Our block arrived with OE powder metal pain caps and installed cam bearings.


Our bores’ center distance measured 4.400″, which met the OE spec.


This blueprint drawing shows cam-to-crank centerline as 124.08mm (4.885″) and deck height (crank centerline to deck) as 234.7mm (9.240″).


The rear of the block features already-drilled and tapped bellhousing bolt holes (10mm x 1.5) to ­accommodate the new style transmissions, but the block will accept an old-style Gen I Chevy bellhousing by drilling and tapping one extra hole in the blank boss, seen at the upper right in this photo.


The right rear of the LS2 block features a cast-in identification of “6.0L.”


The five main caps are powdered iron and are secured with six bolts each (four center 10mm x 2.0 bolts and two 8mm x 1.25 side bolts).


All of the LS engines feature the thrust bearing location at the center (No. 3) main cap.


The main cap center bolts are 10mm x 2.0 pitch. The inboard bolts feature a conventional hex head, while the outboard bolts feature a stud tip to accommodate the oil baffle. The outboard (stud-tipped) bolts feature a main shank length of 85mm, while the inboard bolts are 100mm long.


In our case, the main caps were easily removed by hand. A previously assembled block (where the main cap side bolts have been torqued) might require the use of a spreader bar to allow cap removal without damaging the cap-to-block mating surfaces.


Since the main caps use 8mm side bolts, in addition to the primary vertical cap bolts, pay extra attention to the care and cleanliness of the side bolt-mating surface on both the caps and the block.


A machined flat is featured on the block inner walls where the main cap accommodates the side bolts. Be careful to avoid damaging this soft aluminum surface.


This close-up shows one of the main cap side bolt holes. Flanged 8mm x 1.25 bolts are used here.


We checked each cylinder bore for roundness and were pleasantly surprised at the bore uniformity.


Our block apparently experienced a core shift during casting. Note how the cylinder liner shown here was bored off-center relative to itself. Bore centerline location seems fine relative to the block and main centerline. This simply indicates that we may not be able to overbore as much as we had hoped.

Pressure-equalization slots (crossover passages) are featured between the cylinder bottoms and the main webs.



We examine the block.

by Mike Mavrigian

all photos by author


Our new LS2 aluminum bare block was purchased from Scoggin-Dickey Parts Center. It arrived via UPS in a wood crate.


We’re starting this project build using a new LS2 aluminum block purchased from Scoggin-Dickey Parts Center. The LS2 block features a 4.002″ non-Siamesed bore that we’ll hone out, likely 0.020″ or 0.030″ oversize (the folks at Scoggin-Dickey recommend a max of 0.020″ overbore, while GM Performance suggests a max of 0.030″ overbore). In OE form, this is a 6.0L (364 CID) engine. Of course, we plan to bump displacement up a bit.

According to GM, the cylinder liners were designed for a stock stroke of 3.622″ and stroker combinations should be limited to “around” 4.000″. Depending on our piston design, we plan to stroke to either 4.000″ or 4.125″.

All LS series blocks feature 6-bolt mains (four vertical and two side bolts each), 9.24″ deck height, 4-bolt-per-cylinder head bolt pattern, standard GM bellhousing bolt pattern, 4.40″ bore spacing and 0.842″ lifter bore diameters. The thrust
bearing is located at the center bearing location.

Our block, although listed as “new,” looked like it had been through a battle-dings, nicks, scratches, scrapes, etc. The as-cast aluminum surfaces also had a distinctive brown tinge, like it had been stored in a dirty warehouse uncovered. No biggie since we’ll wash and clean up the various boogers.

I measured our block just to see where we were as compared to OE specs. The deck height varied by about 0.008″ (taller on the left bank). Cylinder bores measured, on average, 3.998″. During bore gauge checking, no discernible out-of-roundness or taper was found, so that was nice. One very noticeable oops that we found was evidence of an apparent casting core shift, as several cylinder liners were thin on one side and thick on the opposite side (front to rear). Basically, it looked as though the core shifted about 0.045″ with the liners then factory-bored on true center. As a result of a few thin liner walls, we’ll need to be careful when it comes to overboring. We may not be able to go much over the OE spec of 4.000″.

In order to prep our block, we’ll need to align-hone the main bore, deck the left bank, bore/hone cylinders to accommodate our pistons of choice and kiss-hone the lifter bores.

I could be wrong, but my guess is that these new bare blocks that are sold to the
aftermarket are units that didn’t pass muster during factory inspections. It’s not that big of a deal as all of the glitches we found are easily correctable. The only thing that bothered me was all of the surface scratches on machined surfaces,
especially on the front timing cover mating surfaces, and an out-of-round starter bolt hole, the obvious result of the boss being smacked (maybe the block was dropped on the floor at some point). But, again, this is easily fixable by running a drill through the hole.

Note: The rear of the block features drilled and tapped holes to accommodate late model bellhousings. Old style bellhousings can be installed, but a blank boss on the upper right side of the block’s rear face must be drilled and tapped. Why GM didn’t go ahead and perform that job is beyond me. The blank boss is already in the casting. I’ll need to install a bellhousing to provide an index point and I’ll drill/tap that hole to 10mm x 1.5 before the short block is assembled.

We also noticed that the top of the cam bearings were slightly damaged, resulting in an out-of-round condition at the cam bearing inner diameters. My guess is that someone along the line inserted a bar through the cam tunnel to lift/move the block. Since we plan to perform machining on this block, this doesn’t present a problem; we’ll need to remove these original cam bearings to properly wash the block anyway.


GM P/N 12568950 available from Scoggin-Dickey for a suggested retail price of $999.99
This LS2 6.0L block is considered a Gen IV block, which is basically the same block with minor changes as compared to the LS1 and LS6 5.7L Gen III blocks. The cam position sensor has been moved to the front timing cover and there is no knock sensor provision in the engine valley. The LS2 features a knock sensor location on the side of the block.

