Father Son 75 CB200T Rise From the Ruins
- Thread starter Texasstar
- Start date
barnett468
Coast to Coast
Texasstar said:
LOL, actually my "handle"is just as insignificant as i am . . i am not affiliated with barnett clutches or the motocross racer mark barnett etc. i actually came up with it because i needed one for ebay when i signed up because i didn't want my real name all over the internet because there are some crazy people out there, so i was at my friends auto shop when i signed up and i was racing vintage motocross at the time so i thought of barnett clutches, but ebay also required that numbers as well as letters be used for passwords or something so i thought of the infamous gm engine that could run on 4, 6 or 8 cylinders and just combined that with barnett.
back with tungsten info very shortly.
barnett468
Coast to Coast
deviant said:Hopefully a big block.
Do you mean like a stroked Ford FE engine or one of those girly orange chevy things?
By the way, did you get yours fixed?
HAPPY HOLIDAYS!
Ha. Mine runs great now. It's a '79 Chevy C10. Just installed new headers. For you, I guessed a Shelby 468. I know of your silly Ford leanings.barnett468 said:Do you mean like a stroked Ford FE engine or one of those girly orange chevy things?
By the way, did you get yours fixed?
HAPPY HOLIDAYS!
barnett468
Coast to Coast
Texasstar said:
xlnt, however, he will be able to shift if MUCH faster if you have all the tranny gears and shafts "REM finished . . this is a metal finishing process that makes then smooth and slick and it too is used by Nascar and many World Cu Cars in their transmissions and differentials, and is also reasonably affordable.
the ultimate may be to have them coated with tungsten afterwards, however, it would have to be determined it the application technique for the tungsten would damage and re finishing.
it will also shift slightly easier if the preload on the shift drim is reducedm howeverm if it is reduced too much, it can allow it to fall out of gear so this will need to be tested to see where that threshold is.
Also, if he doesn't roll off the throttle quite as much when he shifts and simultaneously pulls the clutch in just a little when he shifts, it will shift much faster guaranteed . . shifting without the clutch on that particuar bike is actually a slow method . . no clutch shifting works much better on a 2 stroke motocross bike.
if he wants to learn how to shift faster, which i think is an xnt idea, he should dedicate some time to focus on absolutely nothing other than that for a while and maybe only shift thru 4th [what he apparently calls third] so he can get back and start again sooner.
barnett468
Coast to Coast
deviant said:
C10 stands for how many minutes it takes to do the quarter mile
deviant said:
Why, it won't help cuz its a chevy.
deviant said:
That would be a 428 MR Chevy guy, and yes, I actually have been fortunate enough to have a few but too poor to keep them.
deviant said:
Yes, I learned that if I actually wanted to get where I was going without the help of a tow truck, I needed to drive a FORD!
OMG Barn...that's brutal...and the day before Christmas!
barnett468
Coast to Coast
Texasstar said:
I actually have no idea what Yoshinura does on anything and have never read a book by him or Smokey etc, so I looked Yoshimura up on google and found the link below . . Also, REM and ISF is just a different name for the same thing, and I think the guy that invented the process did it so he could use it on silverware or something odd like that . . It was later that racing teams figured out it would be beneficial to gears.
http://www.yoshimura-rd.com/content/services-isf-process.asp
I'll look again but I think it might only be around $200.00 to do your entire tranny . . It goes into what looks and acts similar to a vibrating polisher with a bunch of mildly abrasive sponges, but a chemical is also added which is what fills in the low spots and is what I want to know if the tungsten process will damage.
Another option that will work well is to simply have the trans lightly stone polished in a tumbler for maybe $100.00, then have it tungsten coated for around $130.00 and I guarantee you it will then be slipperier than heck AND far more durable because of how hard the tungsten is . . This is the method I would use if I wanted the sickest shifting trans on the planet.
barnett468 said:C10 stands for how many minutes it takes to do the quarter mile
Why, it won't help cuz its a chevy.
That would be a 428 MR Chevy guy, and yes, I actually have been fortunate enough to have a few but too poor to keep them.
Yes, I learned that if I actually wanted to get where I was going without the help of a tow truck, I needed to drive a FORD!
OMG Barn...that's brutal...and the day before Christmas!
Ha. You can't burn me more than the family already does. The wife's an Allison and her father was a Ford engineer. My uncle worked under Iaccoca way back in the 60s. Nothing fancy. I have owned a Galaxie 500 way back when I was young, though. Truth is, you'd probably make fun of me more if you knew what I had in the garage. It's not Ford or Chevy.
barnett468
Coast to Coast
deviant said:
Well at least you had one good car in your life, and if the one in the garage is eiher a Mopar or a Rambler then your alright in my book . . I have had both.
Texasstar
Can't is a four letter dirty word
Watching the competitive shooters on YouTube who use Tungsten Disulphide on their bullets.
In this paper on tungsten disulfide it said the military used it and Two Stroke Motorcycles http://www.usbr.gov/research/projects/download_product.cfm?id=312
Sent from my iPhone using Tapatalk
In this paper on tungsten disulfide it said the military used it and Two Stroke Motorcycles http://www.usbr.gov/research/projects/download_product.cfm?id=312
Sent from my iPhone using Tapatalk
Yep. 1970 Roadrunner. Numbers matching 383. My grandmother raced them, so it's in the blood.barnett468 said:
barnett468
Coast to Coast
deviant said:
ok, you can be my friend now . . i have had a 69 and 70 . . my 70 was a 383 with the 3 foot long pisto grip shifter.
had some cudas and challengers too . . love them mopars.
barnett468
Coast to Coast
Texasstar said:
yeah and now the gov wants to use it to prevent muscles from attaching themselves to underwater structures , lol.
i'll post more info on it shortly.
barnett468
Coast to Coast
Texasstar said:
LOL, Nope, and the dipping process they used to apply the tungsten to the piston and rings is definitely NOT Barnett468 OR NASA approved, however, I would gladly give it my blessing for a pizza pie pan.
"Then it was off to Cosworth for Norman-designed two-ring pistons, which then had their short skirts and rings dipped in tungsten disulfide (WS2), titanium nitride (TiN) and other concoctions by Brycoat to lessen internal friction."
