connecting rod & rod length too stroke info



Re: bore to stroke ratio

Postby enigma57 » February 21st, 2011, 6:53 pm

:D Grumpy, your words are wise. I have a little time before I will be ready to assemble the short block and I have been giving plenty of thought to this aspect as well.

This being an engine for a road car...... With the 427 combo, I can give up some HP up top by re-gearing and keeping RPMs down, whilst still making gobs of torque in the low to mid range.

Back in 2003, a friend forwarded me a copy of a computer dyno program (Dyno 2003). Not sure how accurate these things are, but figured it would at least give some meaningful comparisons of changes in cam profile, so I ran calcs of various cam profiles having IVC that fell within a range of 8 degrees (ranging from 2 degrees earlier IVC than would result in 8.0:1 DCR to 6 degrees later IVC).

I ran cam profiles from one extreme to the other, from the very mild (and ancient) Isky 3/4 race E-4 solid flat tappet grind to the UltraDyne solid flat tappet short track cam I had Harold Brookshire grind me with wider 110 degree LSA some years back...... And several hydraulic and solid lifter cams in between.

By way of comparison, I plotted TQ and HP from 2,000 RPMs through 6,000 RPMs and also plotted average TQ and HP from 2,000 RPMs through 4,000 RPMs.

Geared as it is at present, my car will be turning 2,500 RPMs in 6th gear OD cruising at 70 MPH. HP and TQ averages for all cam grinds in the 2,000 - 4,000 RPM range were within 10 HP and 15 ft./lb. of one another...... Though peak TQ ranged from 3,500 RPMs to 4,500 RPMs and peak HP ranged from 5,000 RPMs to 6,000 RPMs, depending upon cam grind. (Given the use of the car, I have placed a self-imposed 6,000 RPM redline on this 4" stoke combo to limit piston speed to 4,000 fpm max at redline as you suggested some time ago.)

What changed most was where the engine made the most power. As expected, the milder cams made more power down low.

Interestingly, all were within 11 HP and 15ft./lb. of TQ at 3,500 RPMs. And as expected, the more aggressive grinds began pulling away from the milder cams which had made more power below 3,500. And from 3,500 RPMs to their respective redlines, each of the more aggressive grinds made up the difference between their lack of power at lower revs...... The averages from 2,000 - 4,000 RPMs being very close.

From 4,500 RPMs to their respective redlines, the more aggressive grinds pulled harder and the higher they spun, the more HP they made...... Whilst shifting TQ peak higher, but with ft./lbs. remaining very close to the numbers achieved at lower revs with the milder grinds. All in all, it was a very interesting comparison.

Here are some examples, beginning with the little Isky E-4 solid lifter grind from the '50s that makes the 427 into a low RPM stump puller......


Solid flat tappet......

------RPMs----- HP ---TQ

2,000 RPMs - 189 - 497 - Cam, Isky E-4 solid, flat tappet, 216 deg. int./216 deg. exh. @ 0.050", 108 deg. LSA, 0.425" lift at valve (both)
2,500 RPMs - 242 - 508
3,000 RPMs - 296 - 519 - Peak HP - 421 @ 5,000 RPMs
3,500 RPMs - 349 - 523 - Peak TQ - 523 ft./lb. @ 3,500 RPMs
4,000 RPMs - 391 - 513
4,500 RPMs - 421 - 492
5,000 RPMs - 421 - 442 - Avg. HP - 2,000 to 4,000 RPMs, 293.4
5,500 RPMs - 408 - 389 - Avg. TQ - 2,000 to 4,000 RPMs, 512 ft./lb.
6,000 RPMs - 376 - 329

------RPMs----- HP ---TQ

2,000 RPMs - 182 - 479 - Cam, Isky RPM-300 solid, flat tappet, 228 deg. int./228 deg. exh. @ 0.050", ground on 108 deg. (not 112 deg.) LSA, 0.448" lift at valve (both)
2,500 RPMs - 235 - 493
3,000 RPMs - 292 - 511 - Peak HP - 458 @ 5,000 RPMs
3,500 RPMs - 349 - 524 - Peak TQ - 527 ft./lb. @ 4,000 RPMs
4,000 RPMs - 402 - 527
4,500 RPMs - 439 - 513
5,000 RPMs - 458 - 481 - Avg. HP - 2,000 to 4,000 RPMs, 292
5,500 RPMs - 450 - 429 - Avg. TQ - 2,000 to 4,000 RPMs, 506.8 ft./lb.
6,000 RPMs - 427 - 374

------RPMs----- HP ---TQ

2,000 RPMs - 184 - 483 - Cam, Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve
2,500 RPMs - 239 - 501
3,000 RPMs - 297 - 520 - Peak HP - 517 @ 5,500 RPMs
3,500 RPMs - 359 - 538 - Peak TQ - 548 ft./lb. @ 4,500 RPMs
4,000 RPMs - 417 - 548
4,500 RPMs - 470 - 548
5,000 RPMs - 505 - 530 - Avg. HP - 2,000 to 4,000 RPMs, 299.2
5,500 RPMs - 517 - 493 - Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.
6,000 RPMs - 509 - 446

------RPMs----- HP ---TQ

2,000 RPMs - 170 - 446 - Cam, UltraDyne solid, flat tappet, 252 deg. int./252 deg. exh. @ 0.050", 110 deg. LSA, 0.525" lift at valve (both)
2,500 RPMs - 223 - 468
3,000 RPMs - 282 - 493 - Peak HP - 546 @ 6,000 RPMs
3,500 RPMs - 348 - 523 - Peak TQ - 555 ft./lb. @ 4,500 RPMs
4,000 RPMs - 413 - 543
4,500 RPMs - 475 - 555
5,000 RPMs - 519 - 545 - Avg. HP - 2,000 to 4,000 RPMs, 287.2
5,500 RPMs - 546 - 522 - Avg. TQ - 2,000 to 4,000 RPMs, 494.6 ft./lb.
6,000 RPMs - 546 - 478


Hydraulic flat tappet......

------RPMs----- HP ---TQ

2,000 RPMs - 176 - 463 - Cam, Comp Cams #CS NX268H-13 ground on 108 LSA, hydraulic, flat tappet, 224 deg. int./236 deg. exh. @ 0.050", 108 deg. LSA, 0.477" int./0.490" exh. lift at valve)
2,500 RPMs - 230 - 483
3,000 RPMs - 285 - 500 - Peak HP - 497 @ 5,500 RPMs
3,500 RPMs - 345 - 518 - Peak TQ - 522 ft./lb. @ 4,500 RPMs
4,000 RPMs - 402 - 527
4,500 RPMs - 447 - 522
5,000 RPMs - 484 - 508 - Avg. HP - 2,000 to 4,000 RPMs, 287.6
5,500 RPMs - 497 - 474 - Avg. TQ - 2,000 to 4,000 RPMs, 498.2 ft./lb.
6,000 RPMs - 486 - 425


