Standard 351 Cleveland push rods are 5/16” diameter and 8.41" long, but when the block is decked, when the heads are milled, when factory head gaskets are replaced by gaskets having a different compressed thickness, or when parts like the camshaft, lifters, valves or rocker arms are changed the required length of the push rods shall change as well.
Push rod deflection can cause many seemingly unrelated engine performance problems; they are the weakest link in an overhead valve type valve train. It is important to use push rods in any application that are rigid enough for the spring forces, for the weight of the valve train components, and for the engine speeds involved. The canted valve Cleveland valve train splays the push rods off to either side of the intake port; these push rod angles expose the 351C push rods to angular bending forces not encountered in the valve train of in-line valve motors; the 351C needs a sturdier push rod. The push rod is not the appropriate component to use for reducing valve train weight or saving money. Using push rods that are “overkill” for their application is my way of insuring the push rods are perfectly rigid and there’s no possible way they contribute to any valve train related reliability or performance issues. Push rods should be manufactured from seamless chromoly tubing. The use of chromoly tubing alone will guarantee a more rigid push rod. The larger the OD of the push rod the more rigid it shall be also, increasing the wall thickness of the tubing does not increase push rod rigidity as much as increasing the outside diameter. Push rods being specified for hydraulic tappet applications should have a 0.040" restriction in one end to control the amount of oil flowing to the valve train. Of course, restricting oil to the valve train via the push rods is not a concern if a motor is equipped with tappet bore bushings having 0.060” orifices.
5/16” push rods made from 0.080” wall thickness tubing are considered adequate for a relatively stock motor but I recommend a more rigid push rod. 5/16” push rods with 0.105” wall thickness are a step up in rigidity. 5/16" push rods made from 0.116” to 0.120" wall thickness tubing are a favorite choice of mine for hydraulic flat tappet applications (spring force up to 330 pounds over the nose) because the passage in the middle of the push rod is only 0.072" diameter. The small passage acts as a restrictor to control the amount of oil flowing to the valve train in lieu of a restrictor in the push rod's tip. The most rigid recommendation however is a 3/8” push rod with 0.080” wall thickness; this has been a common recommendation for 351C applications for decades.
Smith Brothers of Redmond Oregon (800-367-1533) and Manton Pushrods of Lake Elsinore California (951-245-6565) are shops specializing in custom made push rods. Manley Performance Products and Trend Performance are also good places to shop for push rods.
The factory rocker arms are suitable for the hydraulic tappet applications being discussed. There are two common warnings in using the factory rocker arms: (1) Use only steel 4V fulcrums (the 2V fulcrums are made of aluminum). (2) Beware of factory rocker arms that have “lugs” along the edges immediately above either side of the fulcrum area. There is a problem with push-rod clearance when using those rocker arms with camshafts lifting the valves 0.550” or higher, therefore they should be replaced. Sealed Power #R-855 is a recommended replacement for the factory rocker arms.
Beyond those warnings the factory rocker arm has three potential weaknesses: (1) fulcrum bolt stretch, (2) push rod cup wear and (3) the quality of the valve stem contact patch (a rocker arm geometry issue).
Fastening the rocker arms to the pedestals with ARP #641-1500 bolts (4 packs) and #200-8587 washers (2 packs) is recommended to improve the strength of the fulcrum bolts and reduce the possibility of them stretching. The 1/8” thick washers are necessary because the ARP bolts are 1/8” longer than the factory bolts. With the fasteners thus improved the factory rocker arm is good for up to approximately 400 pounds over the nose and it can accommodate applications lifting the valves up to 0.615” off the seat. 0.615” valve lift was Ford’s recommended limit for the production rocker arms based on push rod clearance.
If you wish to upgrade to adjustable valve train and your motor’s factory iron cylinder heads are equipped with unmodified slotted rocker arm pedestals the Scorpion #3224 rocker arm can be bolted directly to the unmodified pedestal and provide push rod cup type valve lash adjustment. This is a high quality billet rocker arm that operates like an individual shaft mount rocker arm. Keep in mind the 5/16” fasteners limit this rocker arm to spring force of about 400 pounds over the nose.
If your motor’s cylinder heads are milled and tapped for 7/16” stud & guide plate type rocker arms the Yella Terra YT-6321 rocker arm is the hot tip. This very rugged rocker arm also performs like a shaft mounted rocker arm therefore it requires no studs, guide plates or hardened push rods. Internet pricing for the Yella Terra YT-6321 rocker arm is in the range of $785 US dollars for a set of 16.
The next step up in price is the T&D Machine individual shaft mount rocker arm, which is available in steel, this is its main benefit. Whereas billet aluminum rocker arms are good for about 10,000 miles, a steel rocker arm is a better choice for an engine planned for high mileage.