I'm getting closer to posting the info you requested...really...i swear.
barnett468
Coast to Coast
TUNGSTEN DISULFIDE APPLICATION TECHNIQUE
This coating thing is all incredibly complicated and is basically impossible for me to explain but in general, there are a few ways to apply it . . Two different techniques are brushing or spray painting it on at low velocities which is also done with some other types of coatings also like the ones companies like Wiseco, Mahle and JE put on their pistons nowadays, however they apply these coatings using a silk screen process which should only be used on parts that do not have high loads or critical tolerances, because this process uses binders/glue so it won't stick any better than the glue will, which means it can be scrubbed off of parts that see high pressures like camshafts if the oil barrier is broken etc . . The thickness/surface of the tungsten is also irregular when applied with the brush or spray paint method which increases its cf factor.
As I understand it, there is a certification/standard number Military Spec DOD-L-85645 & AMS2530 that has been established for the application of tungsten disulfide which NASA requires vendors to meet to get their business, and as far as I know, there is only one place in the US that has this certification . . Many parts used in aviation are not required to have the tungsten applied using the certified process/method because those particular parts are not used in critical areas.
The certified process which is called "impinging", was established by the "inventor" of the product in 1966, but at that time there was not certification certificate established for it.
Impinging is a complicated scientific phenomenon that i cant explain well other than it means that the material is applied in a vacuum environment with high air speed like around 700 miles per hour and no binders/glues can be used . . one reason no glues can be used is because they can emit gasses into NASA's two bazillion dollar space craft which can affect its instruments thereby possibly causing them to guide it to Afghanistan instead of mars etc, and since its not a bomb, the government sees no reason to send it there . . Anyway, this process causes the tungsten to actually create a molecular bond with the metal which is a bazillion times stronger than the bond glue can provide . . In other words, for all intents and purposes, it becomes one with the metal and the only way to remove it is to grind it off.
CHARACTERISTICS
As I mentioned, it works best when it is applied to two mating surfaces which is the case with many anti friction coatings, however, this is not always practical nor is it always necessarily a good thing . . One of the interesting things about many of these types of products is that the more force there is between the parts when these products are applied to both surfaces, the lower the coefficient of friction becomes.
Many of these products are referred to as "dry lubricants" which basically means that no lubrication is required for them to prevent galling, however, this theory only works to a point because I'm confident that if even the best one was applied to an engine and was run with no oil, it would still seize, however, part of the reason for the failure is not entirely due to lack of lubrication, it is also partially caused by the friction between the parts getting so high the anti friction coating fails . . i would actually love to see a dry load test of the material using extremely high pressure where the parts could be kept cold enough so the temp does not reach the failure point to see just how long it would last.
Also, applying "soft" anti friction coatings like Teflon to high pressure areas like a cam won't work because the pressure will destroy it.
FRICTION REDUCTION AND PARTS PROTECTION
These coatings reduce friction because the oil slides off of them easier than it does on bare metal . . These coatings should not, and will not actually ever contact an adjacent surface as long as their is a layer of oil between the surfaces which is exactly how engines etc are intended to work, however, since it is not a perfect world, and not all oils are created equal [the wife and I prefer Wesson to canola], this is is not always the case, which is one reason why engines do not last forever.
The ZDDP or Sulphur or Antimony or Boron that some oils have is only to protect parts if the oil barrier becomes too thin because they sort of stick to the parts, therefore some of it will remain on the part even if the oil does not . . Castrol even has an oil now that has microscopic particles of Titanium in it as a wear inhibitor.
They even have a coating that can be applied thick on a piston that is too small/loose for a bore and apply enough to create a slight interference fit and the coating will wear to the tolerance it needs without damaging the bore, however, the coating will continue to wear whenever the oil barrier is insufficient . . i personally wouldn't use this type of coating in an engine..
WHY DO HONDA 160, 175 AND 200 CAM BEARINGS REALLY GO BAD.
The reasons are a few, however, the main one is due to insufficient oil upon initial start up . . The the reason they have insufficient oil upon start up is because it simply drains back down the oil hole that runs thru the cylinder when the engine is shut off and because the oil pumps suck, it takes a few seconds to get oil pressure to the top of the engine and the surfaces depend upon actual pressure to cause the cam to "float" on the oil film
When the engine is turned off, the oil pressure is gone and the pressure of the cam being forced down onto the cam bearings by the valve springs, squeezes out most of the oil, therefore, the only protection these parts have for that 1 or 2 seconds until the pump provides oil and oil pressure to them is a microscopic layer of whatever cheap ass oil the cheap ass owner put in his bike.
Here's a simple experiment one can do that will help illustrate this point . . get a piece of smooth flat clean metal, then place a clean finger on it with a moderate amount of pressure, then try to drag your finger across it . . next put some 20w-50 oil or wheel bearing grease ec on it and do the same thing.
automotive systems have what is called an anti drain back vale in them that is located close to the oil pump . . this prevents the oil from draining back out of some of the system which in turn creates pressure in the system and supplies oil to the parts fairly quickly . . if the 160 and 175 and 200 bikes as well as many others had an anti drain back valve, the cams and bearings and rockers would last much longer.
DLC COATING (Diamond like carbon)
Info coming up.
CAM, ROCKER AND BEARING WEAR CURE
info to come
This coating thing is all incredibly complicated and is basically impossible for me to explain but in general, there are a few ways to apply it . . Two different techniques are brushing or spray painting it on at low velocities which is also done with some other types of coatings also like the ones companies like Wiseco, Mahle and JE put on their pistons nowadays, however they apply these coatings using a silk screen process which should only be used on parts that do not have high loads or critical tolerances, because this process uses binders/glue so it won't stick any better than the glue will, which means it can be scrubbed off of parts that see high pressures like camshafts if the oil barrier is broken etc . . The thickness/surface of the tungsten is also irregular when applied with the brush or spray paint method which increases its cf factor.
