A problem confronting many engine builders is making the best choice of cam and valvetrain components. Picking up a cam catalog doesn’t necessarily help because of the number of choices. On a single page you may see several cam specs given, and many may seem fine if you base your choice on the application description. Maybe one has a little more lift than another, but a little less duration. Another may have slightly less duration, but be on a tighter lobe centerline angle than yet another possibility. Sure, you can ask people what’s worked in their motors. Although that may seem a simple solution, let me assure you it can easily fail to produce within 30 hp of optimal results in 95 cases out of 100, and I’m talking about a 350 here. If the engine is bigger, the discrepancy can be even greater.
Why so far off? Ask yourself if the person you are seeking advice from has successfully built a motor like you intend to build. If so, do you intend to closely follow the specifications that were used? How do you know this person was qualified to give advice?
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For the record, I’ve found that even among successful, professional, race-winning engine builders, real camshaft expertise is rare. I hate to sound like I am blowing my own trumpet here, but this is a subject I teach at University level. Along with that, I have also redesigned a whole line of cams for several cam companies to very good effect. In two instances this was done to the extent that the engines in question made more power from my hot street cams than they had previously seen from their race cams! The advice you are getting here is based on 40 years of testing cams in general plus well over 10,000 combinations (8,000 in one six-month-long session alone for Crane) for small-block Chevys.
My recommendations are based on obtaining the maximum output, especially in terms of torque, for a given intake duration.
There are five parameters to be addressed when selecting a cam and valvetrain for best performance. In order of importance they are: overlap, the lobe centerline angle (LCA), advance/retard setting, lift, and duration. Understanding how each is affected by engine spec is the key to getting the best cam for the job. Some engine configurations respond better to added lift than to increased duration, and some the reverse. Some engines need a tighter LCA than others, and so on.
Investigation reveals that cam characteristics are tied to cylinder head air- flow capacity and the size of cylinder it has to supply. This implies that a flow bench is not only a tool for developing heads, but also for selecting a compatible cam design. To see how the factors of flow and displacement affect the cam spec required, let’s start with the seemingly disconnected issue of fuel octane.
Fuel Octane and Dynamic CR
Fuel octane influences the compression ratio that can be used and an engine’s CR must, for best results, be tied to cam duration. The key factor is the dynamic CR based on the intake valve closing point. This is closely related to the cranking pressure seen on a compression gauge. Obviously the piston will be well up the bore if the valve closes late as per a long- duration cam. At low RPM at least, the cylinder will trap less air and therefore produce a lower dynamic CR. This causes reduced cranking pressure and consider- ably less torque, especially at the low end.
Fig 7-1 illustrates just how this lower cranking/low speed compression comes about. We need the CR to be as high as possible without incurring detonation. Detonation is brought about by a combination of time, heat, and pressure. For most engines, detonation occurs at low RPM because of the greater length of time the end gases are exposed to the advancing flame front. When a long cam is used, the compression pressures at low RPM are reduced, thus not only allowing—but requiring—a higher CR to be used.
The static CR must be based not only on fuel octane but also intake valve closure point. This means establishing the dynamic CR required. For the record, most successful race engines run dynamic CRs of about 9:1, and a good street motor on service station fuel about 7.5:1. If the compression ratio isn’t in line with the cam duration, the closure point of the intake valve, the dynamic CR, and cranking pres- sure, will consequently be low. This not only kills torque everywhere in the rev range, but also significantly reduces the engine’s low-end output. In other words, it emphasizes the undesirable long-cam features that we’re trying to avoid.
By putting more compression into the engine, it’s possible to regain most of the low-end output lost by the use of more duration, and all of it in the case of shorter performance cams. A good test to establish that an engine has a high- enough CR is to do a cranking test. Assuming that rings and valves seal perfectly, then even a modest street motor should show at least 180 psi if cam and compression are roughly right. However, in my opinion, that is a worst case. For a regular-use performance street motor I’d expect to see 200 psi. For a street/strip motor, this should be 220, and 240 for a race-only engine. Assuming no pressure is lost to leakage, anything below these levels indicates too low a CR. For these reasons, it’s important to follow the CR requirement given for each cam in the selection charts. If for some reason the required CR for a cam can’t be attained, choose a shorter duration cam; you’ll like the result better.
I want to make this next point absolutely clear. It is not the selection of a long cam duration that kills drivability (as opposed to low-speed torque) and vacuum, but overlap. Overlap is the time around TDC when both the intake and exhaust valves are open, and it’s a function of both duration and lobe center line angle. Overlap is crucial for obtaining high output from a race engine. Though good for power, it is, if overdone, destructive to low-end output, drivability, and mileage. This makes the decision as to how much should be used for a given application a critical decision. It’s tempting to add a few more degrees in the hopes it won’t cost much at the bottom end, but in reality, it’s like playing Russian roulette with numbers instead of bullets. For example, selecting a cam that has 40 degrees of overlap instead of 30 degrees may only appear to be adding an extra 25 percent, but this isn’t the case. The critical factor is the overlap area brought about by the combined degrees and lift. Adding more overlap also adds lift during the overlap period. What appears to be only an extra 25 percent of overlap actually ends up being about 80-percent more. The bottom line is, don’t be tempted to select a cam having too much overlap.