ROCKER ARM GEOMETRY
There are six variables which impact the geometry of a rocker arm; (1) the amount of camshaft lobe lift, (2) the design of the rocker arm, (3) the height of the rocker arm's fulcrum, (4) the rocker arm's lateral distance from the valve stem, (5) the height of the valve stem and (6) the length of the push rod. Optimum rocker arm geometry minimizes side thrust on the valve stem and guide which has two substantial benefits; (1) it minimizes the wear of parts AND (2) it minimizes the rocker arm’s contribution to oscillation induced valve train problems.
Geometrically ideal rocker arm geometry will set the rotational axis of the rocker arm at the same height (perpendicular) as the valve tip when the valve is 50% open. That’s just on the rocker tip side, there is also geometry on the push rod side, but getting close to the correct geometry on that side depends upon the rocker arm being designed with that as a consideration, and designed for the amount of lobe lift employed by your motor’s camshaft. When the geometry is correct on both ends the rocker arm will impart the most possible lift to the valve, this will not occur unless the geometry is correct at the rocker arm tip AND the push rod. This indicates the valve train is following the motion of the camshaft lobe most precisely, which is one of the primary goals of a high performance valve train.
Correct geometry at the rocker tip will place the sweep of the rocker tip nearest the rocker arm at fully closed and fully open, the sweep will be furthest from the rocker arm at 50% open, and the rocker tip shall be in the middle of its sweep at approximately 25% and 75% open. This geometry will always result in the narrowest sweep pattern, although there is nothing beneficial about a narrow sweep pattern, it is just a method of evaluating the rocker arm geometry. This description of sweep pattern will be in direct opposition to many of the rocker geometry instructions you shall run across. A few of the camshaft companies are notorious for promoting bogus rocker geometry instructions. The hot rod industry teaches home mechanics (and professional mechanics too) to focus on setting the rocker arm's contact patch on the valve tip, by manipulating the rocker arm's height and the push rod's length, at the expense of other concerns. This may achieve the most rudimentary aspects of rocker adjustment, and it may be convenient, but it cannot possibly result in an ideal adjustment. The most rudimentary aspects of rocker arm adjustment simply keep the operation of the rocker arm within four parameters; (1) the rocker arm should not contact the valve spring retainer when the valve is fully closed, (2) the rocker arm should not contact the push rod when the valve is fully open, (3) the rocker arm slot should never "bottom-out" against the fulcrum, saddle or stud at either extremity of its motion, and (4) the rocker arm tip should never bear down upon an edge of the valve tip; its sweep pattern does not have to be perfectly centered on the valve tip but it should contact the valve tip in the middle half of the valve tip's surface.
As you assemble a cylinder head you can detect rocker geometry and push rod length problems early on by paying attention to the valve stem heights; the valve stem heights should be equal across the cylinder head. If the valve stem heights are unequal, or if one particular valve stem is higher or lower than all the others, you SHALL run into problems.
There are two types of rocker arm designs to consider, the first is the stud mounted, push rod guided type of rocker arm. The height of stud mounted rocker arms is set by the lash adjusting nut (aka the poly lock). Adjusting lash with this type of rocker arm alters the rocker arm's height, and impacts the rocker arm's geometry. In order to maintain consistency in push rod length stud mounted rocker arms are best adjusted mounted on the engine in conjunction with a fixed length push rod.
The other type of rocker arm is the fixed-pedestal mounted type of rocker arm that fastens securely to the cylinder head's rocker arm pedestal. A fixed-pedestal mounted type rocker arm can provide lash adjustment just as easily as the stud mounted variety, by employing a push rod cup style adjuster. The factory rocker arm and the two Yella Terra rocker arms are all of this second type of rocker arm. The high-end T&D and Jesel shaft mount rocker arms are also fixed-pedestal mounted rocker arms. All fixed-pedestal mounted rocker arms are in fact a type of individual shaft mounted rocker arm; they are more stable and contribute fewer rocker arm induced problems to the valve train as long as the saddle/fulcrum is rigid enough. The height of most fixed-pedestal mounted rocker arm is raised by shimming the rocker arm fulcrum/saddle; it is lowered by removing material from the fulcrum/saddle or by removing material from the pedestal cast into the cylinder head. However Yella Terra offers saddles of varying height for their premium YT-6321 rocker arm. This is a very attractive feature of those rocker arms. Increasing valve length also has the same effect as lowering the rocker arm. This type of rocker arm makes it possible to adjust the relationship between the rocker arm tip and the valve tip independent of the push rod, with the cylinder heads sitting on your work bench.
If you are using the factory rocker arms and determine their geometry requires adjustment, a good starting point is to set the height of rocker arm to position the fulcrum’s pedestal approximately in the middle of the rocker arm slot at 50% valve lift. Do not Tufftride the factory rocker arm parts until after the rocker arm geometry has been sorted out.
It is popular to test rocker arm adjustment with the heads assembled on the short block by coloring the valve tips with a felt tip marker, assembling the valve train with the push rods set to zero lash, hand rotating the crankshaft through two revolutions and inspecting the contact patch pattern on the valve tips. As far as I am concerned, the contact patch does not need to be centered on the valve tip, it just needs to stay away from the edges.