As I understand it, there is a certification/standard number Military Spec DOD-L-85645 & AMS2530 that has been established for the application of tungsten disulfide which NASA requires vendors to meet to get their business, and as far as I know, there is only one place in the US that has this certification . . Many parts used in aviation are not required to have the tungsten applied using the certified process/method because those particular parts are not used in critical areas.
The certified process which is called "impinging", was established by the "inventor" of the product in 1966, but at that time there was not certification certificate established for it.
Impinging is a complicated scientific phenomenon that i cant explain well other than it means that the material is applied in a vacuum environment with high air speed like around 700 miles per hour and no binders/glues can be used . . one reason no glues can be used is because they can emit gasses into NASA's two bazillion dollar space craft which can affect its instruments thereby possibly causing them to guide it to Afghanistan instead of mars etc, and since its not a bomb, the government sees no reason to send it there . . Anyway, this process causes the tungsten to actually create a molecular bond with the metal which is a bazillion times stronger than the bond glue can provide . . In other words, for all intents and purposes, it becomes one with the metal and the only way to remove it is to grind it off.
CHARACTERISTICS
As I mentioned, it works best when it is applied to two mating surfaces which is the case with many anti friction coatings, however, this is not always practical nor is it always necessarily a good thing . . One of the interesting things about many of these types of products is that the more force there is between the parts when these products are applied to both surfaces, the lower the coefficient of friction becomes.
Many of these products are referred to as "dry lubricants" which basically means that no lubrication is required for them to prevent galling, however, this theory only works to a point because I'm confident that if even the best one was applied to an engine and was run with no oil, it would still seize, however, part of the reason for the failure is not entirely due to lack of lubrication, it is also partially caused by the friction between the parts getting so high the anti friction coating fails . . i would actually love to see a dry load test of the material using extremely high pressure where the parts could be kept cold enough so the temp does not reach the failure point to see just how long it would last.
Also, applying "soft" anti friction coatings like Teflon to high pressure areas like a cam won't work because the pressure will destroy it.
FRICTION REDUCTION AND PARTS PROTECTION
These coatings reduce friction because the oil slides off of them easier than it does on bare metal . . These coatings should not, and will not actually ever contact an adjacent surface as long as their is a layer of oil between the surfaces which is exactly how engines etc are intended to work, however, since it is not a perfect world, and not all oils are created equal [the wife and I prefer Wesson to canola], this is is not always the case, which is one reason why engines do not last forever.
The ZDDP or Sulphur or Antimony or Boron that some oils have is only to protect parts if the oil barrier becomes too thin because they sort of stick to the parts, therefore some of it will remain on the part even if the oil does not . . Castrol even has an oil now that has microscopic particles of Titanium in it as a wear inhibitor.
They even have a coating that can be applied thick on a piston that is too small/loose for a bore and apply enough to create a slight interference fit and the coating will wear to the tolerance it needs without damaging the bore, however, the coating will continue to wear whenever the oil barrier is insufficient . . i personally wouldn't use this type of coating in an engine..
WHY DO HONDA 160, 175 AND 200 CAM BEARINGS REALLY GO BAD.
The reasons are a few, however, the main one is due to insufficient oil upon initial start up . . The the reason they have insufficient oil upon start up is because it simply drains back down the oil hole that runs thru the cylinder when the engine is shut off and because the oil pumps suck, it takes a few seconds to get oil pressure to the top of the engine and the surfaces depend upon actual pressure to cause the cam to "float" on the oil film
When the engine is turned off, the oil pressure is gone and the pressure of the cam being forced down onto the cam bearings by the valve springs, squeezes out most of the oil, therefore, the only protection these parts have for that 1 or 2 seconds until the pump provides oil and oil pressure to them is a microscopic layer of whatever cheap ass oil the cheap ass owner put in his bike.
Here's a simple experiment one can do that will help illustrate this point . . get a piece of smooth flat clean metal, then place a clean finger on it with a moderate amount of pressure, then try to drag your finger across it . . next put some 20w-50 oil or wheel bearing grease ec on it and do the same thing.
automotive systems have what is called an anti drain back vale in them that is located close to the oil pump . . this prevents the oil from draining back out of some of the system which in turn creates pressure in the system and supplies oil to the parts fairly quickly . . if the 160 and 175 and 200 bikes as well as many others had an anti drain back valve, the cams and bearings and rockers would last much longer.
DLC COATING (Diamond like carbon)
Info coming up.
CAM, ROCKER AND BEARING WEAR CURE
info to come
barnett468
Coast to Coast
.
BIKE WEIGHT REDUCTION
Remove starter, starter gear and the one way clutch on the back of the flywheel then plug the hole with item 163.
http://www.cappellinimoto.it/index.php?option=com_djcatalog2&view=itemstable&cid=4&Itemid=11&lang=en
CAM SHAFT TIMING
If you have shortened your barrel and/or milled your head, your camshaft is now retarded from the mfg's intended installed position . . changing the installed position will change how it performs, and the installed position is NOT always the best for every engine, therefore, to get the optimum performance from any non custom cam, it is really necessary to try different positions on the dyno . . they way you can do that is by installing an adjustable timing gear which you will find on the page in the link below.
http://www.cappellinimoto.it/index.php?option=com_djcatalog2&view=itemstable&cid=4&Itemid=11&lang=en
PISTON PINS
You can reduce reciprocating weight by using tool steel or titanium pins . . have them tapered on the inside . . if you use titanium ones, have them coated with tungsten . . this should increase rpm by maybe 80 to 100.
http://www.aperaceparts.com/wristpins.html
these ape wrist pins were in the worlds fastest drag bike when it grenaded but they didn't break . . your pins wont need to be quite so strong, so tell them your app so they can make them as light as possible.
http://www.cycledrag.com/ape-wrist-pins-withstand-mcbrides-horrific-explosion
CAM CHAIN ROLLER FOR 160, 175 &200
The roller in the bottom of the photo takes a beating and can be replaced with a metal one for around $100.00
P/N 154
http://www.cappellinimoto.it/index.php?option=com_djcatalog2&view=itemstable&cid=4&Itemid=11&lang=en
BIKE WEIGHT REDUCTION
Remove starter, starter gear and the one way clutch on the back of the flywheel then plug the hole with item 163.
