V4 Blueprint. A primer sort of.

Thanks. Since I have 8 rods I should be able to figure out the intended small weight hopefully.

A discussion about the broken valve collets. And shimming in general. I had 5 broken collets, 4 of which were intakes. These all failed within the ducati valve train operational specs but all were looser than assembly specs. The closer clearance range on the broken ones was 0.005 - 0.008 inch (0.125-0.20 mm). The valve stems neck down slightly in the section where the shims seat but as they were measured it was clear that the valve stems were essentially identical. Then the stem holes in all the closer shim were measured. These are also being held very precisely. But they do have clearance. For the collets to break in the middle like they do, the middle has to be able to be distorted relative to the ends which would have to be constrained so that force can be applied to the middle. Two ways this can possibly happen is the ends of the collets are slightly bigger than the middle or the clearance between the closing shim and the valve stem allows the closing shim to .... slightly on the stem and if the collet is oriented such that the middle of the collet is in line (90 degrees to the centerline of the rocker face) with the repetitive forces of the valve bouncing off the seat relative to whatever clearance is there. Recognize also that the opener clearance is based on the change in valve stem length. And that the closer clearance gets bigger as the stem lengthens.
The valves that had broken collets were reassembled with unbroken (but run in on the other valves) collets so the change in closer clearance contribution from the broken collets could assessed. This was found to be between 0.001-.0015 inch (.025-.038 mm). I'll do this again when the new collets show up. This will allow understanding of the contribution from running in new collets. When I reassemble, I intend to orient the collets such that the gaps are in the same plane as the rocker face. This may or may not help as the possibility exists that the whole bit rotates in operation. But at least they have to find their way there first.

Valve shimming. I remember being part of a discussion on here about valve shimming and how to deal with ones outside of the operational specs. Showed a total lack of understanding of the valvetrain dynamics. The bottom value of the assembly specs (1103) on these are 0.004 (0.1 mm) inch for intake openers and 0.005 inch (0.125 mm) exhaust openers and 0.002 inch (0.05 mm) for all the closers. If you want longterm valvetrain reliability these should be adhered too. I intend to take the closer springs off the heads so I can actually assess the needed closer clearance. The closer clearance is primarily driven by the clearance needed to ensure that there is not any operation drag caused by interference of the opener and closing cam lobe profiles. If the machining allows me to run the closers at 0.001 inch I intend to do this as I have with my other bikes (have to look at some thermal characteristics first). Loose closers allow the valve to bounce off the seat as it closes. Bigger gap bigger bounce and more bounces. Ducati is using coil springs with the coils stacked on one another instead of hairpins on the closers because the drag between the coils faces provide a level of damping but the banging about still happens.

A couple of other points. When the opener clearance are larger than the minimum you get less valve lift by whatever the additional clearance is. But that isn't really what you're giving away, what you're really giving away is duration. When the valve is lifted beyond one quarter it's diameter the curtain area between the seat and the valve now exceeds the port choke. So the port is flowing at maximum relative to the pressure drop it's seeing. This is the duration you're giving away. Maybe not alot but the motor's making less torque than it could (blueprinting is about getting it all). And anyone who has ever simultaneously timed the cams (I've never had a Ducati where the cams were actually completely correctly timed from the factory) and set the valvetrain back to minimum assembly specs knows this is magic. Particularly with the twins, they vibrate less and are simply happier and you can feel it in the throttle response. When I timed the cams and reshimmed the valvetrain on the clowncar it was already so smoooh I thought I'd never feel it. Wrong same effect. Just got crisper with less vibration.

Bearing sizes are sorted out so I can buy parts. A comment about bearing clearances. The factory bearing clearance range is 0.04 to 0.07 mm for the mains and .035 to .0.065 for the rods. With my bearing selection my hope (there's variability in the bearing size classes of 0.005 mm) is I'm real close to 0.04 mm for both the rods and mains. Excessive clearances consume more oil pump volume. I want to note that in the end I will use the identical bearing colors (sizes) as the factory did in the same locations for the mains with the exception of the upper center bearing goes from a black to a yellow (more clearance). That being said it's right on the edge so I'll buy an extra black upper and actually try both for clearance in situ. What happened is weird. If the motor had been assembled with a crank that had all the same main journal size (class B) for all the journals the factory main bearing choices would have correct. But it wasn't. It's as if the crank (the center main also has run out remember) was picked from the abby normal pile.