Push Rod Length
Sorting out the rocker arm geometry is a prerequisite for determining push rod length. Due to the age of Cleveland series motors, (1) the original manufacturing tolerances can result in dimensional differences, (2) parts have been mixed and matched over the decades, or (3) some parts have already been refurbished once or twice and worked on by many hands of various skill level. For these reasons you may find each cylinder head requires a different push rod length. The actual length of the push rods you shall order for the engine shall be the sum of the length of the longest or shortest “zero lash” push rod plus a small additional amount. This small additional amount added to the length of the push rod establishes the hydraulic tappet adjustment; i.e. the amount you plan to compress the hydraulic tappet plunger.
The factory fixed-pedestal mounted rocker arms, and stud mounted rocker arms require consistency in rocker arm height amongst the all the rocker arms on each cylinder head (both cylinder heads if possible) so that the push rod length required to set all of the rocker arms at zero lash is within a few thousandths of an inch per cylinder head. The length of the longest push rod required to set all of the rocker arms at zero lash shall be the basis for determining what length of push rods to order.
The fixed-pedestal mounted rocker arms equipped with push rod cup adjusters (such as the Yella Terra rocker arms) do not require as much consistency because the adjustable push rod cups will make up the differences. Start with the push rod cup adjusters screwed all the way into the push rod tips and find the rocker arm requiring the shortest push rod to achieve zero lash. The length of this shortest push rod shall be the basis for determining what length of push rods to order. There is a limit to how far you can screw the adjusters out, so keep an eye out for big differences and resolve any problems.
Hydraulic Tappet Adjustment
One rule of thumb for adjusting hydraulic tappets is to compress the plunger 1/2 of the plunger’s available travel; however it is important to measure the travel of the plunger if that’s your plan. The plunger of a modern hydraulic tappet does not compress as much as the plungers did decades ago. The plunger travel of a Crane roller tappet is only 0.062”. The plunger travel of a 1995 Johnson HT900 tappet I have on hand is 0.125”, whereas the plunger travel of a 1970s vintage HT900 is tappet is 0.187”. I am told the plunger travel of a typical modern hydraulic tappet is in the range of 0.060” to 0.080”. Nowadays the recommended range of hydraulic tappet adjustment when using stud mounted rocker arms is 1/8 to 1/2 turn of the adjusting nut beyond zero lash when the engine is hot (adjustable rocker arms are usually mounted on studs with 3/8-24 or 7/16-20 threads). Decades ago the spec for adjusting small block Chevy tappets was one full turn beyond zero lash, which was supposed to set the tappet plunger in the middle of its travel! This means the plunger in Chevy’s tappet had 0.145” of travel.
Some tappets (such as the Morel HLT hydraulic roller tappets) utilize intentionally limited plunger travel as a method to increase the rpm capability of the tappet. This requires adjustable valve train, and push rod length should be determined following the instructions of the tappet manufacturer.
Longer valves are sometimes required (or at least convenient) for solving 3 problems that crop up when installing a higher-lift camshaft. Longer valves (1) increase the distance between a valve spring retainer and the top of the valve guide, they (2) provide the additional height needed for valve springs which have an installed height that is higher than the installed height of the OEM valve spring, and they (3) raise the height of the valve tip which can be a better choice than lowering the rocker arm when adjusting rocker arm geometry.
Manley severe duty stainless steel valves for the 351C are available off the shelf in +0.100” lengths;
FLAT TAPPET CAMSHAFT SECTION
Flat Tappet Camshaft Issues
On occasion a flat tappet camshaft fails prematurely, usually during break-in or soon after a motor is placed in service. This is something we must consider when choosing to use a flat tappet camshaft. Decades ago we installed flat tappet cams in our motors and never thought twice about the possibility of premature failure. When flat tappet cams fail prematurely today there is a logical reason behind the failure. I believe the failures must boil down to one of four conditions:
(1) A lubricant issue (i.e. insufficient wear protection)
(2) A quality control issue (i.e. the parts were not made of the same materials used decades ago or were not surface hardened properly)
(3) A performance issue (i.e. the flat tappet lobes of today’s street cams have faster lift rates and utilize more valve spring force, therefore the cams wear like race cams did decades ago)
(4) Improper break-in (camshaft lobes are splash lubricated, in order to insure adequate lubrication during break-in the motor must be run above 2000 rpm as soon as it is started up)
Considering those 4 possible reasons for premature flat tappet camshaft failure, my strategies to prevent possible failure are:
• Avoid using the highest lift rate camshaft lobes or unreasonable valve spring force. Valve lift no more than about 0.570" theoretical (i.e. about 0.550" factual). Hydraulic intensity no less than "about" 52, i.e. stay away from Comp Cam's Extreme Energy cams and Lunati's VooDoo cams. Major intensity (solid tappet) no less than "about" 44. Valve spring force no more than 130 pounds seated or 330 pounds over the nose.