http://www.cappellinimoto.it/index.php?option=com_djcatalog2&view=itemstable&cid=4&Itemid=11&lang=en
CAM SHAFT TIMING
If you have shortened your barrel and/or milled your head, your camshaft is now retarded from the mfg's intended installed position . . changing the installed position will change how it performs, and the installed position is NOT always the best for every engine, therefore, to get the optimum performance from any non custom cam, it is really necessary to try different positions on the dyno . . they way you can do that is by installing an adjustable timing gear which you will find on the page in the link below.
http://www.cappellinimoto.it/index.php?option=com_djcatalog2&view=itemstable&cid=4&Itemid=11&lang=en
PISTON PINS
You can reduce reciprocating weight by using tool steel or titanium pins . . have them tapered on the inside . . if you use titanium ones, have them coated with tungsten . . this should increase rpm by maybe 80 to 100.
http://www.aperaceparts.com/wristpins.html
these ape wrist pins were in the worlds fastest drag bike when it grenaded but they didn't break . . your pins wont need to be quite so strong, so tell them your app so they can make them as light as possible.
http://www.cycledrag.com/ape-wrist-pins-withstand-mcbrides-horrific-explosion
CAM CHAIN ROLLER FOR 160, 175 &200
The roller in the bottom of the photo takes a beating and can be replaced with a metal one for around $100.00
P/N 154
http://www.cappellinimoto.it/index.php?option=com_djcatalog2&view=itemstable&cid=4&Itemid=11&lang=en
crazypj
Split personality, I fake being smart
barnett468 said:
Not strictly true or correct.
The reason they fail is the same reason CB350, CB360 and others fail, oil pressure gets TOO HIGH
This is why 360's are so bad - better oil pump than the plunger pump on earlier models - more pressure at lower rpm
Modifying the oil feed to cams and using a THINNER oil pretty much solves the problem of cam bearing failure.
Wear on start up may be slightly higher over a long period of time (or if bike is left for weeks between start up as oil has more time to drain down) but the cam will not seize into bearing (which only happens at high rpm in my experience)
Also, oil drain over time, have you ever stripped a running engine that has stood for weeks or months (or years in some cases) and found cams completely dry? ( they may look dry but unless there is rust on the cam lobes, they still have some sort of residual coating) Personally I've only seen it on car engines where head gasket had failed or bike's that had been 'abandoned' outside for 10yrs (or more, even 30yrs under cover seems OK)
I had 175 cam bearings seize on me in 1976 but doing a lot of thinking and looking at stuff I came up with similar fix to what I tell everyone to do on 360's. Only problem then was, revving 175 (if I remember right with Honda 90 pistons in 160 block 8) ) the big ends failed at about 80mph in third gear chasing down buddy on a Suzuki GT500 (on out way to Stonehenge, summer 1977
)barnett468
Coast to Coast
crazypj said:
With all due respect, it actually is one factor . . I also did not say the oil was completely gone, and yes, in my 45 years of doing this, out of which around 30 of them were spent doing engine rebuilding on many occasions in conjunction with restoring classic cars full time for a living, I have most certainly disassembled my share of old engines both automotive and motorcycle that had been sitting for long periods.
I have also disassembled many that were not sitting for long periods, including many when I was a project engineer and ran part of the R and D department for Kawasaki Motors because we had to assess and document any damage they incurred after we were done testing them which was typically upwards of around 300 hours, and in fact, although I certainly don't know it all, I do have a fairly good background in oil and the affects of oil pressure or the lack thereof on engines.
crazypj said:
Using thinner oil not only reduces oil pressure, it also gets thru the engine faster after the engine has been sitting and the oil has drained away from some parts, which in fact, means that the time the engine is running with no oil pressure at all is greatly reduced, and in most cases, the greatest amount of engine wear occurs during start up for exactly the reasons I stated, however, this does not mean it is the only reason . . Also, since thinner oil does in fact get to the various engine parts faster, how can you separate this effect from the effect of oil pressure?
Another odd phenomenon is that an engine that runs at 190 degrees will wear less than one that runs at 160 degrees if all other factors remain the same.
In addition, just in case for any reason you happen to think that the quality of oil has nothing to do with engine wear, I suggest you do some research on oils if you are interested in the subject . . You can get a lot of interesting info on it on the site below . . Yes, there is a site dedicated solely to those whom have an oil fettish of which I am one, lol.
http://www.bobistheoilguy.com/
crazypj said:
I didn't mention seizing . . The bearing and cam journals get galled/scored, yet they can continue running or a long time like that, and in my experience, a galled cam and/or bearing would be considered to be trash and replaced, especially if it was on one of our 170 mph factory works Superbikes.
crazypj said:
I would also be extremely interested to see any technical papers or imperical evidence garnered thru controlled testing that describe exactly how high oil pressure can damage a bearing, because I personally have never seen such a thing, and all my experience and what I have read contradicts that statement, and it also seems to me that even simple logic would lead one to believe that that is incorrect.
I also worked at Norm Reeves Honda, Kawasaki, Bultaco, and Triumph around late 74 thru part of 77 and there were no service bulletins mentioning excessive wear caused by high oil pressure.
This is a blower drag race engine with well over 1000 hp and around 60 runs on it which include revving the bejesus out of it in the burn out box . . It has 120 psi of oil pressure and you could easily reuse those bearings in a street car for another 100,000 miles.
Here's a brief article on how oil pressure affects bearings.
http://www.machinerylubrication.com/Read/1772/oil-pressure-bearing-wear
The following is an excerpt from the article.
"The following are a few of the common causes of bearing faults:
1. overloading or shock loading
2. contamination of lubricating oil
3. overheating of lubricating oil
4. overheating of the bearing
5. misalignment or incorrect assembly of the bearing
6. insufficient pressure and/or volume of the lubricating oil
The secret to long bearing life, after installation and operational problems have been corrected, is to ensure that the bearing is supplied with the correct grade of oil in sufficient quantities, and that the oil is clean and runs at the correct temperature.
