instant death to the cam lobes and the rocker pads.

If the cam lobe is a tight squeeze on the rocker pad, then the first order of business is to make certain that the rocker pad contact patch is as correct as it can be, and rocker arm geometry must be forced to a secondary position.

Ideally, the rocker pad contact patch should be centred on the rocker pad. Assume that the opening flank of the cam lobe comes within 1/32-inch of one end of the pad and that the closing flank of the cam lobe comes within 3/32-inch of the other end of the pad. This is wrong. The rocker arm position must be raised at the pivot end of the rocker so that the contact patch is equalised at 1/16-inch from each end of the rocker pad. There is a good visual method of checking and measuring the rocker arm contact patch, and there is also a way to "cheat" a bit to gain a higher effective rocker arm ratio in some cases…but that comes later.




But back to more basic stuff. How are valve opening and closing points and valve lift measured? A great deal of confusion exists here and the Datsun service manuals are not too helpful. They do show valve opening and closing points are measured. By correlating their data with a chart of figures used to construct a valve lift curve, we concluded that they measure valve opening and closing points at one-half millimetre (0.020-inch for all practical purposes) valve lift with zero valve lash and with the rocker pad contact patch centred. Any other measurement method not only doesn't make sense but cannot be correlated with Datsun data. So we have adopted the Datsun measurement method intact, but with one exception: We measure valve opening and closing points at a more Americanised 0.025-inch lift figure. Sure, it's arbitrary - just about any method is - but it gives four separate and distinct points of reference: Intake valve opening and closing points, and exhaust valve opening and closing points.

So why not measure these points at operating valve lash? Simple. There is too much room for error because a lot of crankshaft rotation results in only a very small amount of valve lift and the probability of error is too great at these particular points on the cam lobes. So we picked a point where the valve motion is not only quicker and can be measured much more accurately with repeatability - but on that also has much more significance to a functional engine. In other words, effective valve opening and closing points. Translated, this means that the valves are far enough off their seats so that gas flow into and out of the cylinders will have some influential and significant effects upon the engine operating characteristics.

I can hear it now, loud and clear. What about the split overlap and the so-called centreline methods of camshaft installation? They are both junk and should be buried and forgotten, but they can't be condemned without reasons and explanations. Both methods may tell you what you're looking for or want to hear, but they can't tell the engine what it must know to function properly. These are discussed in the sidebar titled Definitions.




Before you dash to your local Datsun high-performance parts outlet, hesitate long enough to do some soul-searching related to the following questions.

1. What do I want from my Datsun that it doesn't have now?

2. Can I afford to trade off some fuel economy in return for better performance?

3. Must the idle and low speed characteristics be civilised or doesn't it matter?

4. Are exhaust emissions likely to be a problem?

5. Do I want a stump-puller at low and mid range engine speeds, or do I need better power further up the RPM Range?

6. Am I willing and able to do the installation myself, and correctly, or is there a reliable local source for such installations?

7. Am I willing to notch the pistons, if required, to obtain the right piston-to-valve clearances for my application?

8. Can I get all the right pieces the first time from one source?


Chapter Seven



The split overlap method of camshaft installation is defined as having both intake and exhaust valves of one cylinder off their respective seats exactly the same amount with the piston at exact top centre of the exhaust stroke; that is, during the valve overlap period when the intake valve is opening and the exhaust valve is closing. This method is valid ONLY if the following conditions are met: (1) The valve lift curve is exactly symmetrical and exactly the same for both intake and exhaust valves. (2) If the engine likes a given cam lobe profile in a statically determined split overlap position. It was earlier stated that some good things happen to the Datsun cammers when a basically symmetrical cam lobe profile is used to make the valves open more slowly than they close. If the split overlap method is used in such a case, the camshaft position in relation to piston position will be retarded - a position that nearly all-small engines and some big ones as well dislike intensely. So what happens if circumstances such as a lousy cylinder head job, or a lousy exhaust system. Or both, demand that the exhaust valve duration be longer than the intake valve duration? Same thing: The camshaft is retarded. Unless the two above conditions are met, which is very highly improbable, the split overlap method of camshaft installation can only lead to false indications and therefore false conclusions.

The so-called centreline method reflects its inherent stupidity in the worst possible choice of words. The centreline of what? The term actually refers to maximum valve lift points in relation to the TC; piston position. Maximum valve lift of the exhaust valve occurs at some point before top centre; maximum valve lift of the intake valve occurs at some point after top centre. As a camshaft installation method it leaves a lot to be desired because it tells you nothing about the camshaft or its characteristics. However, if it is used in conjunction with the opening and closing points of both valves of one cylinder, then there are six points of reference instead of four: Intake valve opening point, maximum lift point, intake valve closing point. And the same for the exhaust valve. But if the so-called centreline method of camshaft installation is used by itself, forget it. Effective valve opening and closing points are left to fall where they may, and his is not at all in the best interests of engine performance. In addition and to prove something to myself, I have intentionally moved the point of maximum valve lift by as much as five crank degrees in each direction, without changing effective valve opening and closing points. Results: The engine couldn't car less where maximum valve lift occurs, at least within plus-or-minus five crankshaft degrees. But if the camshaft is advanced or retarded a similar amount, there is an immediate and very measurable change in operating characteristics. So the engine does recognise effective valve opening and closing points, but the maximum valve lift points are relatively meaningless. Inasmuch as there is some physical effort involved in a camshaft change, do it right, do it once, and leave the guesswork to amateurs. It does take additional time to do it right, but it has to be considered as time well spent.




Tools required for a camshaft installation - aside from the normal hand tools and the valve train hardware - are a fully-degreed crankshaft damper or a degree wheel - the larger in diameter, the better - to bolt to the crankshaft nose, or a fully-degreed flywheel that is easily accessible and visible; a rigid pointer located for easy reading of the degree wheel, damper or flywheel; a good one-inch stroke dial indicator with a large dial graduated in thousandths of an inch; a rigid magnetic base with rigid attachments for holding the dial indicator (flexible "snake" type attachments are not reliable for repeatability); a flat steel plate large enough to accept the magnetic base and one or two holes in on end so the cylinder head hold-down bolts can locate the plate on the cylinder head. Surface grind the plate on one side to provide a stable flat surface for the magnetic base. Two mechanical fuel pump springs are required to replace the valve springs for one cylinder, and an assortment of valve lash pads of different thicknesses. Time, patience, good temperament and cheer are also likely to be quite essential before it's all over.




Now for camshaft selection. Clear your mind of all romance, hogwash, myths, old wives' tales, phase of the moon, etc. You want a camshaft that works for your application, regardless of the degrees duration and/or overlap, valve lift, or whatever. It was stated earlier that small displacement engines need all the torque they can get, particularly at lower engine speeds, and if this is a factor, long duration, very high lift camshafts are O-U-T. They're great for strictly race engines. But for a street-driven vehicle they'd need a road map to fall out of a tree…and probably a push to get them started.




Let's begin with a street application basically a stock engine, where idle characteristics, throttle response, general drivability and exhaust emission levels are all contributing factors. Based on measuring effective valve open duration at 0.025-inch valve lift with zero lash, duration should be in the low to mid-240 degree range with no more than about 25 Degree overlap. Datsun cammer engines respond very nicely to valve lift; it actually helps low and mid-range torque, as well as maximum power. But for this application, lift should be in the 0.430 to 0.450-inch range. A couple of years ago we proved two points:

(1) performance level could be improved and (2) emission levels reduced with nothing more than a mild camshaft.

Average road performance level in the 2,800 - 6500 RPM range was improved by 7+% with a maximum of 10+% at the higher engine speeds. Average exhaust emissions of unburned hydrocarbons, carbon monoxide and oxides of nitrogen were reduced by approximately similar percentages. The vehicle was an otherwise bone-stock 1971 Z-car with about 25,000 miles on the clock. A few carburetion modifications were indicated to help emission reductions even more but the primary objective was to learn what the camshaft-only change would accomplish. An incidental advantage was that fuel economy was increased by about 4%, all of which showed that the thermal efficiency was better than stock. The camshaft had 250-degree effective duration with 0.440-inch lift. With the lower compression ratios of the later Datsun engines, camshafts for strictly street applications must be very mild indeed. The drop in compression ratio means a loss in cylinder pressure, which is contrary to improved performance. The gams plan here is to capture as much cylinder pressure as possible, yet retain normal combustion with pump-type fuels having very small amounts of tetraethyl lead, or none at all. This suggests very short duration camshafts in the mid-230 degree range with from 14 - 18 degree overlap. In 1972 and later engines, a camshaft like this by itself will usually wake up an engine to match the performance level of earlier stock engines with higher compression ratios. In most cases, such a camshaft can use all stock Datsun pieces, with the exception of valve lash pads, so long as maximum engine speed is kept within the 6,000 - 6,400 RPM range.

The next step up for the fours is for someone who is willing to bolt an extra stock two-throat progressive Hitachi carburettor onto a good aftermarket intake manifold, and still keep within existing exhaust emission limits for the year of vehicle concerned. Unfortunately, no such manifolds existed as this book went to press. In this case, a higher performance level can be expected, as well as higher average engine speeds. A camshaft for such and application should have an effective duration in the high-240 degree to low 250-degree range and from 34 degree - 38-degree overlap. For the privileges of higher power output in conjunction with considerably improved performance, and the realisation that this guy will usually have his foot buried a bit deeper in the carburettors when the occasion permits, a higher price must be paid, but not necessarily all in dollars or yen. A rougher and perhaps a faster idle and not much muscle below about 2,8000 - 3,000 RPM can be expected. The rougher idle brings with it less manifold vacuum at idle and the lower engine speeds, which can adversely affect the power braking system. So if you lean on the throttle harder, you can lean on the brake pedal harder. A camshaft assembly of this type will usually include special valve springs, spring retainers and lash pads that will permit a 7,000-plus RPM maximum safe engine speed.

It should be pointed out that the standard Datsun four-speed gearbox is not equipped with ideal (whatever that is) intermediate ratios. The first-to-second spread is OK, as is the third-to-fourth. But the large second-to-third gap is a factor that must influence camshaft selection because after the two-three shift the engine must have enough torque to pull itself out of the hole caused by the two-three gearbox ratio spread.

This is best translated into five simple words: DO NOT OVERCAM YOUR ENGINE! If questions arise about the suitability of two or more camshaft profiles, ask these questions of someone qualified to give the beast answers related to your particular application. If there is still some indecision, pick the milder camshaft, accept its limitations, and be glad you made the right choice.


Chapter Eight




So far, no mention has been made of other engine modifications, and for a purpose. For vehicles that serve as basic point-to-point transportation on freeways and surface streets in areas of high vehicle population, as well as more rural parts of the countryside, efficient and enjoyable vehicle operation pivots about on work: Drivability. Along with providing such basic transport at a relatively nominal price, Datsun's are "fun" cars to drive. A mild camshaft, perhaps with a more efficient induction system, actually improves drivability. Modifications such as larger valves, large intake and exhaust ports, tricky competition-type exhaust header systems, etc., all have one point in common: Singly or in combinations, they seriously inhibit drivability by taking too much torque away from the engine in the most frequently-used engine speed ranges, and simply do not belong in engines caught at stop lights, bumper-to-bumper traffic or short-hopping. Again - either singly or in combination - these modifications will cause poor idle, poor throttle response, poor part-throttle operation, so who needs 'em in a street driven vehicle? You don't if your vehicle fits this category. There is one other exception: Compression ratio. High compression ratios equate to higher thermal efficiency, higher torque and power outputs. However, with the quality of available pump gasoline steadily diminishing, the advisability of raising the compression ratio is questionable, except possibly at higher altitudes, or where one is still blessed with the availability of decent fuels. A higher ratio does help, but there is a very fine line indeed between balancing the highest useful compression ratio with valve timing and average fuel quality. This says nothing for the seemingly insignificant details that must combine to make a higher compression ratio function properly such as reasonably constant (and correct) air/fuel mixture ratio (no lean spots under full load), accurate and consistent total spark advance, correct spark plug heat range, adequate duel delivery system, proper spark advance curve, acceptable engine oil and coolant temperatures, and absolutely zero detonation and/or "dieseling" (when you have to beat the engine with a club to make it quit running after the switch is turned off). These are the trivia that will make relatively high compression ratio engine function as it should, or can break it into more tiny bits than you'd car to count for a strictly street engine, the small increase in compression ratio that could be tolerated probably isn't worth the effort.


Weekend Warriors


Enter the "Weekend Warrior." He has one vehicle and a strong desire to compete in motorsports events. These could be slaloms, gymkhanas, rallies, road races, drag races, hill climbs, etc. He knows a bone-stock engine hasn't a prayer of being competitive in any type of off-road event where modifications are permitted. He also knows that he must sacrifice some drivability and loss of operating economy on the street in order to be reasonably competitive in the off-road event of his choice. And he is willing to accept these sacrifices --up to a point. But where is that point? A good question and a difficult one, and one compounded by increasingly stringent exhaust emission limits. He may be called upon at any time for a roadside emissions check by state or city authorities, a practice that is not only entirely legal but one which is increasing in frequency all over the US. It's a very safe bet that if his engine is modified to the point of being marginally streetable, any exhaust "sniffer" will turn thumbs-down on the exhaust emission levels. Then he had best be prepared to convert the engine to stock condition and submit his vehicle for a recheck, or face the consequences and penalties of the law.

The best advice is to keep all emission control devices hooked up and operative when the vehicle is driven on any public road. These devices do help reduce exhaust emissions and if nothing else, they give tangible evidence to any vehicle inspection officer that the guy's implied intent was conformity with the law, even if the exhaust emission levels are unacceptable. Guy: "I've been thrashin' it pretty hard lately. Guess it needs a sharp tun-up." Inspection Official: "Yeah. Sign your citation here, and do visit our humble inspection station again. Within the mandatory 30-day period, of course. And with the emission levels right." Sound ridiculous? Perhaps. But it's happening every day!

The problem is further complicated by trying to explain to some aspiring John Morton or Bob Sharp that he has to use his Datsun as a transportation hack for six days o week to compete in an off-road event on the seventh day. This very strongly suggests conservatism for any and all internal/external engine modifications. In the area of camshafts for such applications, effective duration should be in the low-to-mid-260 degree range with 44 degree - 48-degree overlap. A camshaft assembly must include special valve springs, spring retainers and lash pads of the correct thickness. Usually, this type of speed of around 7,500 RPM; but it may require notching the pistons for the required piston-to-valve clearance. Before we depart for more exotic worlds, a few more words about exhaust emissions. It has been conclusively proved that a relatively mild camshaft can do two things in Datsun engines: It can improve road performance and reduce exhaust emissions at the same time. Similarly, a good aftermarket intake manifold could do good things in both areas by a vast improvement in cylinder-to-cylinder air/fuel mixture distribution and by maintaining relatively high mixture velocities throughout the entire induction system. A mildly but expertly modified cylinder head has proved beneficial in both areas but it does more for performance than it does for exhaust emissions. A good set of street-type steel-tube exhaust headers, with provision for air pump injection nozzles, will also help but the engine noise level is increased over the stock exhaust manifold. With these mods, a slight increase in compression ratio becomes feasible. Don't get carried away: A good, honest, genuine measured 9 to 9.5 to 1 should be considered adequate because the extra heat generated by higher compression ratios increases oxides of nitrogen (NOx) emissions. Then come the little things that count such as correct carburettor calibration, correct and stabilised ignition advance curve, perhaps a breakerless magnetic impulse ignition system, a good set of wire-type radio-shielded secondary spark plug and coil cables, and so on. Each item individually will improve performance and most of them will reduce exhaust emission levels, while the remaining few will at least not be detrimental to exhaust emissions. So what have we got? With all of the preceding done with precision and moderation, we have a combination; an engine assembly capable of producing a very good level of performance, again in the most frequently-used engine speed ranges…without extremes in any direction. We also have a more efficient engine assembly, on which at least has the potential of reducing exhaust emissions significantly. Whether it does or not is up to the individual more than it is to judiciously-applied and moderate modifications.


Race Engines


Now let's go on to fully-modified Datsun cammer race engines. Generally, these engines are quite sensitive and respond in a most gratifying manner. This doesn't mean "biggest" or "most" is always best in any single component or combination. Again, the strictly competition engine that is most successful in a given application must necessarily be compromised in one or more areas. Assume that you're stuck with a four- speed gearbox. This means that the ultimate, last-gasp, maximum-effort engine will be out of place because the gearbox ratios won't let the engine run where it is happiest and can do its best job. Compromise: You need better torque at lower engine speeds to overcome the gearbox ratio penalty, which usually involves some sacrifice in maximum power output. Solution: Use a camshaft with slightly shorter effective duration, retain as much valve lift as possible, and perhaps advance the camshaft three or four crankshaft degrees, assuming no piston-to-intake-valve hang-up. In addition, a slightly smaller diameter exhaust header primary pipe from four to six inches longer will also help this condition, as will longer carburettor air horns. Secondary solution: If the bankbook and rulebook have no serious objections, purchase the optional Datsun intermediate-ratio five-speed gearbox with the necessary drive shaft, and the appropriate ring and pinion ratio, the right diameter tyres - and go race.

