Valve Train Dynamics

There are a lot of urban myths re vintage valve train components, not necessarily well understood. As I am in process of rebuilding my XPAG, it seemed best to understand the dynamics of the valve train and its components (valves, springs, lifters, pushrods, etc.) in order to take advantage of modern performance and reliability related advances in design and materials that contribute to improving valve train stability. Valve train stability is the end game necessary to generate horsepower at higher RPM range when trying to improve reliability or performance. This discussion has been put together to consolidate previous, possibly disparate fragments of earlier information into one comprehensive discussion.

Remember:

Max HP: occurs at max engine airflow and max fuel burn at stoichiometric fuel air ratio.

Stoichiometric (in lay terms): the correct amount of a reactant, in this case fuel, to completely chemically react/burn with another reactant, in this case air, so there is no leftover residual unburned air fuel mixture remaining.

Max Torque: occurs at max unit air charge in the cylinder. This is best valve timing for max unit air charge. 

Also remember, as vintage engines they need all the help available from modern technology to reduce unwanted stresses/wear from being introduced into the engine. Keep it alive for another 70 years!

1. LIFTERS: Many out dated myths revolve around premature lifter base wear with warnings to stick with OEM spring strengths, even to using only single outer spring(s). (This is a bad idea because the inner spring, wound in opposite direction of the outer spring, acts as a dampener mitigating harmonics of a single outer spring). Removing the inner spring and using only the outer spring should be limited to goal of reducing spring pressure on initial break in start up after a rebuild to avoid high point load scuffing damage potential between lifter base(s) and cam lobe in first few seconds of start-up. Vintage flat tappet auto engines traditionally used lifter bores marginally offset from being located under center of the cam lobe to generate the necessary lifter/pushrod rotation required, to more widely spread point loads of the cam lobe on the lifter base. 

http://www.tcmotoringguild.org/techinfo/TClinic-39.pdf

Confirm that lifters and their pushrods start to rotate immediately as engine is initially turned over without spark plugs installed, prior to break-in start up.

Lifter & cam lobe failure became an issue due to several reasons;

http://www.compcams.com/Base/pdf/FlatTappetCamTechBulletin.pdf

A) QC issues from lifter bore centers being located out of tolerance spec. when originally machined at Factory.

B) High Mileage, worn lifter bores, not addressed on rebuilds with bronze sleeves or oversize lifters. .010” oversize are available from Europe, under the County brand. (County is a major supplier to resellers of vintage parts, possibly now trading under a different name).

C) Poor QC chilled iron lifters or depth of case hardening/nitriding from various aftermarket sources. Note, that Moss outsources production of the lifters they offer. Two versions are offered:

i) One for OEM cams with OEM flat base lifters

ii) Some of the above are re-profiled to make the lifter base convex so they are compatible with modern Crane et. al. tapered lobe cams. Moss’s re-profiling, lauded as step in the right direction may reduce the thickness of the case hardening by “x” amount which could lead to premature lifter failure, also causing premature failure of cam lobe with things going downhill from there. Unknown if Moss has lifters re-hardened after crowning to regain full case hardening thickness.

D) EPA mandating reduced levels of ZDDP in engine oils to protect catalytic converters. Over time, aftermarket valve train suppliers started resolving the issues that were barriers to increasing spring pressures associated with lifter base & cam lobe wear. These new wear complications were limiting higher open valve spring pressures deemed necessary to gain horsepower via higher RPM. Reliable, wear free, road car flat lifters, with open spring loads of many hundreds of #’s, and open spring loads in race engines (700#-1000#+) are now common place, though the service life of springs is impacted as one moves to heavy duty race applications with higher seat and open spring pressures. I.e. moving to high cost, heavier roller cams, less stable valve train is not an end all solution.

EVOLUTION OF SOLUTIONS:

With worn lifter bores, accompanied by variations in tolerances of the bore center not being optimally offset from being under the Center of the cam lobe, and with reduced levels of anti-wear additives starting in mid to late 90s, premature wear began to occur on both chilled iron lifters, nitrided lifter bases and cam lobe. New solutions started to evolve as above causes were identified and understood. An interesting side issue is that wear reducing ZDDP additives were not introduced into motor oil ‘till after WWII. This begs the question of extent of flat lifter base & cam wear prior to introduction of these additives. As offset lifter bores were the pre WWII solution to wear, it seems likely that the lifter/cam lobe wear never became an issue as most cars from the 20s – early 40s had a very short shelf life & were ready to be junked at 30-40k miles. Henry Ford would send engineers out to auto scrap yards with instructions to inspect junked cars to see what parts were still in serviceable condition. He would then ‘dumb down’ these parts to save money. Clearly he was not a disciple of W.E Deming (the QC guru who is credited with turning around post WWII Japanese industry from producing “Jap Crap” to making quality products which would have a worldwide market).

