Relative oil distribution orifices

I have a computational problem that I would like to pose and solicit some help with.

I have a rocker shaft 11.8 inches in length, that serves eight rocker arms. The oil feed hole is 0.080mm in diameter and is located 4.5 inches from one end of the shaft. Assume that the eight holes are equidistantly spaced along the shaft, and the gallery diameter along the length of the shaft is 0.200 inch. Assume that the clearance between the rocker arms and the shaft is consistent at 0.0015 inch. What diameter would each of the holes feeding the 8 rocker arms have to be, in order to provide an equal amount of oil flow to each rocker arm?

Note: The stock rocker arm has eight rocker feed holes, approx. 0.060 inch in diameter. From the wear patterns on the shaft it is obvious that the damage is progressively worse, the further you go from the feed hole, indicating a differential in oil availability.

Any help will be greatly appreciated.

On push rod engines normal practice of UK and USA manufacturers was to meter the oil flowing up to up through the oilway to the rocker gear by passing it through a drilling through one of the camshaft journals. By doing this the rocker gear gets pulses of oil each time the camshaft rotated rather continuous full full pressure flow.

I don't the inner workings of your engine but if you examine a camshaft you can see if this is the case and work out a way to increase the oil flow to the top end.

How I'd approach it is that the flow through an orifice generates a pressure drop PO, and the flow through the bearing generates a differnet pressure drop PB, and the bearing exhausts to atmospheric pressure (ish).

There's also an internal pressure drop in the tube that joins all the orifices, call it PT, that may be very small in the scheme of things

First thing is to find the change of each P with volumetric flow rate (this is the hard bit)

One I had that I'd start from atmospheric at the far end of the camshsaft and add up all the Ps and work out the flow rate and pressure at each station.

Then mess about with the relative orifice sizes until the overall deltap is the same and the flow rate is the same but equalised for each orifice.

This will be wrong in detail if britishtrident is right about your engine but will still guide you in the right direction.

Incidentally if your dimensions are input dia =.080" and you have 8 output holes of .060" then no wonder you have a problem. The first two rockers can take the entire oil supply (ignoring the backpressure from the bearings)

Trident,

The Fiat block is similarly configured. The upper end oiling originates from the center cam journal, which is in turn fed by the center main journal. The make up of this is such that the center cam bearing has a circular groove behind the bearing in the block journal. This makes the oil supply to the upper end a constant flow, as opposed to a pulsed flow.

The oil travels to the deck of the block, where it is diverted in a groove in the head to one of the head studs. The oil then travels through the head around one of the head studs (specially relieved for this purpose and then through a diagonal orifice ending up at #2 rocker arm stand. This is where the entry orifice is for the rocker shaft.

Through these various passages there is a pressure drop of approx 40%. So, at speed with say 70 lbs of oil pressure there would be 40 lbs of oil pressure presented to the 2mm rocker shaft entry orifice. I will have to try and find the formula for computing the pressure drop through an orifice, to estimate what the actual pressure to the rocker shaft gallery is.

This is what piqued my interest about the resultant flow distribution. If for instance the flow through a 2mm orifice were to be around 1 litre per minute @40 PSI, then in order to provide equal distribution of the oil to the 8 rockers, the cumulative leakage flow of the eight rockers must be less than 1 litre per minute, or there will be no pressure buildup. It would appear that there are two likely consequences. 1) It is unlikely that there will ever be any pressure at all at the rocker/shaft boundary interface. 2) The amount of oil presented at each rocker/shaft boundary interface will be inconsistent, with the ones closer to the entry orifice getting proportionally more oil, and the further outlying orifices perhaps getting much less oil flow.

I am I visualizing this correctly?

Paul

Greg,

Thanks for your posting. I think from my answer to Trident that I have the concept pretty well in hand. Now comes the interesting part of computing the various pressure drops and the resultant impact on flow rates. It should be said, that as long as there is some pressure there, perhaps even as low as 10-15 PSI, as long as the in-flow rate is more than the out-flow rate of the 8 rocker arms combined, then the rocker arm/shaft interface will be something in between a hydrodynamic and boundary interface.

Yes, I agree that the current orifice sizes are not conducive to a successful lubrication scheme when used at much higher performance levels, than the factory originally intended. By my calculation, the "oil flow" clearance for a rocker with a .591 inch bore running on a .589 inch shaft (.001 inch clearance per side) would be .0038 sq. inch (this accounts for the fact that oil can escape on both sides of the rocker). Thus for 8 rockers the total out-flow surface area to atmospheric would be just under 0.029 sq. inch.