If mid-range and upper mid-range output is a major factor, effective duration should be kept in the high-280 degree to mid-290 degree range with from 70 degree - 75-degree overlap. Valve lift should be in the 0.580 to 0.610-inch range. If otherwise rightly equipped, such an engine will be strong from about 4,800 RPM on up to about 8,000.

If a maximum-effort engine seems to be the correct plan, and vehicle is properly geared so that minimum engine speed doesn't drop below about 6,000 RPM then the skies (almost) the limit. In this case, effective duration should be in the 310 degree - 320+ degree range with from 0.620 to 0.650-inch valve lift. With such a camshaft, the best effective engine speed range would normally be from about 5,800 to 6,000 RPM through at least 8,800. But a word of caution here: L-16/L-18 four-cylinder engines are better equipped to handle extreme engine speeds that L-24 sixes. The L-24's weak link is the 50% longer crankshaft, therefore torsional oscillations of the crankshaft are more severe. Consequently sustained maximum engine speeds in L-24 engines should be limited to 8,000 to 8,200 RPM, and even then the crankshaft damper, the crankshaft damper bolt, flywheel, clutch pressure plate bolts and camshaft retaining bolt must be constantly checked for torque and replaced, preferably before something breaks. Four-cylinder engines are not as bad in this respect, although there are some unbalanced secondary forces at work, so they are not completely immune from similar problems.




This brings us to another definition, that of displacement angle. Displacement angle is defined as the angular relationship from a given point on the exhaust lobe opening and closing flanks to the same given point on the intake lobe opening and closing flanks of the same cylinder.

Displacement angle is expressed in CAMSHAFT degrees (exactly one-half the amount of crankshaft degrees). Previously-shown camshaft examples have a displacement angle of around 108 degrees as an average figure. With the longer and taller competition camshafts, the Datsun cammer engines like to have the displacement angle squeezed up to about 102 degree to 105 degree. Example: Assume an effective duration of 280-degree with a displacement angle of 108 degree. The nominal valve timing event, in CRANKSHAFT degrees would be: Intake opens 32 degree BTC, closes 68 degree ABC. Exhaust opens 68 Degree BBC, closes 32 degree ATC. Add 32 degree to 68 degree plus 180 degree and the effective duration is 280 degree for both intake and exhaust. Now add the intake opening point (32) to the exhaust closing point (also 32) and we have 64 degree. Next subtract the degrees of overlap (64) from the effective duration (280) and we have the displacement angle of 216 CRANKSHAFT degrees. Finally, divide the 216-degree number by 2 to obtain the displacement angle of 108 in CAMSHAFT degrees.

Now advance the camshaft one camshaft degree (two crankshaft degrees). The valve timing then becomes: Intake opens 34-degree BTC, closes 66-degree ABC. Exhaust opens 70 degree BBC, closes 30 degree ATC. Three things remain exactly the same: Intake valve duration, exhaust valve duration, and overlap. All we have done is to change the valve opening and closing points. If the camshaft is retarded the same amount from the original figures the valve timing becomes: Intake opens 30-degree BTC, closes 70-degree ABC. Exhaust opens 66 degree BBC, closes 34 degree ATC. Intake and exhaust valve duration and valve overlap period still remain exactly the same; all we have done is to move the valve opening and closing points in the opposite direction. Next assume a camshaft with an effective duration of 310 degree and a displacement angle of 105 degree. Nominal valve timing would be: Intake opens 50-degree BTC, closes 80-degree ABC. Exhaust opens 80 degree BBC, closes 50 degree ATC, these numbers again being expressed in crankshaft degrees. As before, add 50 degree to 80 degree plus 180 degree and the effective duration is 310 degree for both intake and exhaust. Again, add the 50-degree intake valve opening point to the 50-degree exhaust closing point for the overlap period of 100-degree. Subtract the 100-degree overlap period from the effective duration for the displacement angle of 210 crankshaft degrees, then divide by 2 for the displacement angle of 105 camshaft degrees. The advance-retard game can be played here as well and again, the valve opening and closing points are the only items that will be changed, and many times this is certainly enough.

The basic idea is that displacement angle is a fixed value. Once the camshaft is made, the displacement angle cannot be changed unless the camshaft is reground, and then only within fairly close limits.

Now let's take our same 310-degree effective duration camshaft and squeeze the displacement angle up from 105-degree to 102-degree. The nominal valve-timing event in crankshaft degrees will be: Intake opens 53-degree BTC, closes 77-degree ABC. Exhaust opens 77 degree BBC, closes 53 degree ATC. The intake and exhaust valve duration remain the same at 310-degree. However, the valve opening and closing points have been changed, and the valve overlap period has been increased from 100 degree to 106 degree. By applying the above equation (duration - overlap divided by 2) the displacement angle is 102 camshaft degrees. There are a couple or three messages here. The first seems pretty obvious: As effective duration increases, the displacement angle should be decreased, within reasonable limits, of course. Why? Primarily because the effective intake valve closing point can be kept at a sensible number, which will keep mid-range torque from dropping absolutely dead. Secondarily, assuming the engine can breathe well during the overlap period, the increased overlap will let it do jus that. The third factor that rears its ugly head as effective duration is increased, or displacement angle is decreased, or a combination of both, is maintaining adequate piston-to-valve clearance through the operating cycle.


Chapter Nine




With a lot of luck, many prayers, and maybe some intelligent letters to companies involved in the manufacture of such items, there may be a decent aftermarket piston available some day for the Datsun cammer. Venolia markets on forged type now and Interpart utilises a TRW piston forging. Both of these have terribly high domes. In any case, the design criteria for an acceptable piston, aside from a good stiffness-to-mass ratio, proper skirt design to prevent skirt collapse, good piston ring placement, minimum piston pin exposure between the pin bosses in the piston, etc., should be:

(1) Acceptable compression ratio, particularly for modified L-16 and L-18 and "open" type L-24 combustion chamber cavities, but without a high piston crown resembling a misplaced Alp.

(2) Adequate space around the spark plug to give a good strong point for combustion propagation.

(3) More-than-adequate crown thickness in the area of the valve reliefs, so that the valve reliefs can be made not only deeper but with radii somewhat larger than the valves, and without weakening the hot-strength of the piston structure.

(4) A piston designer with brains enough to recognise the importance of the fact that the air/fuel mixture and/or exhaust gases simply DO NOT and will not flow properly around sharp edges and corners. Perhaps as a very minor minority of one, I would like to see a piston in which the intake and exhaust valve reliefs are joined together to form a single "trough" type valve relief, similar to the original equipment Chevrolet 302 Z-28 piston.

Item 2, 3 and 4 are essential for proper cylinder breathing, particularly during the valve overlap period, and particularly at extremely high engine speeds, say above about 8,500 RPM on up. However, these items are contrary to Item 1, acceptable compression ratio. Nearly everyone has heard of some ridiculously high compression ratio numbers bandies about for these engines, but I can assure you that it is extremely difficult indeed to get and honest, genuine measured compression ratio of over 12 to 1, without shrouding the valves, the spark plug, or both, and still keep the power curve relatively flat above about 8,500 RPM. Again, this refers to the "open" L-24 type combustion chamber with valve unshrouding modifications, but it also applies to similarly modified optional L-16/L-18 large-valve heads. If the engine is going to run well and produce very good power in the 8,500-plus-RPM range, it must have breathing room to do so, even if it means a sacrifice in compression ratio.




Lots of piston-to-valve clearance does more than allow the engine to breathe well at extreme engine speeds. It also provides the engine with a "cushion" of space in the event of a missed shift or broken driveline component when the engine speed would tend to go completely out of sight, and thus minimises any engine overspeed damage. It also permits the camshaft to be advanced or retarded to suit conditions of the moment.

Given a food dynamically stable cam lobe profile, Datsun cammer engines have been known to run up to well over 10,000 RPM without damage of any kind, but no one in their right mind would do this intentionally. Peak power in a well and properly modified maximum-effort Datsun cammer engine usually falls in the 7,600 - 8,000 RPM range, so an overspeed condition to 8,800 or 9,000 RPM is acceptable. For once, this is a case where the valve train is not the speed-limiting factor of the engine. Happily, this is because the camshaft is well-supported in the cylinder head and is stiff by itself, the rocker arms are stiff, and the reciprocating mass of the rocker, valve, vale springs, spring retainer, lash pad, etc., is quite low.

We normally recommend a minimum piston-to-valve clearance of at least 0.090-inch between the piston and the intake valve at their closest point, and at least 0.100-inch and preferably more - between the piston and the exhaust valve when the piston-to-valve clearance is measured by rotating the engine by hand. In an operating engine at full blat, these numbers are usually reduced by about half because the crankshaft bends, the connecting rods stretch, the piston pins bend and the pistons stretch, and it all happens just before and after top centre when the pistons change direction, the least opportune period for maintaining adequate piston-to-valve clearance. The condition is made worse during a closed-throttle down-shift, or similar circumstance when the cylinder (under a pressure vessel) is changed into a vacuum vessel by the closed throttle, at which time the high vacuum in the cylinder tends to draw the pistons and valves together more closely than when the engine is under load.

The reason why additional piston-to-exhaust valve clearance is called for is because any camshaft drive tends to permit the camshaft to retard itself in relation to the crankshaft, thereby bringing the exhaust valve closer to the piston. This is particularly true with a chain-driven camshaft because as engine speed increases, centrifugal force acting upon the chain pushes the chain away from the sprockets and the chain rollers contact the sprockets increasingly closer to the outer edges of the sprocket teeth, and all the while the camshaft is resisting rotation due to friction, valve spring loading, etc. Chain stretch compounds the problem. It is a function of the load applied to the chain and of course the number of links in the chain. Datsun chains are quite long, about 42 inches in circumference, with 110 links, which means that keeping the valve timing exactly right at all times, is nearly impossible. It is therefore better to start out with a slightly advance camshaft - say by 3 or 4 crankshaft degrees - because there is no way in the world that it can be kept from retarding itself as the engine is run, particularly at high engine speeds. This is also the reason for the three different timing marks and three different dowel pin holes in the Datsun camshaft sprocket; they make provision for advancing the camshaft in 4-crankshaft degree increments, but they make no provision for retarding the camshaft. It does that by itself with no outside help required!

A high dome, high compression ratio piston that shrouds the valves and the flame front isn't all bad; it really wakes up the low and mid-range torque, but don't expect the engine to have good breath control at 9,000 RPM. It won't!

Naturally, any increase in cylinder bore diameter and/or crankshaft stroke will raise the compression ratio with a given combustion chamber cavity volume. In fact, a couple of 2.5 TransAm Datsuns were L-18 engines bored and stroked to just under 2,000cc's.

Obviously the cylinder head can be milled to gain compression ratio, but with the Datsun engines, there are some not-so-obvious factors involved. Milling from 0.060 to 0.070-inch is considered a "safe minimum," and then there may be head gasket sealing problems. If there are sealing problems, the cylinder head should be O-ringed with 0.030 to 0.040-inch diameter soft copper armature wire or soft stainless steel wire, leaving about 0.010 to 0.012-inch of the wire exposed from the cylinder head face to give the required seal around the cylinder bores.

O-ringing the cylinder block is not recommended unless the final cylinder bore honing operation is done with a honing plate, cylinder head gasket and O-rings installed and torqued down. If this is not done with O-rings in the block, the O-rings will cause the tops of the cylinder bores to shrink about 0.002 to 0.003-inch on the bore diameter, and will affect cylinder bore diameter for about and inch down from the block face when the cylinder head is bolted to the block. We really don't need sticking pistons. When a Datsun cylinder head is milled, on and possibly two other items must be put right. With the camshaft in the cylinder head and the head milled, the centre-to-centre distance between the crankshaft and the camshaft is shortened, which means there is extra "slop" in the camshaft drive chain. To correct this condition, the lower end of the left chain guide (viewing the engine from the front) must be moved to the right. Chain guide bolt holes are slotted for this purpose, but it may be necessary to slot the lower hole more to take up the additional slack. This also causes the chain tensioner piston to extend itself further from the chain tensioner housing. If the piston extends from the housing more than about ½-inch, it may become stuck in the bore, unable to move in either direction, and worse, may not be able to take up the chain slack, perhaps causing the chain to skip one tooth or more on either sprocket, at which point the entire engine is in deep trouble. A sure fix for this is to make a thin-wall steel sleeve 1-1/4 to 1-3/8 inches long to press over the existing piston, then bore out the tensioner housing so the sleeve slip fits. Sleeve wall thickness should not be more than 1/16-inch. Sounds like a lot of monkey-motion, but it may be essential if the cylinder block face is machined for minimum deck clearance and if the cylinder head is milled the maximum amount. One more point: If the cylinder block and cylinder head are milled, it is mandatory that all cylinder head bolts be checked in their respective tapped holes in the cylinder block to be certain they do not "bottom" in the holes rather than clamp the cylinder head.

If you have a 68-69 four-cylinder engine with the ratchet-type chain tensioner, throw it away. Replace it with the later type or the FIA type in the competition parts listing. Do not attempt to interchange L-24 tensioners with those for the L-16/L-18 or vice versa.


Spray Bar Details


When installing the L-24 spray bar onto an L-16.18, it is cut in the centre, shortened and rejoined with a silver soldered sleeve. The two centre cam towers must be faced off on one side in a mill to provide a flat surface for mounting. A 10-32 helicoil is inserted in the cam tower to mount the spray bar and the flat-head mounting screws are drilled to allow oil to flow to the spray bar from the cam journals. Ray Gruss at BRE uses the single-drilled-screw method for each of the mounting bosses. Dolf van Kesteren mills the towers and drills two holes to make an installation very similar to that used for the 240Z. He points out that the spray bars have a nasty habit of breaking from fatigue. They should be checked after every race to make sure that no cracks have developed. If any cracks are found, trash-can that assembly and install a new one. There is no use trying to repair the one that is starting to fail. The 240Z spray bar assembly is 13100-E3003 and will be listed as "assembly, oil-tube cam" in your friendly local dealer's parts book. Incidentally, the holes in the tubes are located correctly so you do not have to redrill new ones.


Chapter Ten




The best method for installing a camshaft in a Datsun cammer engine is to have the cylinder head on the bench. This allows the valve seats and valve faces to be ground, as well as any other machine work required for the installation. Therefore, it is best to be armed with the Datsun shop manual for the engine concerned. If the rest of the engine is in pieces, that is one thing, but if the engine is to stay in the chassis, it is another matter. In the latter case, first loosen the camshaft sprocket-retaining bolt. The bolt is (should be) usually very tight, so just barely loosen it - DO NOT remove it. Then the engine must be rotated until the first (bottom) bright (or marked) link on the camshaft drive chain is engaged with the tooth that has the timing mark on the crankshaft sprocket (which you can't see) and the second (top) bright (or marked) link on the chain is engaged on the tooth with the timing mark on the camshafts sprocket. If the engine is not taken apart, you can count 41 links (not link plates) from the second bright link engaged with the timing mark on the cam sprocket until the first bright link engages with the timing mark on the sprocket, then reverse the procedure to get the bright link back down on the crank. That way you can be sure that you have the cam timing correct, assuming that the chain was correctly engaged with the correct tooth on the crank sprocket to start with. At this point, both sprocket timing marks will be nearly horizontal on the right side of the engine when viewed from the front, with the camshaft sprocket mark above horizontal and the crankshaft sprocket mark below horizontal, and the crankshaft key and the camshaft dowel pin will be pointing upward nearly vertically. This will (or should be) top centre of the compression stroke for number 1 cylinder (both valves closed), and the TC mark on the crankshaft damper should be in line with the point. If there is any disagreement in the reference marks, poke a long screwdriver through the spark plug hole in number 1 cylinder so the end of the screwdriver contacts the top of the piston. With one finger lightly pushing the screwdriver upward at the end, rotate the engine in both directions just enough so that the screwdriver tells you that the piston is at top centre as close as you can feel it. The longer the screwdriver, the more sensitive an indicator it will be, and this will be close enough for the moment. Before removing the camshaft sprocket, us Datsun tool number ST-17420000, a hardwood wedge. Slip it gently but firmly down between the timing chain links to keep the timing chain correctly oriented with the timing marks and better yet, to keep the chain tensioner piston from falling out of the chain tensioner housing and into the oil pan. If this happens, the oil pan must be removed to retrieve the piston, and it will only be a stroke of sheerest luck if the piston can be replaced in the housing without removing the entire front cover assembly, including the crankshaft damper, water pump, oil pump, ignition, etc. With the wedge firmly in place, it is now safe to remove the camshaft sprocket and continue with removing the cylinder head assembly.