The effect overlap through-flow area and cylinder displacement size have on the power curve are closely linked. For a given displacement, the greater the through-flow area is (overlap degrees times lift), the more RPM the engine needs before it starts to perform. The bigger the engine is for a given cylinder head, valve size, or overlap through-flow area, the lower the RPM at which it starts to function. This means the selection of the overlap must be based largely on the lowest RPM that the engine is required to run from.
Fig 7-2 gives an indication of what’s required for a small-block Chevy equipped with an exhaust system with near-zero back pressure and an unrestricted intake.
To put numbers to these, we’re talking about an exhaust with less than 0.5-inch of mercury (0.25 psi) and less than 1.5 inches (Hg) of intake manifold vacuum at peak power. If exhaust back pressure increases, the amount of overlap must be decreased. If the intake is restricted, the situation becomes complex and is influenced to an even greater extent by the exhaust used. All the cams in the cam selection chart are based on the premise of enough overlap for the job and no more.
Within both the industry and the hot rodding community at large, the most difficult aspect of cam speccing seems to be determining what LCA will be the most effective. My intention here is to present a basic understanding of what’s required so you can get an idea of how the optimal LCA is influenced by the spec of the rest of the engine. In doing so you will also begin to understand why the cam that may have worked so well in your friend’s 327 won’t be remotely close for your 383.
One basic rule that can be applied across the board is: The more an engine becomes under-valved, the tighter the LCA needs to be. Normally, the LCA required is based on the cylinder head’s flow capability and the size of cylinder to be serviced. Unfortunately, only a minority group of builders are in a position to get cylinder-head flow figures. This means adopting the next best method that produces credible results. As it hap- pens, adopting intake valve size instead of flow gives almost equally good results.
What we look at here is the number of cubic inches each square inch of valve opening area (lift x valve circumference) has to feed. The more cubes per square inch, the tighter the LCA must be. This is why the optimum LCA for a hot 302 is too wide for a 383. Using a cam that worked in the 302 will prevent the increased torque potential of the 383 from being fully realized.
Before finishing with this subject, we need to deal with the effects this angle supposedly has on idle. Many off-the-shelf cams have wider LCAs to preserve idle quality. Many cam companies grind their cams on LCAs a little wider than optimum for maximum output. Intentional or not, this does save the hotrodders from them- selves. How? Because almost all of them work on the Stroker McGurk theory: If some is good and more is better, then too much must be just right. If you’re building a street motor and idle quality is important, you may be pressured from sources other than this book into considering a cam with a wider LCA. However, you will only need to go this route for a quality idle if you’ve been over-enthusiastic in selecting the overlap to be used in the first place. In other words, a wider LCA than the engine actually needs is a second-grade Band-Aid fix for your enthusiasm in terms of duration selection!
We are now getting to yet another important factor that relates engine size to cam selection. Read this twice if you have to, but be sure you understand what I say. Adding valve opening duration has always been an accepted technique for making more power upstairs, but it has its pros and cons. Adding duration is easy on the valvetrain and makes for reliability in that department. But the price of added top end is almost a direct trade off, because it reduces low-end output. Just how much it reduces low-end output depends on the ratio of valve opening area to cylinder volume. Small-displacement big-valve engines rapidly lose low-end output com- pared to under-valved larger units. The bottom line is that if you’re using a smaller version of an engine, you need to be more conservative with duration. If the engine has been stretched by boring, stroking, or both, a little extra duration proves an asset.
By extending cam timing, we cause the torque curve peak to move up the RPM range. More cam timing doesn’t necessarily mean more torque, just more RPM at which it occurs, and this translates into more horsepower. Fig 7-3 shows the effect of more duration while the CR remained the same throughout the test. The cams No. 1 through No. 3 are nominally 270, 285, and 300 degrees of seat duration. It doesn’t take much studying of these curves to see how much of a direct trade off added duration can be. However, because the CR wasn’t increased for the longer cams, these tests show a “worst case” situation. Adding the appropriate CR would redeem half of what was lost at the low end while adding to the top end output.
Before deciding to use a long duration, carefully consider the RPM range you want the engine to have. Also be aware that longer duration cams become fussier as far as the rest of the engine spec is concerned. This means it’s far better to err on the short side, especially on a street motor, as you’ll like the results far better.
Many cams have different intake-to- exhaust duration. For high-compression high-RPM applications, this is desirable because of the reduced time period within which the exhaust must be expelled. Usually about two-thirds of the extra opening period, usually from four to 10 degrees, is on the opening side of the exhaust, thus giving the cylinder more time to “blow down.” On a purpose-built performance street motor, this extra blow down can cut mileage and low-end torque by as much as 5 percent— sometimes more.