------RPMs----- HP ---TQ

2,000 RPMs - 183 - 481 - Cam, Lunati 268 VooDoo hydraulic, flat tappet, 227 deg. int./233 deg. exh. @ 0.050", 110 deg. LSA, 0.489" int./0.504" exh. lift at valve)
2,500 RPMs - 237 - 497
3,000 RPMs - 294 - 515 - Peak HP - 479 @ 5,500 RPMs
3,500 RPMs - 355 - 532 - Peak TQ - 537 ft./lb. @ 4,000 RPMs
4,000 RPMs - 409 - 537
4,500 RPMs - 455 - 531
5,000 RPMs - 478 - 502 - Avg. HP - 2,000 to 4,000 RPMs, 295.6
5,500 RPMs - 479 - 457 - Avg. TQ - 2,000 to 4,000 RPMs, 512.4 ft./lb.
6,000 RPMs - 455 - 399

------RPMs----- HP ---TQ

2,000 RPMs - 187 - 491 - Cam, Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)
2,500 RPMs - 243 - 510 - Note: 1.6 rockers raise peaks by 4 - 5 HP & ft./lb. TQ, but power at lower RPMs remains the same
3,000 RPMs - 302 - 528 - Peak HP - 478 @ 5,000 RPMs
3,500 RPMs - 362 - 543 - Peak TQ - 547 ft./lb. @ 4,000 RPMs
4,000 RPMs - 417 - 547
4,500 RPMs - 458 - 535
5,000 RPMs - 478 - 502 - Avg. HP - 2,000 to 4,000 RPMs, 302.2
5,500 RPMs - 468 - 447 - Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.
6,000 RPMs - 446 - 390

I believe you are right. Better to go ahead and build the larger displacement version. The key to keeping this long stroke engine together and keep it running for many, many miles will be to gear the car to keep RPMs down and build the engine for low and mid range torque.

Of the cams I ran through the desk top dyno, these three came out on top. I have ranked them below based upon how each did when comparing average TQ and HP produced between 2,000 and 4,000 RPMs (the engine speeds this engine will see the most out on the highway). Surprisingly, the little Isky 274 Mega short track hydraulic flat tappet grind came out on top with my engine combo....... Followed by the Isky 530-A short track solid flat tappet grind and the Lunati 268 VooDoo hydraulic flat tappet street cam......

Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)

Peak HP - 478 @ 5,000 RPMs
Peak TQ - 547 ft./lb. @ 4,000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 302.2
Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.


Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve

Peak HP - 517 @ 5,500 RPMs
Peak TQ - 548 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 299.2
Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.


Lunati 268 VooDoo hydraulic, flat tappet, 227 deg. int./233 deg. exh. @ 0.050", 110 deg. LSA, 0.489" int./0.504" exh. lift at valve

Peak HP - 479 @ 5,500 RPMs
Peak TQ - 537 ft./lb. @ 4,000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 295.6
Avg. TQ - 2,000 to 4,000 RPMs, 512.4 ft./lb.

* The Isky 530-A solid lifter grind made the most HP at peak and nearly the same TQ at peak as the Isky 274 Mega hydraulic.

* The Lunati 268 VooDoo hydraulic made nearly the same HP at peak as the Isky 274 Mega hydraulic.

* What I like about the 274 Mega is that it not only made more average power where I will need it most out on the road...... It made peak TQ and HP numbers very near to the others, but at 500 fewer RPMs.

What do you think?

Best regards,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 22nd, 2011, 5:39 am

the isky looks good but Id be looking for a bit more duration and a bit tighter LSA,all your selections will work with obvious various power curve changes, but Id concentrate on maximizing the 3500rpm-5800rpm torque band
try these cams with 1.6 rockers, with that bore stroke and displacement and knowing what your intended use is Id be looking to find a cam with about a 105-107 lsa.236/242-to-239-245 duration and about a .490-550 lift to make a good compromise between power and durability while maintaining low valve train stress
keep in mind that longer stroke and large displacement will react well to a tight lsa and a good deal of overlap but youll want to select the duration to maximize the mid and upper rpm power as the low rpm power should not be much of an issue, yet limit the duration to maintain at least decent low rpm drive ability,remember there's a big difference between a N/A cam used in a naturally aspirated engine and a super charger or nitrous design cam,that would have a wider lsa.
as always why not talk to the cam manufacturer and get their input and talk to them about gearing, rpm power bands etc.and your better off going with a slightly milder cam than over camming the combo, but a 427 sbc needs to breath and it takes some duration and overlap to do that efficiently,keeping in mind your goals, Id suggest and air gap dual plane intake, 750-800cfm cfm carb and long tube 1 3/4 headers
your not building a 350 so you'll need a good deal more cam, than a similar 350 build would require, but don,t get crazy, seeking peak numbers, its maximizing BOTH average torque and peak hp that matters most, don,t forget its the headers,heads, and intake that will also have a noticeable effect on the combo results.
and it only takes about 60hp or less to cruise, at 70mph at part throttle, and you don,t need off idle tire smoking torque, but you do need impressive acceleration from about 3500rpm up to 6000rpm

http://www.crower.com/products/camshaft ... -3956.html

http://www.crower.com/products/camshaft ... -3967.html


BTW
http://www.crower.com/cam-card-finder/

click, enter part number


related info
viewtopic.php?f=52&t=1070

viewtopic.php?f=52&t=480

viewtopic.php?f=55&t=624&p=11125&hilit=port+sectional#p11125

BTW the DD DYNO software tends to favor the wider LSA selection more than real world experience suggests is valid in my experience and over state the low rpm tq, but the peak hp numbers are usually reasonably close
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
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Re: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 10:34 am

Effects of a longer Rod
* Less rod angularity reduces wear.
* Lower piston velocity and acceleration reduces tensile loading of the rods.
* Less ignition timing is required which resist detonation.
* Compression can be increased slightly before detonation is a problem.
* Less intake runner volume is required and high rpm breathing is improved.
* Reduces scavenging at low rpm (weaker low RPM power).
* Longer TDC dwell time. (high RPM efficency).


Effects of a shorter Rod
* Increased rod angularity increases wear.
* Increased piston velocity and acceleration increases tensile loading of the rods.
* Increases scavenging at low rpm (increased low RPM power).
* Reduced TDC dwell time. (Reduced high RPM efficiency).


What they forgot to mention about the long rod is that the positive gain at TDC, is partially offset by wasted dwell time at BDC. The other problem is that on some piston designs, is the close proximity of the wrist pin hole and the bottom oil ring rail, on the piston, can cause flexing in the ring and less effective oil control.
below They used a 3.5" stroke for both rods, which is very close to a 350's 3.48" stroke.
As the graph shows, even 2" longer rod does not perform miracles, the better rod ratio has some effect but nothing thats going to totally make or break a combos effective power curve provided you keep the piston speeds well within the component strength limitations.
as stated before with basically stock components 4000fpm in piston speeds a good compromise, and with all forged and balanced components 4200fpm is doable, even 4500fpm for a second or two can be tolerated in some combos but stress goes up rapidly as rpms increase and stress is cumulative.
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IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
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Re: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 11:03 am

http://www.iskycams.com/techinfo_index.html

Rod Lengths/Ratios: Much ado about almost nothing.