• Purchase the cam from a trustworthy grinder. There is one manufacturer who is (in my estimation) the source of 95% of all failed valve train parts.
• Custom order the cam requesting the cam grinder's best surface hardening treatment (nitriding) and best lobe polishing.
• Cam cores come in different quality levels, the lobes are narrower in some cases, and lobe taper may vary. A quality cam should have .002" taper on the lobe to aid in lifter rotation and break-in. Yet economical cams may only have about .0005" taper. So when you're custom ordering the cam touch upon the subject of core quality and lobe taper. Make sure you specify the best quality cam core, and specify 0.002" lobe taper.
• Use flat tappets manufactured in North America or Australia with trustworthy quality (Johnson HT900 for instance).
• Use motor oil having very high wear protection properties for both break-in AND normal operation (Valvoline VR1 for instance).
• NEVER use break-in oil because break-in oil has low wear protection properties. Break-in oil is not intended for cam lobes, it is intended to help rings seat, but modern rings and modern cylinder honing techniques preclude the need for break-in oil.
• NEVER use an oil additive; the high zinc ZDDP additives diminish the wear protection properties of a good motor oil.
• Run the motor above 2000 rpm during the entire 30 to 45 minute break-in period of the camshaft to insure the camshaft and tappets are "splash lubricated" adequately.
Distributor Gears for Flat Tappet Camshaft Applications
Iron camshaft cores, such as the cores used for all 351C flat tappet camshafts, are compatible with the original equipment distributor gears found on both factory and aftermarket 351C distributors.
Hydraulic Flat Tappets
The Speed Pro (Johnson) HT-900 hydraulic flat tappet has been a reliable choice for decades. Johnson once boasted of the superior heat treatment of their tappet, they also claimed decades ago when the 351C was a popular motor that their tappet metered oil properly for the 351C. The tappet is sturdy enough for performance usage and sturdy enough for the weight, the valve spring forces, and the canted valve geometry of the 351C 4V valve train. It is also available as an anti-pump-up lifter, part number HT-900R, which requires adjustable valve train.
Valve Springs For Flat Tappet Camshafts
The 351 Cleveland is equipped with a “big block” style valve train composed of large - heavy valves, large-heavy springs and spring retainers, and high ratio rocker arms. According to Crane Cams achieving the best compromise between performance and acceptable valve train wear with a flat tappet “big block” valve train such as this requires setting the valve spring force between 115 to 130 pounds on the seat and no more than 330 pounds over the nose.
The best valve spring for flat tappet street applications I am aware of at this time is Crane Cams #99839, which is a single spring with damper style valve spring. This spring was designed for AMC V8 applications, which is a motor with a “big block” style valve train similar to the 351C valve train.
ROLLER TAPPET CAMSHAFT SECTION
Hydraulic Roller Tappet Valve Train Issues
A Crane hydraulic roller tappet is 44% heavier than a Johnson HT-900 hydraulic flat tappet (148 grams verses 103 grams). Taking into account the Cleveland 1.73:1 rocker arm ratio the heavier roller tappet is predicted to reduce the rev limit of a motor by 650 rpm. Hydraulic roller camshaft lobes also lift valves open at a higher lift rate than the lobes of hydraulic flat tappet camshafts. Lifting a heavier valve train component (i.e. the roller tappet) at a faster rate increases the inertia of that component and makes the roller tappet more likely to lose contact with the camshaft lobe at maximum lift when the camshaft lobe’s nose stops lifting the tappet (objects in motion tend to stay in motion). These are the reasons why hydraulic roller cams can negatively impact the high rpm capabilities of a motor, why hydraulic roller camshaft valve trains have more instability problems, and why they require more valve spring force (both seated and at maximum lift) to maintain valve train stability.
The roller and roller axle of a hydraulic roller tappet are splash lubricated as opposed to pressure lubricated. The amount of splash lubrication occurring at idle or low engine speeds is insufficient for the roller and axle to sustain heavy loading, therefore the wear rate of those parts increases as spring forces increase. Although a hydraulic roller cam valve train requires additional spring force to maintain valve train control and stability, the amount of spring force that can be applied is not limitless. With a high-lift hydraulic roller cam we walk a line between applying sufficient valve spring force for good control of the valve train yet keeping that force light enough for acceptable roller tappet wear.