From mid-range RPM on, extra exhaust duration starts to help power. Some off-the-shelf street cams are spec’d this way because the cam designer anticipated their use in engines with standard or inadequate exhaust ports and systems. If a true street engine is equipped with a modified head and a free flow exhaust system, a single-pattern cam will often be the best choice. A dual-pattern really starts to pay off when dealing with engines with higher compressions, higher RPM capabilities, and/or nitrous. We won’t be dealing with blown or turbo motors in this book, but we will be going into nitrous motors in some detail. A nitrous motor can have significantly different cam requirements.
Lift Versus Output
The number-one question concerning lift is how much is needed before any further increase becomes pointless. This might seem tricky, but fortunately there is a simple answer for a small-block Chevy. Because flow goes on increasing to high lift, and because the intake valve is too small for the cylinder concerned, we can say with a great deal of certainty, “The more lift the better.”
It’s revealing to compare the output of two valvetrains (Fig 7-4) generating similar valve opening lift area in terms of inch/degrees.
Although each produces about the same horsepower, the short, high-lift design makes more torque, and as a con- sequence produces its power with fewer RPM. This means less RPM-induced stress on bottom-end parts. Ultimately, it’s RPM that breaks bottom-end parts. A prime goal is to make horsepower without excessive RPM. That means building torque. By making a better choice of valvetrain, we delay the point at which heavier-duty bottom-end parts are needed.
From the curves in Fig 7-4 we can clearly see that cam No. 1 produces the most usable output—especially for street use. The implication is that we should specify a valvetrain that delivers the most lift possible within the dynamic constraints imposed. Unfortunately, building a high-lift/short-duration valvetrain is more difficult than a longer- duration/lower-lift one. Achieving our objective means opening and closing the valve and lifting it higher in a shorter time. We can’t just add spring to control the situation because this means inexpensive parts wear too quickly and costly ones are out of our financial reach. To achieve our goal of power on a tight bud- get, we have to make smarter choices from the less expensive hardware avail- able. That boils down to choosing the right parts and assembling them correctly. Details such as valve, retainer, pushrod, and lifter weight are important, but the principle factors are the valve springs’ capability and cam profile dynamics.
For a street or even a semi-race application, the lift numbers for best output are far higher than can be achieved with available hardware, unless we are dealing (against my advice) with a small- displacement engine. Only for opening durations in excess of about 300 degrees does it become vaguely possible to approach maximum desirable lift. This alone should indicate that maximizing lift, especially for a street motor, should be considered a priority.
Assuming you have or are going to have a decent head on your engine, going for a little more duration and as much extra lift as possible is the way to go. For pushrod engines this means a serious look not only at the cam profile, but also the rocker ratio, which is a little low at 1.5:1 on a stock small-block. The amount of lift that can be ground on a cam is limited. The shorter the cam, the less the lift. The slower the acceleration (as in production style cams), the less the lift. Attempting to establish which of two similar duration profiles has the fastest opening is difficult, but in most cases it will be the cam with the most lift, so lift once again becomes a predominant factor.
Since we can’t get enough lift into our small-block’s valvetrain, we must adopt the highest-ratio rocker our budget allows. Higher-ratio rockers not only pro-duce more lift, but also increase opening rates, thus presenting the cylinder with breathing area faster and sooner. This, for the most part, means adopting 1.6:1 rockers. Unfortunately, indiscriminate use of high-lift rockers in conjunction with a high-lift cam and heavier springs can bring about wear problems unless you know what you’re doing.
This chapter emphasizes lift as a per- formance parameter. Also recommended are cams that are on tighter LCAs than usual, although shorter in duration. These factors combined, especially on big- ger-inch engines, means that the piston- to-valve clearance is used up quicker. Do yourself a favor: always do a trial check of piston to valve clearance. You do not have to wait until final assembly to do this. Just mount a piston, without the rings, on a rod and install it along with the cam and a head but with just the valves of one cylinder installed. Put some clay on the pistons and turn the assembly over. The minimum clearance you are looking for is 0.080. If it is less than this, have your machine shop cut the piston’s valve pock- ets appropriately. Be aware, though, that most pistons have plenty of clearance for big street cams in 383- to 400-inch engines with the cams I recommend.
Lift Versus Wear
Controlling a higher-lift valvetrain almost inevitably means stronger springs. This, coupled with the greater motion of the valve, means more side loading on the valvestem and guide. This can cause more wear. If a stock cast-iron guide is used, then anything over a 0.5-inch lift and about 270 pounds of over-the-nose spring force can wear production cast- iron guides at an unacceptable rate for the street. For instance, the guides may be well worn to the tune of costing 20 to 30 hp after as few as 30,000 miles (although this situation gets better as oils get bet- ter). The fix is bronze guides, optimizing valvetrain geometry and/or roller-tipped rockers, and—most important—keeping the required spring force to a minimum for the job. Installing bronze guides and acquiring roller-tipped rockers is straight- forward and needs no further explana- tion, so let’s deal with the budget, valvetrain geometry, and springs.