Any engine to be used strictly as a race engine should obviously be completely disassembled and rebuilt with the race application in mind. Similarly, a street engine, or a dual-purpose engine with lots of mileage on it should also be completely rebuilt. In the latter case, a suggestion is in order. Stock Datsun pistons are permanent-mould aluminium castings with integral steel struts to control piston skirt expansion available in oversizes up to 0.060-inch (1-1/2 mm). These will give much better oil control than an aftermarket forged piston because you can use tighter piston skit-to-cylinder bore clearances. However, they are only available in a flat-top configuration without valve reliefs. If the engine is to be compromised more toward street operation, use 'em. The price is right, too, and you don't have to fight the battle of full floating piston pins, which can be much more bother than they are worth.


Cylinder Head


Any cylinder head, new or old, should have the valve seats and valve faces ground and lapped, as well as checking and correcting the valve stem-to-valve guide bore clearances. If any port and/or combustion chamber work is to be done, now is the time. Before the cylinder head is completely disassembled, it's a good plan to measure and record and dimension from the valve stem tip to the valve spring pocket in the head with the valve closed for all valves. This is a pretty critical dimension, so a "yardstick" measurement isn't good enough; it must be made with a depth micrometer. The best way to do this is after the camshaft, rocker arms, valve springs, lash pads, retainers, etc., have been removed from the cylinder head assembly. Next, with all valves held closed, lay a straightedge along the line of valves in the head to see if there are any major discrepancies in valve stem length from one end of the head to the other. Ideally, the straightedge should be parallel with the cylinder head gasket face with the straightedge contacting all valve stem tips. There will usually be differences in valve spring pocket depths in the head, so the valve stem tip-to-valve spring pocket dimension will apply to only one valve. The valves must be numbered in the proper sequence so they can be reinstalled in the same guide bores.

These days, nearly every cylinder head artist has his own "trick" method for doing valve jobs, and some of these are OK, but the purpose for which the engine is intended must be considered. Thinning the valve heads, fully-radiused valve seats and other exotica are not for a street or dual-purpose engine where longevity is an important factor. A multiple-angle cut in the valve pocket, beneath the valve seat, is fine as long as the cuts are consistent and conservative, with a "topping" cut of no more than 15 degrees and a flatter angle may be better. Intake valve seat width should be from 0.070 to 0.075-inch and exhaust valve seat width should be from 0.085 to 0.090-inch. Valve seat and valve face angle should be 45 degree. Shallower seat angles may be OK in race-only engines but they don't seal as well or last as long as 45 degree seats. In any case, it is important to keep the valve seats as close as possible to the cylinder head face. In other words, DON'T sink the valve seats in the head. This practice generally ruins airflow characteristics, particularly at the intake valve and port. The depth of the intake valve seats from the cylinder head gasket face must all be equal as close as it is possible to make them. Ditto for the exhaust seats, except they won't necessarily be the same as the intakes due to the difference in valve sizes. Once the valve seats and valve faces have been ground - and preferably lapped - and any other new pieces installed such as new valves, new valve guides, new valve seats, and all related work completed, make another measurement from the valve stem tip to the valve spring pocket with the valve seated for all valves and also make another straightedge check along the line of valves in the head. There is no doubt that the valve stem tips will extend further than they did originally. It is now necessary to find the valve stem that projects the least amount from the topside of the cylinder head. This will be visible with the straightedge check assuming the straightedge is held exactly parallel to the cylinder head gasket face. Next, shorten all remaining valve stems to match the shortest one. This must be done with a valve-refacing grinder with a fixture for grinding valve stem tips. The grinding wheel must be clean, sharp and dressed so the finished valve stem tips are square to the valve stem axis, flat and with a smooth surface finish.

The purpose of this exercise is to equalise all the valve stem tip lengths, but in addition, it permits the same thickness of valve lash pad to be used on all valves. If this procedure is not followed, very likely each valve will require a valve lash pad of different thickness, which is a pain in the neck and elsewhere. This is itchy work because any variation in valve stem tip length is reflected as a valve lash pad thickness requirement as a function of rocker arm ratio. If one valve stem tip extends 0.010-inch more than its neighbours, then the valve lash pad thickness requirement is changed by 0.015-inch, assuming a rocker ratio of 1.5 to 1. Therefore, the closer a zero-tolerance condition is approached, the better.




Before we leave cylinder head modifications, some additional points must be made. If, due to the application, modifications to the intake ports and pockets, exhaust ports and pockets, valves, combustion chambers, etc., are indicated as they usually are even in engines with milder states of tune, it is highly recommended that the cylinder head and valves be sent to a cylinder head expert equipped with an air flow bench and the intelligence to use it correctly. Port, valve and combustion chamber design, to say nothing of induction and exhaust systems, have become extremely sophisticated within recent years, and nothing can kill a cylinder head quicker and deader than gouging it out in the wrong place by someone who may have the best intentions in the world, but who lacks the know-how, experience and air flow measuring equipment, all of which are required these days to do the job correctly. So go to an expert in the first place and save yourself the cost and frustration of having to do the job again after it has been bungled. Be very explicit in telling him your exact requirements for the engine and let him decide what is appropriate in the area of valve and port sizes, etc. The same expert will usually be equipped to equalise the volumes of the combustion chamber cavities in the head. All this may not be cheap but it will be cheaper than having to do the job twice in order to get it done once, and right. If the guy is sharp and knows his Datsuns, it probably won't be necessary to tell him that all L-series engines need more help than the intake ports, and also that total exhaust air flow should fall within the range of 75 to 80% of total intake air flow.

There are two separate and distinct approaches to the proper reworking of L-series Datsun cylinder heads and the correct route depends upon the application. For street or dual-purpose engines, shoot for as much airflow as possible at relatively low valve lifts and let air flow at maximum valve lift fall where it falls. The air flow curves should be good and fat at low valve lifts, on both intake and exhaust ports, without dips or "holes" in the curves, and if air flow doesn't increase much beyond valve lifts of 0.450 to 0.475-inch, who cares? When this is done correctly, gas velocity will usually be quite high for both intake and exhaust and the engine will show it by being very throttle-responsive throughout the normal engine speed range, but will be at tis best in the mid- range and upper mid-range. If air flow through the ports more-or-less "signs off" at say, 0.460-inch valve lift, and the actual valve lift is in the 0.470 to 0.480-inch range, so much the better. This means the valves will be at maximum air flow rate for a longer period of crankshaft rotation and this will keep engine performance alive and well in the higher engine speed ranges. In such cases, camshafts with longer effective durations will not be as strong at lower engine speeds as those with shorter effective durations in conjunction with fairly healthy valve lift numbers. But don't get carried away; those 0.600-plus inch lift numbers are not for street or dual-purpose engines.

Strictly race engines are of another planet. However, airflow at lower valve lifts cannot be abandoned in a search for the highest possible airflow at some ridiculously high valve lift figure. L-series Datsun cylinder heads can be modified to produce very good air flow figures at valve lifts in the 0.650-inch range, but this happens at the expense of air flow at lower lifts and has a detrimental effect on performance in the mid-range and even upper mid-range engine speeds. This occurs because the valves may reach the point of maximum air flow for only the shortest period of time, if at all, so the ports and air flow numbers, as magnificent as they may be, cannot be utilised effectively and the whole episode could easily dissolve into an exercise in futility.

By far the better plan, and one that really works, is to make every attempt to retain as much air flow at lower valve lifts as possible without giving anything away at higher lifts; not necessarily at the most extreme lifts, but at some reasonable and acceptable number say, in the range of 0.575 to 0.600-inch. Then the cam lobe profile can be made with enough additional lift so the valves are at or above the point of maximum port flow for a considerably longer period of time and/or crankshaft rotation. In this way, the engine does not give nearly as much away in the mid-range speeds, yet maximum power will very likely be at least as good, but more likely better than if extremes are attempted. I have seen some L-series Datsun ports that flow some extremely impressive numbers up to and including 0.750-inch valve lift but attempts to make use of these ports have been the most total, dismal and dreary failures imaginable, and for two very good but separate significant reasons. First, airflow at valve lifts below about 0.480 to 0.500-inch has never been enough to blow the dust off your desktop. Second, the L-series Datsun rocker arm pads are simply too short to accommodate the type of cam lobe profiles necessary to generate such enormous valve lifts without running off both ends of the rocker pad, but there are secondary, nevertheless important, mechanical considerations as well. And the Datsuns are of relatively modest piston displacement, and they don't really need, nor can they use, all the valve lift in the world. Besides, you'd have to drain the oil to get the valves open.


Chapter Eleven


Bore Notching


It should be pointed out the combustion chamber cavities in all Datsun L-series engines are longer than they are wide, and the long dimension of the chamber cavities is parallel to the crankshaft and measures about 3-7/16 inches. It is larger than the stock cylinder bore. This becomes visible when a stock cylinder head gasket is laid on top of the cylinder block and located by the two large dowels in the block. Matching the cylinder head gasket to the combustion chamber cavities is, or should be, a foregone conclusion in order to gain as much space around the valves as possible. This will lengthen the chamber cavities about another 1/16-inch to bring the chamber length to about 3-1/2-inches, or about 0.230-inch larger than the stock cylinder bore in L-16 and L-24 engines. Most of the chamber overhang is biased toward the intake valve end of the chamber and represents a very definite breathing restriction. Of course, the larger the cylinder bore, the less the restriction, but even with the largest feasible cylinder bore, the ends of the chambers will overhang the cylinder bore somewhat.

The cure for this is pretty obvious. It's called "eyebrowing" or bore notching and is discussed in the blueprinting and assembly chapter. For best cylinder breathing, this operation should be performed to the cylinder bores beneath both the intake and exhaust valves, although it is more critical on the intake sides of the cylinder bores.


Valve Springs


Stock Datsun valve springs are quite light and generally are not suitable for valve lifts in excess of 0.440 to 0.460-inch. Besides they stack solid at a lift of about 0.500-inch. They're fine for stock camshafts, but they are the engine speed-limiting factor with camshafts having higher valve velocities. The type and number of valve springs, and consequently the valve spring loading, depends upon the type of camshaft and the application. For street or dual-purpose applications, we usually advise a single damper-type outer spring with a load of 70 pounds with the valves seated and 210 pounds at 0.500-inch valve lift. This spring will accept a valve lift of 0.600-inch without overstressing the spring wire, so it has a lot of latitude without causing and overload condition between the cam lobe-rocker pad interface. As a comparison, the stock L-series valve springs (they are the same now for all late L-series engines) have a combined inner and outer spring loading of 67 pounds with the valve seated and 165 pounds at a valve lift of 0.410-inch (stock L-24). The above-mentioned spring with a load of 70 pounds with the valve seated has a load of 188 pounds at 0.410-inch valve lift, so the difference in loading really isn't out of line. The stock L-series Datsun spring retainer is a very sloppy fit in this spring and with the stock retainer, the assembled spring length comes out about 3/16-inch too short, therefore we make a special retainer that is an interference fit in the spring and will also set the springs up at the correct assembled length and load.

For race-only engines, we advise a three-piece outer spring-damper-inner spring assembly. The springs are selectively-fitted to obtain an interference fit and the special retainers are an interference fit with both the inner and outer springs, all of which keeps the springs from waltzing about. With this combination, combined spring loading with the valve seated is 85 pounds with 370 pounds at the valve lift of 0.660-inch. While the latter figure may seem like lots, the 0.660-inch valve lift figure is lots, too.

Some special spring assemblies require that the valve spring pockets in the cylinder head be machined so that the inner and outer springs seat on the same level in the head. This does not imply that the spring pockets should be machined deeper in the head; it simply means that both the inner and outer springs seat on the same plane in the head, and the original spring pocket surface just barely cleaned. Deepening the spring pockets is courting disaster by puncturing the rood of and intake port, or by a highly-undesirable "ventilation" hole in the water jacket above and exhaust port, or both. Don't.

Because most special valve springs are both larger in diameter and longer, there are a couple of points where a hang-up could occur. Fist, in L-16 and L-18 engines, the valve spring and/or spring retainer may interfere with the cam towers on all four intake valves and number 1 and 4 exhaust valves. The corrective measure is to take a slight cut from the cam towers with a rotary file so there is no chance of interference at these points. This is usually not a problem with L-24 engines. The second point of possible conflict could be between the underside of the rocker arm and the edge of the spring retainer when the valve is seated. This is more likely to occur with longer assembled valve spring lengths, but is should be checked out in any case. To correct for interference here, use a fairly large diameter sanding drum to relieve the rocker arm from actual contact, plus an extra 0.015 to 0.20-inch, then round off any sharp corners or edges on the rocker to prevent generating a stress-rise. Or, use the late-style rocker shown and referred to elsewhere in this chapter.

One of the nicest features of the L-series Datsun engines is the absence of pushrods, those rubbery objects that usually do more damage to the theoretical valve lift curve than anything else imaginable. All the L-series engines need is a good, dynamically stable cam lobe profile with enough valve spring load with the valve closed or open to prevent separation of the rocker arm tip and the lash pad at high engine speeds. The term "dynamically stable" simply means that the actual valve lift curve at the red-lined engine speed (maximum safe engine speed) deviates as little as possible from the theoretical valve lift curve, and still has a few hundred revolutions as a "cushion" against engine damage from a momentary accidental overspeed condition.

Thus with the L-series Datsuns, "overkill" with excessive valve spring loading is not only nnecessary but undesirable from the standpoint of accelerated wear of the cam lobe/rocker pad interface but also because higher valve spring loadings chew up horsepower. In additional, these is no way in the wide world that raising the valve spring loading can make a good cam lobe profile out of a bad one. A dynamically bad bump shape is just as bad - maybe worse -with a ton of spring load as it is with none. Any reputable manufacturer of Datsun camshafts and valve train accessories will (or should) have enough variety of valve springs, spring retainers, lash pads, etc., to satisfy must about any requirement within the realm or reason. This doesn't mean that you should blindly accept his word for such things. Check out the spring load and assembled spring length dimensions to be absolutely certain that there is no possibility that the spring, or springs, will be stacked solid at or near full valve lift. There is nothing that is so quickly and completely destructive to the camshaft and valve train equipment as on or more stacked valve springs - and all it takes is an honest mistake in the valve spring load and/or assembled valve spring length dimensions on the timing card that accompanies the camshaft. Or, perhaps worse, the wrong valve springs with the correct specifications for the right springs.

In any cast, all inner valve springs (if used) should be checked at full valve lift, by themselves first, because it is not possible to see through the outer spring coils and damper coils to determine visibly if the inner spring is stacking solid at or near full valve lift. And some inner springs will stack before the outers, and some damper coils will stack before anything else. It therefore becomes mandatory to check each individual valve spring and damper coil to insure that stacking the valve spring assembly is not even a remote possibility. While you're running through this exercise, remember that the assembled spring length of the inner spring is always shorter than that of the outer spring by the thickness of the shoulder on the spring retainer, and in case stock Datsun springs are used, also by the inner spring shoulder in the cylinder head. A spring tester is the best way to handle this chore because spring length versus spring load can be measured simultaneously. If nothing else is handy, a drill press or even a bench vice can be used to measure spring lengths at the valve seated and full lift dimensions as a safeguard against stacking. Also, remember that only the springs are to be measured. Do not measure the spring retainers. Valve springs, particularly race-quality springs, are among the most highly-stressed components of any engine assembly, but nothing lasts forever, so anything that can be done to ease the stress conditions of the springs will automatically make life easier on them and add to their longevity as well. Therefore, RESIST and AVOID the temptation of pulling the springs down to the last jillionth of an inch before they stack solid. All this accomplishes is an undesirably high spring load condition, but worse, causes premature spring fatigue, which ultimately results in spring breakage, and this is as bad as an armed hand grenade in the sump. Another bit of kindly consideration to the valve springs would be to remove the rocker arms if the engine is not to be used for a while to relax the springs at least to their installed lengths, then pour clean engine oil over them and drop the cam cover on to keep out dirt and moisture. If the engine is to be stored for any length of time, it's best to remove the springs and store them in a can of clean and covered engine oil. A valve spring that picks up a spot of rust will break. There is no argument or speculation about that; the only question is-when? Probably when you need it the most in one piece.


Valve Spring Detailing


A couple of other refinements are worthy of mention and consideration. The first is absolutely essential. Before the valve springs are installed, but after the valve stem oil seals have been installed, install the valves and assemble them with spring retainer and the top of the valve stem oil seal. If there is any conflict between this dimension and that of maximum valve lift, the seals must be removed and the tops of the valve guides shortened so that there is a least 1/16-inch between the bottoms of the spring retainers and the tops of the valve stem oil seals. There is usually no hang-up at this point unless the valve lift is 0.600-inch or more. Next, it's a very good plan to remove the sharper corners and edges from the end coils of both ends of all valve springs and damper coils. This is done only at the points where the springs and dampers would contact a flat surface. Use a small, fine grit-sanding drum in a hand grinder or drill press. It takes only a touch to relieve the sharp edges and corners, which would otherwise gnaw away at the spring retainers and the steel shim that must be used between the bottoms of the springs and the aluminium cylinder head. When this is done, thoroughly wash the spring assemblies in clean solvent of lacquer thinner to remove all traces of abrasive dust, blow them bone dry with compressed air and dunk them in clean engine oil for the moment.