In order for a single feed orifice to equal this total "bleed off area", in terms of surface area, the diameter of the orifice would have to be .200 inch ( area of 0.0031 sq. inch). At this diameter the entry orifice should have the ability to flow slightly more oil than the total of the 8 rockers arms can discharge to atmosphere.

Does this make any sense?

Thanks
Paul

Paul you are living in a static world, put the holes at the bottom of the shaft and they get blocked off each time the cam lobe comes around and in the meantime the unloaded rockers are getting prepped for their turn.

As you decided to use bronze bushes lube isn't such a big worry and remember the less oil to the top means less oil that has to find its way to the sump losing you HP.

And yes your calculations make sense, but work out how many holes are actualy able to supply at any given time.

Cheapracer,

Having thought about what you said, I am persuaded that the oil flow mechanism within the rocker shaft is indeed quite dynamic in nature.

When a rocker is under tension, it is up against the bottom of the shaft and, there will be no oil flowing into the rocker/shaft boundary interface pressure point (presuming that the feed hole is on the bottom of the shaft). Considering that each rockers is under tension approx. 250 degrees out of every 720 degrees engine rotation, it can be assumed that for the remaining 470 degrees of engine rotation oil will flow though the rocker/shaft boundary interface clearance (approx 0.002 inch). This flow would re-establishing a lubrication supply to the rocker/shaft boundary interface in preparation for the next 720 degree engine rotation cycle for that particular rocker arm, and, most importantly, carrying away heat generated during the previous cycle at this boundary interface.

According to my observations there are 4 rocker arms in tension (to some extent anyway), at the rocker/shaft boundary interface, at any given point in time. During this "tension period" all the clearance is at the top of the shaft, with virtually nil clearance at the boundary layer pressure point (also the oil feed hole) on the bottom of the shaft. As 4 of the eight rocker feed holes are occluded, either partially or completely, any oil flow would be diverted to the remaining four rockers not under tension, with equidistant circumferential clearance.

Based on my earlier computations of cross-sectional flow area, each rocker's 0.070 inch feed hole will flow about 10% greater volume than the spill volume of the 0.002 clearance between the rocker and shaft. The spill volume computes to an area of .004 sq. inch per rocker arm. as such,the aggregate spill volume for 4 rocker arms would be 0.016 sq. inch. The single oil supply hole to the shaft must be 0.145-0.150 inch in diameter ,to supply sufficient flow, to service any four non-occluded rocker arms at any given point in time.

Conclusion: The previously discussed changes in metallurgy for the shaft and rocker arm incorporating a hard-chrome plated, 4130 shaft with a bronze bushinged rocker arm, will be adopted. The rocker shaft oil delivery orifice for each rocker arm, currently sized at 0.070 inch is of sufficient size to provide adequate oil supply for a rocker arm, running 0.002 clearance to the shaft (input orifice cross-section exceeds the spill orifice cross-section). In order to maximize the delivery of oil to the rocker arm/shaft boundary layer interface pressure point, moving the oil delivery hole to the bottom of the shaft (6 o'clock position vs 3 o'clock) would appear prudent. In addition, scoring the bottom of the rocker shaft, in the vicinity of the oil delivery hole, with an "X" groove pattern will provide better distribution of available lubricant. Finally, in order to provide sufficient input flow to the rocker shaft (oil coming from the engine oil pump), to provide sufficient oil to any 4 rocker arms, not in tension at any given point during an engine rotation, will require a hole with a diameter of 0.145-0.150 inch. The input orifice will therefore be slightly larger than the aggregate spill flow orifices of the 4 rocker arms not in tension at any given point in time.

I would welcome any comments if there are any holes in the argument that I present.

Originally posted by Paul Vanderheijden

I would welcome any comments if there are any holes in the argument that I present.

Ok, I see the pun

My opinion you don't need to "X" across the hole, just decreases surface area.

Thanks everyone for your ideas and contributions. Once I get the parts made and installed on a mule motor on the dyno I will let you know the outcome.

Regards,
Paul

Paul

Metinks that your problem is less a computational problem and more a practical problem. Let me explain.

One can come close to solving it using computational methods.