Why do people change connecting rod lengths or alter their rod length to stroke ratios? I know why, they think they are changing them. They expect to gain (usually based upon the hype of some magazine article or the sales pitch of someone in the parts business) Torque or Horsepower here or there in rather significant "chunks". Well, they will experience some gains and losses here or there in torque and or H.P., but unfortunately these "chunks" everyone talks about are more like "chips".

To hear the hype about running a longer Rod and making more Torque @ low to mid RPM or mid to high RPM (yes, it is, believe it or not actually pitched both ways) you'd think that there must be a tremendous potential for gain, otherwise, why would anyone even bother? Good question. Let's begin with the basics. The manufacture's (Chevy, Ford, Chrysler etc.) employ automotive engineers and designers to do their best (especially today) in creating engine packages that are both powerful and efficient. They of course, must also consider longevity, for what good would come form designing an engine with say 5% more power at a price of one half the life factor? Obviously none. You usually don't get something for nothing - everything usually has its price. For example: I can design a cam with tremendous high RPM/H.P. potential, but it would be silly of me (not to mention the height of arrogance) to criticize the engineer who designed the stock camshaft. For this engine when I know how poorly this cam would perform at the lower operating RPM range in which this engineer was concerned with as his design objective!

Yet, I read of and hear about people who do this all the time with Rod lengths. They actually speak of the automotive engine designer responsible for running "such a short Rod" as a "stupid SOB." Well, folks I am here to tell you that those who spew such garbage should be ashamed of themselves - and not just because the original designer had different design criteria and objectives. I may shock some of you, but in your wildest dreams you are never going to achieve the level of power increase by changing your connecting rod lengths that you would, say in increasing compression ratio, cam duration or cylinder head flow capacity. To illustrate my point, take a look at the chart below. I have illustrated the crank angles and relative piston positions of today's most popular racing engine, the 3.48" stroke small block 350 V8 Chevy in standard 5.7", 6.00", 6.125" and 6.250" long rod lengths in 5 degree increments. Notice the infinitesimal (look it up in the dictionary) change in piston position for a given crank angle with the 4 different length rods. Not much here folks, but "oh, there must be a big difference in piston velocity, right?" Wrong! Again it's a marginal difference (check the source yourself - its performance calculator).

To hear all this hype about rod lengths I'm sure you were prepared for a nice 30, 40, or 50 HP increase, weren't you? Well its more like a 5-7 HP increase at best, and guess what? It comes at a price. The longer the rod, the closer your wrist pin boss will be to your ring lands. In extreme situations, 6.125" & 6.250" lengths for example, both ring and piston life are affected. The rings get a double whammy affect. First, with the pin boss crowding the rings, the normally designed space between the lands must be reduced to accommodate the higher wrist pin boss. Second, the rings wobble more and lose the seal of their fine edge as the piston rocks. A longer Rod influences the piston to dwell a bit longer at TDC than a shorter rod would and conversely, to dwell somewhat less at BDC. This is another area where people often get the information backwards.

In fact, this may surprise you, but I know of a gentleman who runs a 5.5" Rod in a 350 Small Block Chevy who makes more horsepower (we're talking top end here) than he would with a longer rod. Why? Because with a longer dwell time at BDC the short rod will actually allow you a slightly later intake closing point (about 1 or 2 degrees) in terms of crank angle, with the same piston rise in the cylinder. So in terms of the engines sensitivity to "reversion" with the shorter rod lengths you can run about 2-4 degrees more duration (1-2 degrees on both the opening & closing sides) without suffering this adverse affect! So much for the belief that longer rod's always enhance top end power!

Now to the subject of rod to stroke ratios. People are always looking for the "magic number" here - as if like Pythagoras they could possibly discover a mathematical relationship which would secure them a place in history. Rod to stroke ratios are for the most part the naturally occurring result of other engine design criteria. In other-words, much like with ignition timing (spark advance) they are what they are. In regards to the later, the actual number is not as important as finding the right point for a given engine. Why worry for example that a Chrysler "hemi" needs less spark advance that a Chevrolet "wedge" combustion chamber? The number in and of itself is not important and it is much the same with rod to stroke ratios. Unless you want to completely redesign the engine (including your block deck height etc.) leave your rod lengths alone. Let's not forget after all, most of us are not racing at the Indy 500 but rather are hot rodding stock blocks.

Only professional engine builders who have exhausted every other possible avenue of performance should ever consider a rod length change and even they should exercise care so as not to get caught up in the hype.

if you do a bunch of research, or connecting rod length to stroke ratios and piston speeds etc., youll find a general consensus that the ideal rod ratio is near 1.9:1 on the stroke length,so something like a 3' or 3.25' stroke with a 6" connecting rods near the ideal , i also have built enough engines like a 406 chevy with a 5.7' rod or a 540 big block CHEVY with 6.385" rods that have closer to a 1.5:1 ratio that made exceptionally good power,to know that I read thru this thread and many other similar threads about rod to stroke ratios, that the phrase's

"TEMPEST IN A TEA POT"
"MOUNTAIN OUT OF A MOLE HILL"

BOTH COME TO MIND!"
yes theres measurable advantages to be gained in well matched component selection,and careful assembly, but I think the degree of concern here seems a bit excessive compared to the potential gains you might expect from the difference in parts being selected, compared to several other areas that potentially effect engine power that you might be concerned with.
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
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Re: bore to stroke ratio

Postby enigma57 » February 23rd, 2011, 12:31 pm

:D Thanks for the cam recommendations and timing card location info, Grumpy! I ran the Crower cams through Dyno 2003 with my 427 engine combo specs. Used valve lift with 1.6 rockers as suggested.

Of the 2 Crower grinds, my engine combo favours the larger of the 2 cams (00273S). Here are the figures I got.......


------RPMs----- HP ---TQ

2,000 RPMs - 178 - 466 - Cam, Crower 00258S hydraulic, flat tappet, 236 deg. int./242 deg. exh. @ 0.050", 106 deg. LSA, 0.479" int. / 0.475" exh. lift at valve
2,500 RPMs - 228 - 479 - Note: lift shown is with 1.6 rockers
3,000 RPMs - 282 - 493 - Peak HP - 425 @ 5,000 RPMs
3,500 RPMs - 338 - 508 - Peak TQ - 508 ft./lb. @ 3,5000 RPMs
4,000 RPMs - 387 - 508
4,500 RPMs - 416 - 486
5,000 RPMs - 425 - 446 - Avg. HP - 2,000 to 4,000 RPMs, 282.6
5,500 RPMs - 419 - 400 - Avg. TQ - 2,000 to 4,000 RPMs, 490.8 ft./lb.
6,000 RPMs - 395 - 346

------RPMs----- HP ---TQ

2,000 RPMs - 170 - 447 - Cam, Crower 00273S hydraulic, flat tappet, 239 deg. int./248 deg. exh. @ 0.050", 107 deg. LSA, 0.539" int. / 0.547" exh. lift at valve
2,500 RPMs - 225 - 473 - Note: lift shown is with 1.6 rockers
3,000 RPMs - 285 - 499 - Peak HP - 510 @ 5,500 RPMs
3,500 RPMs - 352 - 528 - Peak TQ - 546 ft./lb. @ 4,500 RPMs
4,000 RPMs - 415 - 544
4,500 RPMs - 468 - 546
5,000 RPMs - 496 - 521 - Avg. HP - 2,000 to 4,000 RPMs, 289.4
5,500 RPMs - 510 - 487 - Avg. TQ - 2,000 to 4,000 RPMs, 498.2 ft./lb.
6,000 RPMs - 502 - 439