So I'm to the heads. If any additional power it will be found it will be in the heads. I don't expect much but there's always some (if you do the right things). As I mentioned above when the valve is lifted to greater than 1/4 it's diameter the curtain area between the valve and seat exceeds the cross sectional area of the port choke. I should qualify this by saying if the port was designed appropriately and nobodies gone in and screwed it up. You can always open the port up (bigger cross section) and add some flow volume but if you increase the cross section a greater percentage than you increase flow volume the ports velocity goes down and you may get a few topend HP but this is at the expense of midrange torque and generally acceleration will be worse. On these heads which have a lot of flow volume anyway the plan is increase flow volume while not increasing cross sectional area. This gives you more torque everywhere. The most important part of the port with regards to prep is the last 10 or 15 mm of the port closest to the valve seat. For best performance this needs to be flawless. Originally, I was just going to get rid of any casting defects and perfect the approaches to the seats but as I'm seeing some difference in port centering of the intakes with respect to the valve guide centerline it makes sense to flow test the ports to see if any gain is possible. Doing this is actually pretty simple as all you really need is a shop vac and a makeshift manometer (I'll take pictures when I get there). Commercial flow benches typically are run at 28 inches of water (1 psi). And this is regardless of the valve lift point being tested. I've always considered 28 inches wrong with the exception of a wide open port. When the port is flowing with the valve closer to the seat when the motors actually running the pressure drop will be higher than an unrestricted port. So when the flow bench is built I won't bother to put a secondary restriction to allow me to always adjust the drop to 28 inches as I want to measure low lift flow at higher pressure drops (particularly the exhaust which in real life is operating at much higher pressures). And there's a second reason I want to flow these. On the exhaust side a properly designed port will have a venturi formed after the valve seat. The ratio between the exhaust valve head diameter and the cross sectional area of the venturi is a critical tuning factor. When the venturi is properly sized the torque curve will be optimized. Too big and you'll get some topend hp at the expense of torque and the motor will act more cammy due to worse scavenging in the midrange. Too small and you'll get more low end torque at the expense of pretty much everything else. I find these somewhat smaller than expected but the real measure is the flow volume ratio of the exhaust to the intake. Measured at 28 inch drop the ratio of an effective head is exhaust flow equals about 75% of intake flow typically. The intakes will be flowed with intake manifolds attached and the exhaust will be flowed by attaching the header and pulling a vacuum from the end of the header.
Pictures of the valves from both heads. The rear head valves I've polished on the drill press but the front are untouched. The valves are polished to slow down crud attaching to the valve heads as this adversely effects low life flow (keeping the flow attached with minimum turbulence helps low lift flow). This motor had 14k miles on it. The condition of the valves shows two things; clearly no oil was going down the guides and from the color of the exhausts the motor was running a bit lean as a bit more tan would be more to my liking.

The heads are really beautiful from an airflow standpoint. So any hacking will be pretty minimal. I bought a complete head with a trashed cam saddle for $200 off fleabay to mess with. I have noted some minor stuff particularly in the intake ports closest to the camdrive in both cylinders. I'll have to cut a little than flow (repetitively) to see if I get any meaningful gains.

A couple of weeks ago there was a no start thread where the front head was replaced (and mistimed) due to a toasted cam saddle at the saddle where the chain drive is attached. I asked him to post a picture but never happened. Since the motor being worked on showed some signs of distress I purchased a front head off Ebay that was advertised as having a trashed cam saddle on the chain drive side. The rest of the head appears to be low mileage. I wanted a extra casting to fuss with the porting on anyway so this would also provide an opportunity to see another example of the failure. Unfortunately I don't know the heads history.
Here's the camdrive saddle on the front head from the engine I bought. Remember this one is known to have suffered an oil pressure loss.
On neither head could I pick up any increase in saddle diameter. Thinking this was a purely surface thing I used Q-tips and some aluminum polish on the front saddle of the head from the motor I bought. So the surface discoloration is readily dealt with. And both heads should be serviceable.
Note that the two failures are different. On the Ebay head the discoloration is closest to where the chain pulley attachment but on the other it was more towards the back of the saddle. The Ebay head issue is due to chain forces (probably) and the other was due to the oil pressure failure. Remember this motor had declining oil pressure for a while so the oil pressure being applied to the cam drive chain adjuster would have been lower. And the force being applied to the saddle by the chain would be lower. A couple of pictures of the chain adjuster.