Plain bearings, in particular, are sensitive to oil volume and pressure. Insufficient pressure will normally result in insufficient oil volume being delivered to the bearing. The decreased oil volume causes the bearing to wear out faster, due to increased operating temperatures and contact between the journal and bearing."
Here's some more info you might find interesting . . The authors credentials/qualifications and experience are summarized in the link below.
http://www.epi-eng.com/epi_general_information/ceo_resume.htm
Here's a brief excerpt.
Jack Kane founded EPI, Inc. in 1994, for the purpose of developing and producing high-performance vehicle propulsion systems. He is also the CEO and chief engineer of the company.
Jack was a dean's list student in the Electrical Engineering program at the US Air Force Academy. Later, he earned Bachelor of Mechanical Engineering (cum laude) and Master of Science degrees (summa cum laude) from civilian universities.
Jack has been actively involved in piston engine development since the early ‘50’s, and has configured, built and modified successful engines for a wide variety of specialized applications and winning race cars.
Since the founding of EPI, Inc., he has been responsible for the design and development of liquid-cooled aircraft engines, gearboxes, and accessory drive units for various fixed-wing and helicopter applications. His technical articles published in RACE ENGINE TECHNOLOGY magazine have received acclaim from knowledgeable experts at the highest levels of motorsport.
Beyond engineering, he is an accomplished machinist, a commercial pilot, and has done development, modification and overhaul work on certified (Lycoming, Continental, Orenda) aircraft engines. In his younger days, he was a winning driver in a variety of automobile racing categories including rear-engine formula cars, sports cars, midgets and stock cars, and has won several championships.
Thursday, December 24th, 2015
THIS IS AN EXPANDED VERSION OF AN ARTICLE BY Jack Kane WHICH APPEARED IN ISSUE 030 of RACE ENGINE TECHNOLOGY MAGAZINE
.................................... - Hydrodynamic Bearings - ............................................
........ Insight into how those seemingly-magic rod and main bearings work .............
............................................ INTRODUCTION ..................................................
Most bearings can be described as belonging to one of four classes: (1) rolling element bearings (examples: ball, cylindrical roller, spherical roller, tapered roller, and needle), (2) dry bearings (examples: plastic bushings, coated metal bushings), (3) semi-lubricated (example: oil-impregnated bronze bushings) and (4) fluid film bearings (example: crankshaft bearings).
Aside from an occasional tangent like the Porsche 1.5 litre flat four engine of the sixties and certain radial-configuration aircraft engines, almost all piston engines use fluid film bearings. This is true for the crankshaft and sometimes the camshaft, although often the latter runs directly in the engine structure. He we put the spotlight onto fluid film bearings.
The point of the whole discussion is (a) to explain how fluid film bearings work (which is sometimes counterintuitive) and (b) to demonstrate how engine designers are reducing friction losses through bearing technology.
Fluid film bearings operate by generating, as a by-product of the relative motion between the shaft and the bearing, a very thin film of lubricant at a sufficiently-high pressure to match the applied load, as long as that load is within the bearing capacity.
Fluid film bearings represent a form of scientific magic, by virtue of providing very large load carrying capabilities in a compact, lightweight implementation, and unlike the other classes, in most cases can be designed for infinite life.
Fluid film bearings operate in one of three modes: (a) fully-hydrodynamic, (b) boundary, and (c) mixed.
HYDRODYNAMIC MODE
In fully hydrodynamic (or "full-film") lubrication, the moving surface of the journal is completely separated from the bearing surface by a very thin film of lubricant (as little as 0.0001" with isotropic-superfinished {ISF} surfaces). The applied load causes the centeline of the journal to be displaced from the centeline of the bearing. This eccentricity creates a circular "wedge" in the clearance space, as shown in Figure 1.
Hydrodynamic Wedge
Figure 1
The lubricant, by virtue of its viscosity, clings to the surface of the rotating journal, and is drawn into the wedge, creating a very high pressure (sometimes in excess of 6,000 psi), which acts to separate the journal from the bearing to support the applied load.
The bearing eccentricity is expressed as the centreline displacement divided by the radial clearance. For example, if a bearing which has 0.0012" radial clearance (0.0024" diametral) is operating with a film thickness of 0.0001", then the eccentricity is (.0012 - .0001)/.0012 = 0.917.
The bearing eccentricity increases with applied load and decreases with greater journal speed and viscosity.
Note that the hydrodynamic pressure has no relationship at all to the engine oil pressure, except that if there is insufficient engine oil pressure to deliver the required copious volume of oil into the bearing, the hydrodynamic pressure mechanism will fail and the bearing(s) and journal(s) will be quickly destroyed.
It is interesting to study the pressure distribution in the hydrodynamic region of a fluid film bearing. The hydrodynamic pressure described above increases from quite low in the large clearance zone to its maximum at the point of minimum film thickness as oil (essentially incompressible) is pulled into the converging "wedge" zone of the bearing. Figure 2 shows a representative sketch of the radial pressure distribution in the load-supporting area of the bearing.
Hydrodynamic Pressure Profile - Radial
Figure 2
However, this radial profile does not exist homogeneously across the axial length of the bearing. Figure 3 shows a sketch of the axial pressure distribution profile for fully-developed hydrodynamic lubrication with a non-grooved bearing surface (insert). As the picture shows, the pressure drops off rapidly at the edge of the bearing, because oil is leaking out of the edges under the influence of the high hydrodynamic pressure. Moving inward from the edges, the pressure rises dramatically. If the bearing has sufficient width, the profile will have a nearly flat shape across the high-pressure region.
Axial Pressure Distribution - Non-Grooved Bearing
Figure 3
Long ago, it was standard practice to use fully-grooved main bearings, the thought being that the groove would provide a better supply of oil to the conrod bearings. A quick study of the axial profile of the hydrodynamic pressure distribution for a grooved surface (insert), shown in Figure 4, demonstrates how any interruption of the smooth surface of the bearing in the load-carrying region will severely degrade the capacity of the bearing.