1) Likely the 1st “fix” was to crown the base of the lifter a bit.

2) While this may have worked marginally as ZDDP levels were reduced, the next step was to further enhance lifter rotation by increasing the lifter crowning and also tapering the cam lobe laterally (side to side). With case hardened or chilled cast iron lifters, lifter failure still occurred when associated with low service life and especially heavy duty, higher spring open loads as EPA lowered ZDDP levels further.

3) Another more recent evolution to prevent wear at higher pressures, has been to use tool steel lifters that are hardened through & through to Rockwell 64. Should these lifters show signs of base erosion from dirt or pitting oxidation from moisture laden oil, they can have their crowned base(s) re-profiled indefinitely.

4) Most recently, for more extreme applications and to maximally extend service life, new lifters can be provided that are DLC coated, and can be expected to stop wear problems indefinitely. (Diamond Like Coatings),

http://www.calicocoatings.com/coating-data-sheets/calico-d-1-diamond-like-carbon-dlc-coating

These are for new steel cams, lifters, etc. & cannot be re-profiled without grinding through coating.

5) Additional current “work arounds” to provide more oil to lifter base cam lobe interface include:

a) An oil capture band around the middle of the lifter. Pressurized oil is presented through the lifter bore to the band where a cross drilling to the center of the lifter intersects a vertical drilling from base of lifter. The pressurized oil accumulated in the “band” is fed through drillings to the lifter base & cam interface.

http://www.compcams.com/Products/CC-%27Solid%20Mechanical%20Lifters%27-0.aspx

b) This desirable feature, although not incorporated in the XPAG oiling system, can be replicated to an extent by broaching a slot in the lifter bore to gravity feed oil from the pushrod gallery to the lifter base/cam lobe interface.

http://www.compperformancegroupstores.com/store/merchant.mvc?Screen=PROD&Store_Code=CC&Product_Code=5007&Category_Code

c) Another recent solution revisits a 30s solution used in the Rolls Royce Merlin series of aircraft engines. A UK vendor is offering XPAG camshafts with an axial oil passage through the length of the cam. At each cam lobe there is a cross drilling to the central oil passage in the cam enabling direct pressurized lubrication to cam lobe lifter base interface. A system also used by Porsche on some engines for many years.

d) Moss 433-365 lifters for newish Crane cams with oversize core have incorporated elongated slots in lifter – see lines 54-77 of the link.

The TC Motoring Guild article http://www.tcmotoringguild.org/techinfo/TClinic-39.pdf includes some outdated & incorrect discussion. Mainly the omission of the evolution to laterally tapered cam lobes to further enhance lifter rotation.

Also, there is the assertion that .904” XPAG compatible aftermarket lifters are no longer available. Not so. There are many modern valve train suppliers that supply .904” solid and roller lifters for various Chrysler, Ford, & AMC engines. Some of today’s of today’s aftermarket valve train suppliers also offer modern crowned base, tool steel lifters hardened through & through. e.g. see http://trendperform.com/c-1149649-lifters-tappets.html

Ok, we are on the right trail, but of course, nothing is without issues that must be worked around!

1) These non-proprietary domestic aftermarket .904” lifters are shorter & considerably lighter than OEM XPAG lifters. The distance 1.88” (seat height) from the base of the lifter to the bottom of the lifter cup for the lower end of the pushrod is shorter. The XPAG lifter seat height is mol 2.122, i.e. .242” higher, enough that longer pushrods are needed. Also note that these shorter lifters weigh 84 gm compared to 94 gm for OEM XPAG lifters. (Note that new shorter push rods would also be required if a new cam with Crane style oversize core is used.). XPAG push rod lower ends are not compatible with the receiving socket in modern .904” non XPAG specific lifters, so new push rods are also required. This is works because Modern 5/16” push rods are significantly stiffer and stronger as they are chrome moly with .080” wall, and are rated stiffness (oscillation) stable to 350 hp @ 6000 rpm, a good thing & very desirable.

2) The last issue is that the top cup of current aftermarket push rods is not a match for the ball on the pushrod side of the XPAG rocker adjustor. However the Lifter Supplier advises that they could provide a top cup specific to XPAG adjustor screw 10mm ball.