Checking Installed Spring Height and Other Details


Now we can install the valve springs, right? Wrong! There are a number of operations that must be carried out first that are best done with the head on the bench and without the valve springs. The first is to measure the assembled valve spring length in the cylinder head for each valve assembly, which for the moment consists of a valve, spring retainer, valve locks and steel spring shim. Keep each assembly separated from the others to prevent a parts mix-up during final assembly. Place the spring shim in the spring pocket, install the correct valve in the correct valve guide bore, slop the retainer over the valve stem, install the valve locks in the valve stem groove then pull up hard on the retainer to seat the valve locks on the valve stem and in the retainer. Apply upward pressure to the retainer to keep the valve on the seat and to keep the retainer and valve locks from coming loose. Use a 1-1/2-inch to 2-inch telescope gauge to measure the distance between the out flange of the spring retainer (for the outer spring) to the valve spring shim in the spring pocket. It is very important to keep the telescoping members of the gauge parallel to the valve stem in both planes. When the gauge is in the right position, lock it up, remove it and measure the gauge with a 1 to 2-inch outside micrometer. Comparing this dimension with the specified assembled valve spring length will tell you if the spring length is too long, too short, or within the specified length tolerance (usually plus-or-minus 0.010-inch). If the spring length is too long, add the correct number of shims to the spring pocket in the head and remeasure the assembled spring length again. This is not the place to goof off, so don't. If the spring length is within the specified tolerance, with the required one shim against the head, take the valve assembly apart, put it aside and repeat the process on the next valve assembly. At first, it feels like two hands aren't enough and three are too many, but two is the correct number. If the assembled spring length is shorter than that specified, the problem can be solved in another way. But remember: DO NOT sink the valve spring pockets in the cylinder head any deeper than the original surface plus a clean-up cut of 0.010-inch, no more, and DO NOT sink the valves in the head. The standard L-series valve stem diameter approximates 5/16-inch and, by some coincidence, so does the Chrysler 426 Hemisphere street and race engines made from 1964 through 1971. Get yourself the required number of 426 Hemi valve locks; 16 for L-16 and L-18 engines and 24 for L-24 engines. These locks have keys that fir the valve stem groove about halfway in their length, whereas the keys in the stock Datsun locks are at the very tops. Using the hemi valve locks increases the valve spring length by about 0.065 to 0.075-inch, which may be enough to get the assembled valve spring length back into the right ball park. However, the Hemi locks are too long at the top and if used as is, the valve lash pads will contact the tops of the locks instead of the valve stem tips, a very unhealthy condition. So shorten the Hemi locks by about 0.040-inch from their top surfaces and assemble them with the valves and retainers to be certain that the valve stem tips project from 0.010 to 0.015-inch above the tops of the valve locks. This works and is entirely safe.

If the valve stem diameter in the Datsun head has been changed to approach 11/32-inch (as for use with FIA-type valves), see your friendly Chevrolet parts dispenser for the correct number of 283-327-350-400 Chevrolet valve locks. The same modification must be applied to the Chev valve locks as to the Hemi; otherwise they will fit and function as intended. That is, except for the possibility that the increase in assembled spring length may cause some interference between the tops of the valve springs and/or retainers and the rocker arms, as mentioned earlier. Install the camshaft in the cylinder head using the proper Datsun service manual as a guide and also as a check for camshaft bearing bore-to-camshaft journal clearances. With the camshaft thrust plate and the camshaft sprocket in place, fore-and-aft movement of the camshaft in the head should be from about 0.008 to 0.015-inch. There are three thicknesses of camshaft thrust plates available so use the one that comes closest to the above fore-and-aft movement tolerance. At his point, it isn't necessary to have the camshaft retaining bolt as tight as it should be on final assembly, so an extra 0.002 to 0.003-inch fore-and-aft camshaft movement is OK for now.


Chapter Twelve


Selecting the Rocker Arms


If you haven't already done so, now is the time to become friendly with your local Datsun parts source. Buy a couple of sets or more of new Datsun L-series rocker arms (Datsun part number 13257-21000) with the understanding that you can return the ones you don't need. NEW rocker arms are an absolute necessity with a new camshaft. The purpose here is to select a set of rocker, plus some spares that are as equal in all areas of measurement as possible.

Place on valve in the cylinder head with a light mechanical fuel pump spring to hold the valve closed, the correct spring retainer, valve locks and lash pad. Place a dial indicator spindle on the spring retainer with the spindle parallel to the valve stem in both planes, and in such a way that rocker arms can be removed and replaced without disturbing the indicator. Pre-load the indicator to something more than maximum valve lift. Install on rocker arm pivot and locknut in the head for the valve to be used, and run the pivot down in the bushing then rotate the camshaft until the nose of the cam lobe for the valve to be used is pointed away from the head. Install a clean, dry and new rocker arm by engaging the valve end of the rocker in the lash pad slot first, then move the rocker under the cam lobe and engage the socket end of the rocker with the rocker adjusting pivot. Adjust the pivot height so the valve is off the seat about 0.002 to 0.003-inch and tighten the pivot locknut a bit tighter than finger-tight. Rotate the camshaft to be certain the cam lobe does not run off wither end of the rocker arm pad. This is easily done by daubing a very thin coat of Prussian blue paste on both flanks of the cam lobe, leaving the rocker pad clean and dry. Rotate the camshaft one full turn by hand and observe the traces of Prussian blue paste that have been transferred to the rocker pad. If the cam lobe runs off the valve stem end of the rocker pad, a thinner lash pad is required, and conversely, a thicker lash pad is required if the cam lobe runs off the pivot end of the rocker pad. It may take a couple of tries with lash pads of different thickness, but the cam lobe contact patch should be pretty well centred on the rocker arm pad, and at the moment, "eye-ball close" is OK. For each attempt, the rocker must be removed and the rocker pad wiped clean of all traces of the Prussian blue paste, and the cam lobe re-blued. Just use a touch of the stuff; a little goes far.

When the cam lobe contact patch is centred on the rocker arm pad and the valve is off the seat 0.002 to 0.003-inch with the centre of the heel of the cam lobe contacting the rocker pad, rotate the indicator dial to zero and lock it in place lightly. Rotate the camshaft by had until maximum valve lift shows on the indicator, log the number and rotate the cam until the indicator shows zero again, remove the rocker arm and number it to coincide with the maximum valve lift figure. Do not move the rocker pivot adjustment of the indicator dial. Now install another rocker arm and repeat the process. I f the indicator needle moves more than about 0.002-inch either direction from zero, reject that rocker and got to another. A wide variation here means that the relationship between the socket, pad and tip of the rocker is inconsistent, but a tolerance of plus-or-minus 0.002-inch is acceptable. Don't forget to re-zero the indicator dial, if necessary, for each rocker arm. Also, rig some kind of a pointer aimed at one specific sprocket tooth when the rocker pad contacts the cam lobe in the centre of the heel of the lobe. Mark the sprocket tooth so the same starting point is used for each rocker arm. At this stage, the actual rocker arm ratio is not important. What is important is to find a full set of rocker, plus a few spares, that are as close as possible to being identical. It is also important to use the same valve, the same retainer, the same valve locks and the same cam lobe when checking rocker arms for consistency. Number all rocker arms in their correct sequence from front to back. If one or more rockers are removed, make certain they are replaced on the same cam lobes. DO NOT mix them up.


Getting the Geometry Right


Observation of the rocker arm geometry is next, but this cannot be done without a cutaway lash pad, like that shown in the photos. Rocker arm geometry can be considered correct when the centre of the rocker arm tip radius coincides with the centreline of the valve stem at exactly half valve lift as previously explained. It may take a small variation in lash pad thickness to make this condition come about, but when it is done. There will be some "free" power (the best kind!) due to the reduction of friction in the valve train, as well as lowering the wear rates of the valve guide bores and valve stems. Unfortunately, the rocker arm geometry must be an "eyeball" check, a condition that not only calls for good light pointed in the right direction, but also for critical examination of the surfaces of the rocker arm tip and the lash pad to be certain their relationship is correct and a low-power magnifying glass will help. For this check, it is safe to assume that the centrelines of the lash pad and the valve stem are coaxial. Because the lash has "ears" on two sides, you may want to do as we did for the accompanying photos. The ears can be ground off and a cutaway retainer helps to visualise what is happening.

A point should be made here related to valve lash pads. We make our lash pads from SAE 52100 steel, the exact same material from which are made the best ball and roller bearing assemblies. Heat treat specification on the lash pads is from 52 to 55 Rockwell "C" scale; hard enough to be very wear-resistant, yet extremely tough without being excessively brittle. After finish machining the pads are 100% Magnaflux inspected. The lash pads furnished with our camshaft assemblies are selected for thickness by (1) having the cam lobe contact patch as close to being centred as possible on the rocker arm pad with zero valve lash for a given cam lobe profile and (2) based on a new cylinder head and new valves.

This means that when the valve seats and valve faces are ground, and the valve lash set properly, the lash pads should be slightly thinner than as-supplied. However, these pads are machineable to some extent. The heat treat process is a "through" heat treat and not a "case". This means that they are within 1 Rockwell "C" scale point at their centres as they are on the outer surface. This gives you something to play with because it is always much easier to remover metal than it is to add it, once too much has been removed. And the use of loose shims in the L-series valve train is a no-no. These lash pads can be machined in a lathe with a tungsten-carbide cutting tool; ain't easy, but it's possible. The best method is to surface grind the flat surface that contacts the valve stem tip. Any decent machine shop will have a surface grinder, but make certain that the grinding operation is done with lost of coolant. If it is done "dry" there is a very good possibility that grinding cracks will be encountered or that the temperature of the pad during grinding will be high enough to anneal (soften) the pad in just the areas where it should be hardest. Of course, we do exchange lash pads if the ones supplied are too thick or too thin, but we must know the correct lash pad thickness requirement. If a large variation in lash pad thickness is required to make the rocker arm geometry right, it is absolutely essential a recheck be made of the cam lobe contact patch on the rocker arm pad to be certain the cam lobe doesn't run off one end of the rocker pad. It if is not possible to satisfy both conditions at the same time, it becomes necessary to concentrate on making the cam lobe contact patch fit correctly on the rocker pad as a primary action but secondarily, with the least amount of compromise to the rocker arm geometry.

Just in case you think this lash pad setup procedure is going to be super simple or fast, consider that the contact area may move completely off the pad with a change of only 0.015-inch in lash pad thickness. If the rocker arms have been selected as described and if the valve stem lengths have been equalised, then the rocker arm geometry and the cam lobe contact patch exercises need only be performed on one valve. If there has been no selectivity in picking out rocker arms and/or the valve tem lengths have not been equalised, then you will have on hell of a time optimising each valve to each rocker arm, and you will very likely end up with a lash pad of different thickness for each valve, a condition that should be studiously avoided.


"Improving" the rocker arm ratio


If there is enough latitude in the contact patch on the rocker pad, there is a way to "cheat" on the rocker arm ratio a bit, as long as the rocker arm geometry isn't loused up too badly in the process. This involves using a thinner lash pad to move the contact patch closer to the socket end of the rocker pad. The contact patch should never be closer than 0.010-inch, and preferably 0.015-inch from the end of the pad in any case, but moving the contact patch closer to the socket end of the rocker does increase the rocker arm ratio slightly. Obviously, this practice shouldn't be carried to extremes because of the detrimental effect on rocker arm geometry.

Early in 1973 some new rocker arms started appearing from the factory, and there is no easy way to identify them from the previous ones. To all intents and purposes they look identical to the previous "two-piece" rockers. But they are not. The centre of the tip radius has been moved away from the pivot (toward the valve). This allows using more rocker ratio and practically eliminates most interference problems between the rocker arms and retainers because the pivot end is raised enough to avoid these conflicts. With the previous rockers you could sometimes fudge the ratio out to 1.165 to 1, whereas the newer ones can be relied on for about 1.5438 to 1-a 1.8% increase - on some types of cams.


Checking Valve Timing and Valve-to-pistons Clearance


Now for a valve timing check and if we're clever and lucky, a piston-to-valve clearance check at the same time. Using number 1 cylinder for convenience, install the intake and exhaust valve assemblies and all related pieces in the cylinder head using a light mechanical fuel pump spring in place if the valve springs for each valve. Engage the camshaft dowel pin in the number 1 hole in the camshaft sprocket to start and tighten the sprocket retaining bolt finger-tight. Set the valve lash to zero. Rotate the camshaft so the dowel pin is vertical, above the cam bolt. Timing mark 1 toward the outer edge of the sprocket will be on the right side when viewed from the front, three teeth above a horizontal plane. In this position, the valves correspond to the TC piston position of the compression stroke for number 1 cylinder with both valves seated. Exact TC for number 1 piston must now be located as closely as possible if it hasn't been done before. This is most accurately determined with a dial indicator reading in "tenths" (1/10,000-inch graduations). The so-called "positive-stop" method of locating top centre is not recommended. The stop usually isn't positive enough, and the piston motion curve is not the same on either side of top centre if the piston pins are offset in the pistons, like stock pistons, so exact top centre will not be exact, and can vary by more than one crankshaft degree. We want it exact. Bolt the graduated crankshaft damper or degree wheel to the nose of the crankshaft or a flywheel with a graduated rim to the back of the crankshaft and run through it enough times with the dial indicator to be certain that it will repeat and that it is dead accurate, then leave the piston at exact top centre and adjust the pointer until it, too, shows exact top centre. With number 1 piston at TC of the compression stroke, the crankshaft key will be vertical on the topside of the crank nose, and the crank sprocket timing mark will be on the right side when viewed from the front, about a half a tooth below a horizontal plane.

Without moving the crankshaft or camshaft, install the cylinder head with a used but usable head gasket of the exact type you'll be using on the final assembly. The gasket is necessary to take up its own space; otherwise the chain tensioner will retard the camshaft when it very likely shouldn't be. Torque the cylinder head bolts in steps to no more than 50 pounds-feet. Remove the camshaft sprocket; install the timing chain and lotsa luck! This long thing has more ways of getting caught, hung up, entangled, etc., than you'd believe. Now there are two bright (or marked) links (count 'em) on the short side of the chain. Holding the chain vertically, the bright (or marked) links must be on the right side and must face forward when viewing the chain and engine from the front. The lower bright (or marked) links must be on the right side and must face forward when viewing the chain and engine from the front. The lower bright (or marked) link must be engaged with the crank sprocket tooth that has the timing, mark while the top bright (or marked) link must engage the tooth for the number 1 timing mark on the camshaft sprocket. When this mess is sorted out, reinstall the camshaft sprocket and make certain the camshaft dowel pin is engaged in the pinhole 1 in the sprocket, then tighten the sprocket bolt fairly snugly. Without moving anything double-check the chain installation to be certain the bright (or marked) links are engaged with the proper sprocket teeth, at both the crank and at the cam. This is no time to be off a tooth or more in either direction. One tooth is 18 crankshaft degrees and the broadest possible tolerance isn't broad enough to cover an error of that magnitude.

If everything is OK, or seems that way, mount the dial indicator on the spring retainer of number 1 intake valve, make certain the indicator spindle is parallel to the valve stem in both planes, and pre-load the indicator to something more than maximum valve lift. Rotate the crankshaft slowly and gently in the normal direction of rotation (clockwise from the front). If there is any above-normal resistance to rotation no matter how slight, S-T-O-P!! And investigate the cause. Aside from the piston striking the cylinder head and/or gasket, about the only other possibility is the piston tagging a valve, so move both valves by pushing down on the spring retainers. Now you see one reason for using light springs instead of valve springs.

Assuming there is no conflict (yet), rotate the crankshaft until the dial indicator shows a maximum valve lift and make a mark on the degree wheel, damper or flywheel at which the pointer is aimed. Rotate the crankshaft exactly one full revolution, right back to the same mark, zero the indicator dial. Then rotate the crankshaft slowly in the same direction until the indicator shows a valve lift of 0.025-inch, or whatever the check height for valve timing may be. This will be the intake valve opening point and, unless the camshaft is of extremely short effective duration, this will occur at some point before TC. Read the degree wheel and log the number. Again assuming there have been no collisions between piston and valves, continue to rotate the crankshaft, observing the cam lobe and rocker arm, past maximum lift and down the closing flank of the cam until the dial indicator shows that the valve is 0.025-inch from zero or the correct check height. This indicates intake valve closing point and it will occur at some point past BC. Read the degree wheel, log the number then add the intake valve opening point to the intake valve closing point plus 180 degrees to arrive at the intake valve duration. Compare the valve opening and closing points to those shown on the cam's timing card. The duration should be right there; just what the card says, within plus-or-minus one crankshaft degree. There may be some disagreement in the valve opening and closing figures, and the comparison will give you a pretty good idea how much and in which direction to go to correct the valve timing. Run through the procedure enough times to be certain the figures repeat then transfer the dial indicator to the exhaust valve and go through the operation again. When the intake opening and closing points and the exhaust opening and closing points have been determined, then you have a pretty good fix for amount and direction for any corrective action that must be taken to make the valve timing fall in place.