Also ran the 2 Crane solid lifter cams mentioned on your link here......

viewtopic.php?f=52&t=1070


------RPMs----- HP ---TQ

2,000 RPMs - 163 - 427 - Cam, Crane 110921 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 106 deg. LSA, 0.518" int. / 0.536" exh. lift at valve
2,500 RPMs - 217 - 456 - Note: 1.6 rockers raise peaks by 4 - 5 HP & ft./lb. TQ, but power at lower RPMs remains the same (used 1.5 rockers for simulation)
3,000 RPMs - 275 - 482 - Peak HP - 509 @ 5,500 RPMs
3,500 RPMs - 343 - 514 - Peak TQ - 539 ft./lb. @ 4,500 RPMs
4,000 RPMs - 407 - 535
4,500 RPMs - 462 - 539
5,000 RPMs - 492 - 517 - Avg. HP - 2,000 to 4,000 RPMs, 281
5,500 RPMs - 509 - 486 - Avg. TQ - 2,000 to 4,000 RPMs, 482.8 ft./lb.
6,000 RPMs - 504 - 441


------RPMs----- HP ---TQ

2,000 RPMs - 166 - 435 - Crane 114681 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 112 deg. LSA (per DD2003), 0.518" int. / 0.536" exh. lift at valve
2,500 RPMs - 220 - 463 - Note: 1.5 rockers used for simulation
3,000 RPMs - 279 - 488 - Peak HP - 511 @ 5,500 RPMs
3,500 RPMs - 345 - 518 - Peak TQ - 546 ft./lb. @ 4,500 RPMs
4,000 RPMs - 413 - 542
4,500 RPMs - 468 - 546
5,000 RPMs - 500 - 525 - Avg. HP - 2,000 to 4,000 RPMs, 284.6
5,500 RPMs - 511 - 528 - Avg. TQ - 2,000 to 4,000 RPMs, 489.2 ft./lb.
6,000 RPMs - 504 - 441


And here are the figures I got for the 2 Isky cams that came out best in prior calculations.....


------RPMs----- HP ---TQ

2,000 RPMs - 187 - 491 - Cam, Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)
2,500 RPMs - 243 - 510 - Note: 1.6 rockers raise peaks by 4 - 5 HP & ft./lb. TQ, but power at lower RPMs remains the same (used 1.5 rockers for simulation)
3,000 RPMs - 302 - 528 - Peak HP - 478 @ 5,000 RPMs
3,500 RPMs - 362 - 543 - Peak TQ - 547 ft./lb. @ 4,000 RPMs
4,000 RPMs - 417 - 547
4,500 RPMs - 458 - 535
5,000 RPMs - 478 - 502 - Avg. HP - 2,000 to 4,000 RPMs, 302.2
5,500 RPMs - 468 - 447 - Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.
6,000 RPMs - 446 - 390

------RPMs----- HP ---TQ

2,000 RPMs - 184 - 483 - Cam, Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve
2,500 RPMs - 239 - 501 - Note: 1.5 rockers used for simulation
3,000 RPMs - 297 - 520 - Peak HP - 517 @ 5,500 RPMs
3,500 RPMs - 359 - 538 - Peak TQ - 548 ft./lb. @ 4,500 RPMs
4,000 RPMs - 417 - 548
4,500 RPMs - 470 - 548
5,000 RPMs - 505 - 530 - Avg. HP - 2,000 to 4,000 RPMs, 299.2
5,500 RPMs - 517 - 493 - Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.
6,000 RPMs - 509 - 446


The Isky cams were run with 1.5 rockers, as I tried 1.6 on intake only, then 1.6 on exhaust only and then 1.6 for both. Power remained the same over the lower RPMs and showed only a small increase (4 - 5 HP & ft./lb. TQ) at peak TQ and HP with 1.6 rockers on both.

The Crower cams were run with valve lift calculated for 1.6 rockers as suggested.

The Crane cams were run with 1.5 rockers, as I tried 1.6 on intake only, then 1.6 on exhaust only and then 1.6 for both. Power remained the same over the lower RPMs and showed only a small increase (4 - 5 HP & ft./lb. TQ) at peak TQ and HP.

I have ranked the 2 Isky, the 2 Crower and the 2 Crane cams below by how they came out based upon average HP and TQ between 2,000 and 4,000 RPMs (the engine speeds my engine will see most out on the road)......


Isky 274 Mega hydraulic, flat tappet, 226 deg. int./226 deg. exh. @ 0.050", 108 deg. LSA, 0.490" lift at valve (both)

Peak HP - 478 @ 5,000 RPMs
Peak TQ - 547 ft./lb. @ 4,000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 302.2
Avg. TQ - 2,000 to 4,000 RPMs, 523.8 ft./lb.


Isky 530-A solid, flat tappet, 242 deg. int./246 deg. exh. @ 0.050", 106 deg. LSA, 0.530" int./ 0.535" lift at valve

Peak HP - 517 @ 5,500 RPMs
Peak TQ - 548 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 299.2
Avg. TQ - 2,000 to 4,000 RPMs, 518 ft./lb.


Crower 00273S hydraulic, flat tappet, 239 deg. int./248 deg. exh. @ 0.050", 107 deg. LSA, 0.539" int./0.547" exh. lift at valve

Peak HP - 510 @ 5,500 RPMs
Peak TQ - 546 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 289.4
Avg. TQ - 2,000 to 4,000 RPMs, 498.2 ft./lb.


Crower 00258S hydraulic, flat tappet, 236 deg. int./242 deg. exh. @ 0.050", 106 deg. LSA, 0.479" int./0.475" exh. lift at valve

Peak HP - 425 @ 5,000 RPMs
Peak TQ - 508 ft./lb. @ 3,5000 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 282.6
Avg. TQ - 2,000 to 4,000 RPMs, 490.8 ft./lb.


Crane 114681 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 112 deg. LSA (per DD2003), 0.518" int. / 0.536" exh. lift at valve

Peak HP - 511 @ 5,500 RPMs
Peak TQ - 546 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 284.6
Avg. TQ - 2,000 to 4,000 RPMs, 489.2 ft./lb.


Crane 110921 solid, flat tappet, 244 deg. int./252 deg. exh. @ 0.050", 106 deg. LSA, 0.518" int. / 0.536" exh. lift at valve

Peak HP - 509 @ 5,500 RPMs
Peak TQ - 539 ft./lb. @ 4,500 RPMs

Avg. HP - 2,000 to 4,000 RPMs, 281
Avg. TQ - 2,000 to 4,000 RPMs, 482.8 ft./lb.


As you predicted, of the 2 Crane cams, the Dyno 2003 program favoured the wider LSA version of the same cam.

Not sure why the Isky cam grinds made more comparative power than did the Crower and Crane grinds (in the RPM range my engine will be run at most), but my engine combo seems to respond better to the Isky cams, at least in computer simulation. And as you say, real world performance may vary.