This block is a direct replacement for 2005-2007 LS2 Corvette, SSR, GTO 6.0L and Trailblazer SS.

The block is cast from 319-T5 aluminum and features iron sleeves, 6-bolt iron main bearing caps, a (supposed) 9.240″ deck height and 4.00″ cylinder bores. This block will accept LS1, LS6, LS2, L92 or LS3 cylinder head designs.


Note: The LS2 block upper valley features a series of tower bosses with oil holes drilled through to the crankcase. In case you wondered what purpose these serve, these drilled towers were part of the original LS casting for engines that used GM’s displacement on demand system. On the LS2 block, these serve no purpose at all. The LS2 valley cover underside features o-rings that simply seal these towers off. Also, you may notice that the LS2 valley cover features a raised boss that’s drilled and tapped. This hole accepts the oil pressure switch. The underside of the cover plate also features a large and lumpy black plastic housing, feeding to a metal tube on the top side of the cover. This is for the PCV system.


If you don’t have an LS engine core lying around, you’ll start off with a bare block. Since a bare block is, well, bare, even if you plan to use all aftermarket components, you still need a few OE-only parts. The thoughtful folks at Scoggin-Dickey recognized this need, so they assembled a “block completion kit,” P/N KITLS2CK-1, that includes:



When the above kit arrived, I discovered that I didn’t have the camshaft plate or the rear seal housing. Upon contacting SDPC, I discovered that these parts are to be ordered separately (no big deal, somebody assumed I had an LS1 or LS6 block from which I could scavenge parts). I just wanted to point this out because if you don’t have a donor core from which to steal stuff, you’ll need everything I’ve listed here.

All-told, you’ll spend in the neighborhood of $400-600 or so (depends on current pricing) for the complete block completion kit.
Scoggin-Dickey Parts Center (SDPC) can be contacted at (800) 456-0211.


(Differences and interchangeability of parts)

Note: When using an OE LS2 block, be aware:
• LS2 blocks feature the knock sensor on the side of the block, while LS1 and LS6 blocks feature the knock sensor in the lifter valley.
• The LS2 cam sensor has been moved to the timing cover location, which requires the use of the LS2 cam gear.
• LS2 knock sensors are dedicated for LS2 and are not interchangeable with LS1/LS6 knock sensors.
• An LS1 engine-management computer will control an LS2 engine.
• LS2 blocks feature an original 4.00″ bore, while LS1/LS6 blocks feature a 3.898″ bore.
• Crankshafts, connecting rods, pistons and cylinder heads are interchangeable between LS2 and LS1/LS6 engines.
• LS2 cylinder heads are made from the same raw castings as LS6 heads.




We’ll provide the proper tightening sequence for the main cap bolts in the next article installment.

Note: Always follow the same sequence when tightening main cap fasteners. Follow the same sequence each time the caps are tightened (test fitting, final fitting, final assembly, etc.) to avoid altering bearing clearances. Also, be sure to always use the same torque wrench to avoid variables that could affect bearing clearance. Note that in some cases, a small threaded spreader bar (right- and left-hand threads with turnbuckle) may be needed to slightly expand the block sides to accommodate main cap removal without damaging the main cap to block mating surfaces. In our case, all caps removed easily by hand by gently rocking them fore/aft. A spreader bar may be more commonly needed when servicing the caps in a block where the 8mm main cap side bolts had once been torqued. By the way, the main cap bolts (inners feature a conventional hex head, while outers include a protruding stud for pan baffle mounting) are 10mm x 2.0 (very rough pitch). The inner bolts are 100mm long and the outers are 85mm long.

Note: Because GM reportedly individually fits each bearing (bearing-matching per individual engine), when you plan to use aftermarket main bearings, carefully measure the bearing thickness and crankshaft journals, and align-hone the main bore as needed to achieve proper bearing oil clearance.

I’ve heard that some builders tighten using torque-only, applying 60 lbs./ft. to the inner bolts and 50 lbs./ft. to the outer stud nuts. If you’re
using new OE fasteners, follow the torque-plus-angle method. If using aftermarket fasteners, adhere to the values recommended by the fastener maker.


In our next issue, we’ll complete the short block by fitting our crank, rods and pistons.

This will include checking bearing clearances, sizing bores for our pistons, performing any stroker clearancing that might be needed and balancing the rotating assembly



The camshaft is spun-up on the ­balancer similar to any other rotating shaft.


Example of a camshaft out of balance. Notice that the left side of the ­camshaft is 1.2 in./oz. out of balance and the right is 1.67 in./oz. out. Also note that the unbalanced units are not opposed to each other. This will cause the shaft to combine the out-of-balance forces, causing a bending ­moment as opposed to a canceling effect.

In this example the camshaft is running at 4,000 rpm (half of the 8,000 rpm crankshaft speed). The left side of the camshaft is generating a force of 34.3 pounds and the right side is 47.6 pounds, and both are ­hammering at 66+ times per second. This may be enough to cause the camshaft to ­become excited and set up a vibration pattern that will motivate the roller lifter to bounce on the surface of the lobe. Ultimately, the valve will respond to all of this activity and most likely this will cause the valve to follow a path that is not equal to the designed cam lobe profile. This may also cause the valve spring to become excited, generating an inconsistent travel pattern ­commonly known as valve spring float.


The corrected report shows that the out-of-balance has been ­minimized and the residual unbalance is in-couple (meaning that the forces are opposed to cancel each other).


If imbalance is excessive, it may be necessary to add one or two external weights. This approach of course, is dictated by the cam design, where space permits, such as on a 4.6L Ford. The example shown here features two counterweights.


External counterweights are two-piece and clamp onto the shaft.


Any camshaft, ranging from the cheapest cast OE to the most exotic aftermarket grind, may feature an imbalance condition, from mild to wild. It doesn’t hurt to check.


Correcting a camshaft imbalance condition can result in a more efficient (optimized) valvetrain operation.