I know I can write a programm that would come close. It would only take the better part of a day. Problem is that computational methods only will get you part of the way without rigorous back testing lest there be some unforseen problem or measurement not properly seen.

Or I guess one could look for somebody elses validated programme. Let's see that might only take a day or two of searching and adopting.

Or you could approach the problem from the practical way, the way I would.

I would set the head up on a bench with a drill operated pump supplying oil to the source orifice. I would then observe and measure the flow out of each valve's system. No doubt you will find that the fifth hole from the inlet will get the least as it is the furthest away. The fourth hole on the same side will get the next least. The two valves, fourth on the one side and first on the other side of the inlet will get more and so on.

The problem is that the holes closest to the inlet are scavinging more than their share, what do they say about piglets sucking hind teat? They are the smallest.

I would then make up small wooden or plastic plugs to drive into all but the furthest valve oil holes. Drill small holes into each. You will probably get too much oil to the furthest valve. Keep drilling larger holes until the flows even out. If you make a hole too large replace that plug and re-drill.

Once you understand what size holes you need where then decide whether you can peen some of the original holes smaller or need to set in small brass plugs correctly drilled.

I know I can solve the problem in about half a day's bench work. You can to!

Too much science spoiled the soup. Grandma made soup with her eyes closed, even before computers.

Regards and good luck.

Joe,

Thanks for your practical insight. I did in fact do something very similar to what you suggested. Rather than using the tiny orifices, I constructed a rig from PVC tubing with larger orifices that I could more easily observe and simply used a water hose to provide the pressure/flow. From this I came to understand the "hind teat" syndrome better.

Fiat originally constructed the rocker arm assembly with the rocker feed orifices at 3 o'clock. As such, even with the rocker under tension the oil could (would) flow and the hind teat syndrome would definitely apply if outflow were greater than inflow. In my redesign I moved the feed orifices to the 6 o'clock position. Now of course, when the rocker is under tension the feed hole is partially/completely blocked, unless the pressure at the port is sufficient to overcome the pressure of the rocker against the shaft. (One of the other areas that I am working on is running the engine at lower oil pressures [but still with sufficient flows] to reduce parasitic losses due to pumping.)

On my test model I replicated the various combinations of open/closed feed holes. As only half the number are actually now capable of flowing oil, these all had relatively equal flow so long as the input flow was sufficient to service the cumulative requirement of the four output flows. If the input flow is too low, then one or more of the outputs would suffer.

From my bench model (actually driveway model) I determined that the input hole, feeding the rocker shaft should be enlarged. Once having determined what worked on the driveway model, I scaled the results back to my actual rocker shaft and adjusted the shaft input orifice accordingly.

I am making the parts at the moment, and will have one of my own motors available to test the new parts later this month. Unfortunately getting customer stuff out has to still continue to pay the bills.

Regards,
Paul

"Horsepower has this tendency to break things. If you're not breaking anything you're not going fast enough."

Originally posted by Paul Vanderheijden


Now of course, when the rocker is under tension the feed hole is partially/completely blocked, unless the pressure at the port is sufficient to overcome the pressure of the rocker against the shaft. later this month.

Unfortunately getting customer stuff out has to still continue to pay the bills.

."

Not a hope in hell with 5 other holes exposed.

Don't you hate that

Paul

If your solution works we will all be happy for you.

If not:

1. Cheap's insight might very well be accurate.
and/or
2. Having your oil source at the point of greatest pressure flys in the face of all lubrication engineering theory. I will be happy to discuss the theory if your solution fails.

Regards

Paul, just wondering, what kind of tooling one would use to make that X score inside the 0.60" bore of your bushing? I looked at some old rockers, I see some do have single grooves inside the bush, they appear to be saw cuts (cut through bush wall from the outside of the bush). I didn't press one out to look.

Originally posted by Joe Bosworth
Having your oil source at the point of greatest pressure flys in the face of all lubrication engineering theory.

Agree. I would be wary of moving the feed point from the 3 o'clock specified by the OEM. If you do this it will -
a) Reduce the bearing surface area at the critical load-carrying location.
b) Provide a convenient escape route for oil to be squeezed back into the gallery when the valve is being opened.

Seems to me the original design relied on bush to shaft clearance as the restriction to distribute oil evenly between rockers. Once they have worn beyond a certain point, the distribution goes out the window. If you restore the clearances to original specs you shouldn't have any problems. Don't use orifices significantly smaller than the existing one. These may be prone to blockage.

