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-assuming this is your car and engine-
Sorry to hear about this George.
The engine sounded good, like it was about 50-Cubic Inches Larger than a 388..lol

What RPM's were you shifting at?
What times and mph did you get?

I noticed that the pistons were flat tops with gas ports.
What was the 'Quench Gap'?

I ask this because of the burn pattern I see on the piston
you show at 4.05 minutes into the video.

I see what appears to be a lot of 'Squish Velocity'.

-note-
Squish velocity is the speed at which air is displaced from between
the piston and head surface as the piston approaches TDC.
It is the result of three factors: clearance, area and piston speed.


Also, what type of fuel?

How did they set the deck height and quench gap?
Rod length, piston pin location, deck the block, etc.

See, you post a video and we play 64-Questions..:p

--------------------------------------------------

You said your going to build another 388 cid.
Nice engine combo. . . .

You said your going use a better con rod, an I-Beam.
Good decision.

You said your moving up to an LS7 Head.

Good decision as you can generally get more air / cfm through
the cylinder head by convergence lift, on up to 0.600" valve lift.

With the above in mind; if you desire too, use the 2.204" valve and
bore the engine to 4.155". Have the combustion chamber fitted for
the 4.155" bore and shaped / contoured to get as much air into
the cylinder by convergence lift as you can. Convergence lift for
a 2.204" intake valve will be 0.551" of 'Valve Lift'.

-note- make sure you have the block checked for core shift
before boring it to 4.155".

Just for the record. . .
The earlier blocks were crap in this department,
but the new blocks are good. The blocks they are making today
can be bored to 4.185" and are good for 1,500 fwHP.

Or consider a Dart block instead. . .

If you do this correctly you can get about 385 cfm to 395 cfm
through the head by 0.551", to 0.600" valve lift.

Don't put a camshaft in it with over 0.665" valve lift.
The heads simply go turbulent by about 0.635" lift on
a 'Running Engine. And that is with a good head!

Lesser heads won't flow past about 0.585" of valve lift.

That's one of the reasons so many camshaft grinders don't
grind cams with much more valve lift.

0.665" gross lift, less all valve train flex and geometric
losses, will most likely generate a 'Net' valve lift of
about 0.635". Effective cam lift would be 0.87 multiplied
by 0.665". That would be 0.579" of valve lift. That will
put you right about where you need to be with valve
lift versus cylinder head flow as I have specified above.

Without using something like one of those Comp Cams
that opens and shuts the valve like a 'Hammer', get
as much 'Area' under the curve as you can.

If it's okay with your engine builder, I can put you in
touch with someone who can do that for you.

Faster 'Flank' and opening rate and, more area under the curve
without acting like a hammer.

Be careful with that 12.0:1 static compression ratio.

Make sure the intake valve closing point is 'Dead On'
and the fuel is specified correctly. Also, make sure
the 'Quench Gap' is sufficient.

I might 'Reconsider' moving up to that high of a static Cr.
Just might not be worth it. . . .

If you want to contact me again to discuss this engine,
as we did the previous one, just let me know.

You build expensive engines and I myself would like
to see the next one last my friend.

Take care George!

Cheers,
Bruce
 

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ouch!
 

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death by boost... Sounds like the new block is going to be a beast!! Quick question... Why is an I beam rod better than an H? Just curious for my own knowledge...
My new engine is being built I beam rods also...
 

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death by boost... Sounds like the new block is going to be a beast!! Quick question... Why is an I beam rod better than an H? Just curious for my own knowledge...
My new engine is being built I beam rods also...
You mean you don't remember the 'Great Con Rod' debate
between Matt / GP Tuning and I...lol

I will find the info and post it.

But that is exactly how an H-Beam con rod breaks
most of the time.

The only place the H-Beam is stronger, is up around
the wrist pin area.