After looking at various sources to upgrade the valve train, I am leaning toward Trend performance push rods: http://trendperform.com/c-1149673-push-rods-stocking-push-rods.html with a .080 wall chrome moly @$6.00 + ea. and .904 lifters, tool steel, heat treated through & through to Rockwell 64, Convex base lifters, micro polished, & suitable for cast iron cams.

http://trendperform.com/c-1149649-lifters-tappets.html

@$ 17.11 ea. DLC coated + $65.00 ea. Bnz. lifter bore sleeves for worn .904” lifter bores are also available from Trend. Comparable lifters & p/rods available from Comp Cams, et. al.

Modern valves, desirable beehive valve springs, titanium spring retainers, etc are available from Ferrea, Supertech, Isky, Comp cams et. al.

Every gram saved in the valve train is a plus.

II SPRINGS

A) SPRING LOAD LIFE:

Springs going “soft”; spring wire has a finite fatigue load duty cycle life as they are exposed to high levels of cyclic stress. Over time, the fluctuating, cyclic twisting motion forces on the coil wire, stretching and relaxing the spring as it opens and closes can eventually fatigue fail the wire. If springs are losing installed pressure & go “soft,” the spring(s) have suffered plastic deformation, (failed in yield strength), or overheated beyond heat treatment limit. This occurs as fatigue inducing cyclic loads eventually exceed the yield strength (service life) of the coil wire. When the wire is stretched and its yield strength exceeded, plastic deformation stress levels are reached, the wire will not return to its original form, as the wire material has been stressed beyond its elastic yield limit. The installed spring pressure will be reduced at the installed spring height. Worse yet, over time the fluctuating cyclic forces stretching and relaxing the wire coil may cause the wire to break when the fatigue inducing cyclic stress forces exceed the UTS (Ultimate Tensile Strength) of the coil wire. The spring engineer’s task is to design the spring to operate within the predicted stresses to be encountered over its service life. Predicting spring fatigue life would be a function of determining dynamic loads involved, # of load cycles, & stress levels to be encountered. Design errors in this area could seriously affect the end-users’ pocketbook.

B) VALVE TRAIN DYNAMICS, HARMONICS;

Too complicated except for some generalities:

Basically; light weight, stiffness, low inertia on lifter and valve side of the valve train, along with most stable valve train possible is the end game. The service/fatigue life of the spring will be extended as the valve mass cycling the spring is reduced and the spring doesn’t have to work as hard.

Spring Surge; (The tendency of the spring to operate in & vibrate in an undesirable resonate condition). Springs and all components of the valve train that do not have infinite stiffness will vibrate, i.e. resonate in more than one direction and each at a different frequency. If a component harmonic coincides with other component frequencies, especially the spring/cam harmonic frequencies, the additional stresses can excite & add to surge causing spring to operate outside its design dynamic load capabilities, and fail. For instance, valve springs are held in place stable at their seat & retainer at top, as the middle of the spring is unsupported in column it can oscillate out of control, valve float is experienced and the valve head may hit piston.

III VALVES

In the past there have been dire warnings to avoid using stronger XPAG valve springs for fear of breaking valve heads off. Possibly this is true if these were aftermarket valves forged in a third world country by hand. Many current engines in daily drivers use valves with stems less than 6mm dia. & open spring pressures up to 200 plus psi. Virtually all motorcycles use valve stem diameters in the 4-5mm range at RPMs to 10-14k. Some with hollow heads (a practice introduced in aircraft engines pre WWII) & hollow stems. Valve heads break because the operating loads that valves are exposed to exceeded predicted stresses encountered.

http://www.supertechperformance.com/technical

All of the above is predicated on the assumption that the valve guides are not worn out of Spec; should the guides have excessive wear the valve stem may not translate in a perfect axial direction resulting in one edge of the valve impacting the seat under high load before the rest of the circumference of the valve face. This off axis point load could exceed design dynamic loads the valve stem was intended to cope with and actually break either the edge of the valve rim or the stem, especially if the valve has an undercut stem toward valve head.

For more detailed discussion of Valve Train Dynamics see:

http://www.stangtv.com/tech-%20%20%20stories/engine/ferrea-helps-explains-valve-flow-dynamics/

Exhaust Valves: Modern Inconel & Nimonic alloy exhaust valves have basically eliminated “burnt” exhaust valve problems from exhaust manifold air leaks, or when using superchargers. No need for sodium filled valve stems, just leave stems & heads hollow to save weight. Also modern quality replacement valves have a .001” taper toward valve head to accommodate heat expansion into valve stem under valve head. As to future evolutions of valve train control w/o camshaft see:

WSH

TC 4926 01/01/16-10

Ed’s note:

This discussion paper by Bill Hyatt of Odessa, Florida, USA was recently published in the February 2016 issue of The Sacred Octagon, Since publication, Bill has revisited the text and has made a number of amendments.

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