At this point let’s assume that you don’t have the cash for new rockers and are forced to use the existing stock ones. This is acceptable if the ball and rocker are in good condition and the tip isn’t worn. If the head and block have been decked, the stock pushrods are now too long. This will push the rockers up in relation to the valve. This in turn causes the rocker tip to rotate farther down at the full-lift position, and any wear ridge that was present will now run onto the valve tip and generate noise and extra side thrust. If tip wear is less than 0.005 inch and the budget is tight, you can sal- vage rockers and replace them at a later date. To make them usable, dress out the wear with emery paper wrapped around a flat file. Don’t take off any more metal than necessary so as to retain whatever case hardening may remain.
Also, unless they are the “long slot” variety, you may find that the greater range of rocker motion causes the rocker to run out of slot length and bind against the rocker stud. Check for this and grind the slot longer, as required.
To maximize the tip life of these and, indeed, any non-roller rockers, hardened valve-tip lash caps should be used. These serve two purposes: first, they spread the rocker tip wear over considerably more area, and second, they help compensate for too long a stock pushrod. In setting up the valvetrain geometry, especially with non-roller followers, we’re attempting to minimize the rocker’s contact patch sweep across the valve tip. Unfortunately, in addition to the sweeping action across the end of the valve there’s also a part rolling action due to the radius on the rocker tip. Combining these two aspects and attempting to determine what gives the least side load is far more complex than you might expect. Simple rules such as the half-up/half-down rule that work with roller followers do not deliver in the case of a non-roller follower. This being the case, here are my recommendations for the budget motor with machined heads, block, or both, utilizing non-roller rockers of either 1.5 or 1.6:1 ratio.
First, always use a lash cap. These are usually 0.050 or 0.080 inch, and if the budget allows use of pushrods 0.050- to 0.060-inch shorter for cams up to 0.500- or 0.525-inch lift. Pushrods as much as 0.10-inch shorter may be needed for cams lifting to 0.60-inch, but valvetrains in this category are moving well out of the budget class. Consider it mandatory that anything over 0.550-inch lift should really be using a roller rocker.
If roller tip rockers are to be used, the situation gets easier to deal with. If the pushrod lengths are set as per the recom- mendations just given, the sweep of the rocker across the valve tip will be accept- able for most combinations of heads and rockers, but it still needs to be checked. This should be done with an adjustable pushrod available from any performance parts house or speed shop.
As far as rockers for a basic budget motor are concerned, I use plain-stamped 1.5s and 1.6s since these are about as cheap as can be had. Sources to check for them are Crane, Summit, and Jegs. If the budget stretches for about two- to two-and-a-half times that of the plain stamped rocker, then I use the PRW or COMP Cams’ ball-pivot, roller-tipped stainless Magnum rockers. These are stiff and usu- ally produce more lift than the theoreti- cal ratio they’re supposed to. The bottom line is they usually make at least 5-hp more than even the stamped 1.6s. Also, if they’re not used with more than the recommended 350 pounds over-the- nose spring force, these rockers are about indestructible.
At about 40-percent more money, my next choice for a fully rollerized rocker is the “Energizer” rockers from Crane. There are other brands, but when the situation prevails, I use super-high-ratio rockers from COMP Cams and Crower. With big- inch engines (383 on up), the only way to go is 1.8 to as high as 2:1 ratio, but that’s not within this budget.
Making a Cam Selection
Most publications concerning high performance give you a very basic run- down on cams, to the point where you are sufficiently informed to fall just short of making an effective decision on the cam spec your engine should use. It has been a great frustration over the years to see the public misinformed on this subject time after time by books and magazines alike.
It pays to consider here that making the wrong cam spec decision costs just as much money as the right one. The biggest issue is usually which LCA is used. Getting this wrong by just 2 degrees can typically cost 20 ft-lbs and 20 hp!
Calling a cam company might seem like an easy way to get the right cam, but the bottom line is: calling half-a-dozen cam companies will get you half-a-dozen quite different answers. To avoid this sit- uation I am going to do what no other performance book has ever done—give you what must be as near absolutely fool- proof cam recommendations as is humanly possible.
You are very likely to get some oppo-sition here from so-called experts who will claim that the cam specs don’t fall in line with their experience. If it lets you rest any easier with the choice of going with my recommendations here, just ask the person in question if they have tested in excess of 10,000 combinations in small-block Chevys. If they have not then maybe you should be putting your confidence in the cams listed at the end of this chapter.
I teamed up with Lunati to do these. You will find that, unlike most catalogs where you will ponder for half an hour or more over a number of selection pos- sibilities (and still get it wrong eight times out of ten), making a selection can be done by the average 12-year-old with almost zero automotive knowledge. As for the functionality of the selection, it will be as good as if done by, say, Pro Stock engine builder/racer Warren “The Professor” Johnson. So when you are ready to make a physical selection of your cam, just turn to the last few pages of this chapter.
The spring looks as if it’s the simplest part of our valvetrain to get right. In prac- tice, it’s the single-most critical and com- plex component in terms of striking compromises in the whole valvetrain. As far as the performance of the valvetrain is concerned, it’s solely governed by the spring’s capability. Selecting inappropri- ate springs will limit what can be done. Too much spring force, especially over the nose, wears out flat-tappet cams fast; too little, leads to loss of control.