As a camshaft lobe turns beneath a tappet, lifting the tappet at the same time, it imparts a side thrust force against the tappet, in effect trying to push the tappet against the wall of the tappet bore. This is due to the fact that the ramps and flanks of a camshaft lobe are ground at angles; a lobe does not contact a tappet in such a manner as to push it perfectly upward inside the tappet bore. The angle of the side thrust (and therefore the strength or magnitude of the side thrust) acting upon the tappet is dependent upon the radius of that part of the tappet that contacts the camshaft lobe. A flat tappet’s face is ground on a 50” radius, whereas the roller of a roller tappet has about a 0.35” radius. This is a critical difference between flat tappets and roller tappets. A flat tappet has very little side thrust acting upon it because the angle of that thrust is practically parallel to the axis of the tappet bore. A roller tappet on the other hand has a significant amount of side thrust acting upon it pushing the tappet directly against the tappet bore, making the roller tappet’s body prone to distortion. The internal parts of a hydraulic tappet are some of the most precision manufactured parts in the entire motor, the clearances are critical, the tappet cannot function properly if the body distorts. Thus it is critical that the body of a high performance roller tappet is made sturdy enough to prevent its distortion even when subjected to higher valve spring forces and higher engine speeds.
Distributor Gears for Roller Camshaft Applications
Steel camshaft cores, such as the cores used for roller cams ground by Crane Cams and Bullet Racing Cams, require a compatible steel distributor gear. Crane Cams manufactures the steel roller cam cores used by all the cam grinders, and they also manufacture the proper steel distributor gear for use with camshafts ground on their cores. Crane #52970-1 is the gear for 0.500” distributor shafts; Crane #52971-1 is the gear for 0.531” distributor shafts. The gear for 0.531” shafts is also available via Ford Racing Performance Parts under part number M-12390-J.
Hydraulic Roller Tappets
Although the pricing is tempting the Ford factory 5.0 HO hydraulic roller tappet is not recommended for use in your 351C. The Ford tappet has been problematic in 351C applications. There are four reasons for this: (1)The 351C valve train is heavier than the valve train the 5.0 tappet was designed for; (2) the 351C valve train utilizes higher valve spring forces than the 5.0 tappet was designed for; (3) the 351C valve train geometry and splayed push rods subject the tappet to side thrust forces greater than the forces the 5.0 tappet was designed for; and (4) the waist machined into center of the 5.0 HO tappet is too high, it has been found to rise above the top of the lifter bore at maximum lift and dump the engine’s oil pressure in some 351C blocks.
The aftermarket hydraulic roller tappets sold by Crane Cams and Morel are known to operate reliably in Cleveland applications. The waists machined into the center of these tappets do not rise above the top of the lifter bore at maximum lift, and the tappet bodies are thicker and therefore resist distortion (with the penalty of increased weight).
Crane Cams manufactures one hydraulic roller tappet for 351C applications, part number 36532-16, and it’s a good one. Crane’s tie-bar style roller tappet is machined from 8620 steel billet and it is heat treated. A precision fit plunger assembly is used to provide the proper bleed-down rate, permitting high RPM use in properly set-up engines. The strength of the heat treated 8620 material prevents distortion of the lifter body, thus permitting more consistent operation in high spring pressure and in high RPM applications, due to the consistency of the plunger to tappet body clearance. Crane hydraulic roller tappets weigh 148 grams (that’s half the weight of a pair). Internet pricing for a set of Crane’s tappets is in the range of $635. This is my preferred hydraulic roller tappet.
Morel does not sell their tappets directly to the consumer; their tappets are sold via a network of retail businesses several of them being cam grinders, including Lunati in the US and Crow in Australia. I have not been able to verify the weight of Morel’s tappets. Morel manufactures three 0.875” OD hydraulic roller tappets for 351C applications:
(1) Morel hydraulic roller tapper #5323. This tie-bar style roller tappet is described as a “street” tappet with an upper rpm limit in the range of 6200 rpm to 6500 rpm. Internet pricing for a set of these tappets is in the range of $380.
(2) Morel hydraulic roller tappet #5327. This tie-bar style roller tappet is described as a hydraulic-limited travel tappet (i.e. HLT). I assume this means they are designed for higher rpm and that they require adjustable valve train. Internet pricing for a set of these tappets is in the range of $505.
(3) Morel hydraulic roller tappet #5879. This tie-bar style roller tappet is described as a “pro” high rpm HLT tappet. It is designed for oil viscosity no greater than 5W/40. Since it is a “HLT” style tappet I assume it requires adjustable valve train. Internet pricing for a set of these tappets is in the range of $830.
Valve Springs For Hydraulic Roller Tappet Camshafts
PAC Racing Springs is a small division of the Peterson American Company (i.e. PAC) the largest spring manufacturer in the USA. They manufacture the ovate wire beehive valve springs that have become so popular in the performance industry. The ovate wire beehive valve springs are manufactured in two series, the 1200 series and the 1500 series. The 1200 series valve springs are the budget springs. The 1500 series valve springs are nitrided, polished and nano-peened; they are easily identified by their GOLD COLOR. If your bee hive valve springs are not gold colored, they are not the springs I am recommending, and you must live with the consequences of YOUR choice. The 1500 series valve springs cost about 25% more than the 1200 series valve springs, they are the springs I recommend. There are also low priced substitute beehive springs on the market … buyer beware.
