Tech Tips
BHJ Adapter Ring shimming

BHJ main bearing tip

When using BHJ style 2 inch adapter rings that may be undersized for the application some metallic tape will shim them out to the correct size. Most rolls of tape are of uniform thickness and can be easily cut. It stays where it is applied and comes off easily.

Small Block engine building comments with Roger King about his United Kingdom 289 Cobra.  Proof read 18 Sept 2016.

Good morning all, and thanks for this comprehensive document, Ladd.   

I get the strong feeling this should be published somewhere, as it clearly has relevance for a lot of what happens now with the ‘old car’ hobby.  The background context and way in which this science has developed is fascinating and makes for a very interesting, and informative read for those with a technical interest.  Particularly love the Smokey Yunick story:  first-hand records like that are both entertaining and valuable.  
I have no remarks for correction of this (how could I have?) and think it would be a very valuable contribution to the record for those that are looking for advice, so very happy for this to go on the forum if possible.

Best wishes to all


Good morning, I’m the fellow who provided parts and machine work to Roger, who’s engine is the subject of this thread. My name is Ladd. I wish to offer some thoughts to your group about engine break in and oil as they pertain to this build because many of you are doing similar work which may be improved upon. It seems there is a lot of interest in rear main seal issues too. I’d like to approach this from a perspective of posing then answering questions based on decades of conversations with leading experts in various fields of automotive and aircraft engine performance. And from mention of historical data that leads us to the technology used today. While I’m just a lone guy in a “hole-in-the-wall- shop; I’m also a guy lucky enough to have been mentored by noteworthy men and intuitions who took time to invest in me with answers to many questions so my work could become better. Perhaps this bit of writing will be of interest to you while given in that spirit.  

Engine break in was historically needed as a “finishing” operation because OEM production machine work accuracy and surface finish quality were not acceptable for full power operation “off the production line”. This situation existed from the beginning of combustion engine technology until recently, perhaps arguably until the last decade. Thereafter improvements in OEM manufacturing virtually eliminated need for post assembly mechanical break in art to mate and seal engine components.  Along with benefits of greater power being available sooner, great increases in engine longevity have been realized. Elimination of a wearing process to mate parts for operation generally called “break in” directly yields greater functional life.

The question of importance to us now, as we renew half century and older engines is: how do we apply new technology to obtain the same benefit of more power, sooner and longer? And if that benefit is built into a remanufactured engine, how does one hang onto it?  I believe answers to those questions are fairly simple to understand, once presented for discussion, and are found in three general areas. They are; manufacturing processes and materials, lubrication, and operation practices.

The small block Ford engine designed with knowledge accumulated by engineers in the late 1950’s, then produced starting in the very early 1960’s was considered “modern” back then. It was designed around improved technology, notably materials of cork / rubber and simple flat metal / carbon based center filler for gaskets, bi layer bearings, and improved alloy fully machined piston rings compared to engines of the prior decade. Prior engines used felt / leather, flat copper sheet or asbestos, Babbitt bearings, and low alloy poured cast ring technology.  Those newer designs also used precision manufacturing process, including select fitting instead of prolonged break in.

However legacy practices needed to produce longer lasting engines and better power from older designs lead to unnecessary and sometimes damaging events when applied to the then new design. For example; running new engines for short time intervals with cool down cycles, use of very low temperature thermostats (or none at all), and introduction of abrasive compounds to “set the rings” became obsolete, yet were (and are still) part of some mechanics considerations when dealing with a rebuilt modern engine. Retrofitting solid copper gaskets and leather oil seals sometimes had significant issues back then, and would be nearly unthinkable today.

Break in oil and oil additives were also marketed and used frequently. It is my opinion most break in oils of the 1940’s to the 1970’s fell into several groups. They were (for the USA market) mineral oil based, had limited additives, and lighter viscosity compared against standard 30w oil. They functioned by providing limited lubrication to the cylinder wall and ring set so those components would wear into a functional sealing condition (hopefully without seizing) while holding manufacturing debris / grit, in suspension. Holding debris and grit in suspension was not good for bearings so that oil was drained as soon as “normal” compression pressures were obtained. These products were similar to current Sunnen honing oil now used in cylinder finishing machines worldwide. The next group was viscosity enhancers which claimed to thicken and/or improve anti seizing risks in bearings when added to normally used oil products. STP is a brand name which comes to mind. It was not beneficial for seating new rings but was sometimes mixed with mineral oil in an attempt to gain advantages advertised by both products.

Another group was chemically active additives to enhance properties of normally used oils. They contained anti-scuffing and detergent chemicals to bond to grit and debris and “high quality” base oil stocks. It made these products expensive but beneficial. Most major vehicle and equipment manufactures had proprietary blends sold through their parts distribution networks. At the risk of oversimplification of a complex market environment and technical research it can be said this class of break in product evolved into additive packages now in modern oil.

Because these products were so effective as to become adopted worldwide in regularly used oils need for them as a break in supplement declined until some of them were removed for protection of emission control devices in the 1990’s. Resurgence of break in product marketing followed this genuine need for anti-scuffing agents, as did resurgence of folklore and legacy practices. I should credit Roy Howell, a Cornell graduate appointed as Chief Chemist at Red Line Synthetic Oil Corporation, Dema Elgin a camshaft grinder and instructor at De Anza College, and Joe Mondello who founded the Mondello Technical Institute for this understanding gained in many conversations over many years’ time.

We are now faced with continuing technology advancements leading to production of today’s modern engines. Gaskets are now neoprene / specialty rubber blends or can often be of coated embossed steels.  Specific adhesives and thread locking / lubricating products are common place. Tri-layer and flash coated bearing are typical parts while moly faced and low tension / multi material ring packages are also very common.

In conversation with John Erb, Chief Engineer, KB Pistons in the 1980’s he related a story about building pistons for Chrysler. They wanted to know what to expect by way of variation between larger and smaller pistons in manufacturing so they could plan their select fitting procedures. John told them there was not going to be a manufacturing variation significant enough to warrant any select fitting. Chrysler did not believe that claim initially, but in the end found modern piston manufacturing to be so precise a legacy practice of select fitting was no longer needed in their assembly line.  We, as mechanics, also have an even  larger body of legacy practice which worked on the ‘50’s to ‘90’s engines but sometimes doesn’t work when applied to now current designs, just as new parts technology may not retrofit easily into older engines despite apparently fitting mechanically. Prudent selection of which materials work well with specific processes used in engine re-manufacturing is critical for successful power production and longevity. Prudent selection is made by understanding history and changes from an evolutionary perspective. It also includes quality control inspection against known engineering standards. In our worldwide parts production industry manufacturing standards are often conflicted and obscure.

Oil, as a fluid in engine bearings, has two main functions; Lubrication and cooling. We have a tendency to focus on lubrication and “fixing” some falsehoods about how that happens, while forgetting about cooling so I’ll pose a historical scenario and question why that works as a lead in to modern design practice.

