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Pedal Smasher
1973 Opel GT
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MLS head gaskets use steel that has been stamped to create a spring for each seal. This head gasket was originally created by a Japanese company so it could expand a little bit when the cylinder head is pushed upward by combustion. This prevents high cylinder pressures from forcing the head gasket to fail. It is now one of the most common head gaskets used on production vehicles.

That's what my internet research had to say about MLS. Sounds like a head gasket I'd want to use. One result said the compression height was 75% roughly of the thickness, but I can't verify it yet.
 

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Opel Rallier since 1977
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Bob, is Cometic Gasket's part listings for the head gaskets the uncompressed height or compressed height?
Everything I have ever read, and every number I have used for Cometics, is that their specified thickness IS the compressed thickness. The uncompressed thickness all over the map due the multi-layer make up of these gaskets, plus the embossed sealing ridges. One example of an actual measurement done in the field:

Be aware that Cometic spec's a finer head and block milling finish for their gaskets. They have softened their insistence on this in the last few years, but it is something to be aware of. From what I have learned in using them, the older flat grinder methods of finishing blocks and heads (like Blanchard grinders) typically produced a rougher finish that might give the Cometics some sealing issues. Newer milling methods tend to produce a finer surface finish that does not give any problems.
 

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Discussion Starter · #24 ·
While I generally enjoy the trip down the off-topic journey of head gasket materials, compressed thicknesses and such, this station will now return to the original programming...

Definitely get all your hard numbers in place first.

-calculated swept volume (actual bore x stroke)
-combustion chamber volume (using the actual plugs run in the engine)
-valve relief
-piston dome
-deck height (positive or negative)
-top ring land volume (usually about 1 CC but easily calculated)
-compressed head gasket ID and thickness

This is really the only accurate way to get the true compression ratio.
We have now been able to get accurate measurements and volumes from this engine "as built", and here are the results, as summarized in the attached table (input and calculated results from the html calculation file provided by Port City Engines) and related photos:

The stroke measured at the stock 2.75" (69.8 mm).

The bore is 94 mm (a 1.0 mm overbore) or 3.700".

The piston small ends had been bored to fit the floating pins on the Jahns pistons, and the effective rod length is now 5.02"

We used plasticine to measure a combustion chamber, with a spark plug in place, and determined using fluid displacement that the chambers are almost exactly 48 cc. That is with Chevy 1.72" intake valves, and Chevy 1.5" exhaust valves. It turns out that there are hardened exhaust seat inserts installed.

The net piston dome volume was determined by directly measuring the net non-dome volume (the volume above the piston "flat", below the dome crown) using a ring compressor at the crown height and plasticine to fill the void. That was then verified by repeating the process of the dome piston with the top at the block deck height and subtracting that from the volume of a flat-top piston at the same depth. Both resulted in a net dome volume of 12.5 cc. That does not include the head gasket volume, but does include the piston deck net height.

The PCE calculator allows the input of top ring land volume, which measured 0.20" below the deck.

The head gasket that came off this engine was measured at 0.034" thick, compressed, at the sealing rings.

And the envelope please.....

Static compression ratio was 12.36:1

Dynamic compression ratio (with the 2 degrees cam advance) calculated at 9.96:1


So while some folks might aspire to have a CR of 12.4:1, that is simply too high for this street engine.

We are still evaluating how much piston dome to mill off, but the target SCR is ~10.0:1 and a DCR of ~8.0:1. That seems achievable by milling off approximately 4.5 mm of the 6.27 mm (0.247") piston dome. The valve reliefs measured at 1.5 cc. That is all illustrated in the attached PCE table.

Oh, and as an aside, when the cam duration is reduced from 268 degrees to a bizarre but perhaps illustrative 134 degrees (with the cam lobes worn down), the SCR stays at 12.36:1, but the DCR climbs to 12.34:1. Hence the seemingly huge compression test pressures.

Finally, we tested the "installed" and "closed" spring pressures on the dual springs in this engine. They measured at:

Intake (installed/open): 72 lbs / 158 lbs
Exhaust (installed/open): 75 lbs / 155 lbs


The spring rates at open (0.420") for both springs measured at 205 lbs/inch and 190 lbs/inch respectively. I suspect that they are actually the same if we had used more precise measurement equipment.

The exhaust springs have shims installed, which explains the slightly higher installed pressure. The intake springs did not, so we will add a shim to get installed pressures closer to 90 lbs. Otherwise, the springs are no longer considered complicit in the cam/lifter failure, so that is being ascribed to poor break in, and insufficient ZDDP in the oil.