If, during the initial checkout, there proves to be piston-to-valve contact, simply remove the piston and connecting rod assembly from the engine, after exact TC has been located. The piston isn't really necessary for a valve timing check, or for subsequent valve timing corrections. The normal course of events indicates that the valve timing should be checked first; next correct the valve timing, if a correction is required; then the pistons can be notched after the valve timing has been established.


Chapter Thirteen


How to Make Timing Corrections


Assume the specified duration of a camshaft is 282 degrees and also assume that the specified valve timing is: Intake opens 35 degrees BTC, closes 67 degrees ABC. Exhaust opens 71 degrees BBC, closes 31 degrees ATC. Add the intake opening point (35 degree BTC) to the intake closing point (67 degree ABC), plus 180 degree between TC and BC and, sure enough, the duration is 282 degree. Similarly, by adding the exhaust opening point (71 degree BBC) to the exhaust closing point (31 degree ATC) plus the 180-degree between TC and BC and again, the duration is 282 degree. Observation of the specified timing shows the camshaft is advanced 1 camshaft degree (2 crank degrees) from a "split" timing condition which, in this case, would be 33-69 intake and 69-33 exhaust. Putting into practice the earlier lesson of displacement angle by adding the intake opening point (35 degree BTC) to the exhaust closing point (31 degree ATC). The overlap period is 66 degree which, when subtracted from the 282 degree duration and divided by 2, equals a displacement angle of 108 camshaft degrees.

In the engine, suppose the intake valve timing comes out 31-71 and the exhaust comes out 67-35. The duration is sill exactly the same for both intake and exhaust 282 degree; but it can be seen that the camshaft is retarded 4 crankshaft degrees from where it should be (2 camshaft degrees). Swell. How do we fix it to come out correctly? This one is almost too easy. Earlier it was mentioned that the camshaft sprocket has 3 numbered dowel pinholes and 3 correspondingly numbered timing marks. Normally, with a new timing chain and sprocket set, the cam dowel pin should be located in the hole 1 and the top bright (or marked) timing chain link should be engaged on the cam sprocket tooth 1 timing mark. Get the bright (or marked) chain links lined up just the way they were during initial assembly, that is, with the piston at exact top centre of the compression stroke; use the wooden wedge to separate the timing chain links and to keep the chain tensioner piston and spring from falling out. Remove the cam sprocket and engage the camshaft dowel pin in hole 2 and rotate the cam until the dowel pin is again vertical, above the cam. This will place the timing mark 2 near the rim of the sprocket so the top bright (or marked) chain link can be engaged on sprocket tooth 2. This advances the camshaft approximately 4 crankshaft degrees. Replace the cam sprocket bolt, remove the wooden wedge and run through the valve timing exercise again. With the camshaft timing exercise again. With the camshaft advanced 4-crankshaft degrees, the intake valve timing should be 35-67 and the exhaust timing should be 71-31. By engaging the camshaft dowel pin in the hole 3 in the cam sprocket and by engaging the top bright (or marked) chain link on sprocket tooth 3, the camshaft will be advanced another 4 crankshaft degree increments, but no adjustment to retard it except normal chain and sprocket wear.

But suppose the valve timing is too far advanced in the first place. For example, say the intake valve timing comes out 39-63 and the exhaust comes out 75-27. This shows the valve timings is advanced by 4 crankshaft degrees from where it should be. With no stock adjustment to retards the valve timing and after-market offset (eccentric) camshaft sprocket bushing must be used. These are available in various graduations. Installation involves enlarging one of the cam dowel pinholes to about 0.001-inch smaller than the bushing O.D. The bushing works both ways; it can be used to advance or retard the camshaft the same amount in either direction simply by reversing the offset side of the bushing to face the opposite direction. When the cam sprocket dowel pin hole has been enlarged to be a slight interference fit with the offset bushing, the bushing should first be located on the camshaft dowel pin, then the sprocket should be moved around on the nose of the cam until the bushing is started in the enlarged dowel pin hole. Use a soft mallet to tap the sprocket over the bushing until the sprocket is flush against the nose of the camshaft.

Just make absolutely bloody certain that the offset side of the bushing points in the right direction. Too many engines have been broken or severely wounded because and offset bushing was installed backwards. Looking at the front of the cam with the dowel pin vertical above the bolthole, the thin (offset) side of the bushing must point toward the left to retard the camshaft. Conversely, the thin (offset) side of the bushing must be pointed toward the right to advance the camshaft. It is also advisable to determine if the bushings are marked or coded in crankshaft degrees or camshaft degrees. It does make a difference: 5 camshaft degrees equals 10 crankshaft degrees. Ian any case, after the offset bushing installation, it is essential that the valve timing exercise be put to practice again to be certain the direction and amount of valve timing correction are both right. In this area as in others, the very best practice is to assume nothing and to find out all details for yourself. This is the only way to gain positive knowledge; otherwise you'll expend a lot of energy in pure guesswork.

Any set of valve timing figures can be substituted for the ones shown and the same practices applied to determine how the timing card figures can be made to come out as closely as possible to the valve timing figures actually obtained in the engine.




Now about those pistons. For some strange and mysterious reason, nearly everyone associates maximum valve lift with a panic-stricken thought of the need to gouge great holes in the pistons for piston-to-valve clearance. Ridiculous! With most L-series Datsun camshafts, maximum intake valve lift usually occurs from about 102 to about 110 degrees ATC, and maximum exhaust valve lift usually occurs at about the same number of degrees BTC. Now how in the wide world can a valve get tangled up with a piston when the piston is that far down the cylinder bore, out of reach of the valve? Obvious answer: It can't. The problem of piston-to-valve interference simply does not exist at maximum valve lift if the valve timing is correct. The problem occurs much earlier in the cycle with the intake valve and mush later in the cycle with the exhaust valve. Consider: Tracing the piston past TC of the compression stroke and the power stroke starts. Before the piston reaches BC of the power stroke, the exhaust valve starts to open and continues to do so until maximum valve lift is reached, then it starts to close. Meanwhile, the piston has passed BC and is heading toward TC again on the exhaust stroke. At a point before TC, the intake valve starts to open, and for some period of crankshaft rotation, and therefore piston travel, the intake valve and the piston are on a direct collision course. The only thing that prevents a catastrophe is that the piston slows down as it approaches top centre, meanwhile the intake valve continues to open, and the exhaust valve is still closing. As the piston passes TC of the exhaust stroke and heads toward BC on the induction stroke, first the exhaust valve closes at some point past TC while the intake valve is still opening. Later, but before the piston reaches BC, the intake valve reaches the point of maximum lift and starts to close. As the piston passes BC of the induction stroke and starts toward TC on the compression stroke, the intake valve closes at some point past BC. Both valves will stay closed for the remainder of the compression stroke and for the larger part of the power stroke, when the exhaust valve starts to open again to repeat the cycle.

So when are the valves closest to the piston? During the valve overlap period when both valves are open toward the end of the exhaust stroke and the beginning of the induction stroke. Before TC of the exhaust stroke, the piston is sort of "chasing" the exhaust valve closed, but the intake valve is opening directly toward the advancing piston. Therefore, with all but some freak camshafts, the intake valve comes closest to the piston at a point after TC of the induction stroke, and the exhaust valve comes closest to the piston at a point before TC of the exhaust stroke. Let's assume the three worst conditions:

1. Actual piston-to-valve contact.

2. Required depth of valve reliefs in the piston is unknown.

3. Correct location of the valve reliefs in the piston is unknown.

These conditions could be encountered in an engine with stock flattop pistons, and these pistons can indeed be machined for valve reliefs; not the excessively - but some - and probably more than you'd expect. An after market piston may be necessary to satisfy the piston-to-valve clearance requirement. Most of these are forgings equipped with valve reliefs, which can be sunk considerably deeper because the piston crown has additional material thickness.

It must be carefully noted and logged which valve touches the piston and at what point. Observe the degree wheel carefully to determine location, at the same time rotating the crankshaft slowly and pressing down on each rocker arm until contact is made. Assuming both valves tag the piston, the exhaust valve will make contact first at a point BTC on the exhaust stroke. When this point has been located and logged, simply remove the exhaust rocker arm and continue to rotate the crankshaft until the intake valve contacts the piston at a point ATC on the induction stroke and log this point too. This will give two starting reference points, which will change after the piston has been relieved. But for the purpose of actually relieving the piston, this must be done at exact TC.

In any case, number 1 piston must be installed correctly with the piston pin offset (if there is an offset) pointed in the right direction. With stock pistons, the offset side of the piston must be installed on the LEFT side of the engine when viewed from the front. Rotate the crankshaft until the piston is at exact top centre. Remove the number 1 intake and exhaust valve assemblies from the cylinder head. Replace the valves with a pair of junk valves, preferably about 1/8-inch larger in diameter than the service valve. The condition of these valves is unimportant as long as they are not bent. Hold them in place with pieces of masking tape around the stems close to the tops of the valve guides, then place the cylinder had on the block, again with a used but usable head gasket, install the cylinder head bolts around number 1 cylinder and torque them down fairly snugly. If wither or both valve heads are so large they contact the cylinder bore before they touch the top of the piston, grind a flat on the valve heads so they clear the cylinder bore but keep the valves oriented so the lower segment of the valves (the area closest to the piston) are still round. Double-check the degree wheel to be certain the crank hasn't moved. Remove the masking tape from the intake valve stem and let it drop down the valve guide until the valve head contacts the piston top. Give the valve stem tip a whack or two with the hammer; hard enough to leave a visible crescent-shaped indentation in the top of the piston, but not hard enough to drive the valve through the piston, or damage the piston in any way. If the valve stem tip drops too far down the guide bore to reach the tip with a hammer, use a ¼-inch flat-faced pin punch to reach the valve stem tip, then whack the punch with a hammer. Repeat the process with the exhaust valve, raise the valve until they are seated, with the help of a small screwdriver poked through the spark plug hole if necessary, and re-tape the stems to hold the valves in place. Rotate the crank until the piston is about an inch down the bore, then insert a small medical probe light through the spark plug hole and observe your hammer-work. The crescent-shaped indentations must be sufficiently well defined to establish positive location of the required valve reliefs on the piston top. If they aren't, rotate the crank until the piston is at exact TC again, then have another flailing session with the hammer. It helps if the O.D. of the valve head is sharp so less effort is required to leave visible marks, thereby reducing the possibility of damaging the piston. If all is well, remove the cylinder head and the valves, remove the piston and connecting rod assembly and remove the piston pin from the piston. Now remove yourself to a friendly machine shop equipped with a Bridgeport or similar vertical mill and decent clamping devices so the pistons can be held firmly in the ring belt area without distorting or collapsing them. If the valves used to mark the piston were 1/8-inch larger in diameter than the service valves, the cutters that will form the valve reliefs in the pistons should be the same diameter and with a radius of 0.060-inch at the outer edges. Initially, the valves used to mark the pistons can be used in the mill spindle to align the marks in the piston with the mill spindle, if the valves weren't bent during the hammer exercise. If stock pistons, or other pistons with offset piston pin bores are to be modified with valve reliefs, the pistons MUST be oriented in the clamping fixture from the piston pin bores to prevent gouging the pistons on the wrong side. You don't need a spare set of ashtrays. The angle formed by the intersection of the centrelines of the valve stems and the cylinder bores is nominally 12-degrees, so either the clamping fixture or the mill spindle must be at this angle and locked in place. The angle may vary slightly from one cylinder head to another, but the 12-degree number is the place to start.

The piston crown of a stock piston is about 5/16-inch thick near the centre of the piston and slightly thicker toward the outer edge. This means that the stock pistons can be safely machined for valve reliefs to a depth of 0.125-inch from the top surface of the piston crown IF and ONLY if there is a radius of at least 0.060-inch at the bottom of each relief. This leaves a nominal crown thickness of about 3/16-inch at the thinnest point which, of course, is at the bottom of the relief. The crown thickness increases as the relief approaches the top surface of the piston. Normally, the remaining 3/16-inch crown thickness is adequate for any application involving the use of stock pistons. It must be pointed out that the 5/16-inch total crown thickness is an average figure; some pistons may be thinner in this area than others, so it is certainly advisable to make some careful measurements before arbitrarily gouging the pistons 1/8-inch and finding out later (the hard way) that the 1/8-inch was too much.

The purposes for using 1/8-inch larger diameter valves for marking the piston should be apparent. First, this practice permits a radius of 0.060-inch in each valve relief and this radius should be considered minimum and it will normally prevent the service valves from tagging any part of the radius. If a larger diameter cutter is used to form the valve reliefs, the corner radius of the cutter should be increased proportionally. Example: If a cutter ¼-inch larger in diameter is used to machine the valve reliefs, then increase the cutter's corner radius from 0.060-inch to 0.120-inch. The corner radius in the valve reliefs is highly important because it reduces the possibility of generating stress risers in the pistons, particularly at the higher temperatures under which the piston must function. Sharp corners in the piston reliefs are O-U-T, so don't even think about it. The larger the corner radius can be made, the better, as long as the edge of the valve doesn't interfere with the radius. The seemingly larger-than-maybe-really-necessary size valve reliefs also provide extra room around the vales for improved breathing. Remember? Lots of room around the valves? In addition, the oversize reliefs allow for slight variations of the location of the valves in the cylinder head in relation to the reliefs in the pistons. Some after-market pistons leave some doubts about the accuracy of valve relief location, size and depth, so the mark 'em and machine 'em process should be conveniently applied to these as well.

If you're a confirmed do-it-yourselfer and you want to save a few bucks and time is no object, the same results can be accomplished in the following manner. With a short length of air-hardening tool steel 3/16-inch square, grind it to size and shape to form a cutting tool, with a rake angle, relief angle, corner radius, etc. When this has been carefully done, finish off the cutting edges with a sharpening stone, then silver-solder the cutter across the head of a junk but straight valve with the radiused end overhanging the valve head enough so the service valve will clear the radius. Only one cutting edge is needed, from the centre of the valve to the end of the cutter, so the end opposite the cutting edge must be relieved so it cannot contact the piston. Silver soldering should bring the cutter to critical temperature, then let it cool in air and it will be hard enough for this operation. Two of these cutters are required; on for intake valve reliefs, on for exhaust valve reliefs. The cutting edge itself should be located as close to the centre of the valve head as you can eyeball it, otherwise the surface finish of the reliefs may no be as smooth as it should. Install all pistons, making certain the offset piston pin bores are pointed in the right direction. Bring number 1 piston to exact TC, install one cutter in its respective valve guide, orient the cutter to the spark plug side of the combustion chamber and gently drop the cylinder head in place, but don't bolt it down yet. Rotate the cutter stem by hand to see that it clears the piston. If the cutter hangs up on the edge of the cylinder bore, you're out of luck; the valve reliefs must be machined in the pistons with the pistons removed from the cylinder block. Usually, this will only occur with the largest valve head diameters, but may occur with the next-to-largest valve head sizes as well. If the cutter clears the bore edge, bolt the head in place.

Be careful here. You don't want an intake valve-sized relief for an exhaust valve. Or, much unfunnier, you don't want an exhaust valve -sized relief for an intake valve; the intake valve ain't gonna fit. And don't try to use both cutters at the same time; they overlap each other, or should if they're the right sizes.

If there are no entanglements, mark a small spot on the valve spring seat in the cylinder head. Rotate the cutter until the valve stem tip sticks as far out of the valve guide bore as it is going to. Use a depth micrometer to make this measurement from the valve stem tip to the mark in the valve spring pocket. This is not the time for yardstick measurements or guesstimates. Verify the cutter position by peeking through the spark plug hole.


Chapter Thirteen


Piston modification for valve clearance (cont.)


The cutting edge should contact the piston on the far side of the piston and in a plane perpendicular to the crankshaft axes. If all's well, chuck the valve stem in a 3/8-inch or larger electric hand drill, flood the valve stem guide bore with lubricating oil, fire up the frill and take a light cut, no more than 0.030 to 0.040-inch. Release the chuck from the valve stem, verify this cutter position (the cutter must be in the newly formed valve relief), then verify the depth of cut with the depth micrometer, subtracting the last figure from the first one. When making the cut, don't force the drill down; this is an interrupted cut and the rate of feed must be very slow. Frequent stops are necessary to re-flood the valve guide bores with lubricant; you don't need to wear out the valve guide bores before the engine is fired up.