In the 1930’s and prior years many engines used dipper cups, splash, or vapor mist oiling for all bearings and friction points. There was no oil “pressure” as we know it today – zero - because there was no pump or circulations system. Yet these engines were capable of 2500 RPM and sometimes more, while producing a wide range of horsepower outputs, including some supercharged applications. How did oil at zero pressure prevent metallic contact failure? The answer is capillary action as applied to the load capacity film strength of the oil. Let’s do some abbreviated, approximate, and very shallow math analysis. 

Assume a crank rod journal size of 2.123 and a bearing ID of 2.125 by .75 wide. The difference is oil clearance of .002.

That yields a circumference for the crank of 6.67 inches leading to an area of 5.005 square inches.

Calculating a circumference of the rod bearing is 6.68 inches leading to an area of 5.010 square inches.

 The difference in area being .0055 square inches then leads to a volume calculation of .0055 x .002 yielding .000011 square inches oil space volume. This converts to .00018 cc’s of oil in the bearing. That isn’t much to provide cooling and load capacity so intuition says it needs to be circulated rapidly so it doesn’t absorb so much heat it chemically breaks apart into components no longer acting like oil.

While dynamic running pressures against a connecting rod vary quite a bit for many reasons 1200 to 1750 psi in the combustion chamber is a good starting point for conversation and “bench racing” calculation. In a 4 inch bore engine the piston has 12.56 square inches of area [3.14 x (2x2)]. That leads to calculating a connecting rod load of 12.56 x 1500 (average peak pressure) of 18,840 lbs. That is quite a big number.

Assume one half of the upper half of the bearing carries the combustion pressure load. This assumption is offered in place of calculus based on the geometry of the rod. Think- the lower half of the rod bearing isn’t in compression because it is below the load centerline. Just the upper half carries combustion generated loads. Of the area in the upper bearing half, a point at the top could be considered to carry the entire load while points at the side carry none of the load (being in slip sheer instead of compression). In actual fact that point load is spread out by the oil film so about half the area of the upper rod bearing carries combustion loads in an off TDC position.

Calculate half of 2.658 square inches to remove the lower part of the rod big end, and then half of the remaining upper bearing to find the load carrying portion, leaves .6645 square inches to carry 18,840 lbs. This very approximately calculated load on the oil film is then 12,519 lbs. per square inch.

The following excerpt from “Machinery Lubrication”, in an article by Robert Scott, illustrates this point.

“….The mean pressure in the load zone of a journal bearing is determined by the force per unit area or in this case, the weight or load supported by the bearing divided by the approximate load area of the bearing (the bearing diameter times the length of the bearing). …… Automotive reciprocating engine bearings and some severely loaded industrial applications may have mean pressures of 20.7 to 35 MPa (3,000 to 5,000 psi). At these pressure levels, the viscosity may slightly increase. The maximum pressure encountered by the bearing is typically about twice the mean value, to a maximum of about 70 MPa (10,000 psi).”

It is my opinion oil pressure developed by the engine’s pump, wither it be 45 or 95 lbs, when delivered via a .250 dia. feed hole isn’t going to counter balance that load without other factors being involved. So why have increased oil pressure? And why is there so much effort put into raising it? What are the other factors which really make an oil film lubricate a bearing?       

 A great article on rod loading is found at:   It is titled:

 Dynamic analysis of loads and stresses in connecting rods

P S Shenoy and A Fatemi Department of Mechanical, Industrial, and Manufacturing Engineering, The University of Toledo, Toledo, Ohio, USA

The manuscript was received on 25 June 2005 and was accepted after revision for publication on 6 February 2006.


 And an article covering how the oil film works is found at:


Rochester Institute of Technology RIT Scholar Works Theses Thesis/Dissertation Collections

8-8-2013 It is titled:

Analysis of Connecting Rod Bearing Design Trends Using a Mode-Based Elastohydrodynamic Lubrication Model Travis M. Blais


At the risk of doing a huge disservice to the scholarly papers’ authors, and by adding my historical perspective to their technical findings, our discussion of increases in pressure for bearing lubrication can be summarized in the context of my questions:

“Hot Rods” in the prewar era were melting babbitt bearings so needed to improve oil flow for cooling. The quickest and cheapest method to do that was shimming or changing their oil pump pressure relief springs to a higher value. Increased pressure correlated to slightly increased flow but also allowed increased bearing clearance while maintaining a functional oil film, which clearance was a far larger factor in increasing flow and cooling. Bearing clearances were then increased until their OEM pumps and available oil formulations could not maintain an oil film inside the bearing. Balances between oil film load capacity, heat removal from the bearing, running clearance, and pump output were discovered by trial and error. That could be favored by higher pressure, but at the cost of parasitic horsepower pumping losses and generating unwanted additional heat.  Inconsistent manufacturing, variation in oil products, lack of testing instrumentation, and fictional advertising hindered actual comparisons of successful designs.

 At this point real advances in bearing material, pumps, and oil technology were needed to create higher load and speed bearings for high output engines of WWII. This research continues today which has resulted in modern bearing lubrication at far higher loads and speeds than the 1930’s and post WWII era allowed. Oil pump manufactures after WWII started to market oversize oil pumps in high pressure and high volume versions so hot rod engine builders could tip the balance of a stock lubrication system towards maintaining an oil film when high heat removal from bearings was needed.

However better oils, more rigid bearings, improved surface finishes and higher temperature materials really allowed this advance. But our legacy habits of increasing oil pump pressure to attempt to gain a lubrication advantage persist when we should be gaining understanding, then implementing, modern changes to these other factors to upgrade our vintage engines. In my opinion, we as engine builders, have often placed the minor factor of increasing oil pressure into a role of being a major solution which limits our success in horsepower production to the driving wheels. Much of this information and history came to me over two decades time from Major William (Kelly) Owen, USAF, who among other noteworthy life achievements participated in Indy 500 racing from the 1930’s to 1980’s, and was Project Officer for the Cold Weather Test Detachment of the Proving Ground Command in Fairbanks, Alaska where they tried to make ground support and aircraft engines start, then run at full power in subzero temperatures.

An additional factor in oil pump pressure requirements is the effect of stroke of the crankshaft. Think for a moment of the crankshaft as a slinger style oil pump lubricating the crankcase. Oil enters the pumps center along the main bearing feed and is slung out through the rod bearings to the crankcase. If rod bearing clearance and/or the oil exit path from the bearing is larger than the oil feed hole area, and if the slinging force is greater than pump supply volume then pressure inside the rod bearing will fall to zero and the crankshaft passageway can be sucked dry leading to bearing failure. This is generally called “oil starvation”. However oil starvation failure should be divided into two causes, the first occurs when not enough oil is supplied, the second when too much oil falls or is sucked out of the bearing. This is similar to “cavitation” which isn’t mentioned much in terms of connecting rod bearings yet is well understood in inlet side design of oil pump systems.