If George had just completed adding more ignition advance, and
I am correct in surmising the 'Quench and Squish Velocity'
issues, then this 'Might?' have caused the con rod to break.

More later when I have time. . . .
 

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You mean you don't remember the 'Great Con Rod' debate
between Matt / GP Tuning and I...lol

I will find the info and post it.

But that is exactly how an H-Beam con rod breaks
most of the time.

The only place the H-Beam is stronger, is up around
the wrist pin area.

If George had just completed adding more ignition advance, and
I am correct in surmising the 'Quench and Squish Velocity'
issues, then this 'Might?' have caused the con rod to break.

More later when I have time. . . .

I have a broken Ultra I beam that met an unfortunate fate in an LSX 388.... Nothing is immune to failure
 

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I have a broken Ultra I beam that met an unfortunate fate in an LSX 388.... Nothing is immune to failure
Ooooohhhhh Noooooo, not again..lol

Let's just say my experience building many, many engines
has shown me that all con rods break, but the propensity
for H-Beams to break, similar to / as this one did is more
prevalent with H-Beams then with I-Beams.

If you differ with the above, then I will simply say
that you are entitled to your opinion.

------------------------------------------------

In my career I owned an Production Engine Facility
whereby we had 35 employees. We build 8 - 13 engines
per day. I oversaw all machine work and assembly.

In our performance division we built about 3-Performance
Engines a month. On top of that we maintained several
engines for our own cars.

The production engine business extended for over ten years.
The performance engine business extended for about 25-Years.

Total it all up and that is thousands of engines. . . .

The 'Propensity' for H-Beam Rods to beak (as did this one,
and please look at the picture to see what I mean) is far
higher than an I-Beam. . . . Of the same quality material.

In other words; I-Beams won't break in that manner, when
the con rod is made of the same material, and the cylinder
pressure on top of the piston is the same, when the crankshaft
and piston movement, given in degrees, is exactly at the same
position.


I went through this with engineers many years, ago who
ran cyclic tests on con rods, after I lost an engine when
a salesman sold me on H-Beam rods.

So, my experience tells me that all con rods can break.

But H-Beams break in a different manner then
do I-Beams. Also, where an I-Beam will bend, the
H-Beam might break. The only place on the con
rod that is stronger on an H-Beam, is up at the top
of the rod (small end) where the wrist pin is held.

Respectfully,
The Duck
 

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In for knowledge of two of the guys I respect the most.
 

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I never really thought this would happen again, but here we go. . . .
-Some posts from sometime back regarding this subject-

A connecting rod works in two domains:

1) Tension Loading
2) Compression Loading

A connecting rod’s max tension loads are determined by the mass of the parts involved, the rod length, the stroke length, and the max rpm.. . . Nothing More!

So then;
the max tension loads will never change, no matter what ‘Power Adder’ you add to your engine.

That max tension loading occurs at TDC on the exhaust stroke.
This has nothing what so ever to do with the amount of HP being made.

On the other hand; the connecting rods max compression loads are determined by
the amount of HP being made. It’s a simple matter; the higher the HP, the higher
the compression loading on the connecting rod will be.

So if the HP is increased, the compression loading will also be increased on the connecting rod.

The I-Beam rod design has about twice the strength in compression,
making the I-Beam Connecting Rod the best choice.


The above describes why an I-Beam will usually bend
when an H-Beam will break. . . . .

Cheers,
The Duck

-------------------------------------------------------------------

-and a later post on this subject-

By the way; the last I heard; it is mandatory for an F1 Engine to use I-Beam rods!

HP is what determines the compression loads on a rod and,
I-Beams support those loads better than do H-Beams.

Consider an H beam rod under a bending moment.
There are just 2 very skinny flanges that have to bear the weight.

There is not much skin on the surface of those 2 little skinny flanges
to reinforce the con rod in compression.

That is why an H beam is not used in many non automotive applications.


----------------------------------------------------------------

I decided to perform some additional research after your last post.