Axial Pressure Distribution - Grooved Bearing
Figure 4
BOUNDARY MODE
The second mode of bearing operation is boundary lubrication. In boundary lubrication, the "peaks" of the sliding surfaces (journal and bearing) are touching each other, but there is also an extremely thin film of oil only a few molecules thick which is located in the surface "valleys". That thin film tends to reduce the friction from what it would be if the surfaces were completely dry.
MIXED MODE
The third mode, mixed, is a region of transition between boundary and full-film lubrication. The surface peaks on the journal and bearing surfaces partially penetrate the fluid film and some surface contact occurs, but hydrodynamic pressure is beginning to increase.
OPERATION
To further explain the three lubrication modes, let's examine the operation of a journal bearing from startup to steady state. Figure 5 shows a journal bearing at rest. The applied load causes the journal to contact the bearing surface (eccentricity ratio = 1.0).
Journal Bearing at Rest
Figure 5
When motion begins, the journal tries to climb the wall of the bearing, as illustrated in Figure 6, due to the metal-to-metal friction (boundary lubrication) between the two surfaces.
Journal Bearing - Initial Motion
Figure 6
If there is an adequate supply of lubricant, the motion of the journal begins to drag the lubricant into the wedge area and hydrodynamic lubrication begins to occur along with the boundary lubrication (mixed lubrication).
Assuming the load and viscosity remain relatively constant during this startup period, then as RPM increases, the hydrodynamic operation strengthens until it is fully-developed and it moves the journal into its steady-state orientation (Figure 7), defined by the eccentricity (e) and the orientation angle (a). Note that the direction of the eccentricity, and hence the minimum film thickness, do not occur in line with the load vector, but are angularly displaced from the load.
Journal Bearing - Fully Hydrodynamic
Figure 7
There are three parameters which determine the mode (boundary, mixed, hydrodynamic) in which a given bearing will operate: (1) the speed of the shaft, (2) the viscosity of the lubricant, and (3) the applied unit load.
These three parameters can be combined in the following way to form a value we can call "Bearing Operating Condition" (BOC).
BOC = Viscosity x RPM x Diameter x K / Unit Load
(Equation 1)
The Viscosity parameter is in units of absolute viscosity. The "K" value is a factor which converts RPM and Diameter into journal surface speed. The bearing unit load is the applied force divided by the projected area of the bearing (the insert width times the journal diameter).
ZN/P CURVE ("STRIBECK PLOT")
The BOC value will predict the operating mode of a bearing and the expected friction coefficient for that operating condition. The transitions between these different operating modes, and the related friction properties are illustrated more fully in the Stribeck Plot shown in below in Figure 8. This plot (also known as a "ZN/P Curve") shows the bearing coefficient of friction (on a logarithmic scale) plotted as a function Bearing Operating Condition (BOC). The values plotted on the X-axis are nondimensionalized, and are shown as a percentage of full scale.
ZN/P Graph
Figure 8
The two vertical lines in the plot area show the boundaries between the three operating modes. Area 1, from BOC = zero to about 15, is where boundary lubrication occurs. Area 2 (BOC = 15 to 35) is the region of mixed lubrication, in which, as BOC increases, the hydrodynamic pressure is developing and taking over from boundary lubrication. Area 3 is fully-developed hydrodynamic lubrication.
Note that the purpose of presenting this BOC (or ZN/P) curve is to demonstrate the interrelationship between friction coefficient and the BOC (ZN/P) parameters, not to instruct in bearing design.
In the definitive 2001 reference text "Applied Tribology: Bearing Design and Lubrication" by Dr. Michael Khonsari and Dr. Richard Booser (ref-2:6:12), the Stribeck Plot is shown on Page 12 and is described as a "dimensionless uN/p curve relating lubrication regime and friction coefficient to absolute viscosity". That same DIMENSIONLESS curve ("ZN/P") is shown on page 2097 of "Machinery's Handbook, 24th ed." (ref-2:22:2097)
The "BOC" entity (often known as ZN/P) does indeed have units, which depend completely on the units you choose for (a) surface speed converted to RPM and (b) unit load: psi, n/mm², mpa, etc. Various engineering texts use specific portions of the curve and use whatever ZN/P units they prefer. Others retain the dimensionless construct.
The friction coefficient values shown in Figure 8 were taken from both "Machinery's Handbook, 24th ed." and from "Design of Machine Elements", by M.F. Spotts, Professor of Mechanical Engineering, Northwest University (ref-2:2:302). Both reference works agreed that the low-point is about 0.001, the fluid film range is from 0.001 to at least 0.005, the boundary region is greater than 0.1 to as high as 0.03, and the mixed region is between the other two, as shown on the plot.
This curve illustrates that when operating in the hydrodynamic region (Area 3), if the unit load remains constant and either rpm or viscosity increase, the hydrodynamic pressure increases, the eccentricity decreases and the friction coefficient rises, increasing by a factor of 10 as eccentricity approaches zero.
However, if rpm remains fixed and either viscosity decreases or unit load increases, then the BOC will decrease. Friction coefficient decreases logarithmically down to the low point at around BOC = 35. If the unit load continues to increase and/or viscosity continues to decrease, the BOC will move into the mixed lubrication region and the lubrication mode will change from fully hydrodynamic back to the mixed mode and friction will increase dramatically. If the load increases and/or viscosity decreases even further, BOC continues to decrease, and eventually the journal asperities break through the film and the system reverts back to the very-high-friction boundary lubrication mode.
Note the values for friction coefficient. In the area of boundary lubrication, the friction coefficient is similar to that of a dry bearing (0.25-0.35). At the BOC value of 35, the friction coefficient is in the remarkably low region of 0.001, which is 50% less than the friction coefficient of deep-groove ball bearings. As the BOC increases (any combination of smaller load, higher rpm, higher viscosity) the curve shows that friction coefficient rises exponentially, approaching a value of 0.01, ten times greater than the ideal minimum. That fact illustrates why there is so much attention paid to optimizing the bearings for the application, trying to maintain the BOC in the 35-50 range.
In issues past, we have seen that combustion loads can apply forces in excess of 12,000 pounds to a rod journal. If the bearing were operating at a friction coefficient of 0.002, (BOC roughly 50), an applied load of 12,000 pounds would generate a friction load on the surface of one bearing of 24 pounds.