Because lubrication of the crankcase is totally pointless and detrimental to horsepower production elimination of cavitation inside the rod bearing by other means instead of increasing the oil supply is a more desirable method of upgrading an older engine. These observations and comments arise from discussion with Gary Hubback of Los Altos CA, Bill Jones of Taylorsville UT, and Allan Lockheed of Bolder CO, who participated in teams running high power at Bonneville Salt Flats in record breaking cars.

This elusive and complex balance between pressure, flow, materials, and clearance was worked on by many prominent engine builders of the ‘60’s and ‘70’s. It was most notably codified by Smokey Yunick in his “10 lbs of pressure for every 1000 RPM” statement.  This guideline has taken on far more credence than current day engineering might indicate is needed. In a December 2000 address at the Superflow AETC Richard Maskin was asked what oil pressure his ProStock engines developed by a competitor who said he was having bearing trouble at 95lbs pressure in the 9000 RPM range. Richard replied his national record holding small block engines ran 35 lbs of pressure at above 10,000 RPM and his big block engines had 5 lbs more.  Reference AETC 11-13. Technology advances in the 16 years since then have not mandated higher oil pressures, although many people and teams routinely set up engines for 65 to 85 lbs maximum pressure.  In my opinion they are giving up power output advantages to be had at lower pressures when other modifications allow that set up.

Should we think of a traditional oil pressure gauge mounted in the output path of an engine pump as a pressure relief set point gauge? The moment an oil pressure gauge stops rising the pressure relief valve has opened. Further increases in pump speed simply dumps oil, heated by compression in the pump gears or vanes, back into the oil sump. It doesn’t even get into the filter loop. This is lost horsepower. Many engines achieve full pressure by 2000 RPM but are raced at far higher speeds. Reduction of pump capacity may be indicated after testing and specific research, while improvement in oil system circulation return deserves even more attention.

The V-8 small block engines we are rebuilding today from 50 years ago had oil capacity issues. Most performance designs from the OEM would hold 6 or 7 quarts of oil. The distribution at 5000 RPM was approximately one quart in each valve cover and 2 quarts in the valley under the intake manifold, and as much as half a quart in the timing cover.  This left only 2-3 quarts somewhere in the sump for the pump to intake then pressurize. Many engine builders fitted oversize sumps to “correct” this issue. In my opinion, improving oil flow back to the oil pan may have had a better overall return in investment, as seen in many European and Asian engine designs.   

I believe oil pressure should be thought of as a catalyst in formation of a lubrication film, not the principal force enabling lubrication, which is a property of the oil and geometry of the bearing. Once an engine has enough pressure to create and maintain an oil film, more pressure is detrimental. A watchmaker will lubricate bearings with a needle to which clings a drop of oil. When the needle is touched to the bearing oil instantly flows into the bearing clearance and will stay there for decades separating and protecting those surfaces without input of any pressure energy.  The Holy Grail of hot rod lubrication is zero friction, zero pressure, load carrying bearing systems. I believe high oil pressure is just a crutch we need to have fun running our cars today, while we figure out how build engines with near zero pumping losses.  I believe some builders have progressed along that path further than others.

However, I’m reminded of an occasion at a PRI trade show some years ago where Smokey Yunick was promoting “Prolong Oil Supplements”; additives alleged to reduce friction so effectively the oil could be drained from the crankcase and the car would still function. Prolong had bought a new Dodge Viper, treated it with Prolong, drained the crankcase, then had a famous race driver hot lap the car in a TV ad. About a half dozen of us cornered Smokey at a table asking him what happened to the Viper….. After giving us the promotional pitch in a few different ways and our continued pointed questioning he finally broke out with “what the hell do you think happened with no oil”. It was a fun day to be there for that “‘revised” marketing pronouncement.   

With this background information posted I’ll go back into Roger’s email threads to make specific comments in red mixed into his text in black.

I got into the rear main seal.  Glad I did.  The lip is disintegrating, with a couple of nicks here and there, and it feels very brittle and hard.  I can’t remember how old the gasket set was when I assembled the engine, I guess it could have been on the shelf as much as a year.  Then the completed engine sat in its frame for nearly two years due to unavoidable delays and I guess this combined is what has caused the problem. This is very odd and I’ve not seen one go like this before.  I usually clean with a lint-free paper wipe, no solvents, install and then apply a lubricant (oil, assembly lube) before laying the crank in place.   

This isn’t normal. I have neoprene (or whatever that material actually is) rear seals from old gasket sets that are 15+ years old and have worked fine. And when you think about it, seals in service last for longer than that. If you are seeing hardening to the point where it actually feels too hard then I’d suspect some chemical action or heat changed the seal material. Did it get near or in carburetor cleaner? That is a big no-no for seals.

 So, it’s a couple of days of very careful cleaning before it all starts to go back together again.  I have a new standard Melling oil pump and pickup to go in (using the ‘old’, 200-mile, ARP heavy duty drive shaft), even though the ‘old’ pump has only done 200 miles.  I thought I’d take that precaution even though everything is new. As usual I have stripped and measured the new pump as I trust myself more than whoever put it together.   

I’d have no hesitations reusing the old pump. It is qualified now as a known good part, not hurt in any way from 200 miles. The new pump may be defective – or calibrated differently- so has less value towards solving a mystery because it introduces new variables. And the new pump will shear off metal bits breaking in that the last pump has already rid itself of. 

The pressed-steel sump (oil pan?) I have is not a great fit.  I should have worried more when I found I had to use a thick bead of silicone on the pan side of the gasket for the entire surface that fits around the rear main cap - with the pan dry bolted-down with rubber gasket in place, you could see plenty of daylight through the gap where the curved section should have been compressed.  I have another that I have dry-fitted to a spare block and it is very much better, so that is this week’s job.

Oh my. This is serious parts mismatching and can lead to oil system problems. The rubber seal on the main cap depends on compression against the cap to seal between it and the cap. A leak will occur despite bonding of the silicon to the rubber and silicon to the pan metal if there is no compression against the cap.

 A “thick bead of silicone” in that area is also contrary to good engine building practice because it can break off. If any of the silicon is “missing” from your bead upon disassembly it would be wise to strip the entire engine oil galley system looking for it. Start in the oil pump inlet screen. From there, look in the oil pump relief valve. Then some may have gone into the filter or cooler plumbing which is usually the end of migration for a modified oil system. I say “usually” knowing I’ve found bits of silicon in the ends of pushrod tubes plugging up the rocker arm feed holes. I wonder how those gobs of silicon got through a lifter body but they did. If your system doesn’t have a plugged oil filter pressure relief hole (and there is little reason to block that hole in a stock system) then the silicon gobs can go anywhere in the engine.

I’ve had a few oil pans that didn’t fit too. That is so vexing and troublesome I’ll sometimes make a pan or modify one that fits the rails and end seals nicely to avoid buying one.