That’s right, camshaft balancing. It’s all about optimizing and regaining lost energy.


The camshaft is yet another rotating mass, so why would we not pay attention to the subject of balance?

When was the last time you balanced a camshaft? If you’re like most builders, the answer is never, with the answer accompanied by a puzzled look. Immediately following that answer, most folks are likely to add, “And why the hell would we want to?” At the outset of a recent conversation with equipment-innovation guru Randy Neal of CWT Industries in Norcross, Ga., I have to admit that I had much the same reaction. Neal has developed a software program for balancing camshafts, and after about 10 minutes of listening to his theory, I became a believer.

Once you’re able to erase any long-standing notions and establish an open mind, it makes perfect sense. After all, balance is key in any rotating mass, and cams are no exception. Just consider all of the components that are currently considered candidates for balance checking and balance correction: crankshafts, drive shafts, propeller shafts, supercharger shafts, armatures, turbo impeller shaft assemblies, ring gears, wheels, flywheels, transmission shafts, etc. Why then would we blindly ignore camshaft balance? Just because it’s never been addressed in the past doesn’t mean that it isn’t a good idea. It’s called progress and taking advantage of technology. It wasn’t that long ago that most folks didn’t see the need or benefit of engine coatings. Today, taking advantage of thermal barrier, anti-friction and oil-shedding coatings is commonplace, with proven results.


Yes, some would argue that since the cam runs at half the speed of the crank, camshaft balance isn’t an issue worth exploring. With all due respect, I disagree. While an out-of-balance camshaft may not create a vibration that will be felt in the seat of the pants, any out-of-center forces can create a harmonic disturbance, and that disturbance, or excitement, can result in frequency disturbances (OK, vibrations) that can be transmitted through the rest of the valvetrain, including lifters, pushrods, rockers, springs and valves.

If you wish, you can view the camshaft as a tuning fork. An imbalance condition can excite the cam. Correcting the imbalance can mute or eliminate the excitement. Some might say that an camshaft imbalance would be absorbed by the valve springs, but this is exactly the point-why introduce more vibration or unwanted frequencies into the valve springs if it can be avoided? The springs already have enough work to do. Balancing the camshaft may very well “calm” the springs, removing unnecessary distractions, allowing them to concentrate on the job at hand, which is controlling the valves.

We’re not talking about reinventing the wheel here. Rather, we’re talking about making the engine perform at the best of its abilities. When you’re building a high-performance engine and your goal is to optimize engine performance, it just makes sense to include camshaft balance as part of the overall equation. Will balance correction make more horsepower? No. However, correcting an imbalance condition of any rotating engine component is a positive step toward reducing vibration, uneven loading and unwanted harmonics, which can, in turn, free up horsepower. There’s a difference.

Creating balance or, more accurately, reducing or eliminating imbalance, will reduce the negative effects of lost energy, allowing the engine package to function more efficiently. And nobody can argue that increased efficiency will allow the engine assembly to function at its peak potential, both in terms of power and durability.


At this point, I think it’s best to include information obtained directly from Randy Neal. According to Neal, virtually every camshaft made is not balanced. “We have tested hundreds of camshafts and have never found one that was balanced,” he said. “As a matter of fact, we rarely see two camshafts with the same profile that have exactly the same unbalance. Now you may be saying that just can’t be true due to the accuracy of the new CNC cam grinders; but the balance problem does not generally come from the grinding operation. Rather, it actually started with the raw casting or forging.”

“When the camshaft blanks are made they have Center Registers placed at each end of the camshaft,” Neal continued. “The main bearing and base circle of each lobe are established from these registers. The axis of rotation and the Center of Mass is probably not the same. In fact this is generally what forces that camshaft to be out of balance. It’s important to understand that when the Center of Mass is not the same as the Center of Rotation, there will be a balance error.”

“When a camshaft is out of balance, we need to know the amount of unbalanced forces and their location relative to each other,” Neal said. Pictured are visual reports of a 60mm Roller Camshaft that has been inspected on a Computerized Balancing Machine.

The ultimate result is that the engine will not perform as intended. In all probability at an unknown moment, at an undetermined rpm, the engine will make less power. Neal also noted that, “We have seen this on the power curves when plotting the horsepower/torque on the dynamometers. Some engine builders have determined that these anomalies can be related to ignition/fuel and/or flow characteristics of poorly designed intake systems, and they could be right.”

“However, when the intake system is modified and the same rpm related anomaly is re-plotted, there is strong evidence that the valve train is the potential cause and the vibration of the camshaft has the ability to cause these results,” Neal said.
In all fairness, balancing the camshaft may not eliminate all or any of the valve train issues due to the fact that there are several other moving objects that may have natural frequency issues that could become excited from surface speed activities (rocker flutter, valve spring float, push rod deformation and excess end-gap clearances). But by balancing the camshaft you have eliminated a known variable that will not be a part of your quest for the ultimate performance of your engine.

For the skeptics out there who still believe that the camshaft harmonics have no relevance to horsepower gains, I agree that it will not make power. Rather, it simply unleashes it. There is no discussion that can be supported that says balancing any rotating mass will cause any object to perform in a negative manner.


Correcting the unbalanced camshaft can be achieved by several methods. The first is to modify the timing gear by removing or adding material as specified by the balancing machine. The second is to remove weight from the core of the main bearing area or add weight by first drilling holes into the main bearing and replacing it with heavy metal.

Another option is to drill holes in the butt of the cam and possibly add heavy metal, or to add external weight to the shaft itself. This can be achieved by using an adaptor that is added to the rear of that shaft which locates a counter weight that is modified for the require amount to balance the assembly, or, given the cam design and spacing between lobes, clamp-on counterweights between cam journals.

Is camshaft balancing a necessity? For the street car? No. For the race car? Yes, assuming that your intent is to optimize the engine in order to extract all of its potential. It’s just one additional small step forward for the racer. And that’s a good thing.