The first rule is to minimize the amount of spring force needed to achieve desired RPM. The second is to select the best spring for the job. In case you thought this is easy to figure out, let me set you straight. It’s not just a question of selecting a spring that will give the “seat” and “open” forces you think are needed. There isn’t room here to go into the sci- ence of spring design to any depth, so I’ll attempt to keep the theory part short. Ideally, we need a spring that gives the seat and nose forces and has the highest natural resonant frequency of those available to us. What this really boils down to is the selection of a spring with the highest delivered force-to-weight ratio. In so choosing, we’re moving the spring’s propensity for running into spring surge as far up the RPM range as possible and hopefully out of the range we intend to use.
To give you an example of how important spring mass is, recounting some tests I ran with Brian Crower on Crower’s spin test machine will demon- strate the point. The tests involved three springs weighing in at 82,126, and 142 grams. The poundage seat and nose were 118/342 for the 82-gram spring, 162/405 for the second, and 142/352 for the third. These springs ran the RPM num- bers to 8,500, 8,600, and 8,100. As can be seen from the numbers, the RPM achieved, pound for pound, was far bet- ter with the 82-gram spring. Why? Because far less of the delivered force is required to control the spring’s own mass. This leaves more to control the rest of the valvetrain.
With results like this, it doesn’t take much to figure which spring will produce the least frictional loss and the longest cam life. If you have access to a spring tester, here’s how you select the spring. Get all the possible candidates that will produce the desired seat and full-lift force. The spring package to use will be the lightest one of the group. This system works fine until we get to advanced lev- els of valvetrain. At that point it’s possi- ble to select a spring that becomes over-stressed in use. This is unlikely to happen in our budget scenario.
The spring’s job is to constrain the valvetrain to follow the dictates of the cam profile. There are two forms of loss of valvetrain control. These are shown in Fig 7-6. Lofting the valve at high RPM helps generate more power. Winston Cup motors rely on valve lofting, which picks up about 0.030- to 0.040-inch lift, for the last 40 hp of their output. Lofting is a secret weapon reserved for solid flat-tap- pet cams. For hydraulics it leads to what is commonly known as lifter pump up, and for rollers it’s a good way to ensure a short roller life.
Around the turn of the millennia a new player appeared on the hot rod scene. This was the beehive spring. Although it had been around a long time, the beehive spring did not find its way to any great extent in the auto industry until GM started using it in large numbers in the early 1980s. In 2002, COMP took a serious look at this spring design for use in its Chevy LS1/LS6 cam line. The resultant spring proved to be a winner. Because the beehive form produced a spring with a far less clear-cut resonant frequency, more of the spring’s delivered force is used to con- trol the valvetrain instead of its own mass. The result is less spring force for more RPM.
Although a little more money, the beehive spring is so much better that it really does bear serious consideration. For instance, a relatively short, nitrided, aggressive hydraulic-flat-tappet with high-lift rockers will deliver more torque and power than a more-expensive roller setup with conventional springs (and we are talking cam durations less than about 274 degrees seat-to-seat here).
The subject of lifter pump up was mentioned a few paragraphs back and now seems an appropriate time to talk about it. When lifter pump-up occurs, it’s because the lifter is doing exactly what it was designed to do, namely, take up excessive valvetrain clearance. If the valve starts lofting, it’s mostly due to pushrod flex. This in turn may cause it to unseat from the tappet, and this separa- tion looks exactly the same as unwanted lash. The lifter responds by increasing in length. Unfortunately, the lifter can get longer a lot quicker than it can shorten. Any added length results in the valves being held off the seats. You do not need an anti-pump-up lifter to fix this. The problem is caused either by a driver who can’t read a tach and shift at the correct RPM, by spring surge (resonance), or sim- ply insufficient overall spring force. The fix for each of these is self-evident.
So much for the lifter pump-up diversion. Let’s get back to the subject of valve motion problems. Unlike lofting, valveseat bounce is sure-fire death to horsepower. Certain factors affect how we strive to control each type of normally undesirable valvetrain motions. Too much force over the nose not only pre- vents lofting, which might just be the icing on the cake for a solid flat-tappet cam valvetrain, but also leads to an early demise of the cam nose.
Too much seat load hammers out the valveseats, although it may prevent seat bounce. Too little on the seat will allow seat bounce. To get a working compro- mise, we need to have the spring rate (stiffness) such that it produces appropri- ate loads at the valve’s closed and open position. At the same time, we should use no more force than necessary. Dealing with flat-tappet cams, experience indi- cates that, with regular service and prior to the removal of ZDDP from almost all oil blends (it happened about 2006), at no more than 220 pounds at full lift, a hydraulic flat-tappet cam is good for sev- eral hundred-thousand miles. At 240 pounds, about 100,000 miles is all you can count on. If a 75,000-mile life is acceptable, then the nose load can go to about 260 to 270 pounds. At 280 pounds, cam life can drop to 50,000 or less, and predicting gets to be a matter of chance. For an all-out race effort with a regular cast iron (non hard-faced) flat-tappet cam, consider 350 pounds over the nose as a limit.