Latest brief update - Popped the engine out today, mounted it on the stand and rolled it over - it’s not the oil pan, it’s the rear main seal.  I don’t know why a new one has failed so soon.  There has been a spectacular leak since the first trip out, which is now unsustainable, ½ pint in 50 miles with oil dripping from everything.  It’s radiating out from the crank but luckily has not got on to the clutch.  I can see that the faces of the pan gasket are dry so that has been doing its job.
Odd, because I used a two-piece seal offset by around ⅜” with a small dab of RTV on the ends, and a light smear of RTV on the cap faces too.  Lubricated with a bit of assembly fluid on installation.  I have a PCV on one valve cover and an open breather on the other, so I’m not sure why it’s failed.

Let me interject here a few comments about how important proper PCV and KV systems are. Going back in time; before very early days of “modern” crankcase ventilation, prior to the 1960’s, there was one system. “KV” stood for crankcase ventilation and it was comprised of a road draft tube and vented valve cover breather cap(s). The system was designed to evacuate the entire volume of blow by gases at maximum engine speed by way of venturi action of air flowing over the end of the road draft tube from vehicle velocity. It was vented by valve cover “breathers” allowing entry of “clean” under hood air into the engine. It was understood blow by gases and fumes from burnt oil were detrimental to engine lubrication so needed removal.

 If the vehicle was stopped nothing happened- no venturi suction- but hot air convection might allow reversed flow out valve cover breather(s). This led to all sorts of folklore about determining engine condition by observing where and when oil smoke appeared. Because nothing happened at idle, and Army tests showed the single biggest operational wear factor for engines was dirt inducted by “open” unfiltered KV systems, two changes were made starting in the 1940’s. By the 1960’s they evolved into the Positive Crankcase Ventilation (PCV) system and a new Crankcase Ventilation (KV) system.

 In the Positive Crankcase Ventilation (PCV) system idle speed quantities of blow by were reintroduced into fresh induction air by way of small metering valve, and the road draft tube (KV system) was sealed or removed in conjunction with a new system of the same name venting valve cover breathers to filtered low pressure air in the air cleaner body. There was a distinct purpose to each system, a high speed one, and an idle assist system, but both depending on low pressure from manifold vacuum for effective operation. By the 1960’s systems orifices’ and plumbing hoses were sized for 200 to 300 CID engines producing 100 to 200 horsepower and blow by commensurate to those figures for “normal engine conditions”.

 Why do we use the same size orifices and hose sizes on 300 to 400 CID engines producing 300 to 500 horsepower today? Why do we, in some hot rod applications, remove these systems entirely? Or don’t connect the high speed (off idle) KV system to low pressure so it works as designed? It is well understood 7 lbs of negative crankcase pressure is a near optimum value for increased ring sealing in dry sump oiling systems, so why do we run our wet sump systems at just near zero or positive pressure with similar ring packages? Induction charge “purity” is a minor factor compared to gaining cylinder sealing improvements. In my opinion every effort should be made to lower crankcase pressure and promote evacuation of blow by gases from the crankcase, into the valve cover area, then out of the engine's internal space without heating engine oil excessively. Co-mingling and heating of oil from blow by occurs primarily in the crankcase and lifter valley area so a focus should be made on understanding flow patterns in those regions of an engine to promote separation of gases from liquids. Valve guide lubrication is a secondary benefit of proper PCV and KV system operation. 

I’ve stopped work for the night but will loosen the mains caps and remove the rear one tomorrow to have a look at the seal itself.  I should be able to change it with the motor on the stand - I don’t want to remove the crank, as that would mean disturbing the heads and pistons etc.  The question is, why did it fail, though?  We shall see (hopefully)!


On 21 Jun 2016, at 19:56, wrote:
Roger, could I ask you to hold off on re-installing your engine until I get a chance to tell you how to test that seal for sealing while on your engine stand?  Thanks,  Ladd 

A few years ago I did some engine program work for a team running a Toyota off road truck. They were switching from a 22R series 4 cylinder engine to a 5VZ-FE series V-6. They didn’t have a PVC system which would transfer ahead so neglected to install one. At racing speeds the engine started blowing cam and crank seals out of their housings causing massive oil leaks. I helped them build, then get a PVC - KV system installed and working, but rear main oil seal leaks persisted that were difficult  for their team to troubleshoot. Eventually I discussed, and then showed them how to check a rear main oil seal on an engine stand before it is run.

This method is from an old 1960’s A/C Delco emission testing manual and can also be found in Cummins diesel engine service publications.  First you fully assemble the engine with all covers and manifolds in place. Then you block known leak areas like road draft tubes, crankcase breathers, oil dipstick tubes and so on. Then you pressurize the engine to 3-4 lbs with air and spray the rear main (or any other gasket area) with soap solution looking for bubbles. Remember the engine will not hold pressure. It will leak down past the rings so air will be lost out the manifolds. This is normal, but fizzing or bubbles at seals and gasket interfaces indicate a site where oil will be lost from the engine in service.

On old Ford engines with dual valve cover standpipes for breather caps it was very easy to simply cut a bicycle tube in half, hose clamp each end to a valve cover standpipe, and then use the tube’s fill valve to pressurize the engine. I’m sure you can invent something similar for your Cobra. I’ve also used the oil dipstick tube as an air pressurization point.

A bit of trivia from Caterpillar engine company is their engine paint denoted as “Old Caterpillar Yellow” contains lead and other additives designed to seal gasket oil leaks externally. My understanding is workmen on 1970’s engine assembly lines were directed to liberally paint all engines on all surfaces to prevent leaks. Heavy painting as leak prevention against warranty claims was also taught and practiced at the Cummins engine training facilities when I attended there in the early 1980’s.

On Tue, 21 Jun 2016 19:50:38 +0100, wrote:

I have stopped using Felpro since I had a very odd coolant leak on the Mustang’s intake manifold 8 or 9 years ago.  The built-in silicone elements of the gasket had kind of melted, leaving coolant dribbling down the front of the engine, very hard to spot where it was coming from.  Luckily it did not go the other way into the lifter valley.  I only use Reinz now and have had no problems at all with them.

All gasket manufactures had leak problems with silicon “O” ring gaskets in the late 1990’s early 2000 model years. It was caused by incompatible chemistry in antifreeze. The debate continues today long after antifreeze use charts have been published by major manufacturers and gasket manufactures have changed the formulation of their gasket materials. A portion of that problem was different vehicle manufacturing standards for coolant applications in a world market.  

From: [] 
Sent: Tuesday, June 21, 2016 11:26 AM

Subject: Re: Oil leakage

Yes, I think the synth is the best option.  I did remove the crank with head and pistons in place, and even gave my dear wife the privilege of being involved - she guided the (appropriately protected) rod bolts around the crank journals during replacement.  I’ve put a two-part seal back in, a Reinz brand new one, ⅜” offset and a tiny bit of RTV on each end and on the cap mating surface.  Can’t get rope seals in the UK, and I’ve long lost the pin from the rear main cap.

