Crank throw radius is just one-half the 3.48" stroke on a 350, and OE 350 rods are 5.7" center to center distance.

Duke

 

ZD57blue:

You are getting close. For the expression 5.7 + 1.74 * .7071 - 7.44, first multiply 1.74 times .7071, then add and subtract the rest of the terms. You will end up with a negative number, this just means the piston is below TDC.

Your TI calculator is giving you .525321 for the cosine of 45 degrees because the calculator is in the "radians" mode. Radians are just another way to define an angle. There should be a radians/degrees button on the calculator. Make sure it is in the degrees mode and you should be cooking.

 

In that case you would want to express that as:
5.7 + (1.74 * .7071) - 7.44 =
5.7 + (1.23035) - 7.44 = -0.5096

There appears to be something wrong with the formula. Aren't we solving for the length of the leg of the triangle from the crank main centerline to the piston pin centerline, at 45* ATCD? How are you getting the difference, in the pin to the crank main distance, at TDC and at 45*? My memory of trig has left me, and I have no books, so I'm afraid I can't help.

RACE ON!!!

 

CFI is right,
the angle of the connecting rod has to be considered.
since the crank angle is 45deg in this case both leg's of the first right angle are 1.23037" long. (that makes things easier in this case otherwise you need to figure out the other leg with: radius 1.74 X sin of angle).
then you figure another "right angle triangle" for the connecting rod's distance along centerline. where two side lengths are known (the rod length and the 1.23037" from before).
that formula is:
square root of (5.7" squared - 1.23037"squared)

which is 32.49 - 1.5138 = 30.9762/ square root of = 5.5656"
thats the length along centerline.
add that to the 1.23037 found before then subtract from the TDC measurment:
5.5656 + 1.23037 = 6.796"

7.44" - 6.796 = .644" movement down the cylinder bore for 45deg of rotation.

The way you calculate this stuff will change once you go past 90deg's of rotation.

 

I'm not sure, exactly how you got there, but I like the answer, better. We know that at 90* crank rotation, the piston will be at "half stroke", 1.74" down the hole. Due to the "laziness" at TDC and BDC, I would expect the 45* number to a little under half, the 90* number. Intuitively, .644" looks better.

Also, just food for thought; The piston movement will be the greatest per degree of crank rotation and hence the most efficient at transferring power, when the crank throw and the connecting rod are at, and near 90*, to one another.

RACE ON!!!

 

CFI-EFI, ol,RJ:

The formula you guys are looking at is the APPROXIMATE formula for a long rod case. For the exact formula, look further up in the posts! The exact formula takes into account the angle of the rod. The approximate formula for the long rod case does not. I was trying to help out ZD75blue with something that would get him in the ballpark and wouldn't give him brain freeze on the math.

 

RACE ON!!!

 

COOL!

I found the radian to degree button. That is SLICK! I didnt know you could do that!

So the negative number is ok? I dont need perfect, just something to get a ball park...

But the icing on the cake, if I can get this onto paper... is for me to be able to figure out what degree the valves are opening/closing so I can fool around with the Time they are open. Round about way of doing it... yes! But I think in %'s and pieces of wholes... and just seeing a couple different cam choices doesnt make much sense in my head. Gotta figure out what RPM's require how long for the valves to be open... and compare between the "Industry standard grinds" to see which way I'm going!

Thanks guys... Real cool stuff...

 

Gotcha... the long rod lazyness that makes more top end power in race motors...

So I need to figure in a "fudge factor?" For my ponderings?

 

ok, I see what your talking about.
clearly trying to hand type the formula's required doesn't work too well, a trig book is the best way to go.
either way will take alot of brain work, good luck.

 

Cris,

Please let us know if you find any significant differences when you get this project done...
I have yet to see anybody come up with a decent explanation for one rod length over the other...

 

No. Not if you have a good formula. It is the nature of the beast, that piston speeds (and therefore travel per degree of crank rotation) is less at the top and bottom of the stroke. As I mentioned, the piston movement is the greatest, per degree of rotation, when the crank throw and the con rod, form a 90* angle to one another. The long rod/short rod discussions center around the degree of laziness at the ends of the stokes. Regardless of rod length, the piston speed is the least at the ends of the stroke. After all, the piston is changing directions. It has to stop, and then accelerate.

RACE ON!!

 

The theoretical basis is combustion at constant volume, which yields the best theoretical thermal efficiency. The longer the rod the closer combustion is to constant volume, but the differences using normal r/l ratios (crank throw radius/rod center to center length) in the range of the mid 0.20s to low 0.30s, which is typical for commercial engines, is noise level. A 5.7" rod in a 350 yields and r/l ratio of 0.305; a 6.0" rod is 0.29. Big deal! On a 383 it's 0.33 and 0.31; a 302 is 0.26 and 0.25.