Assuming there was piston-to-valve contact initially, the first light cut won't be deep enough, so proceed slowly and cautiously until the relief depth is 0.085 to 0.090-inch. Of course, you won't have much choice because it's a very slow process. Don't move the piston while the cutting operation is in what we laughingly call progress because some of the chips you've made are bound to get caught between the piston and the cylinder bore and the piston will become scored. Remove the head once in a while to blow the chips out, admire your handiwork and to touch up the cutting tool with a sharpening stone. Observe the quality of the surface finish in the machined area. If it's shaggy, the tool needs sharpening and perhaps the rake and relief angles of the tool should be changed. Chatter marks are another story; they can be downright obstinate with a single-lip-cutting tool and an interrupted cut. If things look good, trade cutters and do the same thing for the other valve in the same cylinder, and to the same 0.085 to 0.090-inch depth.

Now for a piston-to-valve clearance check. Remove the head, blow out the chips, clean the gasket face of the head and block and both sides of the head gasket. Remove the piston to clean out the rest of the chips, reinstall the piston but make certain there are no chips on the crankpin or the rod bearing halves or on the mating faces of the rod and cap. Remove the cutting tool from the head; install the service valve assemblies, again using the fuel pump springs in place of the valve springs. Install the head gasket and head and torque the head bolts down fairly snugly. Install the timing chain and cam sprocket and make certain the camshaft dowel pin is engaged in the correct pin hole in the sprocket and that the sprocket pin hole number corresponds with the sprocket tooth number and finally, that the bright (or marked) timing chain links straddle the correct sprocket teeth at the crank and at the cam. Tighten the cam sprocket-retaining bolt fairly snugly. Install the valve lash pads and the rocker arms, set the valve lash to zero and mount the dial indicator on the spring retainer of number 1 intake valve as previously described. Remove the wooden chain wedge. Crank rotation gets difficult with this thing in place. Now, for experience, exercise and to be certain nothing has changed, run through a valve timing check on the intake valve. Rotate the crank and as the piston passes top centre of the induction stroke, go very slowly indeed, particularly if there was piston-to-intake-valve contact initially. If any but normal resistance to crank rotation is encountered, S-T-O-P!! Try to open the intake valve further by pushing down on the rocker arm and watch the dial indicator. If there is no indicator motion, rotate the crank backwards a few degrees and try again, then sneak up on it slowly by rotating the crank in the normal direction and pushing down gently but firmly on the intake rocker arm, all the while watching the dial indicator. When this instrument shows there is but 0.001 to 0.002-inch piston-to-valve clearance, stop and reconsider.

It gets iffy. If the valve timing is correct, if the intake valve relief is properly located and the correct size, if the relief is from 0.085 to 0.090-inch in depth, if actual piston-to-intake valve contact still exists and if stock pistons are being used, one factor announces its ugly self loudly and clearly: Either the stock pistons or the camshaft are (is) O-U-T for this particular application. You'll puncture the piston crown in two places before you can possibly gain adequate piston-to-valve clearance. It's your choice and your decision. If it's a budget operation and/or street or dual-purpose vehicle, or a strange rule in a strange rulebook, then there's not much of an alternative. Keep the stock pistons and exchange the camshaft for one shorter in effective duration, or less intense in valve action, or both. Most camshaft manufacturers will exchange the camshaft and related pieces at little or no cost except shipping charges, provided all pieces show that they have not been operative in a functioning engine. Or go with an after-market piston. These may come equipped with valve reliefs of sufficient depth to put an end to the problem right there. If not, you're up the same creek without the same paddle - but in a different canoe, where there is some piston-to-valve clearance but not enough to be safe. The advantages of most after-market pistons are that they have fairly deep valve reliefs in the first place and, due to additional crown thickness, the reliefs can be readily made deeper safely.

So let's back up a few notches to cover another possibility. Assume there was no initial piston-to-valve contact, but whatever clearance exists may not be enough. After the valve timing has been checked and corrected as required, start with number 1 piston at top centre of the compression stroke and the dial indicator mounted on number 1 intake spring retainer pre-loaded to something more than full valve lift, and with zero valve lash, and with the indicator dial zeroed with the cam lobe in the centre of the heel of the cam lobe. Rotate the crank about 350 degree or until the piston is about 10 degree closer to top centre and push down on the rocker again. Continue to do this until the piston-to-intake valve clearance is about 0.100-inch, then reduce the crank rotation increments to 5 degrees, depressing the rocker arm until it contacts the piston at each stop. Someplace along the line, the piston-to-intake valve clearance each time, as shown by the indicator. At some point, the piston will be as close to the valve as it will get. At this point, push the rocker arm down until the valve contacts the piston three or four times, watching the indicator closely to be certain there is no error. Release the rocker arm, zero the indicator dial, push the rocker arm down three or four more times and read the piston-to-valve clearance directly on the indicator from the zero point. Let's say the minimum piston-to-intake valve clearance is 0.065-inch. To obtain the required minimum of 0.0900-inch, the intake valve relief must be made 0.025-inch deeper than it is. No sweat with most after-market pistons. There is usually no problem with a stock piston that has been previously modified with a valve relief from 0.085 to 0.090-inch deep. Much more and things get shaky, but this is a reasonable figure. It's best not to consider the valve lash when making these measurements; the lash may be very valuable sometime as a relatively slight but extra "cushion" against engine overspeed damage.

All ya gotta do now is - the same thing with the exhaust valve, but allow a minimum of 0.100-inch piston-to-exhaust valve clearance. The extra 0.010-inch clearance from the intake to the exhaust valve doesn't seem like much, and sometimes it may not be enough, but as previously explained, the camshaft retards itself as a function of engine speed and also as a function of normal attrition of the sprockets and chain. And as the camshaft retards itself, gradually, abruptly, or whatever, the piston-to-exhaust valve clearance decreases while the piston-to-intake valve clearance increases.

The piston casting or forging permitting, the intake and exhaust valve reliefs should be joined in a straight line, eliminating any "scallops" between the reliefs to improve breathing during the overlap period at very high engine speeds. This is not advisable with stock pistons relieved to the maximum (about 0.125-inch) depth. Finally, a radius should be ground at the intersection of the top of the piston and the deep side of each valve relief.

When you do-it-yourselfers get sick and tired of doing it yourself, send the) after-market only) pistons back to the manufacturer with explicit instructions regarding the required corrections to the valve relief depth, diameter, bottom corner radius, top corner radius, etc. They may not like it, but they will probably do it for a nominal (?) charge. If the manufacturer gives you any romance, bundle the pistons up and send them to: Smith Brothers' Manufacturing Company (The guys who make pushrods. You remember pushrods?), c/o Hank Smith, 1201 North Azusa Canyon Road, West Covina, Calif. 91790 (213) 338-8026). Give Hank the same instructions and he'll do it quickly, accurately and usually for no more than about $5.00 per piston. The Smiths will also do it to stock pistons but you will have to supply tem with a sample marked piston, the valves used for marking the piston and their relationship in diameter to the service valves. Be certain to mark each piston correctly "front" so the offset piston pin bore won't be reversed. Include the required depth of the valve reliefs from the top of the piston, a plan and cross-sectional sketch of the desired reliefs, the valve-to-cylinder bore angle of 12 degree and last but highly important, a sketch of the valve sequence down the cylinder had from front to back there is a valve sequence reversal in the centre of all L-series cylinder heads. It's a good plan to number all pistons and refer to intake or exhaust reliefs with a felt-marking pen on the top of each piston. After the pistons have been reworked, it is essential that they be rechecked for piston-to-both valve clearances, just to be certain no one goofed. By now you probably wish you had a zipper on the engine. It's a drag but a necessary one. This time, the only difference is that the piston-to-valve clearance check must be performed for all pistons and all valves, and in the manner described with the dial indicator to give a direct reading. A "plus" tolerance is fine, and a "minus" tolerance, as long as it doesn't exceed about 0.005-inch, is acceptable.




With things lookin' good, the lower half of the engine can be assembled and buttoned up. Next, the cylinder head can be assembled, this time using the real valve springs and service valves, lash pads, the correctly numbered rocker arms for their respective valves and the "mousetrap" rocker arm springs.

Now consider the camshaft sprocket-retaining bolt. These things can behave very badly at times by working themselves loose, or falling out altogether, in either case allowing the camshaft sprocket to slip off the nose of the cam. At this point, your world will be filled with many instant problems. In the case of the L-16 and L-18 engines, the mystery of the self-loosening bolt is caused by secondary vibrations inherent in any four cylinder engine configuration. With the L-24, the same effect is caused more by torsional oscillations of the longer six-cylinder crankshaft, transmitted from the crankshaft to the camshaft by the timing chain. A few precautions taken here may save an engine. First, scrub out the threaded hole in the nose of the camshaft with lacquer thinner and a bristle brush and blow it clean and dry with compressed air. Ditto for the bolt itself. Then make certain the bolt doesn't "bottom" in the threaded hole. Place a new split-type lock-washer against the head of the bolt, followed by the fuel pump eccentric-flat washer assembly. Put a few drops of Loctite along the threads of the bolt and make sure it is evenly distributed. Install the bolt and tighten it to 70 pounds-feet of torque. The book says 43-1/2 pounds-feet, but 70 is a better number. That should hold it, but recheck the torque occasionally just to be certain it isn't in the process of working loose.

If the stock fuel pump is to be discarded in favour of a high-volume electric fuel pump for a race engine, a more positive fix can be made. In this case, the fuel pump eccentric is not necessary and should be left off, replacing it with a steel flat washer of about the same thickness. Do everything outlined above, except drill the bolt head through to accept a strand of 0.060-inch diameter stainless steel safety wire. Don't forget the lock-washer and the: Loctite." Tighten the bolt to 70 pounds-feet of torque, then safety wire the bolt head tightly and in approved aircraft fashion to one of the spokes in the cam sprocket in such a way that the bolt cannot work loose. Then make a plate to plug up the fuel pump opening in the front engine cover and bolt it in place with a fuel pump gasket between the plate and front cover to prevent and oil leak at this point.

When the cylinder head assembly is completed, install the head with a new cylinder head gasket and (preferably) new cylinder head bolts that have been Magnaflux inspected. Needless to say, the crankshaft and camshaft must first be positioned correctly, and you'll have to fight the battle of the chain links and sprockets again. The boltholes in the block should be cleaned with the correct bottoming tap and the holes blown clean and dry with compressed air. Install new hardened flat washers between the cylinder head and the bolt heads.

Just to keep in practice, and to be sure the chain and sprockets were installed correctly, and to find out what has been lost by a full valve spring load, run through a valve timing check again, hopefully for the last time before the Great Day of the first engine fire-up. You will probably lose a couple of degrees in effective duration and the camshaft may show as being slightly retarded from the last valve timing check figures. If it's a slight amount, don't worry about it; you will no doubt have to advance the camshaft soon enough. If it's a large amount in either direction, something was installed incorrectly and the reason for the discrepancy must be tracked down and corrected before going any further.


Chapter Fifteen




Now for the valve lash. There is a very good way of doing this and the first time takes the longest, but subsequent valve lash adjustments can be made quickly, easily and accurately. First, find maximum valve lift of number 1 intake valve with the dial indicator mounted on the valve spring retainer, then make a permanent mark on the crankshaft damper at the spot on the rim where the pointer is aimed, using a narrow stripe of nail polish or contrasting paint. Now rotate the crankshaft exactly one full revolution, right back to the same mark. This will place the rocker arm pad in contact with the centre of the heel of the cam lobe of number intake valve. Repeat the process for number 1 exhaust valve but use a different color of nail polish or paint. Then follow the firing order and do the same to the remaining valves. With L-16 and L-18 engines, you will have two stripes for intakes and three for exhausts because each mark will repeat itself at the crankshaft but no at the camshaft due to the 2-to-1 reduction between the crank and cam. For subsequent valve lash adjustments, start with number 1 cylinder and simply follow the marks on the crankshaft damper through the firing order until the valve lash has been adjusted on all valves. It isn't necessary to use the dial indicator again, but it is necessary initially to be certain the stripes on the damper are correctly located. Advancing or retarding the camshaft a few degrees will have no effect, so the original stripes on the damper can be used. "Cold" valve lash, that is with the engine assembly at room temperature, should be set from 0.001 to 0.002-inch tighter than the specified "hot" valve lash. Hot lash adjustments must be made with the engine good and hot with the engine coolant and oil temperatures stabilised at or near their maximum levels. DON'T attempt a hot valve lash adjustment with the engine running unless the idea of drowning in hot engine oil has some masochistic appeal. Because of this, the hot valve lash adjustment must be made quickly, before the engine temperature level has had time to change significantly.

Any valve lash adjustment, hot or cold must be made carefully and accurately with feel gauge blades flat and smooth on both sides. Normally, lash adjustments are made between the cam lobe and the rocker arm pad; therefore any errors at this point will be multiplied as a function of rocker arm ratio at the valve end of the rocker arm. A more accurate method is to measure the valve lash between the rocker arm tip and the valve lash pad, but his means narrowing standard width feel gauge blades down to no more than 3/8-inch wide so the blades will fit in the lash pad slot. If this is done, the rocker arm ratio must be taken into account. If the specified valve lash is 0.012-inch (intake) and 0.014-inch (exhaust). This is the preferred method because it is consistently more accurate and also because any errors in adjustment are not subject to multiplication.




Presumably all systems are now functional and "Go." So go! Gently at first to be certain that everything is right. Get the air/fuel mixture sorted out ("lean and clean") and this includes the idle mixture, accelerator pumps (if any), etc. Usually, a relatively mild camshaft will not require a change in carburetor main metering jets or metering rods. Not too much total spark advance to start with, particularly in an engine with a fairly high compression ratio: 32 to 34 crankshaft degrees total spark advance is considered safe and conservative with an open exhaust system, and 36 crank degrees total advance is about optimum, but "optimum" will vary slightly with an individual engine. Avoid detonation as if it were and epidemic of bubonic plaque. Remember that total spark advance alone does not win races; very likely it's the other way around. At his stage, the wiser course is conservatism.




An accurate fuel pressure gauge, mounted where it is easily visible, can be a valuable instrument in determining if the fuel delivery system is adequate, particularly at the top end of the engine speed range in the higher gears when the load on the engine is heavier and the time the engine is under the heavier load is longer. If the fuel pressure diminishes toward zero, don't tempt fate; shut the engine down, junk the existing fuel pump and install fuel delivery system that is capable of handling more than the required volume of fuel when the engine is in the speed range of maximum power or beyond, and this means fuel lines of adequate diameter and a fuel pressure regulator capacity well in excess of the anticipated maximum fuel flow requirement. Hitachi (S.U.). Mikuno (Solex) and Weber are all pressure-sensitive carburetors, and this is not simply a game of fuel pressure. When the volume of fuel delivered to the carburetors is at least adequate, or more than adequate, which is the preferred state, any fluctuation of fuel pressure will be minimal. When the volume of fuel delivered to the carburetors is not adequate, then the fuel pressure will drop toward zero and the carburetors will eventually run out of fuel, possibly with catastrophic results. And the instrument that can tell you if things in this area are right or wrong is a good, accurate fuel pressure gauge.

The fuel pressure gauge should be hooked into the main fuel delivery line as close as possible to the carburetors and between the carburetors and the fuel pressure regulator. Some sanctioning groups (National Hot Road Association, for one) will not allow fuel lines in the cockpit. This makes sense, but it means that the fuel pressure gauge must be mounted on the firewall in the engine compartment, and the hood notched out for clearance. Other race sanctioning groups (NASCAR, for one) will not permit electric fuel pumps. This makes sense, too, because in the event of an accident, an unconscious or semi-conscious driver cannot be expected to fumble about for a fuel pump and/or ignition switch. So know the rules of your group and build a fuel delivery system with more than adequate flow capacity around them.

So what has all this got to do with camshafts and valve train pieces? LOTS! I wish I had a nickel for each time I've tried to tell some dumbhead that his fuel delivery system, such as it is, is very likely worst than a worn-out stocker because there is no way it can deliver the volume of fuel to the carburetors. Sure! Lotsa pressure! Enough to sink the floats of all carburetors to the bottoms of their float bowls forevermore, but not enough fuel flow to move the vehicle from point A to point B without running the carburetors dry. But the camshaft is really to blame. Not top end power. Sure.

With the fuel delivery system right, you should be able to literally drown the engine with a sloppy rich, dripping air/fuel mixture at the very top end by nothing more than two or three sizes larger main metering jets. This should tell you

(1) that such a condition can be accomplished with nothing more than a change of jets and

(2) that the fuel delivery system is probably adequate, if marginal.




Now for some minor tuning data. The L-series Datsun engines like to run with a sustained engine coolant temperature of about 200 degrees F.; fairly warm, like other liquid-cooled engines with aluminium cylinder heads. Engine lubricating oil temperature should be slightly less, say between 180 and 190 degrees F., otherwise an oil cooler, larger sump capacity, or both may be required. Assuming normal combustion condition, engine oil temperature is a better indication of power output than engine coolant temperature, but this is only true to a point, so don't get carried away and to a point, so don't get carried away and try to boil the oil. It isn't advisable to tune the engine around the coldest spark plug available because when the time comes that you really need a colder plug, you haven't any place to go. Besides, the engine will be sharper, crisper and more responsive with plugs that are one or two steps hotter, assuming the air/fuel mixture, total spark advance, etc., precludes any possibility of detonation and/or pre-ignition. A fairly cold plug is called for with an engine to be run on a dynamometer, where the same engine in a drag race vehicle would use a fairly warm plug, and a road race or circle track engine should use a plug somewhere between the other two.