I believe you need to check a couple of other things too.

A quote from Albert Einstein is “insanity is doing the same thing over and over expecting a different result”.  When I hear of people replacing a rear main seal over and over again I have to wonder if there might be some other cause for a leak than being misassembled.

I know many tens of thousands of rope rear main seals have been replaced with updated lip seal style neoprene assemblies. However when the rope seal engines were produced manufacturing tolerance of the rope seal area were not very tightly controlled. Manufactures counted on the rope seal material to conform to any geometry their tools cut. Now, decades later, seal manufacturers are making a “standard part” which might not conform to the block and cap tightly enough to create an oil proof OD seal while being perfectly acceptable on the ID against the crankshaft. Yet they feel “OK” sliding into place. Or the concentricity of the rope seal area may not be centered well enough for a lip seal to function. And the new seals themselves may have excessive manufacturing tolerances on the OD because they come from plants all over the world that each does some details a bit differently. Perhaps these issues are root causes for oil leakage falsely blamed on joint offset or misapplication of gasket sealer. I may be incorrect in saying no OEM offsets lip seal gaps in production. I believe it is a service procedure, sometimes of limited value, because if the OD crush of the seal is reasonably correct its ends will butt virtually oil tight and do so without any additional sealer.

In the Ford small block engine the rear main cap has a large window for oil drainage off the bearing. Unless that window is blocked by a mis-fitting oil pan, gross misapplication of silicon sealer, or dramatic over filling of the oil pan; oil pressure coming out the edge of the bearing falls to zero within a few thousands of an inch of the bearing edge.

The crankshaft has a slinger ridge which then guides oil from the bearing’s oil exit area into the drain window. Its forces of operation are parallel to the rear main seal lips so don’t contribute energy to the oil in a way to force it past the seal.  Space between the oil slinger cavity and rear main seal is mostly empty - free to drain.

If rear main bearing pressures on the oil are relieved before it gets to the seal what moves  oil across seals lips blocking that path ? What adds pressure to the oil so it seeks a lower pressure area outside of the engine?

A general answer is crankcase pressure from blow by gases. This may be true and is easily tested. However leaks persist even when this possibility is reasonably eliminated. I think it is wise to consider the possibility some pumping force created in the space between the crankshaft slinger and lip seal in the engine overhaul process  acts on the oil so very low or perhaps near zero crankcase pressure is added to and becomes enough to cause an oil leak that didn’t exist before. Or a leak that the more robust OEM rope seal closed successfully.  

I suggest checking and correcting the length and sealing quality of new replacement flywheel bolts may lead to stopping oil leaks in the vicinity of rear main seals. I believe if that bolt is too long it will grab oil in that cavity and spin it around just like a rotor vane pump does. The “pump” inlet becomes the slinger exit area while its exit is the drain window. The upper seal area can become packed with oil waiting to drain out the window which also leaks past the seal. Bolts that are too short cause a similar but less intense pumping action because of the void in the crankshaft threaded sections. This will be an RPM speed sensitive leak.

When we tested the Toyota team’s engine it had signs of oil radiating from the crankshaft on the flywheel clutch side. A bubble test showed leakage out the bolt heads. No sealer had been applied to new grade 8 bolts purchased at an industrial supply house instead of getting OEM bolts that had sealer pre-applied, a flat shoulder under the bolt head, and were “one time use” fasteners. Sealing the bolts was an easy fix for something that had vexed their team for many months.

On Tue, 21 Jun 2016 18:27:06 +0100, wrote:

I like this idea of personal tech service at a high level, I might get used to it. 

Hi Roger, It is a good thing to do for me and for you. About a third of my income comes from consulting now and I've met some really interesting people doing some pretty interesting things. I hope you figured out that the crank can be removed with the heads and pistons intact in the block. I'd put in a rope seal. I'm writing on a longer explanation now.   Go for the synthetic oil. Don't be afraid and never look back.   Lol.      Ladd

I’d have to say that top of my list of questions is, should I go to the full ester synthetic now (200 miles), and not worry about the ZDDP on the flat tappet cam? 
Whilst replacing this rear seal, I have been through the engine completely and it looks great.  It is a tribute to your skills, Ladd, that everything fits so nicely and neatly and all the bearing surfaces have such a uniform polished pattern with no signs of scuffing or uneven wear.  I can live with the oil pressure, although I’d expect it to drop a little bit more with a full synthetic. 

Thank you. I appreciate your work in my shop and conversation about the things we do. I believe you can see that traditional issues of engine break in, bearings, gears, chains, and valve train have already “broken in” in less than 200 miles. I think in other emails you have said the engine runs smoother than any V-8 you have previously owned so my guess is the rings and cylinders are working well also. As I mentioned elsewhere this engine was prepared with modern methods and parts so would “break in” very quickly, say in less than 10 minutes or 50 miles. So go for the full synthetic oil. Don’t hold back.

In the early 2000’s I went to a technical conference sponsored by Joe Gibbs racing where oil (among other things) was discussed in detail. Of the professional engine builders there we all had lost a few camshafts to lubrication failure in prior years. ZDDP was discussed and we were assured the Joe Gibbs product had enough to meet our needs “off the shelf”. This was possible because of a loophole in the EPA laws for small quantity manufacturers of specialty blends. It was also possible because different standards existed for HD truck oils than passenger car lubricants. A consensus was quickly formed that Shell Rotella and Chevron Dello diesel truck oil were acceptable alternatives to expensive additives. And later on I learned that Redline and Royal Purple products were “correctly” formulated for applications with high camshaft to tappet loads. Whatever the equivalent products are in the UK will work fine. I kept your valve spring tensions low deliberately so troubles would not darken your doorway but provided a set of race springs in case you wanted to push your RPM limit up to 7000 after a while.

I am also reminded of a 1980 conversation with Henry Styers, then a regional GM training instructor about small block GM camshaft failures. I’d made some cocky comment about being able to fix any cam problem GM cars had – just send them over to my shop and I’d put new parts in from after market sources. Whew- big mistake. He turned around and told me the problem was bigger than GM so who was I to mouth off that way about things I didn’t understand. I believe he’d been an Air Force DI, then officer candidate instructor in WWII, so his attention to disarming my cocky attitude was detailed, complete, and expertly done with the grace of a southern gentleman.

I learned from Henry that GM was fully aware of camshafts failing because of oil issues some years prior. They understood where government regulations had driven the oil industry, and were already testing valve train durability extensively. GM’s answer was an additive sold over their parts counters and provided to dealership mechanics when replacing engines. He also told me GM produced some blocks where the lifter bores were not angled correctly which caused failure. These were being quietly replaced under warranty. He denied any metallurgical issues with cams and lifters themselves. Henry ended on two points. First, whoever makes a car part can do it anyway they want to – it is their right to innovate, but the market would tell who did the best job, and he believed GM was the premier car company so had the best parts. Second that my invitation (as an independent mechanic durning a GM educational promotion) to attend a single training class was extended indefinitely at his personal recommendation. So I attended GM school as often as I could until Henry’s retirement in 1983 by which time I’d attended nearly every class offered.