The whole rod length issue is WAYYYYYYYY overhyped.

Duke

 

Duke just hit on what sparked the whole idea...

Spark timing! The flame is starting nearly 40 degrees before the piston hits the top of the chamber. The longer you can wait and still fire it off, at least in my mind... the bigger the boom... With a question mark. Is that off base? I know the flame has to take time to burn... but if you get a complete burn do you make more power by starting the burn later?

I'm really liking taylors book, thanks for the recomendation Duke!

 

Didn't you just buy Taylor's two volume set?

Read the Vol I chapters on theoretical heat engine cycles - Carnot, etc., and combustion models - constant volume, constant pressure, etc.

The derivation of piston motion is in Vol II in the chapter on engine balance. You can set up a spread sheet with plug-in values for r and l and differentiate the piston displacement equation with respect to crank angle three times to get velocity, acceleration, and jerk and set these relationships up in a spread sheet.

Then all you have to do is input any r and l values, and you will get the results for every crank degree (or whatever increment you chose), in the blink of an eye.

Here are some other insights from Taylor. At any given speed and load, if spark timing is set for maximum torque (which should also yield maximum thermal efficiency/minimum fuel consumption) the combustion process is essentially concluded about the same number of degrees after TDC as BTC spark timing, and peak cylinder pressure will occur just after TDC.

Compact combustion chambers need less spark advance, so total crank rotation to complete combustion is less, which means both clock time for heat transfer is reduced and combustion is closer to the constant volume ideal. For example vintage SBs need about 38 degrees spark advance to achieve peak power and efficiency at WOT. My Cosworth Vega engine only needs 32 because it has a more compact combustion chamber and a centrally located spark plug, which means combustion is completed faster and closer to constant volume. This shows up as slightly higher thermal efficiency at the same compression ratio.

If you take this to an extreme you run into some practical limitations, primarily peak pressure and rate of pressure rise, which place increasing loads on the engine structure. Most diesel engine designs are up against these practical limits, so injection rate has to be slowed, and some designs even break up injection to multiple events (start-stop-start-stop) to limit peak pressure and rate of pressure rise to acceptable limits.

Duke

 

This is the reason for a tight quench area in a wedge design combustion chamber. The close contact between the piston and the head causes turbulence during combustion, that spreads the flame front around the combustion chamber more quickly than the relatively slow speed of the flame front travel. The quicker, more efficient burn, allows maximum power to be made with less spark advance. The later start and the faster burn produces less pressure on the raising piston, before TDC. The less amount of advance required to produce the maximum power, the more efficient the combustion precess. The guy that brags the he can advance his timing to excess numbers and make more power, is telling you something (if you recognize it) he didn't intend to.

I also agree on the false performance claims made for the "dwell" time of the piston at TDC with a long rod. I prefer to avoid a short rod, to reduce the side thrust of the piston into the cylinder wall.

RACE ON!!!

 

So the fancy way of saying it... if you have an engine that requires less overall spark advance you have a more thermally efficient engine. Making more "power" because of increased capture of the heat, by motion.

Dumb question...

How does a flat head do when it comes to flame travel?

 

Yes.

The flat top is the best. A dish isn't too bad as long as the quench area isn't compromised ("D" cup dish), but a dome will inhibit free flame travel.

RACE ON!!!

 

Are you talking flat top piston or flat head like a flat head V8 ford?

 

If you're talking about flathead engines like the old Ford V8 from the thirties, they have lousy combustion chamber geometry, which is why they are historical artifacts. The combustion chamber is very long, so it has long combustion time, which causes a high tendency to detonate (especially with the hot exhaust valve at the "end" of the combustion chamber that can easily act as a preignition point to start a second flame front, which will lead to a rapid pressure/temperature rise and initiate detonation) and a high surface area to volume ratio, which loses a lot of heat.

Compression ratio potential was poor due to lack of detonation resistance an high surface area to volume ratio. Production ratios were typically in the range of 6-8:1. A "high compression" hot rod flathead was maybe around 10:1.

The wedge chamber OHV V8 developed by Ed Cole and his group at GM in the late forties was a dramatic improvement in auto engine technology. That architecture endures today - in very highly refined form - in current GM OHV V8 engines, and it has been developed to a point where it is competitive in terms of specific output with modern pentroof 4V designs. More displacement can be packed into a given volume package and it will be both lighter, have higher average torque, and have higher fuel efficiency due to lower interal friction with only 5 cam bearings instead of 20.

Duke