Always use the cam assembly lube provided or recommended by the camshaft manufacturer, in addition to using the correct break-in oil.


Nothing changes with regard to prep. Coat each lobe completely with the recommended assembly lube, usually a high-pressure moly-based material.


Even with the lobes coated with moly, using an engine oil that is too low in zinc phosphate can kill an otherwise perfectly good cam.


Another example of a dedicated break-in oil is the Joe Gibbs BR oil. This has been formulated specifically for seating flat tappet cams and lifters.


Among the handful of engine oils currently ­available that are designed specifically for break-in of flat tappet cams is Brad Penn Penn Grade 1. The container is clearly marked as “Break-In Racing Oil.” This is top-notch insurance to ­protect the cam lobes and lifters.


After break-in has taken place (rings seated, cam and lifters properly mated, etc.), it’s still wise to use a high zinc oil for dyno and long term engine use, especially when the engine is equipped with a flat tappet cam.


Offered by:
American Refining Group Inc.
77 N. Kendall Ave.
Bradford, PA 16701
(814) 368-1200

“For use in break-in of high-powered, high-performance race engines … promotes ring seal and provides maximum protection for cams and lifters during initial break-in. Exceptional break-in performance in flat tappet race engines. Can also be used to break-in roller cam engines. Can be used as general purpose dyno oil …” the company says.
Offered by:
Joe Gibbs Driven
13415 Reese Blvd. West
Huntersville, NC 28078
(866) 611-1820

Offered by:
3499 Blazer Pkwy
Lexington, KY 40509
(859) 357-7777


SWEPCO 306 15W-40
Offered by:
Southwestern Petroleum Corp.
P.O. Box 961005
Fort Worth, TX 76161-0005
(800) 877-9372

NOTE: For break-in of flat tappet engines, avoid any engine oil if the container features the API small starburst logo. The star indicates that the oil has been formulated for new engines from an energy-conserving standpoint. That stuff is for passenger car gas engines equipped with roller cams. Avoid the star like the plague, at least for flat tappet break-ins.

1. Don’t use a synthetic oil for break-in. It may be too slippery to assure ring seating and flat tappet lifter rotation.
2. Always apply the cam maker-specified moly paste to the cam lobes and lifter faces.
3. Do not pump-up hydraulic lifters before use. This can cause the lifters to hold a valve open during engine cranking, which will cause low compression and delay engine start-up.
4. Prime the engine’s oil system before start-up.
5. Lubricate lifter walls and pushrods with engine oil.
6. Fire the engine, bringing it to a fast idle of between 1,500 and 3,000 rpm. Vary this idle speed during running, in a slow-to-moderate acceleration/deceleration cycle. Continue this varying-speed cycle for 20 to 30 minutes. This is to ensure proper lifter rotation in an effort to mate each lifter to its cam lobe.



All oils are not created equal.

by Mike Mavrigian

photos by author


Avoiding flat tappet cam failure can be as simple as using the correct oil during break-in.

Chances are you’ve run into this problem first-hand, or know a builder who has shared his horror stories. We’re talking about flat tappet camshaft failures during break-in. In recent years, there has been a rash of cam problems and, in too many cases, the cam makers have unjustifiably shouldered the blame. The problem does not lie with the camshafts. Rather, the problem is caused by the engine oil used during break-in.

Because of mandates by the EPA, a vital element of the oil mix, commonly called ZDDP, has been drastically reduced in standard engine oils that are intended for the late-model street vehicle. In short, if the oil doesn’t contain enough ZDDP, it doesn’t offer adequate anti-scuff protection for the initial break-in of flat tappet cams. Let’s face it, when a customer’s brand new cam gets wiped out through no fault of yours, life can be frustrating to say the least.

ZDDP (zinc dialkyl dithio phosphate) is an anti-wear and antioxidant, initially developed in 1930 as an antioxidant to prevent engine bearing corrosion. ZDDP also features excellent anti-scuff and anti-wear properties. In the 1960s, ZDDP featured a zinc level of 0.07 percent when high-performance flat tappet camshafts were common. At that time, new camshafts were phosphate coated as well and the combination worked well to protect new camshafts and lifters from premature wear, especially during break-in.

In the 1970s, zinc levels increased to 0.09 percent because ZDDP is an excellent antioxidant. As engines became more powerful, oil recipes changed as well, becoming more complex with more functional additives such as friction modifiers, antioxidants, detergents, etc. Friction modifiers gained further popularity to aid fuel economy, with zinc content increasing to 0.2 percent in the 1980s and early 1990s.

By the way, ZDDP is only one acronym for this anti-wear/antioxidant content. It’s also referred to as ZDP or ZZDP. Why? Who knows and who cares?

So if ZDDP is so cool, why has it been reduced to the point where it’s causing flat tappet cam problems? Phosphorous is a well-known contamination source for catalytic converters (some refer to it as converter poison). The limit for phosphorous dropped to 0.10 percent, which means that the zinc level dropped as well. In 2004, with Tier 2 emissions standards, OEM warranties changed to 10 year/100,000 miles, and phosphorous dropped again to 0.08 percent, with zinc down to 0.09 percent.

Engine oils, in general, are vastly superior to oils made in the past, a major factor responsible for some engines being able to last for 250,000 miles or so. Also, today’s metallurgy is better. The issue here is high-performance flat tappet cam lobe wear during the break-in period. Aggressive cams with high spring loads compound the problem. This issue does not affect roller cams, since there’s no scuff wear issue with rollers.

In a nutshell, whether in a direct or indirect manner, the EPA has told the oil makers to ignore older (i.e. flat tappet cam) engines and to make an oil that avoids converter damage (thereby reducing emissions) in late model cars, and the hell with the restoration and performance market. Marie Antoinette once told the French peasants to eat cake. The EPA has basically told car guys to fend for themselves. Either expression is offensive as hell.