So what can be done to improve the chances of a flat-tappet cam’s survival? First, you will need to use oils that have ZDDP in them. Most specialty oils do so because they blend to meet the hot rod- der’s needs, not the EPA’s oil standards. A few suggestions here are Amsoil, Quan- tum Blue, Red Line, Royal Purple, and Torco. (From that list you would think that colored oils are really popular!) You also should check with the cam company for oil additives that address the problem of ZDDP removal from the oils normally sold in auto stores. Another point here is that Lunati (and others) offers a cam nitriding service. This puts a hard layer on the cam lobe and makes for a highly wear-resistant surface. My personal pref- erence here is to always use a nitrided flat-tappet cam and then pair this with hard faced lifters.
For a short cam (less than 270 degrees) motor this makes for a near optimal component combination.
Other than lubrication problems, the most prominent cause of flat-tappet cam lobe failure is spring coil bind. On assembly, make sure the valve can go to full lift and still have at least 0.050-inch more lift available. If the coils clash together during high-speed operation, the springs will lose their force far quicker and may even break. To main- tain good spring performance, never float the valvetrain.
Another fix for regular cam wear, although you might not want to hear it, is a less-aggressive profile. In addition to profile selection, cam life also is greatly influenced by the break-in procedure. Use of the recommended cam lube is a must. Also, the lobe nose, which is the point of maximum wear, experiences the highest loads at idle. The first few min- utes of running are the most critical aspect of breaking in a new cam, be it street or race. Don’t let the motor run below 2,000 rpm for the first 20 minutes. Avoid idling a flat-tappet race cam at low RPM, even if your setup allows it.
If you have opted, as I suggested, for the nitride process on your flat-tappet cam, this will allow as much as 40 to 50 pounds more over the nose without sac- rificing life. This means 260 pounds over the nose, even on a more aggressive pro- file, will last indefinitely in a street motor. For the racer who invests in a quality cam, it’s good insurance that the cam will last longer than it needs to, saving the cost of a replacement. Seat bounce is caused by too high a valve-closing velocity. The best way to deal with it is to set the valve on the seat as gen- tly as possible. In this respect, hydraulic cams can be better than solids, but it’s not an open-and-shut deal. That, we’ll get to later. Next on the list are tight lash cams having up to 0.020-inch lash clearance. These are often, but not always, better in terms of seat poundage required to prevent seat bounce. Last are the regular lash cams, having clearance from 0.020-inch on up.
Retainer weight does make a differ- ence, even on a mild profile. Stock steel retainers weigh in at about 20 to 28 grams. The use of a 13-gram titanium retainer ups the RPM limit of a 6,500-rpm motor about 100 rpm. If the motor ran to 7,800 with steel retainers, the lighter retainer pushes that to about 8,000 rpm.
Remember, part of the advantage of a beehive spring is the 11-gram steel retainer it uses! Aluminum retainers work as long as they have an extra cross-sectional area to make up for the lower strength. If this is done, they can rival titanium at a much lower price. Be aware that you should avoid over-speeding the valvetrain with aluminum retainers, because they become a problem after such an event more easily than steel.
In the order of cost we have the choice of the following lifter options: solid flat tappet, hydraulic flat tappet, solid roller, and hydraulic roller. Unless your engine came equipped with hydraulic rollers, they’re out of our price range.
Solid rollers, however, can just about squeak in if we do our own porting and any other jobs that can, with a little effort, be tackled at home. The most likely and practical choice is a flat-tappet cam. Obvi- ously, the tappet style used reflects the choice of cam. When making a compari- son between a solid and a hydraulic cam, do not compare based on seat and 0.050 duration or lift figures, because they won’t directly cross over. Because the lash and the checking height are rarely com- patible, the duration figures—even at 0.050—can be a little misleading. The eas- iest way to make a comparison of cams from the same manufacturer is to com- pare the lift, which for the solid is arrived at by subtracting the required lash. Assuming the same LCA, a solid acts as if it’s about 5-degrees shorter than its hydraulic counterpart.
Stock Hydraulic Roller Lifters If you are rebuilding a post-1987 engine, then it more than likely has hydraulic roller lifters. Although these are fine in a stock application, where RPM much over 5,000 is irrelevant, they can have dire problems with collapse at higher RPM. On the Spintron I have seen as much as 0.100 inch of valve lift just vanish. In practice such a loss not only causes the lift to be inadequate but it also shortens up the duration seen at the valve. Experience on the dyno has shown this can reduce output by anything from (wait for this…it’s a shocker) 60 to as much as 120 hp! Unlike flat-tappet hydraulics, it seems the reason for the far- greater hydraulic collapse of roller lifters is that they experience a body-distorting side load during operation. This causes a far greater leakage between the internal components and results in the dramatic collapse seen. The question here is: what can we do to reduce or eliminate hydraulic lifter collapse be it for flat or roller lifters?