I just thought I would provide an update for those who expressed an interest in the problem.

I did end making a new shaft with revised orifice sizes and moving the oil supply hole to the bottom of the shaft, as proposed. I had a X groove ground in the shaft, not in the bushing as supposed by Scooperman.

The shaft was made from 4340 thick wall, boiler steam tubing (It was the only stuff that I could get that had a 0.250 wall thickness). It was surplus material from a project at Boeing Aircraft corporation and was specially made for them. It was then centerless ground to 0.005 undersize, and drilled and notched. It was then hard-chrome plated to 0.004 oversize, and then centerless ground back to standard factory dimension, leaving a 0.005 thick hard-dhrome layer. This allows the shaft to be used with standard rockers, standard rockers with bronze bushes, or aluminium roller rockers with bronze bushings, all of which we can supply.





This design has now been implemented successfully in several applications, including a car that successfully completed the Historic Monte Carlo Rally in January 2009.

Originally posted by Paul Vanderheijden
I just thought I would provide an update for those who expressed an interest in the problem.

This design has now been implemented successfully in several applications, including a car that successfully completed the Historic Monte Carlo Rally in January 2009.

Thanks for that Paul, keep us up to date, interesting stuff for some of us old schoolers

Originally posted by gruntguru
Agree. I would be wary of moving the feed point from the 3 o'clock specified by the OEM. If you do this it will -
a) Reduce the bearing surface area at the critical load-carrying location.
b) Provide a convenient escape route for oil to be squeezed back into the gallery when the valve is being opened.
.

a) The critical peak pressure point isn't at the bottom of the shaft, it's around 5 oclock.

b) A joke right?

Originally posted by cheapracer


a) The critical peak pressure point isn't at the bottom of the shaft, it's around 5 oclock.

b) A joke right?

a) or 7 o'clock depending on direction of rotation. Regardless the load carrying area will extend for several "hours" of the clock face.

b) No Joke. Oil pressure at the hole is in the tens of PSI whereas the film pressure necessary to support the loaded rocker is in the thousands. A pressure gradient will occur as you move out from the hole so an area much larger than the hole itself will be operating at lower pressure than the rest of the peak load area.

Gruntguru and Cheapracer,

First thanks for all of the constructive criticism. I helps to vet ideas.

My seat of the pants decision, to use 6 o'clock of the orifice position, was reinforced by a conversation I had with Noel Manton of Manton Engineering. It turns out he redesigned the Top Fuel rocker arm assembly, due to a similar problem, in much the same way as I implemented my solution. The 6 o'clock position can service the swing from 5 to 7 as illustrated by Gruntguru.

As this is a boundary layer interface, the main systematic requirements are for a periodic resupply of the oil film, and an equally periodic shedding of heat that is invariably generated by such rocker/shaft combination. I believe my seat-of-the-pant solution achieved both of these goals. Further, the decision to adopt a bronze/steel interface combination also steers away from the catastrophic possibilities when the metallurgy of the rocker and the shaft are similar in nature and also the fragile nature of any needle roller solution.

While my methodology may not have met strict scientific engineering standards, there is something to be said for "reasoned assessment resulting in quantifiable success".

Originally posted by gruntguru
a) or 7 o'clock

Standing at which end of the engine?

Originally posted by cheapracer
Standing at which end of the engine?

It probably oscillates between 5 and 7 O'clock, but viewed from which end of the engine? Too hard for me.

Hi, I happened across this thread on my way to somewhere completely different, and looked in because I have had a similar problem with an engine in my vintage racer.

There seems to be some unnecessary complication here, and I brazenly suggest a practical view....

What is the problem ?

Assuming you are supplying oil to a rocker shaft at Xpsi (engine running) pressure, and there is too much oil passing through the rocker plain bearings and building up in the rocker cover, therefore going down valve guides/ leaking out etc, you may decide to limit the flow to the rocker shaft by creating more restriction. Unfortunately if you do this at the entrance to the rocker shaft, you are likely to cause a pressure drop along the length of the shaft, causing insufficient supply, and worn bushes.

Solution, create the restriction at each rocker outlet, so the supply will be constant down the length of the shaft.