The #1520 Big Block Chevy beehive spring manufactured by PAC Racing Springs is a good choice for 351C hydraulic roller cam applications, since the Big Block Chevy’s valve train is very similar to the 351C valve train.
PART 8 - PREPARING A 351C RACING ENGINE
The production 351C was never intended for high rpm racing (8000+ rpm) but that didn't stop people from doing so. When the production engine is set-up for racing (excepting the connecting rods) it will withstand those sort of engine speeds and higher for a while before something breaks. The 351C was duty-cycle-tested up to 7000 rpm which was a high rpm duty cycle for an engine intended for mass production circa 1968. Based on my experience, and being conservative, I'd say the production block, crank, connecting rods and cylinder heads are good for many years of racing if engine speed is limited to about 7000 rpm; even the 2 bolt main caps resist "walking" at 7000 rpm! However, for reasons I shall explain below I don't recommend spending money to prepare the production connecting rods for racing unless the rules require using them. There are two caveats regarding the production engine block: the lubrication system and the thin cylinder walls require steps taken to amend their shortcomings. Another consideration, any partially counter-weighted crankshaft that has been designed for maximum bob-weight instead of minimum bearing load increases the loading of the second and fourth main bearings and bulkheads, and cracking of those bulkheads is a possibility. In the end, the durability of a 351C racing motor shall hinge upon the supporting parts that are selected and the time, money and detail invested in preparing it. I don’t claim to be an expert. But if you may find what I've learned over the decades helpful, here's the synopsis. The following set-up info is good for all types of competition excepting drag racing.
I'd prefer to build a racing motor around a heavy duty block designed for that purpose; for instance the US manufactured racing block known as either the SK block or the 366 block, the Australian manufactured XE192540 NASCAR block or the new Tod Buttermore block. The heavy duty block would be a more durable choice, having thicker cylinder walls and thicker bulkheads. These blocks are less likely to fail during the abuse of racing, therefore they make a good insurance policy against wasting the money you've invested in preparing the race engine. The price of replacing racing parts is expensive, as is the price of machine work and the price of assembling a racing engine. If one or two production blocks fail over the course of several racing seasons then using a production block would end up costing more money in the long run. Thicker cylinder walls make it possible to use more compression and higher engine speeds. Sturdier bulkheads make the block more compatible with a less expensive partially counter-weighted crankshaft such as the factory crank or a "sportsman" crank ... although I would still prefer to use a fully counter-weighted crankshaft if it is in the budget.
If I planned to use the iron 4V heads then the block material would be iron as well, whereas aluminum heads can be mated to an iron block or an aluminum block. Besides the weight reduction additional horsepower can usually be coaxed from aluminum high-port racing heads (not because they are made of aluminum, but because they have higher ports and possibly high swirl combustion chambers), but there is a limited selection of intake manifolds available for the high-port heads and they require custom manufactured exhaust headers in many applications. Iron racing blocks are both sturdier and less expensive than aluminum blocks. However a heavy duty aluminum block from Tod Buttermore and a set of aluminum heads shall reduce the weight of a race car by a significant 200 pounds (91 kilograms). Of course the benefits of weight reduction must be weighed against the higher price and the lesser durability of the aluminum block. Some guys argue in favor of an aluminum block by pointing out it is often repairable when damaged whereas an iron block is not.
If I intended to use the production block, I would accept the compression ratio and engine speed limitations inherent in that choice. The De Tomaso factory determined circa 1973 that to avoid failure of their racing engines employing the production block they had to limit those engines to 7000 rpm and 10.5:1 static compression. Operating within those limits the engines produced about 500 horsepower. There are choices however in parts and machine work that shall reduce the possibility of failure or raise the operating limits of the production block.