We could say that this entire problem could have been avoided if we (the engine community) were properly informed about the change in oil make-up. In that case, we could have made a point to search for specific break-in oils that did contain adequate ZDDP levels. Instead, many of us learned the hard way by needlessly wiping out otherwise perfectly good camshafts during break-in runs. At this point, it’s a case of too little too late. Let’s all send a big fat thank you to the government and the major oil marketing companies for making our lives a living hell.


To avoid flat-tappet cam lobe damage during break-in, naturally you must continue to apply the specific cam lobe and lifter assembly lube that’s recommended by the cam maker, plus you can install low-rate valve springs for the break-in.

In addition, you can and should use one of the few currently-available engine oils that do contain sufficient ZDDP. These oils are available, but you need to remember to specifically purchase these oils and dedicate them for flat tappet cam break-ins.

Engine oils that are specifically designed for use in diesel applications will usually feature more zinc than passenger car gas engine oils. However, diesel engines are coming under greater scrutiny as well in an effort to further reduce emissions nationwide. So, while a dedicated diesel oil may be better than a passenger gas engine oil in terms of zinc content, you can’t automatically assume that any diesel oil contains enough ZDDP to protect a new flat tappet cam.

According to the tech boys at Crane Cams, oils that they are currently aware of that are compatible for flat tappet cam break-in include Shell Rotella T, Chevron Delo 400 and Mobil DELVAC. All three of these are classified as diesel oils. Crane did note that Rotella T has apparently been modified with a slight cutback on zinc (rumor has it that zinc was reduced from 1,400 ppm to 1,200 ppm), but that should not be enough to cause problems.

I have heard other unsubstantiated rumors, however, that Rotella’s zinc has more recently been further reduced to 800 ppm (and possibly even further), but we could not get that rumor qualified in time for this article. Crane did note that if a questionable oil is to be used (where the user is not sure of the ZDDP content), a friction modifier such as GM EOS should be added for break-in. However, no additional friction modifier additives should be added to a break-in oil that is documented as flat-tappet-break-in safe.

When we spoke with the folks at Redline, they did mention that while they do not currently offer a break-in oil, they do have plans to introduce a ZDDP additive in the future. So, nothing at the moment, but possibly down the road.

A call to the tech department at Valvoline corrected an issue that is currently misunderstood among some builders. While Valvoline does offer an Off Road 20W-50, that is not the oil recommended for flat tappet cam break-in. Instead, they advised using their VR1 Racing Oil, an SM-rated oil that features 1,300 ppm of zinc. They also noted that while a common misconception is that SM-rated oils are considered unacceptable for this application, that this is simply not the case.

Castrol has recently introduced its new CASTROL SYNTEC 20W-50, which reportedly “contains increased zinc levels for extra engine wear prevention … uses proprietary additives and base oils to reduce metal-on-metal contact of aging engine parts … engineered to increase wear protection for classic cars with flat tappet camshafts.” While this oil may be fine and dandy for day-to-day use in flat-tappet engines, because it is a synthetic oil, in good conscience we can’t recommend it specifically for break-in since it’s never a good idea to break in any engine (flat tappet or roller) with a too-slippery synthetic oil with regard to ring seating and flat tappet lifter rotation.

Lake Speed Jr. at Joe Gibbs Driven noted that they developed their dedicated BR break-in oil specifically to meet the needs of flat-tappet cam applications, which contains a whopping 2,800 ppm zinc. Speed told us that they developed this oil in order to be able to use the same oil for break-in and for complete dyno sessions. Gibbs also offers a special Assembly Grease for lobe and lifter lube during assembly.

Dick Glady, a highly respected racing oil expert who has worked in the racing oil industry for decades, of American Refining Group, makers of Brad Penn Racing oils, informed us that their entire line of racing oils has never been reduced of its zinc content. All of their racing oils still contain favorable levels of zinc and contain special cuts that enhance oil cling and anti-scuff properties.

While many of American Refining Group’s race engine customers use its 20W-50 racing oil for break-in, dyno and competition use, the company has also introduced a dedicated break-in oil for flat tappet cam engines called Penn Grade 1. This is a straight 30W oil with high levels of zinc and special anti-scuff properties, and is specially formulated to promote proper piston ring seating during break-in as well. This is definitely a premium break-in oil.

To summarize, while there are a handful of engine oils out there that are reportedly safe to use for flat tappet cam break-in, the select few for which we have definite approval include the Brad Penn Penn Grade 1, the Joe Gibbs MicroZol BR and Valvoline’s VR1 Racing oils. Coat the lobes and lifter faces with the cam maker’s approved assembly lube, use one of these oils, and you should be good to go.



Both rod and main bearings must feature a ­specified amount of crush area in order to achieve bearing lock-in within the housing.


The bearing shell outer edges (approaching the parting line areas) are larger than the saddle, which adds to the lock-in placement of the ­bearing shells.


Dirt particles or embedded blasting shot left behind in the housing face can push the bearing inward, creating a high spot between the bearing and shaft, which leads to damage of the bearing’s outer layers. This will restrict oil clearance and can cause localized overload, friction, heat and will lead to bearing failure.


Oil film is formed during shaft rotation. As rotation begins, the oil film is generated, which travels around the circumference of the shaft, literally lifting the shaft into a centered location.


This view shows the damage that results from trapped particles behind the bearing. Note the bearing wear on the face (left). A metal chip that was trapped behind the bearing during assembly distorted the bearing wall ­inward against the shaft, causing wear through the overlay into the copper-lead base.


Through the analyzing of elasto-hydrodynamic lubrication (EHL), we see the difference in oil pressure peaks with H-beam (left) and I-beam connecting rods.


During TDC on the exhaust stroke, the bearing housing may begin to ­elongate, creating a greater bearing clearance at the top of the rod bearing.


Bearing cavitation damage can begin to occur when the shaft pulls away from the oil film abruptly, creating vapor bubbles. Once the bubbles break, this cavitation erosion can slowly begin to erode the bearing face.