About the only way and engine normally knows how to respond to a relatively rich air/fuel mixture condition is to demand lots of spark advance and a fairly hot spark plug. A lot of people in the world firmly believe that in order to get a power increase from an engine, it is necessary to really pour the fuel to it, and that it can't produce power without being dead rich. Their thought processes are confused. Obviously, if an engine does show a power increase, it will consume more fuel, but the specific fuel consumption will usually remain in the same range, if everything is clean and sanitary during engine operation. Brake specific fuel consumption, expressed in pounds of fuel per brake horsepower per hour, usually runs from about 0.50 to about 0.52 (pounds of fuel/brake horsepower/hour), based on gasoline as the fuel, with the best engines having the lower specific consumption because they make better use of the fuel. This roughly represents and air/fuel ratio of from about 12.5 to 1 to about 13.0 to 1. A not-so-obvious point is inherent in most carburetors. This is the fact that most carburetors are capable of handling the additional air and fuel required for moderate power increases, but at some point, the air/fuel mixture ratio will become progressively richer because the additional air flow demand by the engine is met with a more-than-optimum fuel supply from the carburetor. It is therefore entirely possible that a moderate power increase may be accompanied by the requirement for a leaner air/fuel mixture, particularly near the top end of the engine speed range. So why try to drown an engine with a surplus of fuel that it can't use and doesn't want? However, if an air scoop is part of the induction system with the intent of supplying relatively cool air to a carburetor air box or whatever, it may be necessary to make a compromise or two in the air/fuel mixture ratio toward a very slightly rich condition at lower vehicle speeds to take advantage of the additional air available to the engine as the vehicle approaches its maximum speed.

But most air scoops are junk, doing more harm than good. It takes some really intelligent and careful work to design and execute a very good air scoop/carburetor air box system, but when it's right, it is certainly worth the time and effort of making a few wrong ones. Depending upon rules, not all vehicles are permitted the luxury of an air scoop/air box system, but if it is legal and acceptable, do it. A good air scoop/air box system will do and/or not do several things:

(1) It will supply all carburetor air, being sealed off from secondary air sources.

(2) It will be removed as far as possible from any source of heated air (radiator, oil cooler, etc.). (3) It will equalise cylinder-to-cylinder air/fuel distribution so that all spark plugs and all piston crowns appear as close to identical as possible under all operating conditions.

(4) It will be located at a safe point above ground level to preclude the possibility of inhaling rocker, sock, rags, beer cans or other trash.

(5) It will provide easy access to all carburetors.

(6) It will not direct high velocity air across the carburetor air horns causing fuel to be siphoned from the carburetors.

(7) It will provide a significant power increase, particularly nearing maximum vehicle speeds.

Sound simple? Try it. A good carburetor air supply system may indicate a camshaft with longer effective duration for even better top end power, if the minimum engine speed is kept well above the stagger-stumble-lurch range.


Chapter Sixteen


Minor Changes at the Track


So now we're at the racecourse and (almost) ready to race, but the indications are that the engine doesn't have enough muscle at lower engine speeds. What can be done to fix it quickly? Several things. Assuming the venturi diameter of the carburetors is not too large, and that the air/fuel mixture is "lean and clean" through the engine speed range (black smoke and the "blubbers" are out), and you're stuck with existing intermediate and final drive ratios, the most direct approach is to advance the camshaft a few degrees. The L-series Datsun engines are sensitive to this treatment and respond well to it. This improves low and mid-range torque by the simple expedient of closing the intake valve earlier, although the entire camshaft must be advanced to capture the beneficial effect upon torque output at lower engine speeds caused by the earlier closing of the intake valve. A "few" degrees means to advance the camshaft enough so the engine can recognise that a change has been made, but not so much as to hill the performance level at higher engine speeds. A good starting point is to advance the cam from 3 to 5 crankshaft degrees (1-1/2 to 2-1/2 camshaft degrees). Presumably, the earlier advice of lots of piston-to-valve clearance has been well taken, so that after this change is made, piston-to-intake valve clearance will not be dangerously diminished. Just make certain the cam is advanced and not retarded. This means that when viewed from the front of the engine, the camshaft must be moved in a clockwise direction in relation to its present position and everything else stays put.

Another change that can be made to improve low-end torque output is to increase the intake valve lash by from 0.004 to 0.006-inch, perhaps a bit more, but this approach is not as direct as advancing the camshaft. Most cam lobe profiles have a clearance ramp long enough to accommodate a reasonable change of valve lash before the ramp joins the cam lobe flank. There are a couple of practical limitations here. The length and shape of the clearance ramp has a direct influence upon the maximum safe engine speed of a given cam lobe profile, so the valve lash cannot be increased to the point where the clearance ramp is by-passed because this can make life very difficult indeed on the cam lobe/rocker pad interface stress conditions. Another factor: If the valve lash is increased too much, the valve lash will improve low-end torque by effectively shortening intake valve duration slightly, and also by a slight reduction in effective valve overlap. Again, make enough of a change so the engine will know a change has occurred, but not enough to ruin the performance level higher up, or to cause unnecessary valve train damage.

If this captures about half the missing torque, it is usually possible to gain some on the other missing half by applying the same trick to the exhaust valve lash. Don't expect to make and improvement of the same magnitude here, though, because the engine is not as sensitive to slight changes in effective exhaust duration as it is to similar changes in effective intake, and by observing the same limitations and precautions.

There are very few races in the world, and still fewer racecourses, where total, maximum; last-gasp top-end horsepower is an absolute requirement for the fastest possible vehicle speed or lap time. Some of these may include the Super tracks like Daytona (excluding the road race section), Talledga, Bonneville (the Great White Dyno), flying kilometer time trials for boats and maybe, if the gearing is just right, Ontario, Indianapolis and LeMans. The vast majority of race courses, ashore or afloat, have one point in common: all other things being equal, the races become a series of short, medium or long drag races as far as engine performance level is concerned. This means that both the engine and the vehicle must have the ability to accelerate - a word that strongly and correctly implies that the engine must work at its very best through an engine speed range. There are no high-performance engines - none - that are run at a constant, unvarying engine speed. If this were the case, an engine could be tuned to produce ultimate power output at a fixed engine speed to the utter exclusion of every other factor, and it would be very much easier to do so. Instead, all engines must operate through an engine speed range. Sometimes the range must be very broad, a condition that demands flexibility, even if it means sacrificing a few horsepower on the top end to gain the required degree of flexibility elsewhere within the working range. And flexibility, even if it means sacrificing a few horsepower on the top end to gain the required degree of flexibility elsewhere within the working range. And flexibility spells combination much more loudly and clearly than it spells maximum horsepower effort.

About the least likely place on earth to qualify, as a drag race is Bonneville, but it's true; it is indeed a drag race. The vehicle must start from zero miles per hour and reach its maximum speed within a fixed distance. If the engine doesn't have the ability to accelerate (that word again) because it lacks some essential factor for the right combination (that word, too), it's a waste of time because it isn't even an acceptable place for a vacation. The fire-time Bonneville competitor invariably arrives with his equipment overgeared, overcammed, underfueled, wrongly-tired, over-carbureted, under-jetted, wrong attitude of vehicle at speed, a totally inadequate air induction system and un-knowledged of the peculiarities and perversities of the place. Besides, he will have forgotten his metric toolbox. A week later, he leaves the wretched place with a truckload of fragments, a junk ex-race car and he has emergency hospital cases of sunburn, dehydration, malnutrition, hypertension, exhaustion, shock and a hangover. But it's a fun way to race, it affords a weeklong opportunity to find the right combination, even if it was left at home, and it does give some insight into the mysteries of tuning for flat-out maximum power. But it is STILL a drag race.

In such a case, as rare as it may be, if better maximum power output is required to satisfy a given condition, you have to rob Peter to pay Paul. In other words, it's a trade-off of low and mid-range torque for better output at the top end of the engine speed range. As on might expect, the correct approach is the exact opposite of that for improving low and mid-range torque; that is, the camshaft should be retarded a few degrees and the valve lash should be decreased, or a satisfactory combination of both. In retarding the camshaft, be careful. The maximum amount advisable initially is 3 crankshaft degrees (1-1/2 camshaft degrees); as previously mentioned, the centrifugal action on the timing chain at higher engine speeds will account for an additional 1 to 2 crankshaft degrees, perhaps more, depending upon the condition of the chain. Of course, this assumes adequate piston-to-exhaust valve clearance after the camshaft has been retarded. Again, make certain the camshaft is moved in the correct direction. To retard the camshaft, the cam must be moved the desired amount in a counterclockwise direction when viewing the engine from the front and all other associated pieces stay put. After the camshaft has been retarded, it is highly advisable that a valve timing check and a piston-to-exhaust valve clearance check be made to be dead certain there has been no slip-up. A relatively small error in this direction and you've got a handful of bent exhaust valves at best, and they don't even make decent paperweights.

Decreasing the valve lash will help top end power slightly but in this case too, retarding the camshaft is the most direct method and the results will be more positive. Moderation is required in squeezing the valve lash closer; initially by no more than from 0.002 to 0.004-inch, then work the engine good and hard, shut it off and very quickly measure the valve lash with the engine temperature as hot as it is likely to get. This is a necessary precaution because when the valve lash is decreased, the effective duration is increased slightly and as a function of this, the valves are seated for less time, therefore there is less time available to transmit the heat of the valves to the valve seats. As a result, valve operating temperatures are increased somewhat with a consequent increase in the thermal expansion of the valves, which further decreases the valve lash. Obviously, there must be some valve lash, even if the entire engine is at the melting point. Titanium and stainless steel valves are affected more by changes in valve operating temperatures than more conventional valve steel alloys. In any case, take it easy in this area and approach the minimum practicable valve lash condition gradually, if for no other reason than the increased thermal expansion of the valves could easily be more than anticipated, in addition to which, there is a point of diminishing returns.


Where does the Power Occur?


Completely and competently modified Datsun L-16 and L-24 strictly race engines are capable of producing 1.90 brake horsepower per cubic inch of piston displacement (115 brake horsepower per litre of piston displacement). These are average figures; some have been slightly better, some not quite as good. The 1.90 figure will certainly make such an engine competitive with other Datsuns as well as with lots of other makes. Maximum power is usually reached in the 7,800 to 8,000 RPM range, while peak torque is reached at about 6,500. Similarly modified L-18 engines reach max power in the 7,500 to 7,700 range with max torque at about 6,200. Average specific power output is a bit lower, about 1.87bhp/ (About 114 bhp/litre. A broader torque range in conjunction with better specific torque output is where the L-18 shines over the L-16 and L-24. This is due not only to the larger cylinder bore and longer crankshaft stroke, but also to the shorter centre-to-centre dimension of the connecting rod. The above figures are with good quality gasoline for fuel.


When to Shift?


This data leads up to the questions of shift points and how fast to run the engine in top gear. Best point-to-point acceleration is usually obtained when the engine is permitted to run from 6 to 8% higher than the speed at which max power occurs. If max power occurs at 8,000 RPM, Shift points in the intermediate ratios should be at about 8,500 to 8,650 with decent intermediate gear ratios. If there is a gigantic hole in the 2-3 intermediate ratios, the shift-point in 2nd will probably have to be increased to about 8,800, perhaps 9,000. If max power occurs at 7,600 RPM, shift points should be in the 8,100-8,200 range. The purpose of over-speeding the engine beyond the point of max power is to be certain that after a shift has been made. Engine speed will drop back to a point between max-torque and max power so the engine won't have to pull itself up by its bootstraps in the lower engine speed ranges. Max-sustained engine speeds in top gear should be about 8,200 to 8,300 RPM with L-16's about 8,000 with L-24's and about 7,800 to 7,900 with L-18's. It doesn't do much good to run an L-18 faster than about 8,500 in any case because the engine has inherently better torque output at lower engine speeds than either the L-16 or L-24. These RPM limits are meant as suggestions for starting points only. Each engine-vehicle combination could very likely require on or more modifications to these limits for best overall performance and the only way to find out for sure is to try different RPM levels in the intermediate gears and in top gear. While Datsun 510-610-Z-cars are very popular for anything from transportation hacks to flat-out race cars, Datsun engines are deservedly filling in some niches very successfully that have been dominated by other makes. A few of these would include midget race cars, small drag race vehicles, dune buggies, sand drag racers, Baja-type vehicles, marine installations, etc. And why not? The engines are very strong structurally as well as in power output, they are hard to break, are fairly light and they have the ability to absorb more punishment and abuse than they deserve. In addition, they have the capability of actually producing better power than other designs that may look better but can't produce.





The foregoing is more-or-less related to Datsuns equipped with one of the several combinations of standard 4-speed or optional 5-speed gearboxes. A special note regarding camshaft selection for any of the L-series Datsuns equipped with an automatic transmission is necessary. The convenience of an automatic gearbox, particularly in bumper-to-bumper traffic, cannot be ignored. But from the standpoint of performance, the automatics used in L-series Datsuns take quite a bit from the engines because the torque converters sop up power like a blotter and the intermediate ratios are awful, the worst being from 2nd to top gear. However, a few tricks can be applied to the 3-speed automatic-equipped Datsuns to help close the performance gap, as compared to similar cars equipped with 4 or 5-speed gearboxes. If better acceleration is desired, and this would almost have to be the first consideration with the automatic, a higher numerical final drive ratio is in order. If the final drive ratio is say, a 3.5, then a 3.9 ratio, which is about a 10% differential, would automatically (no pun) improve acceleration without having the engine excessively buzzy at normal highway cruising speeds. Judicious use of the throttle foot won't destroy fuel economy; in fact, it may improve. An L-18 engine in an L-16 vehicle will give and 11% increase in piston displacement and an increase in torque of about 12 to 14%. Any other internal-external engine mods must be very conservative indeed. The engine must idle, in gear, in a civilised manner and at an acceptable engine speed because the torque converter places a load on the engine, which makes it doubly sensitive at idle and just off-idle, and this is where any performance gain (or loss) will be most noticeable. The message here is to improve torque output in the lowest engine speed ranges. Presumably, an automatic-equipped Datsun is strictly a transportation chug suitable for driving almost anywhere by your wife, girlfriend or mother (or all three), but there is no valid reason for it to be dull. In this case, a compromise is obvious, and this must be in the engine speed range of max power, but with your wife, girlfriend or mother (or all three) aboard, you'll never get a chance to use it. However, acceleration can be pretty zippy (comparatively) if the camshaft is mild, MILD. M-I-L-D! Effective duration should be in the low to mid-220 degree range with from 4-8 degree overlap (no misprint) with lift in the 0.450 to 0.460-inch range so it won't die completely at 6,000 RPM or so. This isn't a budget-busting installation either; the only special items really required are the camshaft and a set of valve lash pads of the correct thickness. New stock Datsun rockers are a must, and a new set of late-type L-series valve springs is highly advisable. There are a couple of other indirect economical and ecological advantages too: Fuel economy will very likely be measurably improved along with significant reductions of exhaust emissions. The only real penalty with a camshaft of very short effective duration and reasonable valve lift is that the valve motion is fairly swift so the engine may be somewhat speed-limited to a safe 6,500 RPM.


Chapter Seventeen




The practice of turbocharging the latest crop of small engines is becoming increasingly popular in an attempt to get the little kids to catch up with the big kids. Supercharging, particularly turbocharging can show some value as an exhaust emissions tool if it is done correctly, as well as giving relatively large performance gains. L-series Datsuns respond well to this treatment, about the only weak link being blown cylinder head gaskets, usually when some ignorant, overzealous lout tries to shove more boost pressure down its throat than it has capacity to digest. O-ringing the cylinder head, or a solid copper head gasket, or a combination of both if one is really serious, should handle this problem.