Any further comments (on-topic, if you please)?
 

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Discussion Starter · #25 ·
I think you might want to include a different gasket as part of your solution, unless you make the pistons flat tops.
Why (aside from using the later non-cork version)? We are milling off most of the piston domes off, which achieves the net combustion chamber volume (and SCR) desired.
 
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Good deal on the springs.... it's unlikely that those pressures would damage the cam IF the ZDDP was maintained with good break-in practice. Make sure you check out the whole head oiling system thoroughly to make sure it is all good. My last spring install was:
  • At an installed height of 1.60 to 1.55, the closed pressure is 120 lbs + or -. (Stock numbers that I have are 85 lb IN & 74 lb EX.) For my .407 lift cam, open pressures are going to be around 210-220 lbs, vs the stock 150-160 lbs.
  • This works out to 235 lb per in spring rate.
It's all surviving well so far.... the lifter were re-conditioned units from OGTS... I trust the old metalurgy more than new. (And I think Gill feels the same.) Break in was with Gibbs oil and I have switched to Mobil1 15W-50 to get a better cold start viscosity to keep down some of the leakdown and start-up lifter rattle. (The higher spring pressures cause more leakdown issues.)

The piston milling oughta work well. Is the deck height above or below the deck? Just be sure of the sign convention used in the calculator.... ditto for the sign convention used for the valve reliefs and domes. This is one spot where folks go astray. Sounds like the piston to head clearance on the piston flat will be 32 to 36 cc (depends on the deck height being above or below block deck ). Either would do well for quench/squish effect to help combat detonation so that is good. I just get a bit nervous down around 0.030"; my last was .028", but so far, no evidence of pistons kissing the head!

BTW, the the cam duration of 268 that you used is the right number for DCR computations if you got it out of the OGTS ads, and would be called 'advertised duration' by US cam companies. Just be aware that the slow ramps of the old Isky hydraulic cam designs will make the duration effectively shorter than 268 in its effect on DCR (and raising the cranking compression numbers). I'd like to hear your cranking compression numbers when it is broken in. A real DCR of 8 at 3000' computes to a cranking compression of around 150 psi.
 

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Discussion Starter · #27 ·
A standard 1.9 gasket usually specs out to .039” new, and .031” compressed with a 94.5 mm ID.
Bob, on the topic of head gaskets, could you please provide advice as to what size of head gasket to purchase?

The cylinder bores on the engine are 94 mm, so exactly in the middle of the 1.9 (93 mm) and 2.0/2.2/2.4 (95 mm) cylinder size. The head gasket that was in this engine was a Fel-Pro, with a measured ID of 95.1 mm. Visually, it appears that the sealing ring is JUST barely larger than the cylinders. It was the 10-bolt style with the cork spacer gasket at the front, which was incorrect, as this engine has a 12-bolt style chain case, negating the cork gasket. Clearly it sealed the cylinders, even at the elevated compression pressures this engine created. Perhaps not so much the high pressure oil passage, due to the cork gasket, as evidenced by the oil found in the coolant.

My inclination is to purchase a 2.0 (95 mm ID) head gasket, to ensure sufficient diameter for the 94 mm cylinders. Or is a 12-bolt (non-cork) 1.9 head gasket a better choice, as it provides less lost volume around the edges of the cylinder?

I will also ask Gil what head gasket he suggests, based on the actual ID of his head gaskets.

TIA.
 

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Bob, on the topic of head gaskets, could you please provide advice as to what size of head gasket to purchase?

The cylinder bores on the engine are 94 mm, so exactly in the middle of the 1.9 (93 mm) and 2.0/2.2/2.4 (95 mm) cylinder size. The head gasket that was in this engine was a Fel-Pro, with a measured ID of 95.1 mm. Visually, it appears that the sealing ring is JUST barely larger than the cylinders. It was the 10-bolt style with the cork spacer gasket at the front, which was incorrect, as this engine has a 12-bolt style chain case, negating the cork gasket. Clearly it sealed the cylinders, even at the elevated compression pressures this engine created. Perhaps not so much the high pressure oil passage, due to the cork gasket, as evidenced by the oil found in the coolant.

My inclination is to purchase a 2.0 (95 mm ID) head gasket, to ensure sufficient diameter for the 94 mm cylinders. Or is a 12-bolt (non-cork) 1.9 head gasket a better choice, as it provides less lost volume around the edges of the cylinder?