The normally recommended tech- nique to set hydraulic lifters is to adjust them until they are into the hydraulic action at about a quarter turn of the rocker adjusting nut. This may be fine for normal use since it ensures indefinite maintenance-free operation, but it is far from the best for high RPM power. Just as with solids, hydraulics can run into valveseat bounce. But in this instance, it can be caused by an increase in closing velocity because the lifter itself has short- ened up due to leakage and collapse. If we adjust the lifter almost to the bottom of its travel, there is now little room for it to col- lapse. Also, the volume of oil needed to refill it is very minimal so its recovery from any collapse is fast. After setting the lifters like this, you need to check the oil flow out of the pushrod because it is possible that one or two lifters out of any batch can restrict oil flow up to the rockers.
If you want to go one step better, then it is entirely practical to strip the lifter and pack it with a small spacer or washers (if you can find any to fit), and limit the plunger travel to about 0.010 to 0.015 inch. With this setup you would lash the lifters about 0.010 from the bot- tom of the travel.
The first reaction most people have when confronted with this lifter adjust- ment/modification option is that the lifter will now pump up. The bottom line is that if any lifter pump up is experi- enced, what it means is that the valve spring is not controlling the valve- train properly and is pumping up because of component separation. The fix is to re- evaluate the spring used.
A successful pushrod is certainly much more than just a lightweight, stiff component. These are requirements, but the pushrod’s ability to remain, as far as possible, unaffected by vibrations is of equal importance. The pushrod will be stiff if the steel is hardened close to its maximum, but it will “ring” far more eas-ily and adverse vibrations will pass back and forth between it and the valve spring.
If it is too soft it will flex and cause loss of controlled motion sooner. Within the structure of what we are trying to do here, a quality 5/16 pushrod will get the job done. Aftermarket chrome-moly pushrods are better and allow more power to be produced if the cam is of the higher intensity type. For all stock, or near-stock applications, a stock pushrod will work but it is a little too long if block and head work has been done.
Timing in the Cam
Here is a real world truth you need to keep well in mind when installing your cam: The more accurately the cam is spec’d, the more important it becomes to time it in accurately. It is easy to see why this must be so.
If we have a cam that is on the wrong LCA, then advancing or retarding the cam could make the intake more effective while making the exhaust less effective, or vise versa. For a number of degrees in either direction, one factor can cancel out the other, so the effect of mistiming the cam is minimal. If all the valve events are on the cam correctly, then mistiming the cam means all the events are off the mark and power will drop much faster. That said, the “correct” cam, timed in a degree or so out, will still make a lot more power than an “incorrect” cam optimally timed.
If the cam is not installed into the engine in the correct position relative to the crankshaft motion, all our work so far will be undone! Surprisingly, there is not a great deal of difference in what is required, whether the engine is near stock or virtually full race. Because of the great similarity of the heads from stock to race in terms of valve-size ratio and intake to exhaust flow ratios, we find that 4 degrees of cam advance pretty much universally hits the mark in a naturally aspirated engine.
OK, I have brought the term “advance” (and by inference “retard”) in here. Exactly what defines cam advance? Before going on here just flip back a few pages and re-familiarize yourself with the cam diagram on page 91.
To establish a cam position in the engine, we need a relevant reference fig- ure by which to gauge the amount of advance or retard a cam has when installed in the engine. The reference fig- ure we use is the cam’s Lobe Centerline Angle. If the cam is set into the engine to produce full lift at the same number of degrees after TDC on the number-1 inlet valve as the cam’s LCA, then the cam is said to be timed straight up. This means it is neither advanced nor retarded. For example, if a cam has a 108-degree LCA and is timed-in so that the intake valve on number-1 cylinder reaches full lift at 108 degrees after TDC, then that is “straight up.” If we advance the cam 4 degrees (that is make all the events happen sooner in relation to the crank rotation) the inlet valve would reach full lift at 104 degrees (108 – 4 = 104). If the cam was retarded 4 degrees (all events happen later in relation to the crank rotation) it would produce full lift at 112 degrees after TDC (108 + 4 = 112).
How important is it to get the cam timed in right? It’s a job you should do when installing a new cam, but there’s a certain amount of leeway within which you can work without losing any signifi- cant amount of power.
If you hit the mark within plus or minus 2 degrees, you are unlikely to feel the difference through the seat of your pants, but the drag strip clocks might just show a difference. It depends on whether you were 2-degrees-too-far advanced or 2- degrees-too-far retarded. For the most part, if a cam is too far advanced by a couple of degrees, the change in the power curve over the RPM band used for racing is almost unchanged. The clocks on the strip would show the same whether the cam was right on or 2 advanced. On the other hand, if the cam is 2-degrees retarded, you will see the car take longer to stop the clocks on the big end of the strip. If you want to contain any possible loss of output from a mistimed cam to about 0.3 percent, then figure that the tolerance you need to work within is +2 degrees of advance to –1 degree of retard over whatever is called for. For example, if the cam was supposed to be 104 degrees for 4 degrees of advance it would be OK from 105 degrees (1 degree- retarded over what it is supposed to be) to 102 degrees (2-degrees advanced over what it is supposed to be).
Installing a Roller Cam
If you decide to go with a roller cam setup in a pre-1987 block, then you will need to take care of the cam’s end float because it will not have the thrust plate that takes care of that as per post-1987 blocks. Be aware that you can install the earlier cam/timing set in a later block, but not the other way around. For non-roller blocks the cam will need a thrust button that bumps up against the timing cover.