This view shows the bearing construction of a Clevite TriArmor bearing, including the steel backing, cast copper/lead layer, babbitt outer layer and the moly/graphite surface coating.


Moly/graphite anti-friction coating needn’t be applied to the thrust faces of main bearings, since it’s common for builders to fine-tune crankshaft thrust by sanding the thrust surface of the bearing.



Cavitation erosion of a bearing occurs when rapid movement of the shaft away from the bearing surface causes vapor bubbles to form in the oil film. When these bubbles break, the resulting force causes erosion of the bearing soft overlay layer. Appearance and location of cavitation erosion will differ with operating conditions due to varying load patterns in different engine applications.

According to Havel, the nickel dam in the H-series and copper-indium intermetallic compound in the V-series helps to resist further penetration. Prolonged exposure will eventually result in erosion of the nickel or copper-indium dam. H-series bearings feature a thicker nickel dam to resist cavitation longer. Eventual penetration of the nickel dam causes copper particles to break loose, enter the clearance gap and become embedded into the bearing surface.

So far, the most effective means of controlling cavitation erosion seems to be a reduction in bearing clearance. This has worked in IRL and NASCAR applications. As an example, Aurora IRL engines running 0.0028″ rod clearance experienced cavitation erosion, but those running a slightly decreased clearance of 0.0020″ showed little or no cavitation erosion.


In order to gain further insight from a race engine builder’s perspective, we spoke with three noted builders. Following are their comments.

We’ve tested cryogenic treatments on main and rod bearings but, while cryo treatment has its place in other areas, they haven’t seen a benefit in terms of bearings. The new coating for the TriArmor bearings is far superior to the original coating and holds up very well. The new coatings have improved a ton. Oil clearances have definitely become tighter, with 0.0011-0.0015″ now becoming commonplace, especially due to smaller journal sizes and the use of synthetic oils (depending on the application, they’re running 0-10 up to 20-50 weights, with the majority using 10-30 and 10-40).

Depending on the application, our engines are running less oil pressure, and less parasitic drag (we’ve taken full advantage of oil shedding coatings to help reduce drag). The only secret I’ll share is our method of installing rod bearings. Instead of finger-pressing them into place, we roll the bearing shells into place in a back and forth motion for improved seating uniformity, which removes any high spots.

Ray Jager
Power Source Racing Engines
Fox Lake, IL

Unlike a few years ago, there are actually quite a few similarities between street high-performance engines and race engines in terms of bearing applications. Race engine bearings used to run gobs of clearance, upwards of 0.003″ or so, based largely on the oils we were using. Today, with thinner synthetics and improved oiling systems, we’re able to run tighter clearances in the neighborhood of 0.0015″ on rod bearings, and sometimes less. Today we’re able to do this while still providing better bearing life and decreasing friction at the same time. Granted, we pre-heat our oil before the engine is started in order to get optimum flow from the start, so it’s a more controlled environment as opposed to what a street engine will see.

In an effort to save weight, our main journals are down to 2″. Our main bearings clearances get down to almost as tight as the rod bearing clearances. We ball-mic each and every bearing to verify thickness and we even check for straightness, allowing us to categorize bearings. However, the Clevite bearings are so precisely made that we can pretty much just run them out of the box. Once in a while, we do mix and match bearings to achieve desired clearances, but we prefer not to.

Sometimes we’ll run a half-under on one side and a standard on the other side. Clevite TriArmor bearings are already coated, but not on the thrust faces. That’s good, since we sometimes lap the thrust faces on a granite block in order to fine-tune out thrust clearance.

As far as crank oil holes are concerned, we simply deburr the holes to break off the edge. Years ago, we used to radius-sweep the holes, but you get too much bleed-off doing that, so now we simply deburr the holes, removing as little material as possible. Since Cup engines run flat tappet cams, we use a mineral oil for break-in. Once we know that everything is seated and ready to go, we then switch to a synthetic for track use. We do pay close attention to the zinc phosphate content of the break-in oil, since inferior levels of the scuff protection can cause severe problems during break-in.

Dennis Borem
Pro Motor Engineering
Mooresville, NC

When we select bearing sizes, we pay attention to not only suggested clearance, but we also take into account the bearing surface from an anticipated load standpoint, as well as bearing speed, based on journal circumference.

In higher-end engines, where you plan to run smaller journals sizes, you really need to pay attention to the load-carrying capabilities.

In order to provide adequate oil delivery, we sometimes drill extra oil holes in the bearings and partial-radius grooves in the housing or saddle area of the mains to create multiple oil supply points. This is especially important in engines that use smaller bearings and will experience higher loads.

We try to run a fairly high crush while maintaining this within an acceptable range. Considering bearing load and journal and housing deflection, we want to make sure that the bearing is securely held in place. Where you have oil films that are in the tenths of thousands clearance, the bearing gets very hot. If you don’t have adequate crush, you won’t get enough heat transfer. Avoid taking housings to their maximum size to avoid inadequate heat transfer.

In many of our high-load builds, we modify the crankshaft journal oil holes in order to drive more oil to the rods. As you shrink the rod journal diameter, the load goes up. In order to get extra oil to the rod bearings, we create a slight teardrop groove to the crank main oil holes. We slightly groove the leading edge (attack side) of the oil hole. As the crankshaft rotates, this slight teardrop-shaped cavity fills with oil and is then force-pumped into the oil hole, increasing boost pressure.

This can cure problems with rod bearings that were otherwise seeing too much load. This can be done with a grinder, but we usually perform this on a CNC machine. However, you need to pay strict attention to the dimensions of the teardrop groove in terms of width, length and depth. Generally speaking, this teardrop groove is usually around 0.300″ to 0.400″ in length. If the groove is too aggressive, you could start starving the mains for oil. The specific profile of this groove controls the amount of oil pressurizing into the rod.