Its the same principle as in a "fuel log" . I drilled and tapped the oil holes in the rocker shaft, and screwed in grub screws (pipe plugs) which I drilled out with number drills till I got the right flow. I have to admit this was done by running the engine with the rocker covers off and monitoring the visible flow !

I suggest you make sure the oil holes in the shaft are facing towards the head, where the pressure on the bush is, and machine oil grooves as in the photo.

Incidentally, this problem occurred in my engine because I doubled the oil pressure. This can lead to all sorts of problems like leaks, oil surge, broken pump drive gears/shafts etc etc.

The old Cadillac V8 in my Allard had all of the above. At one stage I rigged up a crude manometer, ie a clear plastic pipe attached to the sump plug, running vertically up the outside of the block, and was fascinated to watch all of the oil being pumped out of the sump as soon as the engine was cranked over, a condition at its worst with cold oil/engine. Effectively the engine carries about 4 litres of oil everywhere except the sump when the oil is hot, running at fast tick-over! Increasing oil pressure from 25 to 50 undoubtably created this problem. I eventually installed restrictor plugs in the cam bearing oil holes which raised the engine oil pressure by 15 psi, allowing me to wind back the oil pump blow-off.

Hello Tom,

Thank you for your input. The REAL problem came about because of the use of much more aggressive camshafts. The standard camshaft was around 0.260 lift at 230 degrees duration. With 1.45:1 rocker arms this means 0.377 valve lift. Valve spring pressures were fairly mild. In one engine that I examined I could rotate the valve springs with my fingers with the valves closed. Thus the pressure on the rocker/shaft interface was not that great. Even so, if you ran one of these engines for 35.000 miles, the shaft was history and the rocker looked like the army had trooped through. The problem appeared to be two-fold.

1. With the excess shaft wear the clearance became worse and worse, with the oil control situation becoming seriously compromised.
2. As the oil feed was at 3:00 o'clock (Fiat did this so that you could "flip the shaft" and have a new virgin surface to work with) it was always doubtful that proper lubrication of the rocker/shaft interface could be maintained when all was new, and most assuredly not when there was excess shaft wear.

Today, in a competition environment, we are using camshafts with .332 lift at 300 degrees duration at the lifter. This is the ABSOLUTE maximum lift that can be used, given the diameter of the head of the mushroom lifter, without going off the edge. At the valve this means 0.481 lift, or greater than 0.100 increase in lift. The attendant increased lift, and accelleration of the valve train required much higher valve spring pressure, both static and over the nose of the cam.

Even when using standard Fiat components, all to specification with the correct clearances, the fact that the shaft and the rocker were both made of similar grades of steel and hardened to similar Rockwell values made them substandard for the increased level of performance required.

It was not a problem with too much oil in the upper end of the motor, or too little in the sump. Generally stock engines developed 50-60 PSI and competition cars developed 60-70 PSI. As it turns out there is a reasonable pressure drop between the crankshaft and the upper end of the engine, so this has not been identified as a problem.

This is the gist of what caused me to reconsider how the lubrication system worked, and how the rocker/shaft interface could be enhanced.

While changing the orifices of the oil feed ports was "one" adjustment, the re-engineering of the shaft (both in terms of surface technologies and oil feed supply location) and the inclusion of a bronze bushing (thus having totally dissimilar metal at the rocker/shaft interface) were instrumental in insuring that the rockers did not seize on the shaft.

I do like your idea of regulating flow with little plugs. I will have to look at if there is enough room to incorporate this idea, as the shaft is only 0.590 in diameter, with a 0.200 wall thickness.

Regards
Paul

Originally posted by tom walker
H

I suggest you make sure the oil holes in the shaft are facing towards the head, where the pressure on the bush is, and machine oil grooves as in the photo.

Incidentally, this problem occurred in my engine because I doubled the oil pressure. This can lead to all sorts of problems like leaks, oil surge, broken pump drive gears/shafts etc etc.
.

With respect and as I said earlier in this thread, machining oil grooves isn't required and went out around the 70's.

Notice with main and big end shells there is no longer oil grooves - again stopped around the 70's in production that I noticed anyway.

Grooves simply gives less surface area and increases wear for no benefit especially in this situation and oil is quite capable of finding it's way through a 0.005" clearance as is proved every day in millions of engines.

As you stated yourself, increasing oil pressure is just a waste of time and a power looser, increasing volume can however have some benefits for some engines that need it but not many.