Data for the various types, makes and pressurising characteristics can be found elsewhere (in another H.P. Books!). However, all superchargers have some common characteristics and anyone contemplating a supercharger installation in a road vehicle should be aware of these. First, all superchargers are sort of "demand" type pressurising devices. That is, the supercharger doesn't produce any positive intake manifold pressure until the throttle is about ½ to ¾ open and the engine is pulling against a load. At idle and during part throttle cruise conditions, the super charger is there, going along for the ride, but is doing very little, if any, useful work. A supercharger does more for the torque curve than it does for the power curve. A moderately-supercharged road engine, showing a positive intake manifold pressure of 7.5 pounds per square inch, will typically show and increase in max torque of from 37-40% while max power increase will be about 25%. The first 5-7 psi manifold pressure are the ones that really get the job done; anything above this may be frosting on a cake that doesn't need it. The message here is not to try to over-supercharge a road engine. The old hot rodder's maxim "If some's good, more's just gotta be better" is fine in its limited place, but supercharging is like a disease; the more you get, the more you want and it can be a long, hard, expensive lesson learning when and where to quit. So quit while you're ahead; before you melt down one or more engines. Accept a moderate boost pressure, live with it and be glad you've got it. This is particularly true in today's world of diminishing fuel quality. And here's why: A moderately-supercharged engine with say, a maximum boost pressure of 7.5 psi can be said to have a 50% supercharge at sea level. This means that the supercharge overcomes a very slight intake manifold depression (vacuum) at wide open throttle and pumps an additional 50% of air/fuel mixture into the cylinders. This is what gives you the whack in the back when you lean on the throttle. But for this, no matter how good it seems (and is), there's a penalty of increased cylinder heat by a factor of over 100% with gasoline as fuel. With today's fuels, no engine will sit still for long under sustained full-boost conditions before it detonates, pre-ignites, or just flat melts. A supercharged road engine is great fun as long as this major limitation is acknowledged and accepted.

Now for the plumbing. By far the best method of supercharging is to place the supercharger between the induction system and the carburetor(s). This very neatly avoids the complications associated with blowing compressed (and hot) air through the carburetors, the necessary modification of equalising the carburetor float bowl pressure with ambient pressure around the carburetor, and the modification of the fuel system to include another fuel pump so fuel pressure delivered to the pressurised carburetor will be about the same as normal. With the supercharger mounted between the induction system and the carburetor, the fuel itself acts, as a coolant on the compressed air/fuel mixture so there is only a slight increase in mixture temperature. Better yet, the supercharger, particularly a centrifugal type, just flogs, thrashes, whips, beats, churns, mauls and mutilates the living daylights out of the air/fuel mixture, making it into some semblance of a semi-homogenous, semi-vapourised charge. Can you possibly imagine what happens to the air/fuel mixture on the compressor side of a turbocharger when the impeller is "idling" at from 20,000 to 50,000 RPM? Not necessarily producing any positive manifold pressure, you understand, simply chewing hell out of the fuel particles and forcing them into much more intimate association with the surrounding air. In this very much more finely-divided state, the air/fuel mixture is in better condition to enter the cylinders in more-or-less uniform density, probably because it's too weak to fight back. This is how a supercharger can help in the battle of exhaust emissions, even if it isn't "working" as such. This condition also tends to make throttle response astonishingly quick. Understand that a supercharger cannot correct for a poor intake manifolding condition by guaranteeing equal cylinder-to-cylinder air/fuel mixture distribution, simply because no supercharger, blower, or "suckbox" if you prefer, has the inherent capability of doing this. A poor intake manifold condition is just as poor, supercharged or unsupercharged.


Camshafts for supercharged L-series Datsuns.


To be hones about it, I have never really been able to define, to my own satisfaction, just what a "supercharger camshaft" it. Even in the world of the alcohols, nitro-methane, nitropropane, polypropolene, nitrous oxide, hydrogen peroxide and even a touch of oxygen, and back to more mundane fuels such as gasoline, I have never seen that much difference in the valve timing requirements of an engine, gasoline or romantic fuels, supercharged or unsupercharged. There are a few subtly differences but certainly nothing earth-shattering. In fact, I have found that in the vast majority of cases, the camshaft profile that works best in an unsupercharged engine will usually work best in the same engine with a supercharger. This observation has been very nearly infallible with moderately supercharged gasoline-burning road vehicles. So whatever camshaft profile works best in your unsupercharged L-series Datsun will, with about 99% certainty, work best with a mild supercharge. There is one increasingly popular exception.


Enter the turbocharger


The nature of this exhaust-driven supercharger requires a different approach, but not without reasons attached. Normally, effective intake duration can remain about the same as in an unsupercharged engine. If a change is indicated, it will usually be in the direction of later intake valve opening and sometimes later intake valve closing. The biggest change, and one that is relatively enormous, lies in the required effective exhaust valve duration. This should be short, SHORT, S-H-O-R-T! Even a double-throw-down, triple-whammy, flat-out turbocharged race engine for the turbocharger to work most efficiently - should have an effective exhaust valve duration of no more than the mid-270 degree range, very little overlap and as much valve lift as can be conveniently cranked into it, taking the dynamic stability of the exhaust cam lobe at maximum engine speed into consideration. Didn't know that, didja?




A turbocharger works best as a function of the temperature and velocity of the exhaust gases, and the volume of exhaust gases is secondary. This strongly suggests late exhaust valve opening and early exhaust valve closing so that exhaust gases, in their purest form, can work on the turbo impeller, which is directly connected by the same shaft to the compressor impeller. This doesn't indicate that a turbocharged engine won't function with stock camshafts or those with a relatively long effective exhaust valve durations. It will. But longer exhaust valve durations have drastic effects on the average exhaust gas velocity and temperature. In addition, longer valve overlap periods permit fuel from the pressurised induction system to be pumped out the still-open exhaust valve, lowering exhaust gas temperature and density even further.

This is the primary reason why most stock-but-turbocharged engines feel like you've stepped on a wet sponge at low engine speeds; they're just plain soggy until there is enough of a blast of exhaust gases to wake up the turbo impeller which, in turn, wakes up the compressor impeller. So, for a moderately-turbocharged L-series Datsun road vehicle, effective exhaust valve durations should be in the low to mid-220 degree range and possibly a later opening intake valve. If you don't think this will put some instant muscle in your Datsun, try it.




Now lest' consider "standoff" or "pressure reversion" a phenomenon that shouldn't exist to the detriment of engine performance, but does in too many cases. Nearly everyone even remotely associated with high-performance engines is aware of the so-called "ram" effect as applied to engine induction and exhaust systems. This condition is generated by sonic pulses that continually rattle about in an engines induction system, exhaust system and at times, in the combustion chamber area, and a favourable ram effect will occur when these pulses become aligned in direction and magnitude to cause a larger-than-normal charge of air/fuel mixture to be pumped into the cylinder from the induction system. A similar alignment in direction and magnitude of pulses causes a larger-than0normal volume of exhaust gases to be pumped from the cylinder through the exhaust system. In simplest terms, pressure reversion could be defined as diametrically opposite a favourable ram effect, a condition in which the air/fuel mixture is forced away from the cylinder, back up the induction system toward the atmosphere. A similar condition can exist on the exhaust side whereby exhaust gases can be forced back toward the cylinder from the exhaust system. It is quite reasonably felt by some experts in the field that the pressure reversion in the induction system is caused by a pressure reversion in the exhaust system, with the combustion chamber area as the connecting link between the two during the overlap period. However, later indications, are that pressure reversions in either system can occur independently of the other, although their magnitude seems considerably less than a combined induction-exhaust reversion.

Pressure reversion usually manifests itself visibly as liquid fuel or fuel stains on some surface at or near the upstream sides of the carburetors. Sometimes it doesn't get as far as the atmosphere, in such cases being visible as a sort of "ball of fog" extending from the intake manifold and perhaps into the carburetor while the engine is operating through its normal speed range, preferably at full throttle. Sometimes it may not be visible at all due to the intake manifold configuration, which would cause the reversion pulses to be damped and contained within the manifold. Happily, there are times when a pressure reversion condition does not exist at all within the normal engine speed range.

We sometimes like to think of the air/fuel mixture and exhaust gases as smoothly-flowing, but such is not the case. The sonic pulses, or pressure waves, as you prefer, which incidentally, occur in all engines, cause violent disturbances to the air/fuel mixture and exhaust gases within the cylinder and induction and exhaust systems. These pulses represent energy, quite a bit of energy, in fact, and when they can be made to work favourably for and with an engine, as they can in the correct application of the ram principle, engine performance comes to life. However, when they work against an engine, as in the case of reversion, engine performance tales a gigantic nose-dive. If these pulses were one-directional downstream pulses, that is, from the atmosphere, through the induction system and into the cylinder, then from the cylinder through the exhaust system to the atmosphere, things would be lovely. However, for every downstream pulse, there is a reflected upstream pulse of lesser magnitude and these are the ones that do the damage, particularly when they become so unsynchronised or out-of-phase as to cause a pressure reversion, a highly-undesirable, performance killing condition.

The consensus is that these sonic pulses are initially generated by the opening and closing of the valves, although when either or both valves are open, the piston crown cannot be ignored as a possible secondary source of pulse generation. It could also be that the piston is a primary source of pulse generation within the combustion space when both valves are closed. The latest data corroborates earlier findings in that the pulses are basically sonic in velocity. But sonic velocity varies with the density, temperature and pressure of the working fluid; therefore the actual pulse velocity in the induction system will vary greatly from that in the exhaust system, with the combustion space serving as a transition between the two extremes. In addition, there is a thought that downstream pulse velocity should be added to downstream gas velocity, while upstream pulse velocity should have downstream gas velocity subtracted from it. With all this downstream/upstream gas/pulse thrashing about going on simultaneously, there is little wonder that some disagreement exists between experts in the field, but the biggest wonder is that the engine runs at all.

Pressure or pulse reversion exists most prevalently to a performance-damaging degree in engines equipped with individual runner (IR) induction systems where each cylinder has its own isolated carburetor throat and intake manifold runner and there are no interconnections between carburetor throats or manifold runners. This applies to the L-series Datsuns because this type of system is used most frequently for Datsun race engines, and to some extent, for modified street and dual-purpose engines with Datsun-available 44mm or 50mm Mikuni/Solex side-draft carburetors and manifolds, and sometimes with Weber carburetors. The reversion problem shows up at its worst when the induction and exhaust systems appear to be "clean"; that is, when the carburetor throats, manifold runners, cylinder head ports, exhaust header pipes are all nicely matched and blended to their mating pieces. It may be clean in fact as well as appearance, but unfortunately, it is clean in both directions, so reversion pulses have and easy time of it.

Four separate and distinct areas require possible reworking to minimise the effects of pressure reversion, if not eliminate them completely. First, the exhaust system flange and primary pipe should be about 1/8-inch larger on all sides than the port opening in the cylinder head. Second, the intake port face in the cylinder head should be about 1/8-inch (1/4-inch on the diameter) larger than the intake manifold runner, then the port should be funneled down to more normal dimensions as it approaches the intake valve. Third, the intake manifold runner should be about ¼-inch larger in diameter than the carburetor throttle bores, and the runners funneled down to a smaller dimension at the manifold mounting face. The idea is to make deliberate mismatches at these three points.

The reasoning behind this is that there is pretty conclusive evidence that the downstream pulses (the good guys) take the shortest distance to get where they're going, while the reversion pulses (the bad guys) stay close to the walls of the carburetor, intake manifold runner, intake port, exhaust port and exhaust pipe. The deliberate mismatches make abrupt changes in cross-sectional area, which are highly beneficial in damping the unwanted reversion pulses. In addition, the air/fuel mixture traveling downstream is pumped into areas of lower-than-normal pressure, which in itself, helps induce a larger volume of mixture into the cylinder, and the same is true on the exhaust side. Edelbrock Equipment Company has made a couple of prototype manifolds incorporating the mismatch concept for the L-16, L-18 engines with encouraging results for a first attempt in damping reversion pulses.

The fourth are that may require a change is valve timing. By itself, valve timing can have rather dramatic effects upon the presence or absence of pressure reversion.

If a reversion problem exists, the changes should be made one at a time and in the order shown until the problem disappears completely or is at least helped considerably. At the points of mismatches, leave the edges square and sharp. DO NOT ROUND OFF THE SHARP EDGES! Perhaps strangely, there are highly modified L-series engines with no reversion problems at all within the normal operating speed range.


Chapter Eighteen




When the piston displacement of an engine is increased, diameter, crankshaft stroke, or a combination of both, it becomes less sensitive to effective valve timing than it was in its original state. While a bore increase has some influence in this direction, the better de-sensitiser is a stroke increase. This comes about for two reasons: (1) A stroke increase does increase piston displacement. (2) The ratio of connecting rod centre-to-centre length-to-crankshaft stroke decreases. The L-16 rod length of 5.236 inches (133.025mm), divided by the 2.902 inches (73.7mm) stroke length gives a rod-to-stroke ratio of 1.8042 to 1. Substitute the L-18 stroke of 3.0708 inches (78mm). With the same rod length, the rod-to-stroke ratio is reduced to 1.705 to 1, a reduction of 5.81%. True, average piston velocity is higher with the longer stroke, as is rod angularity. These may seem like steps in reverse, but they aren't all bad. Piston velocities across top and bottom centers are faster with the longer stroke but are relatively slower through the mid-part of each stroke because it is an inescapable fact of mechanical life that exactly one revolution of the centre, to bottom centre, and back to top centre again, regardless of the rod length or the piston stroke. However, the faster piston velocities across top and bottom centres have the most significant de-sensitising effect upon valve timing because the effective valve opening and closing points usually fall within the range of these faster piston velocities. This is sometimes convenient because it allows the small luxury of a small error to be made in effective duration without the usual penalty of feeling that the vehicle is stuck to one spot in the pavement.

Now put the situation in reverse; that is, assume it seems advisable to "shrink" the piston displacement of a given engine to make it fit into a specific class governed by piston displacement. Sleeving the cylinder bores to a smaller diameter is a bad scene because it robs the cylinder block of some very necessary structural stiffness and also heat-transfer problems are almost inevitable. The more practical approach is to de-stroke the crankshaft to something less than stock. Well, maybe it seems more practical. This time, we'll reduce the stroke by the same amount the above example was increased, and use the same rod length. Now the stroke is 2.7352-inches (69.474mm), and the rod-to-stroke ratio is 1.9143 to 1, an increase of 7.19% over the original version and an increase of 13.43% over the long-stroker. While the short-stroker figures are not extreme by any stretch, this is where effective valve timing can easily become downright hostile, and for reason diametrically opposed to those that help the long-stroke condition, as one may suspect. Piston velocities across top and bottom centres are slower with the higher rod-to-stroke ratio, and faster in the mid-part of the stroke. This condition demands significantly shorter effective durations, narrower displacement angles and as much valve lift as feasibly without affecting valve train stability, the latter factor being important because the short-stroker will undoubtedly be operating in the super-soprano range of engine speeds. If it isn't, someone made a grievous error in judgement and the engine will need the Red Cross treatment to climb a driveway. These valve-timing requirements are positive for the short-strokers. No ifs, ands, buts, maybes or other qualifications. This is the way it works. Or doesn't, as the case may be.

The L-18 engine is ahead of the L-16 by 10.97% in piston displacement alone. But a couple of other sneaky factors may not be so apparent. Compared to the L-16, the bore increase is only 0.080-inch (2mm) but the stroke increase is over twice that at 0.1688-inch (4.3mm), which reduces the L-18 bore-stroke ratio by 3.34%; not monumental, but it helps. The L-18 rod length is 0.104-inch (2.64mm) shorter than the L-16 and makes the L-18 rod-to-stroke ratio come out to 1.6712 to 1, a decrease of 7.95% from the L-16. This is the item that gets the job done and has a mothering effect by being more tolerant and more forgiving of small errors in effective valve timing, holes in the intermediate gearbox ratios, etc., to say nothing of driver errors. This is the way to go for street or dual-purpose engines. It may not be able to climb trees. Small engines just don't climbe trees intentionally. But it will make the driveways seem flatter, even if they aren't.

The old theoretically alleged "ideal" rod-to-stroke ratio of 2 to 1 surely made a lasting impression. Some guys till cling to it like they were welded. And invariably, these are the guys who do not, cannot, will not, won't, get it through their anvil-thick skulls that they are in a lousy bargaining position to accept what they lose at lower engine speeds for what may possibly (not even probably) be gained at maximum engine speed. They must all belong to some universal idiots' association, for the chant is always the same; "I never see less than 7 thou - well - maybe 65 hun." How can you argue with such beautiful logic? But it's perfectly true. To see it, they'd have to look for it. The 2 to 1 rod-to-stroke ratio is just that. Old. Theoretical. Alleged. Extinct. It does have a place or two, but these are increasingly rare exceptions to the rule. I have a nodding acquaintance with the problem.

But for what it's worth, a much more usable, workable, livable set of rod-to-stroke numbers these days are within the range of 1.85 down to about 1.65 to 1. The higher ratios should be reserved for engines that operate at consistently higher average engine speeds, while the lower ratios work best and are happier at lower engine speeds, and also when the engine must be strong through a broad speed range. It isn't always that simple, but if there is a choice and physical limitations are no problem, that's the way it should be. In spite of UUI (United Universalised Idiots).




That should about cover camshafts and valve train pieces for the L-series Datsuns. The meanderings that may seem far afield from the direct subject matter were included to point out that except in the mildest states of tune, it will nearly always take more than on piece, component or system to transform and L-series Datsun engine into a workable, livable, usable high-performance unit. I have also attempted to show the relationship between other components and systems to the valve timing valve lift-overlap requirements, all of which are so essential in developing a right and winning combination (that word again).

Four final words: DON'T "OVERCAM" YOUR DATSUN!! And don't break it, either.