Understanding the specific engine’s oiling system is key. For example, in OHC engines, where hydraulics are responsible for much of the valve control, you need to maximize oiling efficiency to make sure that sufficient oil gets delivered to the top of the engine quickly.

As far as bearing clearances are concerned, for street engines that see higher loads, we tend to run somewhere around 0.003″ for mains and around 0.0025″ for rods.

We work within a window of about 0.001″ and keep a pretty tight tolerance range.
For engines that will see lots of heat for extended periods, such as endurance engines or marine engines, we tend to run tighter bearing clearances, to compensate for the fact that clearances will loosen under hot conditions.

Mike Schropp
Livernois Motorsports
Dearborn, MI


In an effort to aid engine builders in fine-tuning their bearing clearances, MAHLE Clevite recently introduced half-size H-series performance bearings. New part numbers include 0.009″, 0.011″, 0.019″ and 0.021″ rod and main bearings. These special-size bearings offer greater latitude in choosing bearings for performance engines that feature an undersize-ground crankshaft. By taking advantage of these new bearings, builders can easily achieve specific bearing clearances of +/- a half-thousandths of an inch. For example, using a pair of 0.009″ bearings will reduce bearing clearance by 0.001″ compared with normal 0.010″ bearings.

Similarly, clearance can be increased by 0.001″ by using a pair of 0.011″ bearings. It is also possible to use one 0.009″ bearing shell in combination with a regular 0.010″ shell to reduce clearance by 0.0005″, etc. These special half-size bearings are currently available for w wide range of Chevy, Ford and Chrysler applications. All bearings feature TriArmor construction and MAHLE Clevite’s unique moly graphite coating that’s distributed in a PTFE (polymer) carrier.


If you’re having bearings coated, avoid applying anti-friction coating to the parting line surfaces. A 0.0003″ coating on the parting line faces can increase the total bearing crush effect by 0.0012″, which can create an initially distorted bore. As the coating extrudes from the parting line surfaces, the bearing may then lose its tight fit in the housing. Also, notes Clevite’s McKnight, coating should not be applied to main bearing thrust faces. “Since engine builders will often sand the coating off of the thrust faces in order to achieve the desired end play,” he said.



The bearing gurus at Clevite share insight regarding current and future rod and main bearing development.


Installed bearing sets create a slightly eccentric bore, which promotes oil film capture and enables the rotational movement of the shaft to generate a hydroplane of oil film around the circumference.


Bearing crush is critical to holding the bearing in place. With the cap installed, this exerts radial pressure, forcing the bearing backs outward radially.

When big bucks and series championships are on the line, every nuance is considered. Bearings are no exception.

According to MAHLE Clevite’s Bill McKnight, top NASCAR Cup teams are carefully selecting bearings and measuring each. They will typically order about 500 bearings and check each individual bearing shell for height (in terms of crush factor) and thickness, etc. They’ll carefully categorize each bearing as tight, loose or intermediate, which then allows them to pick and choose bearings depending on the specific engine application.

The trend among the Cup teams is to run tighter clearances, creating higher oil film pressure.

McKnight also noted that some teams are buying Clevite bearings from a proprietary facility in Scotland whose technicians painstakingly measure and package sets guaranteed to be exact matches. “These measured-and-matched bearings are rather expensive, costing around $30 each, so only teams with very healthy budgets are taking advantage of this,” McKnight said. “NASCAR engine developers are also currently testing cryogenic treatment of bearings to determine if this offers any benefit, but the jury is still out with regard to this. All teams are definitely taking advantage of bearing surface coatings, which is no longer considered a questionable approach. The use of anti-friction and oil-embedding coatings have definitely proven their worth.”

McKnight noted that Clevite’s first effort at cryogenically treating race bearings for drag applications was made during the summer of 2007 for the top fuel and funny car teams, the hardest users of bearings we have. Tests continue, improvements look to be small, perhaps 5-7 percent, but when you’re using all we’ve got in a bearing and still asking for more, that may be enough. The treatment consists of freezing finished product at well below -300F for seven to eight hours.

Clevite has made a running change in the TriArmor, coated bearings. Improved coating durability is the result without sacrificing any of the features of the original bearing, except for a darker blue color to the coating. The change was a running production change last spring so most, if not all, of the line is new material when you order parts from your supplier. Clevite also supplies the custom bearing coaters with Clevite race bearings for their production (including HM Elliott, Calico and Polydyne). Coated bearings are very popular in the professional ranks of both circle track and drag racing-friction reduction and resistance to scuffing are two key benefits.

In the NASCAR realm, a few changes have been instituted. Standardizing on a minimum size of 1.850″ for the rod pins and 1.999″ for the mains is now completed. Clevite has numerous choices for these engines, including both H series and V series materials, 2.000″ main sets, main sets allowing teams who ran a 2.017″ main journal last year to use those same blocks with a 2″ crank (.017, .018, .019″-under sets) narrowed rod shells, dowel holes in some sets, extra clearance sets and inventory on the older sizes still used by some teams (1.770″ rods and 2.300″ main, RO7 blocks).

As NASCAR teams continue to narrow the rod bearings in search of less friction and more horsepower, the locating lug on the bearing insert has been placed very close to the edge of the bearing, creating what might be considered a stress point. Clevite is currently in phase-two testing of a new indentless lug insert for those applications. This provides a smooth, uninterrupted surface at the parting line. The company expects to start shifting the production of NASCAR rod shells to that design some time in 2008.

Also for drag racing applications, pro teams, many of whom are using Clevite off-the-shelf race bearings, are starting to take a look at what the NASCAR teams have been doing. Reduced mass, lower friction and coatings are all subjects being talked about at the track. One way to look at it is if your engine bearings look really good at teardown for refreshing the motor, you’re probably leaving something on the table in terms of performance improvements. Now, bear in mind, for the vast majority of racers, there are easier places to get more horsepower than engine bearing modifications but, for those at the very top levels, perhaps not.