In terms of preparing the block for racing I’d sonic check the cylinder walls to establish their thicknesses; insuring the walls are at least 0.120” thick on the thrust sides after boring and at least 0.080” thick on the non-thrust sides after boring. I'd have the crankshaft main bearing saddles align honed. I'd level the block's decks, setting the decks up with a finish compatible with multi-layer steel (MLS) head gaskets. I'd index the boring machine to the crankshaft's axis during the boring process to insure the cylinders are perpendicular to the axis of the crankshaft. A piston trying to stroke up and down in a cylinder that is canted to the front or rear must operate in a “wedged” manner that puts an abnormal load on the cylinder walls and causes floating wrist pins to hammer out their locks. A piston will operate in a cocked manner if a cylinder is canted to the left or right which again puts an abnormal load on the cylinder wall. This abnormal cylinder wall loading contributes to cylinder wall cracking, therefore indexing the boring machine to the crankshaft's axis helps to alleviate cracking of the production block's thin cylinder walls. It also reduces frictional losses and makes more horsepower! The cylinders would also be bored and honed with head plates and main bearing caps torqued in place for the best possible ring seal, which also makes more horsepower. I'd install lifter bore bushings in all 16 lifter bores if the block incorporated the factory 351C lubrication passages (the Buttermore block incorporates a main priority lubrication system). The bushings would have 0.060" orifices. I'd install cam bearing oil passage restrictors at all 5 cam bearings no matter what block I'm using, also with 0.060" orifices. I would use MLS head gaskets for racing. The production block’s rear main seal is a rope seal; I’d replace it with a neoprene seal which requires pulling a small pin from the seal groove in the rear main bearing cap and filling the pin hole with a dab of sealant. The main bearing caps and the heads would be clamped with studs instead of the factory bolts. Limiting the production block to a static compression ratio in the range of 10.5:1 to 11.0:1 (8.0:1 dynamic compression) is another measure that can be taken to prevent cylinder wall cracking. I know guys who would scoff at the idea of racing with the production block set at 11.0:1 static compression, but I'm not sure if they installed full round skirt pistons in their race motors. Building the motor around a racing block having thicker cylinder walls provides more latitude to set the compression ratio higher, if the fuel the motor shall be operating on and the cylinder head combustion chamber design allowed it. I'd prefer to lubricate the motor with a dry sump lubrication system because a wet sump system is not ideal for coping with the g-forces encountered while cornering, accelerating and braking on modern race tires. However, if I intended to use a wet sump lubrication system then I'd plan to use a high capacity oil pan which incorporates baffles with hinged doors, a windage tray and a scraper. The wet sump system would also incorporate a high capacity oil accumulator (i.e. an Accusump). Regardless if the car is equipped with a dry sump lubrication system or a wet sump lubrication system it MUST be equipped with an oil cooler.
The aftermarket stroker crankshafts are manufactured as inexpensively as possible using Chinese castings or forgings, the quality of their machine work is adequate at best (and often inadequate), and their quality control is poor. I would not consider using a crankshaft manufactured to such standards for a street engine OR a racing engine. Nor would I use a crankshaft with more than a 3.50" stroke, the additional crank-arm leverage & piston speed is not beneficial in terms of sports car racing, road racing, track racing or circuit racing. In regards to selecting a crankshaft for a racing engine, there are three viable choices. Choice number 1: The first choice is to utilize the production nodular iron crankshaft. The production crank has a track record of quality and durability. Since the factory crank was externally balanced I would have it internally balanced, which increases the price of using the factory crank. Since the production crankshaft was designed for maximum bob-weight as opposed to being designed for minimum bearing load it heavily "loads" the second and fourth main bearings and bulkheads during ultra-high rpm operation. This means I'd choose to limit engine rpm to about 7000 rpm. Choice number 2: The second choice is to purchase a mid-price "sportsman" style forged steel crankshaft. A forged steel crank should be tougher than a cast iron crank, but realistically we must keep in mind they are based upon Chinese forgings. The good ones are machined in the US by reputable companies. The sportsman crank should come from the manufacturer internally balanced, however like the production crankshaft a sportsman crank is only partially counter-weighted. So before you purchase one shop around, talk to the engineers who designed them, and make sure you purchase one that has been designed to minimize bearing loads. Otherwise it shall heavily load the second and fourth main bearings and bulkheads during ultra-high rpm operation just like the factory crankshaft. Some crankshafts also offer improved rod bearing lubrication passages, which is another topic to discuss with the engineers before you make your choice. Choice number 3: The third choice is to purchase a very expensive, fully counter-weighted, steel crankshaft (forged or billet). The fully counter-weighted crank is best at reducing the "loading" of the second and fourth main bearings and bulkheads during ultra-high rpm operation. The bending deflection across the center main at high loadings and high engine speeds causes measurable power losses in engines equipped with partially counter-weighted crankshafts. Therefore the benefits of a fully counter-weighted crankshaft are less stress on the engine block and the reduction of power losses, i.e. an increase in power output!
Regardless of which crankshaft choice I make, I'd have the crankshaft magnafluxed, tufftrided, polished, and dynamically balanced (remember the production crank should also be internally balanced). If the motor is set up for 10W, 15W or 20W oil then I'd use fully grooved copper-lead alloy main bearings (Mahle/Clevite MS-1010P). Acquiring fully grooved bearings requires using the upper halves of two sets of standard bearings. If the motor is set up for 0W or 5W oil then I'd use 3/4 grooved copper-lead alloy main bearings (King Bearing # MB5169HP). The main bearings would be set-up with 0.0009" to 0.0011" clearance per inch of main bearing journal diameter which is how they were set-up 40 years ago and is still fairly common these days. Using the factory crank with a heavy piston & rod combination will require more rod bearing clearance than what is customary however (0.0011" to 0.0013" clearance per inch of rod bearing journal diameter). I'd use the ATI #918920 neutral balanced steel crankshaft damper. This damper has a reputation for preventing cracking of the second and fourth bulkheads when the factory crankshaft is used. I prefer the durability of a light weight neutral balanced steel flywheel (Yella Terra YT9902N) over the additional weight-loss of an aluminum flywheel.