 Henry’s passion for all things GM was counterpointed by Orion Yando’s passion for all things Ford. He was Ford’s western region general manager. I attended High School with his son Dick so observed first-hand how dedicated he was to advancing Ford products when occasionally in their home. Later I attended Ford Industrial engine school finding the quality of teaching staff excellent, and similar to GM school. A take away lesson I learned is, passion for a company’s product and detailed knowledge of the product go best when hand in hand.           

I have a question on oil pressure.  We went with std main bearing shells on the crank, and you gave me the options of building the rod bearings looser (all std) or tighter (std and +1thou) or tight (all +1thou).  I have gone with the tight option, using the oversize shells upper and lower in the rods.  I’ve now done around 200 miles in the car, but the oil pressure seems a little low at 20psi hot idle (650rpm) and 45psi hot at 3000rpm.  I’d welcome your thoughts on this, although I know these are acceptable figures for a small block Ford.

I read in some forum posts how you struggled with lifters wondering if they caused your lower than expected oil pressure. Your focus seemed to be on the oil passage band around the tappets center. My understanding of how that band works follows:

 It needs to be wide and deep enough that full oil flow and pressure can be passed the entire length of the engine oil galley. And there needs to be enough capacity in the band area to feed upper valve train components more or less equally through a hole in the tappets side. The depth of the band is limited by the diameter of the plunger inside the tappet vs. the body OD vs. strength of the body wall so it doesn’t break in half. The width of the band is limited by the geometric relationship of camshaft base circle diameter vs. lobe lift vs. length and height of the block’s bored tappet holes. Within these parameters the oil band can be located anywhere it doesn’t overlap outside the block tappet bore.

 If the tappet drops too low (towards the camshaft) exposing the band to crankcase volume you will get a high pressure internal oil leak of about the same magnitude as leaving the crossover galley plug out. However the band’s recess will enable side loading from camshaft rotation to nibble at the block’s tappet bore on every reentry. This will go on for a few hundred cycles until the tappet jams in the wallowed out bore, breaks the camshaft while blowing a piece out of the block, and maybe snapping the timing chain. I don’t think you had this problem. If the band comes above the block tappet bore a similar leak will occur but those consequences might be much less dramatic, or not. The tappet will hang up a little bit on each reentry maybe making a bit of a noise. It will take smaller nibbles out of the block but eventually will jam causing that valve to hang open. The open valve may hit a piston which will get your attention. In either case a high pressure or high volume oil pump might mask this issue visually on a gauge for a while.

 I check for tappet oil band fit by inspection in the block. Very early in the assembly process I trial fit a tappet into its bore deep enough to expose the band to the camshaft tunnel. Then I hold it there with a small copper jaw welding clamp and measure the distance from the top of the block's tappet bore or deck to the edge of the tappet body. Subtract .060 or more from that depth to determine maximum tappet drop. Then reset the clamp with the oil channel band exposed at the upper edge and repeat the measurement and calculation.

 The camshaft must not drive the tappet past either limit. If it does, something is seriously mismatched.  And the oil channel band must not move so far up or down as to allow the tappet body to block off the oil galley. This is something easy to see by looking down the galley bore with a flashlight. This design favors a wide tappet oil band channel. If the tappet body blocks the galley bore it will not show up by measuring maximum pressure at the OEM gauge location, which is ahead of this tappet system, because the pressure relief valve is open. Often times (in other engines families) an oil galley bore is offset enough the tappet body cannot ever block it. This design favors a narrow tappet oil band channel.  If your parts selection “passes” these inspections then oil leakage and flow will be “normal” no matter who manufactured the tappet. Problems arise here from non-standard base circle camshafts and incorrect application of tappets.

Oil leakage past the tappet body is important to understand. It can be calculated in ways similar to connecting rod clearance but we don’t know how much oil actually flows past the calculated space so this is a comparison of clearance volumes, not gallons per hour. The Windsor Ford tappet is .874 dia. (smallest) and .8745 dia. (largest) Manufacturing tolerances are very closely held so anything outside this range should be rejected as a defective part. If we accept .0027 as a clearance wear limit, oil at the tappet needs to pass a .048 cc clearance space. Circumference = 2.75 inches, passing approximately .200 inches twice (upper and lower tappet bearing areas) at .0027 clearance. Sixteen tappets would have .768 cc’s clearance volume – not much considering pumps are rated in gallons. However tappets erupt a literal caldron of oil in operation which needs to be returned to the sump as rapidly as possible through over sized drain holes and scraped off the crankshaft.


I am concerned that as the engine beds in these figures may go lower.  I’m running a mineral 20W-50 (high ZDDP) to run in, but was planning to go to a 15W-30 fully synthetic later.  I think that may show an even lower pressure though, so may stick with the mineral.  I’m a born worrier, as you can see!

 Noticing a pressure drop when switching to full synthetic oil from conventional oil is normal. Viscosity standards for modern oil and modern synthetic oil are different than standards for the 30w we grew up with. I believe the words of art used by petroleum companies are “equivalent protection level” when marketing those products. If you were to take a can of vintage oil and pour it out next to a stream of modern oil of the same advertised rating the new oils can would empty far faster. It will run through an engine bearing far faster too which is why it will show a lower pressure indication. But its load carrying capacity is slightly greater so it works, if not degraded. This is a point many people miss. Fuel dilution ruins multi weight oils faster than single weight products. Those oils were developed for cold weather – below freezing - rapid initial circulation lubrication, not for any advantage in more normal climates. Each vehicle operator will have to decide which danger is more important to protect his engine against, a cold start over rev before the engine has full oil pressure and circulation; or fuel dilution from a carb that isn’t quite right or sticking choke or PCV / KV system not balanced correctly. I like synthetic oils a lot for high temperature protection but their multi viscosity part isn’t needed until below freezing. So I’m not certain you need multi-weight oil at all, while sometimes being in the position of not being able to buy single weight full synthetic oil easily.