The heads are fed oil from the mains galley (after the mains are fed). So effectively last. The oil feed for the chain drive saddle is shared with the chain adjuster. The chain adjusters bleed pressure off (while lubing the chain and chain guides) thru the end hole above (1.07 mm). At the head deck interface, the oil feed to the cam to the chain drive saddle is jetted down by a 1.4 mm hole. A second jetted feed goes to the opposite side saddles with a 1 mm jet. The jet to the chain drive side has the twice the area (so twice the oil flow) as the non chain drive side and 70% greater area than it's partner on the oil feed, the adjuster.

Will this be an issue with the heads on high mileage motors? In order to keep as much oil as possible being fed to the crank I don't want to increase the flow to the heads (and the heads are actively scavenged). But some remediation might be needed here. Since the oil feeds to the saddles also feeds the rocker pivots it maybe be useful to restrict flow to the rocker pivots to get more oil to the saddle. The problem is figuring out how to do that. Note that the rear head does not suffer as much of a problem due to the chain drive being on the other side and on this side the oil hole precedes the precedes the load and due to the oil hole not being at the bottom the effective bearing area is increased.

A comment on a possible fix. The chain drive saddle shares the oil galley with the chain adjuster. The pressure applied to the chain is directly related to the oil pressure being applied to the adjuster and the size of its end bleed. The adjuster is fine threaded and is pulled up against the head. Like any other threaded part a gap will exist on the back side of the thread after it's torqued (if not there would be no market for pipe goop and teflon tape). So the oil loss (use) of the adjuster is due to the bleed and the loss down the threads. A little high pressure pipe sealant on the threads will stop that leak. Therefore the bleed at the end of the adjuster can be opened a little bit which would diminish the chain load a little without increasing oil circulation. Then on the front head only the oil jet fitting the chain drive saddle will be opened slightly to increase flow. I think the oil to the rockers can be decreased at the rocker pivots, which are hollow, by putting a jet of sorts into leading rocker pivot.

Before I go out drinking maybe I should finish the thought. If I increase the bleed size to the chain drive saddle to the next drill size the flow to the saddle increases 16%. The total increased flow to the front head would be 10%. Acceptable I think. The heads are actively scavenged but the drain path from the front head is far more torturous. The hole thru the head gasket is 10.5 mm but the oil returns are 9.9 mm. If these are increased to 10.5 mm the area increase is about 15%. This return turns 90 degrees twice, once at the entry to the lower case galley and again at the pump entry. The pump entry is 12+ mm. The outer diameter of the lower case galley is 18.25 mm. This can be increased from 9.9 mm to 11 mm with more than adequate galley wall thickness. This increases the area of that galley an additional 10%. Then where possible where a sharp edge transition exists, like it does from the head to the case, it will be blended into a radius. This will hopefully allow more efficient scavenging compensating for the increased oil flow to the front head.

A picture of the scavenge side of the oil pump.
I stuck the picture here so the head scavenge screens could be seen. So lets consider the scavenge screens. We all know from oil changes and setting the appropriate oil level how much oil is in the heads even at idle. So the valve train starts to check out and puts debris onto the screen until it's clogged, head fills with oil and if it's enough oil volume it takes a rod with it due to lack of oil volume in the sump. I consider the screens pretty silly. On the clowncar I epoxied an alnico magnets (good heat tolerance) into the head bolt hex openings on the headbolts adjacent to the oil returns. So the same on this one but maybe a couple more magnets. Then I'm going to toss the screens, they are a little restrictive. BTW, if the 2025 Pani (which is now the R pump) is backwards compatible I suspect that pump will be one I use.

Small chamfers will be added to exit side of the rod journals.
I've decided to add windage elements to help get the oil off the rod big ends and into the scavenge sections of the pump. I'll post a picture when done.

Above I commented on modifying the cam chain adjusters by increasing oil flow thru it. After some thot I consider that incorrect. I've come up with a superior approach. Since I'm into the heads I've decided not to give away any of my head development efforts (and I've come up with something else that all I can say is that it makes me giggle as I anticipate the result). Thanks for the interest.

Above I commented on a potential additional windage feature. First a couple of pictures of the lower half of the crankcases followed by a couple of pictures of the profiling of the counterweights.