Really appreciate your assistance, Grumpy!

Best regards,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 12:51 pm

I,d doubt you'll be disappointed with any of the better selections, but if you do get the car dyno tested after its tuned it will be interesting to compare the real world results with the dd dyno predicted wild guess the software shows
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Re: bore to stroke ratio

Postby enigma57 » February 23rd, 2011, 3:34 pm

:D I agree, Grumpy! It will be very interesting to see how this works out in terms of real world TQ and HP at the planned operating range. FWIW...... I plan on using this engine as a test bed to see how several intake systems work, including a home brewed IR intake. It will be a while before the car is ready to go, but I will keep in touch and let you know how things are progressing.

There may be some compromises required when fabbing the exhaust system because I do not want the hassle of header leaks on a road car with full length, muffled exhaust. Nor do I want to fry starters nor have engine oil temps increase due to oil filter and header clearance issues.

On the other hand...... As I have made the decision to build this engine for torque and keep RPMs down...... I may be able to get the sizing I need in heavy wall tubing and fit it as I have in mind after all. At present, I am thinking in terms of welded thickwall upswept headers similar in function to the old MOPAR Max Wedge cast headers of the early '60s. These were actually a 4 - 2 - 1 Tri-Y design and I believe something similar would enhance torque all the more.

Best regards,

Harry

BTW...... I am thinking about going to a standard volume big block oil pump. I understand from speaking to several engine builders that Melling and others they make pumps for to sell under their brand name have gone through several production changes in recent years. First, they began utilizing thinner pump body castings sourced from the ChiComs with the result that a high percentage of pumps used in performance and heavy duty applications developed cracks in the pump bodies.

Then they tried to get by with powdered metal pump gears...... Which also didn't work out well, either. I understand that they are working their way back to building pumps as they once did...... But that there are still pumps in warehouses that have the thin body castings and / or the powdered metal gears. My dilemma is where to get a really good and reliable oil pump and shaft that will last the life of this engine. Any ideas?
BOHICA! - How's all that 'change' workin' for ya? Image
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Re: bore to stroke ratio

Postby grumpyvette » February 23rd, 2011, 4:11 pm

http://www.moroso.com/catalog/categoryd ... code=17006

http://www.melling.com/Aftermarket/High ... Pumps.aspx
watch this video
http://www.summitracing.com/parts/MEL-10778/?rtype=4

the melling 10990 is generally a well built pump but the price has gone up significantly since they started building them correctly again.
its basically a big block 5 bolt pump set up to run in a small block application
Ive run them in most of my sbc engines built recently and its in my corvettes 383

Image

http://www.summitracing.com/parts/MEL-10990/

btw

10778 used for big block engines

http://www.summitracing.com/parts/MEL-10778/
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
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Re: bore to stroke ratio

Postby enigma57 » February 27th, 2011, 12:48 am

:D Thanks, Grumpy! That's good info. With the 383 small block applications you build, do you keep the spring that comes in the 10990 pump or replace it with the supplied standard pressure spring?

And have you used the 10990C (version with the anti-cavitation grooves)? On a mostly street driven application, I am a bit leery of the stated 'reduces pressure at idle'.

http://www.melling.com/Aftermarket/High ... Pumps.aspx

10990

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The 10990 is a Big Block style oil pump made to fit the Small Block applications.
* The drive and idler shafts have been extended to allow for additional support in the cover. Additional support eliminates dynamic shaft deflection at increased RPM levels.
* The cover is doweled to the pump housing to assure alignment of the shaft bores.
* The relief valve has a screw-in plug instead of a pin.
* The housing and cover are CNC machined and phosphate coated.
* An additional spring, the original stock replacement is supplied which will reduce bypass pressure if needed.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Patent No. 5,810,571.

10990C (Anti-Cavitation)

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The same as the 10990 except with the addition of grooves machined in the housing and cover. The grooves reduce cavitation effects in high RPM applications.
* Using this oil pump will reduce pressure at idle.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Racing applications only.
* Patent No. 5,810,571.


I have a 7 quart Moroso #20190 pan with kickouts on both sides. It is 8-1/4" deep......

http://www.moroso.com/catalog/categoryd ... code=11002

In the notes, Moroso oil pickup #24320 is spec'd for big block oil pumps, so am I right in assuming that this would be the correct pickup to use with the Melling 10990 pump and this pan?

I see that Moroso now has pans notched for strokes over 3.80" as well. I will hold off until I have clearanced my block, trial fitting the crank and pistons and oil pan to see if my #20190 pan will work. But if not, I will likely go to the #20195 pan that has the same dimensions, but is notched for longer stroke crank throws.

Best regards,

Harry
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Re: bore to stroke ratio

Postby grumpyvette » February 27th, 2011, 8:33 am

enigma57 wrote::D Thanks, Grumpy! That's good info. With the 383 small block applications you build, do you keep the spring that comes in the 10990 pump or replace it with the supplied standard pressure spring?
I use the lower pressure spring

And have you used the 10990C (version with the anti-cavitation grooves)? On a mostly street driven application, I am a bit leery of the stated 'reduces pressure at idle'.

http://www.melling.com/Aftermarket/High ... Pumps.aspx

10990

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The 10990 is a Big Block style oil pump made to fit the Small Block applications.
* The drive and idler shafts have been extended to allow for additional support in the cover. Additional support eliminates dynamic shaft deflection at increased RPM levels.
* The cover is doweled to the pump housing to assure alignment of the shaft bores.
* The relief valve has a screw-in plug instead of a pin.
* The housing and cover are CNC machined and phosphate coated.
* An additional spring, the original stock replacement is supplied which will reduce bypass pressure if needed.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Patent No. 5,810,571.

10990C (Anti-Cavitation)

* High volume performance upgrade for the M-99HV-S.
* Increase in volume of 25% over stock oil pump.
* The same as the 10990 except with the addition of grooves machined in the housing and cover. The grooves reduce cavitation effects in high RPM applications.
* Using this oil pump will reduce pressure at idle.
* Includes intermediate shaft with steel guide.
* Uses 3/4” press in screen.
* Racing applications only.
* Patent No. 5,810,571.


I have a 7 quart Moroso #20190 pan with kickouts on both sides. It is 8-1/4" deep......

http://www.moroso.com/catalog/categoryd ... code=11002

In the notes, Moroso oil pickup #24320 is spec'd for big block oil pumps, so am I right in assuming that this would be the correct pickup to use with the Melling 10990 pump and this pan?

ALWAYS VERIFY WHAT COMPONENTS ARE TO BE MATCHED IN ANY APPLICATION ,WITH THE MANUFACTURERS ,and ALWAYS ASK IF THEY HAVE OTHER RELATED TIPS,INFO, or RELATED WARNINGS but yes your more than likely correct but as Ive stated ..ALWAYS ASK[/b]

I see that Moroso now has pans notched for strokes over 3.80" as well. I will hold off until I have clearanced my block, trial fitting the crank and pistons and oil pan to see if my #20190 pan will work. But if not, I will likely go to the #20195 pan that has the same dimensions, but is notched for longer stroke crank throws.