 I’m glad you set your crank up to the tighter specifications. The photos of your Plastic Gauge clearance checking are ok. There is modest controversy about its use. When I started using Plastic Gauge in the 1960’s the directions were to apply it to clean and dry bearing journals for the most accurate reading. That became problematic in my opinion because what you are trying to figure out is oil clearance not air clearance. And putting Plastic Gauge in dry created a problem getting it out again. I was always putting a dirty fingernail in there against a very clean crankshaft which bothered me a lot. Not to mention how I felt about scraping it off a new bearing shell. So now, a few hundred feet or more into use of that product, I always put it against an oil wet crank. I did noticed your pattern was a bit short though. I’d use a longer piece so it lays over the bearing shell edge on both sides. Seeing taper in a journal with plastic gauge is possible, but so is seeing a twisted rod, bent rod, error from torquing the rod cap (which unavoidably pressures the journal so the plastic gauge needs to be placed at 90 degrees to the cap parting line) and a couple of other bad or possibly illusionary things. Measuring bearing fit with a micrometer has limits of about a half thousands realistically, which is the same as plastic gauge. My opinion is both options have a great place in engine building but don’t trust either one on its own to be a final indicator. And I’ve nearly given up getting Plastic Gauge out. Just leave it in there. It is low temperature material so vanishes when an engine is started doing far less damage than the dirt under my fingernail might.

I noticed in a forum thread your idle speed was 650 RPM. That is too low. Light oil will not pump well at that pump speed. Remember your oil pump runs at half crankshaft speed. And it depends on constant RPM to keep pluses and flow going smoothly. Bump that up to 800 or so. Everything will work better. I’ve noticed many modern engines using light oil have pumps running at crankshaft speed. There are reasons for doing that which we need to think about before we run light oil by just pouring it in older engines.

I noticed in a forum thread a comment about 360 degree oiling not being necessary on a Windsor small block. I agree, but pose these questions: What happens when the oil passages at the main bearing shuts off every 36 degrees (180 degrees divided by 5 main bearings) sending a pulse echoing back to the oil pump? And what happens to cavitation of the rod bearing when smooth flow is interrupted? I think the answer is "nothing good". Which leads to another question: why didn’t Ford put in 360 degree oiling from the start?

An echo back pressure spike will rattle the pump gears and rattle the pump drive which can rattle the distributor gear drive harmonically. All that is part of the "nothing good" answer. An interruption of pressure and flow to the rod bearings means energy is dissipated doing no good while energy is later needed to restart that flow to get back to whatever level it was at. This is wasteful; and risky if conditions are marginal. Ford made a choice about their main bearings. A 180 degree oiling channeled upper shell could supply the rod bearing for stock use. The 180 plain shell lower could carry more combustion chamber pressure load allowing the entire bearing to be designed narrower and cheaper. So that is what happened. For economic reasons 180 degree oiling was chosen, not because it was a better design. But that was then and this is now. To get from 180 degree oiling to 360 degree feed of the rods takes another $60 main bearing set to rob the upper shells from and 10 minutes with a cut off wheel to re-notch the main web for a modified bearing tang location. The question becomes then: is that security and performance benefit worth it to you?

The only thing that was not normal during cam break-in was that the motor got excessively hot - bright red glowing cast headers - which was down to a distributor problem giving retarded timing.  Cam break-in was interrupted a couple of times to try to sort this, which was finally achieved with a different distributor.  I can’t see how this would affect the crank seal, but all is now assembled again by the book and I will report back on progress.

 That kind of heat affects different engines differently. I’d worry that your exhaust valve seats would become annealed, shrink, and fall out. Excessive exhaust back pressure can do the same thing. You’ll know in a few thousand miles about that. Munch-munch and crunch goes the piston top.

Heat in any cylinder head is removed by water and oil and exhaust gases. Oil from the rocker arm drips down the spring and across the head surface moving heat to the oil pan and out to the cooler from there. Some engines are rebuilt so oddly with high pressure and / or high volume pumps they flood the valve covers (and tappet valley area) and run out of oil in the pan to pump to the bearings. These engines don’t last long so to “fix” that issue some builders restrict oil flow to the valve train knowing roller rockers don’t need more than a mist of oil to function…. More oil stays in the oil pan – sort of- It is just pumped in a circle from pan to pump to pan via the pressure relief valve, while gaining heat. The small amount of oil running across the head becomes super-heated and quickly degrades. That heat and degraded oil hardens all engine seals over time.

Some engine builders went so far out on that limb after restricting oil flow to the valve train and fitting over sized oil pumps and sumps they installed bypass lines and spray bars directed at the valve springs to cool them. In my opinion simple and OEM over oiling the rocker arm cannot hurt anything, but under oiling the valve spring is death for the spring. So my position is; don’t add valve train oil restriction, improve oil flow drain back instead.

Evaluation of how much heat goes into a valve spring lower coil from the exhaust port is different for various points in the head. And some heads put the springs in very bad spots for staying cool. I have noticed when testing new spring sets they form a bell curve of pressure distribution. I install them so they do equal work by shimming installed height to nearly equal values +/- .015 then putting the slightly stronger springs found in the bell curve distribution into the slightly tallest remaining positions. This equalizes heat generated by spring compression in operation. I adjust that procedure if I’m using a dual pattern cam. The goal is to make the springs all last an equally long time. If I cannot make a “set” of valve springs from 16 I’ll buy another set to increase the number of springs in a favorable portion of the bell curve distribution. Isky makes really good valve springs at a racer / hot rod level. Comp cams springs seem of inconsistent quality, but getting better. When valve springs are run very hard, hot, removed, and retested a bell curve distribution of pressure has gone away and more random pattern has formed, although sometimes that pattern is distinctly exhaust or intake side denoted. Adding oil deflectors to the top of rocker arms so the lubrication spurt is directed down into the pivot and into the spring coil (instead of up onto the valve cover and wasted) is very helpful. Ford did this on their big block industrial engines.     

Thanks, Best of luck to you going back together and when on the road.   Ladd

Thanks for all your work in preparing this reference work.  It is really helpful and should be there for others to read.
Oh well, it seems we’ve just left Europe - time to get back in the workshop and put politicians out of my mind for a while at least.
I’ll report back, All best wishes to all     Roger


 Dear both {Kevin & Ladd}
, the engine’s back in and running as sweetly as ever.  I’ve done a 40-mile round trip since refitting it and no signs of any leaks so far.  I did your RMS {rear main seal} test with the motor on the stand, and there were no bubbles, so fingers crossed!  I also checked the flywheel bolt length as discussed:  they are ARP bolts which are 5mm short of the back of the mounting flange, so leave a slight cavity here but not, I would think, a significant one.

 The faster you turn the crank the more important it becomes.

 It’s now running on Fuchs Titan Pro R 15W-50 full ester synthetic (used to be called Silkolene), which is what my Mustang runs on.  Interestingly oil pressure has improved at idle to around 25psi, with 45-50psi at 3000, all good.  It certainly runs very sweetly but I do need to get it on a rolling road for carb tuning.  That will have to wait until I’m back from Le Mans next week - I have set idle mixture at around 13:1 and am running 55 (Holley sizes, in Autolite 4100) primary mains as the plugs were just a little bit white on 54s.  Ignition timing all-in at 34° @ 3000rpm, no pinking so far but the rolling road will iron out the fine detail.