Consider the following as well;
The lighter, stronger and more efficient I-beam design is why Auto OEM, piston aircraft, large ship, trains,
and other industrial engines, as well as aircraft, spacecraft, and ship structural frames and of course,
bridges and large buildings, all use I-beams.

"If the H-Beam is better, then why haven't they been used in the above applications"?

-additional food for thought-
The highest cylinder pressures are just after top dead center. The piston has very little leverage at that point so the
connecting rod must take the load. The load increases as HP goes up! <= This partially describes what happened to George

By the time the piston has the greatest leverage ( 75 deg or so ATDC ), the cylinder pressure has been greatly reduced.

So as you can now see; the rod sees its highest 'Compression Load' right after TDC, while at peak torque rpm, not peak HP rpm.

Highest 'Tension Load' is 'Always' at the highest rpm at TDC on the exhaust stroke.

Now, go back up to where I cite all the engineering facts regarding various industries
that specify and use I-Beams.

The
Duck

-------------------------------------------------------------

And for those interested I also found this in my files from the past:

Some General Guidelines regarding Connecting Rod Lengths / Ratios:

1.40-1.60:1 is short
1.65-1.80 is nice for most applications
1.85-2.1+ is good for large bore short stroke, high rpm situations

--------------------------------------------

But as you increase the Rod Ratio, the Piston Dwell Periods at TDC
and BDC are not changed 'Symmetrically', but Asymmetrically.

That is:
As Con Rod Ratios 'Go Up' the Dwell Period of the piston is not 'Increased'
the same amount at TDC, as it is 'Increased' at BDC.

One now can easily imagine how the Inertia Effect of a very dense
Mass Charge would like an 'Optimum' length Con Rod and an intake valve
that closes just a 'Little' late..:eek:

--------------------------------------------

In general, regarding Con Rod Lengths and Ratios, it goes like this:

Rod Ratio Relationships;
Short Rod is slower at BDC range and faster at TDC range.
Long Rod is faster at BDC range and slower at TDC range.

On the Intake Stroke; the Long Rod will draw harder on cylinder head from 90* ATDC to BDC.

On the Intake Stroke; Short rod spends less time near TDC and will suck harder on the cylinder head from
10* ATDC to 90* ATDC the early part of the stroke.

But will not suck as hard from 90* ATDC to BDC as a long rod.

--------------------------------------------

Looking at the entire process;
1) Generate Sufficient Flow through the heads
2) Control Pressure Recovery in Cylinder
3) Maximize Flow at BDC via Inertia Filling.

Respecftully,
The Duck
 

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Just my opinion from what I have seen in George's video,
but here is another component (most probably) that can
contribute to such issues.

Also please highlight the info where I discussed rod angle
in degrees at TDC, whereby the con rod is almost vertical
when the charge is ignited.

Fast burn cylinder heads, such as the LS-Engine series
are in fact fast burn heads.


---------------------------------------------------------

Here we go with another post I made sometime back:
Some general info and considerations regarding quench & squish:

Squish velocity is determined by three factors:
piston to head clearance, squish area ratio _Piston area / _Squish area),
and piston speed. All three must be considered along with ignition timing.

Squish velocity is the speed at which air is displaced from between the piston
and head surface as the piston approaches TDC.

So again; it is the result of three factors: clearance, area and piston speed.

Inlet air temperature and mixture also play a role.

Considering the above, only clearance and squish ratio are fixed by design.

The other variables can be changed, or will
change as a result of engine operating conditions.

An engine with low squish velocity will not complete the combustion
process in the allotted time during high engine speed operation.

An engine with excessively high squish velocity will burn too rapidly,
causing combustion pressure rise at or before TDC. <= This also
happens or is exaggerated when using Nitrous.

---------------------------------------------------------------

Excessively high squish velocity can move the combustion pressure
curve too close to TDC, causing a loss of power and possible pre-ignition.