Six criteria that must be considered in terms of high-performance bearings include bearing construction and materials; proper housing and shaft geometry; proper bearing geometry; proper surface finishes; sufficient supply of clean oil; and adequate oil viscosity.

In terms of construction, cast bearings provide superior strength as opposed to sintered bearings produced from powdered metal. Sintered bearings lack the continuous copper phase that is needed for strength. A high-performance bearing produced using a steel back, a cast copper-lead primary layer, a nickel dam (laid on top of the copper-lead mass) and a lead/tin/copper overlay provide the best performance for bearings used in high-stress applications.

In short, all high-performance bearings are of the cast copper-lead type.
Main and rod bearing shells feature a slight projection area when installed in the saddle or cap. When the cap is installed and fully tightened to specification, this crush height forces the bearing shells to attempt to expand outward, applying radial pressure concentrically around the bearing housing. This creates the proper geometric shape of the bearing I.D. and locks the bearing in place, preventing bearing movement relative to the bearing housing.

Contrary to popular belief, the small locating tangs/grooves featured adjacent to bearing parting lines are not responsible for locking the bearings in place. These tangs and grooves serve to locate the bearings during assembly only. The installed radial pressure is the primary force responsible for locking the bearings in place. Bearing shells are also slightly larger across the open end as compared to the housing (saddle/cap). When forced into the housing, this slight interference fit also holds the individual bearing shells in place during the assembly process.
Bearing shells are slightly tapered (thinner) approaching and at the parting line.

This creates an eccentric I.D. that promotes a pressurized oil film as the shaft rotates, allowing the oil film to compress and expand as the shaft rotates. This creates a hydroplaning effect, providing an oil film to quickly establish between the journal surface and the bearing. This is the reason that rod bearings are not grooved because you want to create a hydroplaning effect for maximum oil travel between the bearing and journal. The bearing I.D. eccentricity also dictates how an installed bearing I.D. is to be measured. The tightest areas (in terms of minimum clearance between the bearing and journal) will always be at the top and bottom (12 o’clock and 6 o’clock) positions, so bearing I.D. must be measured between these two points.

Performance bearings use a maximum amount of crush. If you over-compress the bearing, you’ll create a thick spot on the bearing wall, possibly minimizing clearance and oil travel. If you have insufficient crush, the bearing will be loose, resulting in bearing surface polishing or fretting (metal transfer).

According to Clevite’s John Havel, if you’re faced with slight geometry problems regarding journal geometry or misalignment, a good choice is Clevite’s V-series bearing, which features a 0.0010″ lead/indium overlay and no nickel dam. This bearing offers a softer surface edge that will be more forgiving if you’re dealing with slight misalignment. The P-series features a 0.0005″ lead/tin/copper overlay. The H-series rod bearings feature a 0.0005″ lead/tin/copper overlay, and the H-series main bearings feature a 0.0010″ lead/tin/copper overlay.


Start with 0.0010″ of clearance per inch of journal diameter. For example: 2.100″ journal diameter X 0.0010 = 0.0021″ clearance. For high-performance applications, add 0.0005″. If, for example, initial clearance is determined to be 0.0021″, add 0.0005″ for a final clearance of 0.0026″. From this point, tighten clearance as your experience dictates in specific applications.

Note: Use of a dial bore gauge is always the recommended method for measuring oil clearance. Instead of measuring journal diameter and then measuring installed bearing diameter, zero the bore gauge at the actual journal diameter. When you measure bearing diameter, you’ll obtain a direct clearance reading without the need to perform math procedures, avoiding potential math mistakes.

Havel emphasizes that if clearance modification is needed, do not increase or decrease clearance by modifying housing size outside of tolerance limits. An undersized housing will over-crush the bearing and an oversized housing will reduce crush and bearing retention.

Currently, Clevite utilizes finite element analysis computer modeling to examine the elastic deflections of all bearing-related areas. EHL, or elasto-hydrodynamic lubrication, allows engineers to more accurately determine the affects of dynamic forces in relation to forces and oil clearances. This understanding of loads, metal deflection and affects on clearance has allowed a more precise view of what the bearings are subjected to, and furthers engineers’ ability to develop bearings that will function properly in high-stress dynamic racing applications.



LS1 chamber.

LS1 intake ports. All except L92 and LS7 feature cathedral style intake ports.


Cathedral intake ports measure approximately 3 1/8″ in height.


Cathedral intake ports measure approximately 1″ in width.


Exhaust ports are D-shaped.


The LS6 heads feature a casting number of 243.


A view of LS6 intake ports.


Heads feature pushrod clearance reliefs.


The LS7 heads feature casting No. 8452.


Full view of LS7 rectangular intake ports.


Ported LS7 intake ports.


All Gen III/IV heads feature beehive-style single valve springs and small retainers.


Beehive springs take advantage of coil shape to ­minimize unwanted harmonics, eliminating the need for dual damper springs.


LS1 and LS6 heads require a rail for rocker mounting. The mounting bosses in the head casting are machined flat to accept this rail.


The LS7 needs no rail. Pedestals are machined with a female radius for paired-up rocker shafts.


LS7 intake rockers feature an offset.


In some applications, exhaust valves are hollow with sodium. photo courtesy Katech


Roller lifters are ­registered to prevent rotation in lightweight composite lifter trays.


Each lifter tray holds four lifters. The tray is secured in place to the head with a center bolt.


Camshafts are unique to the Gen III/IV engines. (photo courtesy Katech)


Here’s a close-up of a CNC ported combustion chamber on an LS6 head. (photo courtesy Katech)


Katech’s “Street Attack” 7.0L LS7. (photo courtesy Katech)


Katech’s LS7 crate engine. (photo courtesy Katech)


LS6 equipped with a supercharger. (photo courtesy Katech)


All Gen III/IV engines feature individual ignition coil packs. (photo courtesy Katech)

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