Best regards,

Harry


hope that helps
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
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Re: bore to stroke ratio

Postby enigma57 » March 3rd, 2011, 7:31 pm

Yes, thanks, Grumpy! This helps a lot. I figured the big block style pump with the lower (normal) pressure spring would be the way to go and I will verify the info on pickup tube and screen after I clearance my block and see if my 20190 pan will work with the increased stroke.

Many thanks,

Harry
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Re: connecting rod info

Postby grumpyvette » April 8th, 2011, 10:37 am

http://www.circletrack.com/enginetech/c ... index.html

http://www.carcraft.com/techarticles/11 ... index.html

http://www.eaglerod.com/mosmodule/bolt_torque.html

Fasteners, no matter what type, are the greatest weakness in a connecting rod. However, if installed correctly, most fasteners from quality manufacturers are capable of handling the stresses they are designed for.
most engine failures ,Ive seen that were blamed on connecting rods are not truly the result of the connecting rods failing but either improper assembly or its been the valve train or lubrication system that failed resulting in the rods being damaged, drop a valve and you can,t expect the rod to compress a solid steel valve into a cylinder head without being over stressed and bending, loose lubrication, the bearings cooling and load supporting oil film disappears and the bearings spin, the rod gets slack, the piston starts slapping the heads, and disaster follows in seconds.
The fasteners used to hold the two pieces of the big end of the rod together come in two designs. Thru-bolt designs have a complete bolt and nut to clamp the rod together. A cap-screw design eliminates the nut, instead utilizing threads in the rod for the bolt to thread into. A thru-bolt design requires flat faces to be cut into the big end of the rod for each bolt (one for the head of the bolt and one for the nut). Eliminating the flat for the nut makes the cap screw that much stronger. Additionally, threading the bolt directly into the body of the rod also helps rigidity.
Torque vs. Stretch
The torque spec applied to any particular fastener is merely an estimate of the twisting force required to achieve the correct amount of preload or clamp load. Many times this is the only way to apply fastener load because the bolt threads into a blind hole like in the cylinder block. One advantage to the rod bolt is that both ends of the bolt can be accessed. This allows you to use a rod bolt stretch gauge. This is a specialty tool sold through companies like ARP that will accurately measure the amount of bolt stretch.

The procedure is actually quite simple. Once the connecting rod and cap are installed on the crank, start a nut on the rod bolt, slip on the appropriate-size box-end wrench, and then install the stretch gauge. All ARP connecting rod bolts have a small dimple placed on both ends of the bolt that accurately position the rod bolt gauge pins on the bolt. Next, zero the gauge on the relaxed bolt. Then you carefully tighten the rod bolt until the gauge reads the appropriate stretch amount. For example, a standard 11/32-inch ARP small-block Chevy specs out at 0.0063 inch.


http://www.circletrack.com/tipstricks/4 ... index.html

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Do not assume all the rod bolts will all take the same torque to get to the specified listed stretch

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THE stock and less expensive replacement rod bolts pictured above tend to be made of less expensive materials, rod bolts with knurled shanks tend to be weaker than the ARP WAV LOC designs but check out the rated stress and torque levels
and remember bolts with 7/16" cap screws (ARP 2000) tend to be more rigid than bolts requiring locking nuts
most QUALITY aftermarket rod bolts, and rods with cap screw fasteners will have, the cap screws thread into the main connecting rod body without nuts, or NUTS or bolt heads that will be of the 12 point design vs the stock 6 point nuts, and pull thru bolt design


Image
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IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
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Re: bore to stroke ratio

Postby grumpyvette » May 11th, 2011, 9:38 am

The Confusion Factor - A Collection of Misunderstood Ideas and Terms By B. Rawls



Camshaft Overlap and Compression
Exhaust System Diameter and Engine Horsepower
Lobe Separation Angle and Engine Usage
Custom Ground Camshafts
Old Camshafts Lacked Sound Design Principles
Degreeing Camshafts
Rocker Arm Ratios
Piston to Valve Clearance
Adjusting Lash on Mechanical/Solid Cams
Hydraulic Lifter Preload and Pump-Up
Pushrod Length and Valve Stem Centerline



Camshaft Overlap and Compression- A very common idea, although for the most part incorrect, is that overlap bleeds off compression. Overlap, by itself, does not bleed off compression. Overlap is the angle between the exhaust closing and intake opening and is used to tune the exhaust's ability draw in additional intake charge as well as tuning idle vacuum and controlling power band width. Cylinder pressure is generated after the intake valve has closed, through the ignition process, and before the exhaust opens; in other words the compression cycle and ignition cycle of the 4 cycle or 4 stroke engine. Within practical limits, an early intake closing and late exhaust opening will maintain the highest cylinder pressure. Overlap can be increased by narrowing the Lobe Separation Angle while holding the intake and exhaust lobe duration constant. In doing so, the cylinder pressure or dynamic compression can actually increase as the earlier intake closing pairs up with the delayed exhaust opening. Overlap can also be increasing by widening intake and exhaust lobe duration while the Lobe Separation Angle is held constant. This decreases cylinder pressure or bleeds off compression. In both scenarios the overlap was increased, but the outcome differs as the intake closing and exhaust opening relationships change. The true culprit that bleeds off compression is the whole collection of valve events, not just overlap.



Exhaust System Diameter and Engine Horsepower- A popular idea is to select or size the exhaust system components to the engine's horsepower output. This methodology attributes a header diameter or an exhaust system diameter to a particular horsepower level. To debunk this idea, look at how an engine operates and consider one cylinder. The amount of charge that can enter the cylinder is dependent on inlet flow capability, crank geometry, rpm, and valve timing as a minimum consideration. Likewise, the amount of spent charge exiting the cylinder is dependent on the same characteristics.

An engine's output is usually thought of in terms of horsepower. Actually, an engine produces work, measured as torque on a dynamometer, and the horsepower is calculated through a units conversion. The amount of torque an engine produces is directly related to the amount of cylinder pressure generated. Once again, this is all affected by the same previous characteristics (flow capability, crank geometry, rpm, valve timing, etc). Therefore, an engine's power output is about air exchange and cylinder pressure. Using this line of thinking, reconsider the exhaust path. The exhaust system is more reflective of the engine's ability to move air, as opposed to horsepower numbers. Engine output does not address the breathing aspects of the engine and is not a good criteria to use for exhaust sizing.

There is a very good reason that tuners/engineers/specialist have attempted to assign exhaust to intake relationships around 70-80% for a typical natural aspirated set-up. In non-detailed terms, it is a range that offers a good balance for power capability. Other relationships, such as 1:1, are used and they work very well, but these methods have to be applied and tuned for very specific circumstances. This relationship does not stop on the flow bench on a lone cylinder head; it applies all the way from the intake path opening to the exhaust system termination. In short, try to maintain exhaust sizes that are in line with the intake flow capability. Also, do not stop your analysis at the intake and exhaust paths. If the engine already has the camshaft, look at the valve events. If the specs favor a restricted exhaust (indicated by early and wider exhaust openings with wider lobe separation angles), then size it accordingly by using exhaust components with smaller cross-sections. If the valve timing specs favor the intake, then the engine needs some serious exhaust flow capability, which is only possible with larger cross-sections.