The crankcases are sealed as a pair of twins. The only entrances to each twin crankcase is top ring end gap and the various oil orifices including the jets squirting on the backs of the piston heads. Exit is thru the 2 pump scavenge sections one per side. Each scavenge section is sized at about 132% of the total pressure side output volume (264% for both). This pump is rev 2 I think. The heads have a separate scavenge section. The crankcase scavenges discharge into the sump oil pickup. The separate head scavenge section's output is pointed at the gearbox on (and on the bearings of) the drive side. What surprised me is the although the crank counterweights have a windage feature the rods do not. That would be the fins sitting above the scavenge inlets. These knock the oil off the end of the counterweights and are aided in their function by the crank profiling. I've cogitated on the lack of a rod feature a bit. When the front cylinders are at the bottom their strokes (the crank is sorta parked there for a few degrees of rotation) they're directly above the scavenge inlets. A lot of oil is being constantly squirted on the backs of the pistons and the most natural return path is down the rod to the bigend. The rear cylinders are at the bottom of their stroke 90 degrees forward (crank rotationally) of the scavenge outlet. So I expected there would be a rod bigend windage feature coming off the opposite crankcase wall at 90 degrees plus half the width of the bigends. The rear cylinders are parked at that point, downward inertial forces on the oil are at their greatest so a feature to knock the remaining oil off the rear cylinder bigends bigends right as they start to accelerate upwards would make sense (to me anyway) there. Since you're there you might as well profile the front cylinder bigends too. As a cast feature this could be a straight ramp starting at the crankcase face and angling in to profile the bigends with a greater than 90 degree turn at the bottom to encourage the any oil draining down on it's face to separate and drop towards the scavenge inlet. I've marked up a picture to sorta illustrate location. I can create something in sheetmetal but to secure it I would have to penetrate the outer crankcase wall.
I took the oil pump apart so a couple of pictures of the housings particularly relating to the oil pressure relief valve and the head scavenge section.

A concern is always that the oil being bypassed upsets the oil pump main inlet. Here though (top cavity) the bypassed oil jet is being aimed at the curved portion of that cavity so the force is being dispersed somewhat there. The following picture is the relief plunger and spring. I actually took apart the pump initially to measure the K of the relief spring but I'll have to find a decent load cell to measure it accurately. Then I can calculate the opening pressure. When I do I'll post the numbers. A couple of pictures of the head scavenge section.


The front head scavenge return passages are very convoluted while the rear head scavenge passages drop straight to the pump. Ducati has compensated for the difference somewhat by turning the scavenged oil twice as it comes into the scavenges section (the long cavity to the gear section top picture).
I'm to the heads so I need to build a flow bench. I've decided to post the raw numbers and my numbers at completion but not exactly how I got there.

When I started going thru this engine it was primarily as a replacement for the engine in the clowncar. An assurance that the assembly was correct from a reliability standpoint and maybe net out a few HP. This however has morphed somewhat as I'm planning to go to R throttle bodies. They're 56 mm vs 52 mm on the 1103's. To take advantage of these the cam centerlines will be changed slightly and the exhaust valve bowl diameter will be increased a few percent. The hope is that with some flow improvement in the intake ports the HP peak can be moved upwards 750-1000 rpm (20 HP?). The trick here is not to murder the midrange torque by having slowed the intake velocity with the bigger throttle bodies. Which leads me to a short discussion on using the ETV maps to maintain port velocity by restricting flow thru use of partially closed throttles below the torque peak. The 56 throttle bodies have about 15% more area than the 52's. So even with the throttle WFO the thot is to have actual percentage of throttle opening below 100% (ETV restriction if you will) thru the midrange and being tapered to 100% at the torque peak. I actually use a version of this to about 7500 rpm in the 4,5,6 map of the clowncar which provides the best acceleration in higher gear roll-ons starting at about 4500. Curious what one ultimately finds useful.
In order to travel on the clowncar I geared the bike up to 36/17. Makes for more casual 100 mph cruising and improve gas mileage. I have been studying the various V4 (including Multistrada) transmissions. The clowncar has a 30/24 (1.25) sixth gear. The multistrada uses a 1.08 sixth and a 1.26 fifth. Since the parts are readily transferable the multi 5th and 6th can be used with any other 1-4. For me the best choice would be new Pani 1st (12% higher) and second (6% higher), original 3/4 (they're clustered) and then 5th and sixth from a multi. Then I could go to 39/16 and have the same final 6th drive ratio as 36/17. The first 4 gears would be closer and I'd get some anti-squat.