I'd use piston and rod assemblies with floating pins, my preference being to use 6.00" long connecting rods. The 6.00" rods are gentler on the cylinder walls, they are gentler on the piston skirts, they rock the pistons less in the bores and since they require "shorter" pistons the weight of the pistons is reduced. As long as the rod length to stroke ratio does not exceed 1.72:1 the longer rods will not impair acceleration or create induction system "lag" issues. The 6.00" rods are actually small block Chevy rods; using such rods requires a crankshaft with 2.100" rod journals in order to compliment Chevy diameter rod bearings. You'll find that is the standard journal diameter for aftermarket crankshafts. The standard big-end width of a Chevy connecting rod is 0.940" however. Chevy rods in which the big-ends had been narrowed to 0.831" in order compliment Ford width rod journals (intended for use with Mahle/Clevite CB1227 rod bearings) were once readily available, but that no longer seems to be the case. You may have to order custom rods for this application, but don't let that deter you because there are several excellent choices in reasonably priced custom manufactured rods on the market; the sport rods from Howards Racing Cams are an example.
It has been customary to turn-down the rod journals of the factory crankshaft to 2.100" when using it with 6" connecting rods, but if you must custom order the 6" rods why not specify having the rods machined for standard 351C size rod journals and standard 351C rod bearings? Another option in a 6" long connecting rod (actually 5.956") that is compatible with the standard 351C rod journal diameter (2.31") is the Eagle H-beam connecting rod for the 351W, #CRS5956F3D. The crankshaft’s rod bearing journals do not require being re-ground because the 351W has the same size rod journals as a 351C (2.311”). The 351W uses a Clevite CB831 rod bearing however, which has a different shell thickness but the same width and ID as a 351C bearing. The 351W rod also uses the same size wrist pin as a 351C (0.912”).
However, for those preferring to use 5.78" production length connecting rods (and therefore production 2.31" diameter rod journals) it doesn't make financial sense to use the factory rods for racing unless the rules require them. The factory rods require magna-fluxing, shot-peening, 180,000 psi rod bolts, re-sizing the big ends and installing bushings in the small ends for floating wrist pins. Thus prepared the factory rods still lack locating dowels for the big end caps, and are only reliable to about 7200 rpm. The price difference between setting up a set of production rods in that manner and purchasing a set of Eagle #CRS5780F3D H-beam rods is almost nil, yet the Eagle rods are better quality rods, and are reliable at higher rpm (the Eagle rods use Mahle/Clevite #CB831 351W bearings).
Regardless of the length of the connecting rods, they should be used in conjunction with full round skirt forged flat-top endurance racing pistons. The combination of a 6" rod and a round skirt piston has an excellent track record for preventing cracking of the production block's thin cylinder walls. The Ross pistons are currently available for 4.030" bores in pin-heights for factory length connecting rods or 6" long connecting rods from Summit Racing at a very good price. The Ross #80556 pistons with 1.668" pin height are for production length rods, the Ross #80566 pistons with 1.446" pin height are for 6" long rods; the second pistons also use Chevy diameter wrist pins. Using a 351W rod will require a custom round skirt piston having a pin height in the range of 1.47” to 1.495”.
Assuming the cylinder heads are designed to use 351C valve train parts, I would use Yella Terra YT6321 or T&D Machine #7200 or #7201 rocker arms. It is common these days to employ 1.8:1 ratio rocker arms for the intake valves and 1.7:1 or 1.6:1 ratio rocker arms for the exhaust valves; the T&D rocker arms are available in several rocker arm ratios. On the intake side I'd use Manley's #11872-8 light weight race master 4V stainless steel intake valve with a titanium spring retainer. Of course, if I wanted to maximize the life of the valve train I'd opt for titanium intake valves which would allow me to select softer valve springs or rev the motor to higher rpm. On the exhaust side I'd use Manley's #11805-8 severe duty 4V stainless steel exhaust valve with a chromoly spring retainer. One caveat here, when juggling the weight of the valves its safer to set-up the valve train so the intake valves are the first to float, because it’s usually the exhaust valves that hit the pistons first so you want to avoid floating the exhaust valves. I would not bolt the cylinder heads on the motor until the rocker arm geometry is set-up properly (see my notes on this in the valve train section above). The pedestal mounted rocker arms I've recommended make this possible. I'd operate the valve train with 3/8" OD push rods made from 0.080" wall thickness seamless chromoly tubing. I'd select a camshaft with lobes that are as conservative as possible in terms of ramp design and lift rate, while keeping the motor competitive. It is important to realize that not all camshafts are created equal; some lobes are tougher on the valve train than others. PAC Racing valve springs would be selected to complement the cam, tappets and the application.
I'd set the rev limiter of a race engine using the production block somewhere around 7000 to 7200 rpm. A heavy duty block employing a fully counter-weighted crankshaft can rev much higher. Rock-n-Roll!