Finally Making a Cam Selection
I went to some pains earlier to make a point on the uncertainty of cam selection the performance community generally faces. Since making a selection is close let me reiterate the points made earlier. Most performance books will tell you enough about cam selection to totally optimize the uncertainty of your choice. I don’t intend to do that here. What I have done instead is to team up with Lunati and, using its profile library, spec’d out a range of cams that are as near goof-proof in terms of application optimization as possible. Not only that, but with the way they are laid out there will be no uncertainty as to what you need to use once you have made up your mind exactly what your application is. The profiles I have chosen to use are biased toward reliability rather than out- right power. However, because the valve events are more accurate, these cams will almost certainly make more power than a cam with incorrect events and more aggressive profiles. As far as reliability goes, I am reasoning that, since we are all on a relatively tight budget, none of us can afford to have a motor that even vaguely approaches grenade status. What this means is that these timing numbers, which are the result of thousands of hours of test- ing, are within about 1.5 percent of opti- mum 98.5 percent of the time.
Also notice that all the cams are single pattern. That was done for two reasons. First, in this day and age, street cams need to address fuel consumption as well as power. Extending the exhaust timing to get more power above the 4,000-rpm mark costs mileage. Here a single-pattern cam is better. The second reason for all single- pattern cams is to simplify selection. These charts will deliver an optimal cam spec (for a given intake timing) more than 98 selec- tions out of 100 to within less than 2 per- cent. To put that into perspective, that’s at least 20 times more accurate than the cam industry at large. Don’t believe me? Try this: call five cam companies for a spec for your motor and see how many different recommendations you get. Can they all be right? With most companies you can do this test from their catalog. Just look and see how many vague suggestions seem to fit your motor requirements. If you are not yet convinced, consider this: in 30 years I have never lost a cam shootout on the dyno against cam companies.
If you are looking for some fine-tuning on the cam specs recommended here in terms of dual-pattern specs, more aggres- sive lobes, etc., I am planning to put a more comprehensive selection chart on my web site at MotorTecmagazine.net.
Read This Before You Make a Cam Selection
You don’t want to make a hasty cam selection and order something that, though it has the right valve event timing, doesn’t actually fit the block you have and has to be returned. Here are a few things to tell the tech guy to get a cam that will fit and function as intended.
First, spec the distributor gear. Any cam that is made from a cast-iron or cast-steel blank can run with a stock steel distributor gear. If you opt for a cam that has to be made from steel billet, as per most race solid roller cams, then a bronze gear will be needed for the distributor.
Remember that if you have a post- 1987 roller lifter block, then the cam nor- mally required for this has a stepped nose to accommodate a thrust plate. The tim- ing gear bolt-hole pitch circle is also smaller. This setup will not fit the earlier blocks but the earlier cam/timing gear setup can be used in the later block. In such a case, cam end float needs to be controlled by a thrust button as per the early block. Also, there may be some interference between the timing gear and the block. The photo on page 104 shows this and what to grind away for clear- ance. If you have a roller block, be aware that that the link bar of aftermarket solid or hydraulic rollers may seat down on the block before the roller has actually con- tacted the cam. A fix here is to grind about 1/16 inch off the offending area or use the later taller lifters.
If you are building a stroker motor there is a real possibility of rod shoul- ders hitting cam lobe flanks. The usual fix here is to order a reduced base circle cam for added clearance. If you are using the Scat rods that I recommend, there is far more rod-to-cam clearance than with most other rods so this post- pones the need for a reduced base circle cam until more stroke or a longer dura- tion cam is involved. I do not like to just go ahead and order a reduced base cir- cle cam until I am sure the setup actu- ally needs it because the cam dynamics are degraded and the contact loads increase when the base circle is reduced. For what it’s worth, most 33⁄4-stroke cranks with Scat rods clear most stock base circle cams.
If you have absorbed all this you should now be in a position to go ahead and use the selection charts that follow. To determine what cam is needed, go across the Application row until you hit the column that has your intended appli- cation (Quick Street, Fast Street, or what ever). Go down this column until you get to the row that has your engine’s dis- placement (as listed in the column far left). Where the row and the column cross is the cam your engine needs.
One last point here concerns cam duration and engine displacement. As the displacement goes up, the cam acts smaller. For instance a “true street” cam in a 350 will act like a “street and tow” cam in a 400. You can figure that for each 25 to 30 inches over 350-inches displace- ment, the cam will act as if it is about 6- degrees shorter. This gives you the option of stepping across to the right by one col- umn over what would have been the selection for a 350. Let’s do an example here. Say you want the characteristics of a “true street” flat tappet hydraulic cam (perfect idle, great low and mid-range with a respectable top end), then, for a 350, the chart would lead you to a [DV256-08HFL] spec cam. But you have a 383 and with those extra inches you could step across to a “quick street” spec and pretty much expect to get those same characteristics. This is an option that is open to you but don’t play the game to excess. It is better to have a cam slightly
Written by David Vizard and Posted with Permission of CarTechBooks