This section was written with natural aspirated combinations in mind. However, by using the 'air exchange' rationale, it becomes apparent why forced induction engines typically benefit from increased exhaust flow capability. Also, look at the nitrous combinations. The intake system remains virtually unchanged, yet with the major increases in cylinder pressure it acts like a substantially larger engine on the exhaust side, requiring earlier exhaust openings and/or higher exhaust flow capability.



Lobe Separation Angle and Engine Usage- There are many terms associated with camshafts that get tossed around often. Lobe Separation Angle (LSA) is a term that receives a lot of attention, but mostly incorrectly. For some reason, when cam application and selection is discussed, LSA seems to come first and gets linked to engine usage. Categorically, narrow LSAs are associated with racing applications with narrow/peaky operating ranges. Wide LSA’s are associated with streetability, broad powerband response, and exhaust emissions. A very effective argument to this approach is to inquire about camshafts used in Pro-Stock and Competition Eliminator drag sanctioned classes that utilize Lobe Separation Angles around 114 degrees, and occasionally in the 116 to 118 degree range. These are racing engines that can achieve 3 hp/cid and have powerbands that are often within a narrow 2000 rpm envelope; clearly violating the LSA selection guidelines. Another approach might be to compare a 283 cid Chevrolet Super Stock to a 280 cid Chevrolet Competition Eliminator. The camshaft on the Super Stock application might have a 104 LSA, while the Comp Eliminator has a 114 LSA.

These examples reveal a key piece of info regarding LSA. If you really look at the engine combinations, the more breathing capability a motor has (relative to its displacement), the wider the LSA can end up. This observation indicates that LSA and engine usage comparisons are not globally valid. If it is so easy to point out very well established examples that violate the criteria, maybe the premise of LSA versus engine usage (void of specific engine parameters) is not a valid cam selection criterion at all.

Another key piece of info is that different engine combinations require completely different valve events. Once those valve events are determined, and lobe requirements are established, the LSA is calculated, and the camshaft can be manufactured. Maybe, LSA should be thought of as a camshaft manufacturing term as opposed to a camshaft design criteria.



Custom Ground Camshafts- When optimized performance of an engine combination is desired, the camshaft design parameters are calculated from the engine and vehicle specifications to perform within specific operating conditions. Let me emphasize that last statement, 'within specific operating conditions'. In no way was total maximum power for the engine implied. The intent is to maximize performance within the intended design parameters. If that means taking a pro-stock motor and wanting to run it from 2000-5000 rpm, then so be it.

The camshaft's seat timing events, ramp rate, and lift are directly related to the intake and exhaust flow capabilities, crankshaft geometry, static compression, rpm range, as well as other criteria. A camshaft selected in this manner, becomes personalized to that particular engine combination. Usually a custom grind is selected as an intake lobe and exhaust lobe with a particular phasing to each other (lobe separation angle, LSA) and sometimes a specified amount of advance or retard is built in. Although, it could easily end up having completely reengineered lobe characteristics, requiring new lobe masters with specialized ramp requirements. It is possible for an off-the-shelf camshaft to be a classified as a 'custom'. If the cam design is calculated for a particular combination and an off-the-shelf part number fits the bill, then for all practical purposes that part number is a 'custom' cam (but only for that particular set-up).

Typically, cam catalogs do not specifically list custom ground camshafts, because the possibilities are endless. They stick to particular series or families of camshafts. The super stock grinds come closest to an off-the-shelf grind that is truly optimized for a combination. There will be small differences due to header sizes and engine builders’ secrets, but usually the catalogs are pretty close to a good baseline. Likewise, brand to brand, the grinds will be very similar because of the 'class' dictated combinations and the flow characteristics are so well documented.



Old Camshafts Lacked Sound Design Principles- To quote Chevy High Performance’s Cam-Tastic! Issue from March 2000, Camshaft Basics, “In the old days of camshaft design most cams were designed with exactly the same duration on the intake and exhaust lobes”.

I have seen other articles and books make similar claims. It is time to cut to the chase on this and clear the air of misinformation. Real camshaft design has always addressed the needs of the engine. It’s the high performance marketplace that, for some reason, skirted the whole idea of what the camshaft does, adopting this same intake and exhaust lobe subject line. Therefore, they are the ones staking claim to noticing the change over the past 20 or so years. In short, real engineering design in valve events and camshaft technology has always been around. Here a few examples:

In the 1930s, the Chevrolet Brothers’ Frontenac Stagger-Valve cylinder head conversion utilized 7 degrees more intake duration than exhaust.

In the late 1940s, Offenhauser was using cams with different intake and exhaust lobes on their speedway motors.

In 1959, Almquest Engineering (pioneers in the hot rod mail order business, beginning in the 1940s) offered different camshaft grinds for the flathead Ford V8’s, and some of those utilized more exhaust lobe duration.

In 1966, Ford designed the camshaft for its 289 Trans-Am Program that utilized 14 degrees less exhaust seat duration, to match the 90% exhaust to intake ratio of the seriously hogged out exhaust ports with 1.625 valves.

The Z-28 “Special Off-Road” Camshaft utilized more than 10 degrees of additional exhaust duration.

For an article entitled “Camshaft Basics”, it missed some relevant history. The article claims the design process changed over the years. Technology has certainly advanced, but the design process of matching the camshaft to the motor hasn’t.



Degreeing Camshafts- There is no special magic involved for degreeing a camshaft during installation, but this is not the same thing as random advancing, retarding, or installing the gears 'lined up'. Degreeing a camshaft involves definite known values for valve events. Typically this is specified as an Intake Centerline or as opening/closing events at specific lobe lifts. This is done to insure the cam is installed per specific requirements, such as a recommendation from an engine builder or the vendor's data sheet for that camshaft grind. Manufacturing tolerances and shop practices do not guarantee that the cam matches the data sheet, when installed at crank gear 'zero'. The cam will usually need to be advanced or retarded to the correct location. If it is correct, at crank gear 'zero', then the cam has still been degreed. It just did not require any additional tweaking to meet the requirements. Verifying the installation is what degreeing a cam is all about. A common misused term is the 'straight-up' installation. Typically this is described as installing the cam at crank gear 'zero'. This is 100% wrong. ‘Straight-up’ refers to the intake and exhaust centerlines being the same. In other words the cam will have no advance or retard during installation, regardless of the amount of advance/retard ground in by the vendor. In reality, the cam may have to be advanced or retarded (from crank gear 'zero') significantly to arrive at a ‘straight-up’ position.



Rocker Arm Ratios- What is the role of the valve rocker arms in an engine? Rocker arms redirect the line of action of the camshaft's lifters or followers to the valve stems' centerline. Most rocker arms utilize a design that multiplies the camshaft lobe profile providing an amplified path that the valves follow. This multiplier is called the rocker arm ratio.