I will also ask Gil what head gasket he suggests, based on the actual ID of his head gaskets.

TIA.
You don’t need a 2.0 head gasket for a 94 mm bore.
 
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FWIW, I noticed that the fire rings were sealing well on the pix of the head. As long as they seal, there is no need to go bigger. It is only to make sure that the rings register well around each cylinder.

When I have selected Cometic's in the past for Mopar V-8's I have typically gone to holes .080" (2 mm) larger than the actual bore, but that extra is just to insure that any bore mis-registration is not going to slip under the edge of the fire rings. Sometimes bores DO get registered off-center a bit when cylinders get bored. (And on occasion, that is intentional...)
 

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Discussion Starter · #30 ·
Following this discussion and others, it was decided to mill a portion of the piston domes off. The target would be an honest +10.0:1 SCR, and an DCR somewhat above 8.0:1.

Based on that, John and I set up his mill and modified the piston holding fixture he had built to mill his RallyBob Stroker 2.4 (Chevy 305 Keith Black Hypereutectic) pistons 0.050", to match his block deck (see the recent post about "What Cam for a RallyBob 2.4 Stroker).

These Jahns pistons are sand-cast and then machined to the specific desired design. We determined that, even if we milled the pistons all the way down to flat top, the piston tops would still be almost 1/4" thick. So lots of meat in these pistons.

The domes were almost exactly an additional 1/4" (0.248") above the piston top "flat". With a bit of volumetric calculation, we milled the domes down to ~0.100". In other words, the 6 mm domes would now be ~2.5 mm.

The attached photos show the process. We then repeated the net dome volume determination using plasticine and a ring compressor. The net dome volume had been reduced from 12.8 cc down to 6.35 cc. That reduced the SCR from 12.44:1 to 10.96:1, and the calculated DCR from 9.87:1 down to 8.72:1.

I happened to watch an engine builder's YouTube video the other day that discussed the issue of "What compression ratio is optimum for a street engine using pump gas". His contention, which is strictly an opinion, if an experienced one, is that too many engine builders give up too much engine potential unnecessarily, by limiting CR's to 9:1 or slightly above. He firmly believed that a CR of 11:1, WITH PROPER PISTON QUENCH DESIGN (specifically NOT dished pistons), may be allowable, and achieves higher, more efficient power.

We still have a lot of work to do before this engine is ready to be put back together. But the major impediment, excessive compression ratio for this street engine, is, I believe, behind us.

Comments?
 

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His contention, which is strictly an opinion, if an experienced one, is that too many engine builders give up too much engine potential unnecessarily, by limiting CR's to 9:1 or slightly above. He firmly believed that a CR of 11:1, WITH PROPER PISTON QUENCH DESIGN (specifically NOT dished pistons), may be allowable, and achieves higher, more efficient power.

We still have a lot of work to do before this engine is ready to be put back together. But the major impediment, excessive compression ratio for this street engine, is, I believe, behind us.

Comments?
I’m personally a fan of the Smokey Yunick thought process when it comes to compression ratio and the best way to get it.

In a nutshell, a tight quench, a small combustion chamber and minimal dome ( or flat top, or yes, preferentially a dish) is far more efficient than a big dome and a big chamber with the piston below deck. It’s so much more efficient you can run substantially higher compression without detonation while making better power and fuel economy.

This is one of the reasons I like using the 1.5/1.6/1.7 heads for street engines. The small chambers, after modification of course, are far more efficient.

Regarding the use of dished pistons, while I agree that a large dish is not the most efficient, as long as it has a perimeter ring that is flush with the deck (or slightly above) this will induce some swirl.

Ideally, the best design is a tiny flat chamber with a reverse (mirror) dish piston design, slightly above deck. This creates a large quench pad that interacts directly with the head’s surface. In this regard, the 1971-1975 Opel pistons were actually very well designed.

Taking it one step further, if you designed a piston with a flat narrow outer ring that was above deck (say, 1/4” wide), and then the inboard piston ‘flat’ areas were machined at a slight taper (down towards the piston center), you’d pretty much eliminate all ‘dead’ quench areas and focus the high pressure swirl towards the bore center. It is possible to have too much quench and if you had too much flat area the wave front can run out of time to travel in the brief moment the piston is at TDC.
 