Some believe the above can happen if the 'squish clearance'
is set for less than 0.050" on an NA Engine.

Some also believe that anything less than 0.060"
of squish clearance will cause 'pumping losses'.

Others believe that anything over 0.035" of
squish clearance will cause the engine to
be down on power.

And some state that the 'minimum' squish clearance
should be set relative to the bore size.

They then calculate the required squish clearance stating
that the squish clearance should amount to 0.005"
of clearance per inch of bore diameter.

That would only be about 0.020" for a 4" bore.
The above info came from MIT years back!

-generally speaking-
When the gap / squish clearance is less than 0.050"
the flame is quenched and end gasses are left over.

When the gap / squish clearance is over 0.050"
the flame front can enter, not be quenched
and those gases are burnt. <= Flame front or
propagation here is important!

The downside to using a squish clearance of
somewhere between 0.050" to maybe 0.100"
is the following. . . . .

=>If the flame speed is to low, this allows for the combustion
end gases to heat up to auto-ignition temperatures causing
detonation. End gases aid in causing auto / self detonation.

This is why some (high FI pressure ratios & Nitrous applications)
convert the wedge head (hard head) to a soft head, by blowing
out the chamber, thereby ostensibly converting it to an open chamber.

And conversely;
As I stated above; excessively high squish velocity can
move the combustion pressure curve too close to TDC,
causing a loss of power and possible pre-ignition.


The above once again ties in with the 'almost' vertical con rod.

---------------------------------------------------------------

Now, let's look a look at this from another perspective that might help us understand;

Some years back we found that as we increased the percent
of 'volumetric efficiency' within the higher engine rpm range,
we could not run the fuel that NHRA demanded we use.

We were running a static compression ratio of 15.2:1.
But because the engine was so efficient, the 'trapped
compression' ratio exceeded ~17.2:1.

This is because the trapped compression, or as some
call it; the dynamic compression ratio is increased
in relationship to the amount of fuel the engine
was pulling in. Fuel takes up area and this then
reduces the area of the combustion chamber
when the piston is at TDC, thereby increasing
the 'trapped compression'.

So what do you suppose is happening with a
Supercharged Engine. . . . .

The required 'squish clearance' can be calculated.

I begin with the 0.035" of 'squish gap',which most
use in performance NA engines.

Then you simply factor the above via a simple
mathematical calculation utilizing a portion of
the density ratio.

Since you have contacted Matt (which I was going to recommend)
then the two of you can determine what will fill your needs together.

Happy New Year!
 

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I never really thought this would happen again, but here we go. . . .
-Some posts from sometime back regarding this subject-

A connecting rod works in two domains:

1) Tension Loading
2) Compression Loading

A connecting rod’s max tension loads are determined by the mass of the parts involved, the rod length, the stroke length, and the max rpm.. . . Nothing More!

So then;
the max tension loads will never change, no matter what ‘Power Adder’ you add to your engine.

That max tension loading occurs at TDC on the exhaust stroke.
This has nothing what so ever to do with the amount of HP being made.
ACKTUALLY!!

Consider a turbo motor that makes 40 psi of backpressure. Cylinder pressure never drops to zero while the exhaust valve is open. 40 psi of backpressure on a 4" bore is 502 lbf.
 

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The only real argument I have heard for H beams before (vs I beam) is lower cost and lighter rotating mass.

Lighter mass seems like it would only really make a big difference in high rpm applications but if you look at the Honda S2000 F20C OEM rod it is an I beam...and it ties the Ferrari 458 for highest redline in a production engine!

I have definitely seen a lot of broken H beams and bent I beams (and more rarely broken I beams but usually doing something crazy).
 

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The only real argument I have heard for H beams before (vs I beam) is lower cost and lighter rotating mass.

Lighter mass seems like it would only really make a big difference in high rpm applications but if you look at the Honda S2000 F20C OEM rod it is an I beam...and it ties the Ferrari 458 for highest redline in a production engine!