If the diameter of the journal carrying the 12,000 lb. is 2.50", then the friction torque lost to that bearing will be 24 lbs x 1.25" = 30 lb-in or 2.5 lb-ft. If all 5 main journals carry the same load, then the friction torque lost to the main bearings alone is 5 x 2.5 = 12.5 lb-ft, which at 9000 rpm, absorbs 21.4 HP.
If that journal diameter were reduced to 2.00", one might think that a 20% reduction in main bearing friction torque could be realized. However, for the same bearing width, reducing the journal diameter 20% reduces the projected area by 20%, which increases the unit loading, resulting in a reduced BOC for the same load, rpm and viscosity. Further, reducing the journal diameter by 20% also reduces the surface speed by 20%, which for the same RPM and viscosity, lowers BOC even further. Add in the effect of the very low viscosity lubricants some teams use, and the net effect can be a dramatic reduction of BOC. As long as the BOC stays within the hydrodynamic region, the smaller BOC will yield an even lower friction coefficient, which further reduces the bearing friction losses.
Of course in practice, it's not that big a payoff, because the 12,000 lb. load is not applied for the whole 360° of rotation. But the illustration serves to point out an area that savvy engine designers have been successfully harvesting.
SQUEEZE-FILM LUBRICATION
There is another form of fluid-film lubrication, which adds to the load capacity in applications with oscillating loads (such as a piston engine), known as squeeze-film lubrication. Squeeze-film action is based on the fact that a given amount of time is required to squeeze the lubricant out of a bearing axially, thereby adding to the hydrodynamic pressure, and therefore to the load capacity. Since there is little or no significant rotating action in the wrist-pin bores, squeeze-film hydrodynamic lubrication is the prevailing mechanism which separates wrist pins from their bores in the rods and pistons.
GEOMETRY
Crankshaft bearings are not round. The main bearing journals and crank pins that run within these (conventionally) plain bearings are perfectly round but the bearing surfaces that surround them are not. For a start, the crush that locates a plain bearing in its housing causes distortion of the housing, the nature of which will reflect the material and geometry of the part forming it. On top of this, these bearings are actually designed to be out of round.
If engine load and speed were constant and bearing geometry could always be maintained during operation a perfectly round bearing surface profile would work fine. Of course in the internal combustion engine load and speed do vary constantly and the varying loading imparted to the bearing housing constantly alters its geometry. In fact, the racing engine is an elastic device, to an extent that is not always fully appreciated. Enormous loads go both up and down the con rod, lengthening and shortening it and distorting the shape of its big and small end. In view of this contemporary steel backed plain bearings are designed to be semi-flexible rather than rigid structures.
In The Definitive V8 Engines, we showed that a naturally aspirated 2.4 litre, 750 bhp Formula One V8 running to 20,000 rpm (2006 regulations) is subject to a maximum crankpin load in the region of 13,300 lb while a naturally aspirated 5.86 litre, 850 bhp Cup V8 running to 9500 rpm is subject to about 12,500 lb. Such crankpin loads deform the crankshaft, which in turn transfers deformation to the crankcase through its main bearing journals. Thus in operation, both the rod bearing housing (conrod big end) and the main bearing housings deform.
In practice it has been established that the appropriate static profile for a crankshaft bearing is normally oval, having its minimum diameter in line with the direction of maximum load. Generally this is taken to be at 90 degrees to the parting line. Bearings are therefore typically manufactured with a wall thickness that is greatest at 90 degrees to the parting line, tapering off from that point to the parting line each side by a specified amount. This is known as bearing ovality (sometimes called "eccentricity", but that usage can be confused with the eccentricity essential to hydrodynamic lubrication) and it is tailored to the characteristics of a specific engine. For example, a heavy piston assembly and high rate of piston acceleration will result in high inertia loading at the top of the exhaust stroke that will cause pronounced stretch of the con rod, this in turn significantly squeezing the big end - a high degree of ovality is required to stop the bearing then pinching the crank pin.
BEARING STRESS
While bearings are a source of friction (including consequent shearing of the oil film) and thus heat, they are also a route for heat to escape from the reciprocating/rotating assembly to the stationary structure of the engine and, more importantly, into the circulating oil. In terms of the stress that the bearings see, it is notable that the magnitude, and sometimes even the direction of the loading varies throughout the course of each stroke. How much stress a given bearing experiences is a function of net loading and bearing projected area, which fluctuates accordingly.
Net loading varies dramatically with throttle and rpm, and throughout an engine cycle at any given throttle and rpm. For example, on the power stroke the compression / combustion loading on the con rod is compressive and this opposes the tensile inertia loading caused by piston acceleration. At low engine speed with wide open throttle there is less inertia loading balancing the piston combustion forces and, depending on the engine's torque characteristics this can impart greater net loading to the bearings than WOT operation at higher speeds. Conversely, at engine speeds above peak torque inertia forces come to dominate and the net effect on the bearings is increased loading compared to operation at peak torque rpm. However, the con rod loading that occurs in the vicinity of TDC overlap is extremely high tensile loading because there is very little cylinder pressure to oppose the piston acceleration. That load varies with the square of rpm, and can apply immense loads (and consequent deflections) to the cap-half insert.
Sustained high rpm operation is another threat to the bearings since it causes high temperature running, which in turn can cause excessive oil heating and with that a loss of viscosity. In this respect Cup oval running can be more taxing to the bearings than Formula One road racing.
An article in Race Engine Technology, Issue 20, showed an example of cavitation damage on a big end bearing from the Cosworth 2.4 litre V8 engine of 2006, which was designed to run to 20,000 rpm. As the piston approached top dead center the tendency was for the upper portion of the titanium rod's big end to arch away from the steel crankshaft journal and for the steel backed bearing to distort accordingly. There was thus a cavity formed between the bearing and the journal upon which it ran, creating a low-pressure zone in the oil film, encouraging the formation of vapour bubbles. As the piston reversed direction the combustion pressure took out the cavity, collapsing the bubbles, which added to the loading on the big end. In fact, shock waves were formed that stressed the surface of the bearing, to the extent that material could even be lost from it. Following a problem of this nature at the Malaysian Grand Prix the oil viscosity was increased. This avoided any cavitation damage until design changes could be implemented to address the problem. High shear viscosity at high temperature is critical for bearing duty as this extreme example attests. Oil development through 2006 led to a reduction in the variation of viscosity with temperature ("viscosity index").