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Pedal Smasher
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Your dynamic CR limit with running pump gas is going to be low 8’s. 8.72 is likely going to be too high for pump gas, especially if the engine isn’t EFI. My 2014 Mustang GT had an 11:1 SCR in that 5.0L mod motor but it also had VVT and was properly designed for optimum performance on pump gas. A whole different world compared to the CIH. So I would agree that a street engine can be 11:1 if you have a bunch of tech to support it. For the CIH, I’d error on the side of caution. I wouldn’t want to be in a position where you’re having to add octane booster or run 100 octane gas to help keep the engine happy. 10.96:1 is too high in my opinion unless you go with a really late ICA.

You asked for comments, that’s what I’m thinking. Sure I’m not an expert but it’s easy to research DCR limit. If you keep coming back to the same info, then chances are it’s generally true. At 8.72, I’d want to know for sure that other people have done it on pump gas with no additives. I’d also do a wet compression test to see what kind of cranking pressures you’re getting.

You could try running E85 at a DCR of 8.72. Maybe that would work. E85 has a colder flame temp and can deal with higher cylinder pressure which is why the turbo crowd likes it.
 

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Looks like good work Kwil. BTW, I see on the computation sheets that you have 0.02 degrees advance entered.... should that not be 2 degrees? And be advised.... the cam lobes on the Isky cams are very old designs (late 60's) designs and have slow cam rates near the opening and closing events. This means that the intake closing angle effectively occurs earlier... which means it will act like a shorter duration cam and will raise your DCR. So be aware .....

FWIW.... The energy content of E85 is lower... which is why the cylinder pressures are lower. In essence, it is getting to the same cylinder pressures with a different path. The problem becomes obtaining the E85 in one's area.

I know a very good engine builder in OR who runs street DCR's in the low 9's, and is pushing 1.5 Hp per cubic inch naturally aspirated, and does it on iron heads. But he reeeealy know what he is doing and goes to a lot of lengths to prevent detonation. He does the tricks mentioned above line with quench/squish, and also polishes chambers, runs Evans coolant, and so on. All to avoid hot spots, have a good burn, etc.

The point is that running up in the high 8's generally means you have to be on your toes in assembly and design, and with tuning mixture and ignition timing. You may have to do more; I'd certainly polish the had chambers which will remove maybe a cc of material and help out. Most guys should not push beyond the high 7's or low 8's.... not saying you are 'most guys'... just be aware you are pushing things. But that is part of the fun for some of us.

Of course it is not a disaster to start where it has been proposed.... just retard the cam timing or change cams or get a thicker head gasket like a thicker MLS or old Felpro's and you can move the DCR down easily. None of those things are difficult IMHO. These engines are pretty easy to work on... well, maybe not in a GT LOL
 

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Discussion Starter · #34 ·
...BTW, I see on the computation sheets that you have 0.02 degrees advance entered.... should that not be 2 degrees?
It makes me happy that at least ONE person is reading what I post.

Yes, that is an input error. Effectively zero for cam advance in the computation. Which, given the high'ish DCR, might be where we start. That will likely be effectively cam-retarded, due to historical head and block milling. But it isn't terribly difficult to change the cam off-set bushing, so some experimenting may be in order.
 

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Hey I'm on it! Some background:

This is part of the job that I put a lot of effort into and enjoy, and have put a lot of time in learning the DCR issues, the effects, etc. My first cut at all this was in 1974, my 1st performance engine (Ford 351Cleveland), and before all the computation sheets existed... no internet, and lot of talk back then about 'effective compression ratio'. I carefully did all the SCR computations by hand and made may best guess at SCR and cam combo. A few years ago, I dug out my old notes on that engine's parameters and put them into a DCR computation, and now know that I hit 8.3 DCR. Cranking compression at 1000' was 160-165 psi, which lines up with that DCR so the cam parameters were accurate for DCR.

Just what I wanted for a strong, torquey street engine but that could get 19 mpg .. that engine would pull from 2500 and the shift point was 6500 RPM; the smaller cam, the big breathing heads and intake and large headers, and high DCR made that work. And it could beat factory 440 Mopars all day long LOL.

For detonation, it was fine at the local altitiude of around 1000' but when I would go to sea level, it would want to ping. Faster distributor advance would do the same. So that is one direct experinece with understanding DCR limits and gas, and why I say what I do about your pushing the envelope. Chambers are different so the exact outcome vs. DCR and cranking compression may vary a bit but not by much.

I have been spending a lot more time in the last several years with cams and profiles to try to get the computations more accurate. The intake closing angle is a critical number to the DCR computation.
 

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It makes me happy that at least ONE person is reading what I post.
There's more than that kwilford it's just way above my head so I'm silently following ( as requested) lol
 
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