A camshaft lobe profile is measured in lift per degrees of crankshaft rotation. The typical lobe profile graph will have a vertical axis in lift increments (inches or mm) and a horizontal axis in terms of crankshaft rotation (degrees). This data forms the specs that camshaft lobes are described by. It will be stated as degrees of lobe duration at a specific lobe lift. Similarly, the valves, acted on by the rocker arm, can be measured in degrees of crankshaft rotation at a specific valve lift. This duration can be decribed as valve duration. Altering the rocker arm ratio changes the multiplication effect. The crankshaft rotation for a given valve lift will be changed from its original baseline positioning; in other words, the valve duration changes. Most often, we read that rocker arm ratio changes only affect the maximum valve lift; this is incorrect and impossible. By definition, all points where lobe lift occurs are translated to the valves through the rocker arm ratio. The change in valve duration that occurs due to the rocker arm ratios has to be decoupled from the camshaft lobe specs. These are two completely different measurements, yet intrinsically related because of the rocker arm. The valve duration changes, obviously the camshaft lobe duration does not.



A higher ratio will increase the valve lift per degree of crankshaft rotation. Therefore, the crankshaft will reach a given valve lift opening point earlier and closing point later; increasing the valve duration at the specific valve lift. Increasing the rocker arm ratio increases valve duration.

A lower ratio will decrease the valve lift per degree of crankshaft rotation. Therefore, the crankshaft will reach a given valve lift opening point later and closing point earlier; decreasing the valve duration at the specific valve lift. Decreasing the rocker arm ratio decreases valve duration.

Here is a graphical depiction of the lobe profile versus valve activity with differing rocker arm ratios.



Piston To Valve Clearance- Piston clearance is a function of lobe geometry and phasing to the piston. Cam lift should not be a deciding a factor in clearance issues. Valves will hit the piston in the overlap period, while exhaust is closing and intake is opening. Exhaust clearance problems will typically occur just before TDC and intake just after TDC, not at max lift. Some cylinder head venders and other component manufacturers advertise a max duration or lift before clearance issues arise. This is very misleading. Maximum safe duration is a totally bogus value, and is completely worthless without knowing anything about the ramp rates or actual timing/phasing events of the installation. At least with maximum safe lift, the vendor can apply a ridiculously fast ramp at a very early opening/closing and arrive at a somewhat meaningful measurement, but without knowing the design specifics the information is still next to useless.



Adjusting Lash on Mechanical/Solid Cams- If valve lash changes significantly over time, then something is wrong. Cam wear is very slight, along the order of .002 or less. Lash settings should be taken/adjusted at the same temperature and same order as the previous or original setting. This is the only way to rule out expansion/contraction of the components from temperature changes. This temperature delta is usually the culprit of most valve lash dilemmas. At initial start-up and break-in of a new set-up: cam, lifters, rockers, pushrods, valve job, etc., the lash may move around during the break-in procedure and for a short time after. This is because all the parts are seating into their new wear patterns. Once this occurs, the lash setting should stay steady. If lash settings change more than .005” then there has been a component failure (loosened hardware or actual mechanical failure)



Hydraulic Lifter Preload and Pump-Up- Hydraulic lifters are intended to make up for valvetrain dimensional differences as well as providing a self-adjusting method of maintaining valve lash, or rather the lack of. By setting the valvetrain so the lifter plunger is depressed slightly, the lifter is able to compensate for these differences, making a convenient hassle-free valvetrain set-up. For performance applications, lifter preload is not needed or wanted. As rpm's increase, the lifter has a tendency to bounce over the back of the lobe as it comes back down from the maximum lift point. The pressurized oil fills the lifter body to account for this bouncing. Eventually, after several engine revolutions (fractions of a second), the oil can completely fill the lifter body and the plunger will be pushed up to its full travel (pump-up). Higher oil pressures can amplify this problem. With the lifter pre-loaded, this can cause a valve to run off its seat and can cause piston clearance issues if and when pump-up occurs. By setting the valvetrain preload at “zero lash”, or just beyond, as felt by the hands and fingers during the adjustment process, lifter pump up is prevented and in most cases, the cam will rev higher. This adjustment process will typically end up with about .003” to .007” of lifter preload. Ford tech and tuning articles in the late 60's actually urged 'stock' class racers to run a positive lash of .001”-.003” on hydraulic cams.



Pushrod Length and Valve Stem Centerline- Incorrect pushrod length can be detrimental to valve guide wear. Most sources say that centering the rocker contact patch on the valve stem centerline at mid valve lift is the correct method for determining the optimum pushrod length. This method is wrong and can actually cause more harm than good. The method only applies when the valvetrain geometry is correct. This means that the rocker arm lengths and stud placement and valve tip heights are all perfect. This is rarely the case. To illustrate this, think of the valve angle and the rocker stud angle. They are usually not the same. If a longer or shorter valve is installed, then the relationship of the valve tip to the rocker stud centerline has changed. Heads that have had multiple valve jobs can also see this relationship change. Notice, the rocker length (pivot to tip) remains unchanged, so the rocker contact patch will have to move off the valve centerline some particular distance for optimum geometry to be maintained.

The optimum length, for component longevity, is the length that will give the least rocker arm contact area on the valve stem. In other words the narrowest wear pattern. This assures that the relationship is optimized and the valve stem centerline is tangent to the rocker arm’s circular swept path. The optimum rocker tip contact point probably will not coincide with the valve stem centerline. What is the acceptable limit for being offset from the valve stem centerline? That will depend on the set-up. A safe margin to strive for is about +/-.080" of the centerline of an 11/32 diameter valve stem. No part of the wear pattern should be outside of this .160" wide envelope. As the pushrod length is changed, the pattern will change noticeably. As the geometry becomes closer to optimum, the pattern will get narrowest. If the narrowest pattern is too far from the valve stem centerline, then the valve to rocker relationship has to be changed. In this case, the valve stem length or the rocker arm will need to be changed. This does not imply a change of rocker ratio, but rather the sweep radius.
IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
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Re: bore to stroke ratio

Postby enigma57 » May 18th, 2011, 12:32 am

:D A very interesting read, Grumpy! Thanks for posting!

Best regards,

Harry
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Re: bore to stroke ratio

Postby Randy_W » May 18th, 2011, 6:11 am

I'm about to start the build on a moderate 403 Olds for my 442, talk about over square...
Bore.....Stroke
4.351 x 3.385
The windowed block webbing limits overall horsepower but I only want 350-360, and for big bore short stroke motors, they make great torque if you keep the ports and valves and cam in line with the intended use. I built a very mild one in my '79 Trans Am when it was new, well really I just bolted stuff on. At the time intakes and such were a bit limited, all I did was Hooker headers with true duals (2.5"), Holley intake and 750 vacuum secondary and a set of 3.23 gears. On old Firestone SS 275/60's that thing went 13.60's- 13.80's depending on how well I got it to 60'.
Randy
Don't mess with old men, we didn't get that way by being stupid!
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Re: bore to stroke ratio

Postby grumpyvette » May 18th, 2011, 7:09 am

IF YOU CAN,T SMOKE THE TIRES AT WILL,FROM A 60 MPH ROLLING START YOUR ENGINE NEEDS MORE WORK!!"!
IF YOU CAN , YOU NEED BETTER TIRES AND YOUR SUSPENSION NEEDS MORE WORK!!
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