I have definitely seen a lot of broken H beams and bent I beams (and more rarely broken I beams but usually doing something crazy)
.
You got to be 'Crazy Nuts' to do what we all do with our rides..:D
 

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ACKTUALLY!!

Consider a turbo motor that makes 40 psi of backpressure. Cylinder pressure never drops to zero while the exhaust valve is open. 40 psi of backpressure on a 4" bore is 502 lbf.
Yes, but. . .
Does the back-pressure cancel out all, or just some of the forces involved.

Here is the piston motion and G-Forces for a 376 cid
engine spun to 7,000 rpm's.

Just for 'Kicks..lol
How would you imagine the G-Forces would change?

------- Piston Motion Data -------
Average Piston Speed. . . . . . . . . . (FPM)= 4225.67 in Feet Per Minute
Maximum Piston Speed. . . . . . . . . (FPM)= 6925.896 occurs at 74.65985 Degrees ATDC
Piston Depth at 74.660 degree . . . ATDC= 1.5873 inches Cylinder Volume= 337.6 CC
Maximum TDC Rod Tension. . . . . GForce= 3268.7850 G's
Maximum BDC Rod Compression. . GForce= 1772.1930 G's
 

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You said the power adder didn't matter. But it does change with a turbo with back pressure. I think this is why you see such great success with turbo's stock bottom end LS motors. That backpressure keeps the rod from seeing the same tension loads it would see in a similar NA or blown motor.

I’d like to see the numbers behind those rod tension gforce calculations. G is a unit of acceleration. 32.2ft/sec^2. To get force on the rod F=MA. Mass would be the piston, pin, rings and small end rod weight. Typically a little less than 2 lbs. F=2*(3268.75*32.2)=210,000lbf. Two 7/16 arp 2000 rod bolts are only good for 45,000lbf. total. That acceleration seems off to me, but I make mistakes in math.


I made a piston speed and acceleration spreadsheet back in my IC engines class in college. I’ll see if I can dig it up.

Edit: Maybe it's just supposed to be 3268ft/s^2 and the "G" is redundant. So F=2*(3268)=6536lb. In that case, 500lbf counteracting that makes it only 6000lb tension on the rod, an 8% decrease in load.
 

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What 1/4 mile numbers did you get prior to it letting go?

I just don’t hear of any results from the TT builds.


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Something theory mumbo jumbo to add to the Rod discussion.

Parts like rods, wrist pins, cranks can break from single time overloading or from cyclical loading over time. With the cyclical loading, you get a decrease in life depending on the load. It's typically shown by and SN curve (stress/force vs number of cycles). You can google one, but basically it says that when you cyclically load something above ~50% of it's tensile strength there is a finite life. The closer you get to the tensile strength the shorter the life. Under ~50% it could theoretically go forever. That ~50% of tensile strength is the fatigue life and varies based on material.

The physical way this happens is by crack growth. Some imperfection in or on the part starts a crack and it grows with each cycle. This is why we shot peen rods or check used cranks for cracks. So, you don't have to go above tensile strength to break a rod. But it does need to see significant loading and lots of cycles to break.

Why this matters on the turbo setup, is that difference between compressive and tensile stress is lower due to the backpressure keeping some compressive load on the rod. The average between the two is the number you use in the sn curve. Since it's lower, you get a logarithmically longer life.

We are all making wild guesses based on limited info here. CUIN9SEC has the whole story in front of him, we just get limited pictures from the video. It could have overheated the rings drug the piston pulling the wrist pin out of the piston, the pin could have seized in the piston put some bending load on it. There's lots of possiblities for failure besides just compressive and tensile loading. Failure analysis is hard because it was turning some rpm when it came apart and secondary damage probably hid most everything.

That's all the nerding I can handle for the day.
 

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So if the headers are touching the plug wires ...


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