Hi Roger, glad you are running again. I enjoy hearing my tips helped your efforts succeed. I’m not sure why your oil pressure improved. I’d guess the oil which was drained out had a bit different thickness (actually thinness) so didn’t pump the same. If you get a good KV system installed I’d recommend going onto an oil analysis program for a year or two so you don’t over change expensive lubricants while discovering how seldom that needs to be done in today’s vehicles. We found on the IMCA car, even running in the dirt at very high oil temps of 300+ F, Redline synthetic oil would not break down until half the season was past. Dirt ingesting from open breathers was an issue though. In a half dozen other engines I’ve had on synthetic oil, with monitoring, (BMW, Honda, Ford on Mobile I) sometimes unbelievable (to me) change intervals (8 to 10,000+ miles 3 years) were discovered to be fine chemically. These engines have all gone great distances in high speed/hard service and are on their second or third hundred thousand mile go round without any lubrication wear issues or any “overhaul” service what-so-ever. I believe data gathered in oil analysis can lead to economic savings (and catch early failures) so recommend it while clients get their new engines figured out operationally.

 Some years ago when pump gas fuels started to greatly change in formulation several of us listened to Rick Gold from ERC fuels discuss tuning by spark plug color at an advanced engine theory class. I was also listening to Bill Jones views on that subject. My understanding at that time was, several men, including Bill, were helping Alan Kulwicki with his NASCAR engine program. Alan’s spark plugs would be sent overnight FedEx to Bill who would mill them apart for color inspection, render his opinion, and send them back to Alan so his team could make jetting changes. I saw those spark plugs, examined them, and listened to their opinions carefully.

 Several beliefs were formed in me at that time which have held true over the years since. They are: Jetting by plug color with race fuel is possible if the fuel is always the same and you don’t change ignition around or compression, and you relate jetting changes with other weather, dyno, and engine use. Jetting by plug color for pump gas on the street is impossible because the lead is gone which formed the glaze which gave the additives their colors. There is simply too much “white” zone where data isn’t generated by fuel burning on the spark plug. You can select heat range by electrode condition, but that isn’t the same as selection of jetting for mixture.

 At that time O2 sensors were pretty new for mixture control – maybe a decade old - give or take a few years. And portable analyzers were pretty much out of reach financially for everyone; so there was a period of time when setting carburetor jetting got really hard to do. It became an issue of experience and sensing what an engine needed more than an issue of looking at information from a spark plug and tuning binder records. Now methodology has changed again.

 There are a couple of really good and easy to use portable air / fuel ratio gauges on the market which don’t cost too much. I use that method and like it a lot, but should mention this consideration when setting mixture from an air / fuel ratio gauge vs. a spark plug. The spark plug gave an average reading with its color. It kept you safe if you snuck up on going too lean. The air / fuel gauge gives an instant reading. You can get too lean very fast if you just jump into big changes because the gauge shows some value. This means to me making little thoughtful changes, one at a time in tuning, is more important than ever. But I’m not a “tuner”. There are wizards, scientists, and kids with laptop computers who do that now.             


Rear axle unit’s out now as I am changing the gearing from 3.07 to 3.31.  I have a couple of days to get the new one in and do a test run.
I look forward to seeing Ladd’s contribution to the forum - this kind of information needs to be published for everyone's benefit.

Ok Roger. That is it for me. I think I’ve covered my build theory and practice enough for anyone to wade through. I’ll send this to you first for comments and corrections then maybe post to the Ford engine forum. 


Notes on Piston Ring Sealing    January 11  2014

Hi Roger, Kevin passed this link on to me so I thought I'd let you know we follow your thoughts and ideas over here pretty close. And we hope your build is going well. If I might add a few stories about the engine bearing and glaze to comments you posted, maybe something might "ring true" for the group and move all of us to a better understanding.

In the mid 80's I was running a dirt track stock car under NASCAR rules in the Winston West series. I was very lucky to get firsthand advice from manufactures of rings and bearings through their regional tech specialists. Their position was (is) moly faced rings if put against a properly surface finished cylinder wall "seated" virtually instantly. And engine bearings assembled against a properly finished and sized crankshaft required no break in period. After wrapping my head around those statements I stopped trying to develop a perfect break in process and instead devoted additional effort into improving the surface finish and size control of the parts I put into my (and my clients) engines.

This effort put me squarely into questions about cylinder wall glazing. At the time Smoky Yunick was just near the end of writing for Circle Track magazine so I posed the question "what is cylinder glaze and how do you form and control it" for his last technical column. He answered my question in a partial way saying it was a product of the oils and additives in the engine formed when exposed to heat. However we continued exploring that answer privately the following year at the PRI show. His further comments reveled to me the importance of thinking through some pretty self-evident issues, among them the fact that an engine’s cylinders cannot function well without a glazed or very smooth surface for oil film to work with. And many engines assembled without any piston rings will still pump more than 50% of their designed compression pressure. Therefor controlling formation of glaze is the primary goal of a modern break in process, not the final resizing of rings or bearings.

This put me in a hard spot to reconcile historical engine machining folklore vs. modern manufacturing processes. And got me thinking about how to form a properly glazed cylinder wall. I searched a lot of SAE information and talked to several nationally known winning engine builders finding no single product that everyone successful used. I was lucky enough to trade Joe Mondello some intake manifold machining fixtures for a month with Joe in the early 2000's. This enabled attending his engine building and porting school while picking his brain on this subject (and many others) too. I can say with near certanity that nobody at the national level uses WD-40 to lubricate piston rings or bearings for final assembly, but it does make a fair rinse off product for fingerprints and invisible airborne contamination. I can say that some people at national levels use assembly oils which have the following characteristics: increased clinging power, increased moisture tolerance, enhanced load capacity, and ash free burning. My choice for final engine assembly (except in the valve train) has been 2 stroke outboard motorboat oil for a long time. It seems to work for me and is consistent in properties with what some of the best in the business are doing. And most of that gets flushed out of these vintage engines before starting by an engine prelube procedure using Dello 15-40 diesel truck oil.

Two interesting comments to close with follow: the first from a SuperFlow Engine Conference in the late 1990's by a presenter who builds Pro Stock engines for clients and their companies lease program. He said with hot honing and bore plates; cylinder geometry is pretty much perfect for the rings to seal against. And the rings barrel face and composition are pretty much perfect for the cylinder material. Those combinations are well engineered and known. So the real challenge to overcome to produce maximum compression and power is in perfecting the seal between the piston lands and the ring. So they brought their piston production into their facility away from "outside vendors" and buy box stock rings now. And the last from Smoky Yunick who told me most of the problems attributed to glazed cylinder bores are actually gummed up rings stuck in the piston lands from incorrect assembly lubrication (or clearances) which cannot thereafter seal the bore properly (I think that came from testing either for or in GM's engine lab in the 1960's).

A final point to consider is every manufacturer of engines I am aware of specifies some variety of SAE or API rated engine oil for ring assembly including HD truck, marine, and aircraft. The days of folklore about non-detergent oil and other mystery stuff are long gone.  

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