MATERIALS AND COATINGS
Ideally a bearing material should offer low friction properties, but given that in fully-hydrodynamic operation, the bearing surface is separated from the surface of the journal by a thin film of oil, it is clearly the lubricant rather than the respective surface materials that dominates the friction generated under normal running conditions.
Therefore, if there is an adequate supply of lubrication and a suitable load / speed ratio, the material forming the bearing's working surface is not crucial in terms of frictional losses. Inevitably, however, metal-to-metal contact will occur, particularly on start up. The journal is invariably steel, and copper, for example (used as the sole material for some early bearings) running on steel has a kinetic coefficient of 0.36. However, any metal running on steel given proper lubrication has a kinetic coefficient of in the region of 0.06 (it will vary as shown in the Stribeck curve above).
In view of the unavoidable metal-to-metal contact, low friction coatings are sometimes applied to bearings. For example, one manufacturer has developed an ultra-slippery moly/graphite blend, which is suspended in an inert PTFE substrate, which provides the adhesion necessary to attach it to the top surface of the bearing. This coating, only one thou thick, which is compatible with contemporary lubricants and lubricant additives, is sacrificial - the bearing will outlive it but in the meantime it is claimed to reduce friction and wear. If there is any contact it will prevent scuffing and even absorb debris.
PLAIN BEARING CONSTRUCTION FUNDAMENTALS
Typically the tri-metal plain bearing common to contemporary high-performance engines is formed as a laminated structure having a relatively thick steel backing layer in contact with the housing, a harder, thin middle layer (copper-lead, lead-bronze, aluminum-tin, etc.) and a very thin upper layer of soft material (lead, zinc, cadmium, lead-indium, and a host of others), the top layer forming the actual bearing surface. The maximum applied pressure a bearing can carry is determined by the strength and hardness properties of the upper surface. The maximum relative velocity between the journal and the bearing is governed by the bearing's ability to dissipate the heat generated by the shearing of the oil film.
Except for the rare instances of built-up crankshafts, the plain bearing is split into upper and lower halves, so that it can be fitted over the journal. One half fits into the main structure, the other into the cap. Each half is known as a shell hence this type of bearing can be referred to as plain or shell-type. Normally only one of the main bearings is designed as the thrust bearing necessary to minimise axial movement of the crankshaft.
The multiple layers have been developed to provide the properties required for the specific application. While the backing will invariably be steel, a steel bearing running against a steel journal with no coating on either surface would cause high friction and wear in the boundary and mixed lubrication modes, and would provide little or no ability to allow foreign particles to embed in the material, but would instead capture them and turn them into cutting tools. Therefore, the upper layer is a softer metal, designed for minimum friction with sufficient embedability. The idea is to allow abrasive particles to embed below the working surface and thereby minimise wear. Moreover, the softer upper layers will help the bearing act as a cushion in the face of severe operating forces. In addition to high mechanical strength and high resistance to temperature the composite bearing needs good conformability and good surface properties - it needs 'compatibility' to prevent pick up or even seizure if the oil film momentarily breaks down.
Due to the mechanical properties of the soft bearing material, one might think it would be squeezed out of the bearing due to the forces acting upon it. However, the very thin nature of the soft layer, supported by the much stronger and thicker base layer, prevents the extrusion of the soft material.
The inability of the applied load to squeeze out the soft layer is known as the principle of plastic constraint. Consider a thick layer of clay sandwiched between two plates of steel. If pressure is applied to the steel plates, the clay will deform and squeeze out the edges of the sandwich. But as the thickness of the clay gets ever smaller, it takes an ever-increasing amount of force to squeeze out more clay. Eventually, a thin layer of clay remains that cannot be extruded out without the application of an infinite amount of pressure.
A bearing needs to conform to the shape of its housing; a shape that is constantly in a state of flux since the engine is an elastic device. In view of this the bearing is designed so that when the two halves of the housing are correctly bolted together its parting line surfaces adjoin and the bearing correctly conforms to the housing, leaving the required running clearance between its working surface and the journal. However, when a bearing shell is fitted into its respective housing its edges will stand slightly proud of the housing faces so that when the cap bolts bring the parting line surfaces together there will be a slight gap between the housing faces. When further tightening brings the faces into contact the gap will have gone and the resultant 'crush' means that the bearing is compressed like a spring and applies a radial load to its housing.
Although a plain bearing is thus an interference fit in its housing locating lugs can be fitted to assist positioning during assembly. More typically each bearing shell is retained by a pin projecting into it from the housing. These lugs or pins will help avoid any danger of movement relative to the housing in operation but that is not their primary purpose and the interference fit must be good enough in this respect to ensure reliable operation.
In the case of the big end the interface between the plain bearing and its respective journal normally receives a supply of pressurized lubricant from a drilling in the journal. The relative movement of journal and bearing and the forces involved cause the oil to spread out and form the necessary film throughout the radial interface, before spilling into the crankcase.
INFLUENCES
Crankshaft main journals are subject to extremes of torsional vibration, and that influences their diameter. However, journal overlap and crankshaft balancing techniques are further factors, which may permit the use of smaller diameter and narrower journals. It is notable that the Cosworth DFV 3.0 litre V8 of 1967 had a main bearing journal diameter of 60 mm with a big end journal diameter of 49 mm. By contrast, a third of a century later a 3.0 litre V10 typically had a main journal diameter in the range 40-45 mm, a big end journal in the range 35-40 mm. However, there is also a very large difference between the operating speeds of those two engines. Since main journal diameter is a major factor in crankshaft torsional stiffness, perhaps the reduction in crankshaft torsional stiffness caused both by the reduced diameter and the increased length served to provide a greater separation between the crankshaft torsional resonance point and the much higher excitation frequency of today's engines.
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