We have a new SFI certified non-adjustable billet flywheel for the flathead. This flywheel is stock diameter, weighs 4.05lbs, and has the keyway set at 30°.
The only difference between this flywheel and our Limited Mod adjustable flywheel is that this one does not have the steel adjustable timing hub.
Part #6620 introductory racer price $125.00We made the flywheel that best fits ARC's capabilities. It's the best piece we could make with focus on affordability and maintaining our SFI certification. That's the best we can do for the racer. The design allows you to start your engine with either an electric starter OR the Briggs BS555165 manual pull starter.
Tuesday, February 5, 2008
New ECONOMY Flathead Billet Flywheel from ARC
ARC billet rods for the Vanguard / Mitsubishi OHV 6HP
We have two new billet rods for the Briggs and Stratton Vanguard OHV 6HP
#6264 Stock Length 3.465", uses stock piston $83.95 MSRP
#6263 Long rod 3.850", uses JE piston #EXF4500-120 with approx. .090" cut off the top. It is .385" longer than stock. $83.95 MSRP
Both of these rods use the same bearing inserts as the Honda GX200 (and clone) engines. By increasing the number of rods that use this bearing, we hope to be able to justify undersized bearings.
Does ARC make Custom Rods?
Quoted from http://karting.4cycle.com/showthread.php?t=167398:
Q
How expensive is it to have a rod built to a custom length? A
That depends on the length, and we have to make a minimum of 3 rods. A bore diameter change at either end is $750 (for each end) unless it matches up with a bore of one of our existing rods so we can borrow existing tooling. A length change could be as much as $2000, unless it is only a slight center to center change that will fit within an existing rod profile. The same profile allows the use of existing jaws.
So, for the first 3 special rods, $600-$3500 (for all 3) depending on those factors. Additional rods made in the same run will be about $100 each depending on how many you get. There are other factors that could make it more, or less expensive, such as material cost and availability.
Monday, April 4, 2005
On Rod Bolts, By Mike Gifford
My name is Mike Gifford; I'm "Outrider" in the 4Cycle.com forums, the guy that commented on how happy I was to see a fastener tightening procedure that made allowances for the differences in friction coefficient (and hence applied torque required) when different thread lubes are used on threaded fasteners. At the time, I promised you that I would eventually send you some comments on fasteners and tightening procedures.
By way of background, I started my engineering career with a BS degree in Engineering, an Ensign's commission in the Naval reserve, and an Unlimited Horsepower US Coast Guard 3rd Engineer's License for both Steam and Diesel ships. After several years as a sea-going Marine Engineer, I came ashore in 1970 and eventually went to work for the Navy. I ended up as an engineer in the Submarine Fluid Systems Division, and spent the better part of the next thirty years in various Engineering and Naval Architect billets in the Naval Sea Systems Command, all in the submarine community. Got to ride/play with things most people only see in National Geographic specials and on the History Channel, and got paid for it. And got to retire when I was 55.
Submarines have thousands of bolted joints, many in critical applications, which, in the end, accidentally caused me to build a small empire within the Navy's engineering community as a fastener expert. Amazing how a collateral duty (that's all it ever was) gets to be a big deal.
In the early 1970s, the Navy's submarine community realized that they had a bunch of bolting problems; both with fastener selection and joint assembly, and by 1978 published a Submarine Fastener Manual. In 1981 I switched jobs and, as part of my new job's duties, inherited responsibility for that manual; after that, wherever I went in the submarine community, the job description for my new position would be rewritten to include the Submarine Fastener Manual and all submarine fastener problems as a collateral duty. As one who had been assembling motorcycle and automobile engines since he was 12, and new how to use a torque wrench, I thought I had this one wired. Within 3 days of inheriting the Fastener Manual I discovered that I about had the tip of the iceberg wired, and embarked on a very steep learning curve for the next two months. The learning hasn't stopped to this day, but the curve is rarely that steep any more. My specialty was "in service" engineering - how to select a fastener that would live for a specific application, installation procedures, including procedures for tightening with a torque wrench, use of angular turn, and use of ultrasonic direct stress measurement; review of failed fastener analysis reports and developing solutions for preventing those failures in the future. The less publicly visible (but probably more important) section handled fastener materials, manufacturing processes and QA/QC testing of materials and finished products, and contributed to my ongoing education on a regular basis. So that's how I got to be a fastener geek/goon.
Random comments, in no particular order of importance:
1. While there are many materials used for bolts, cap screws, machine screws and nuts, where there are not problems with corrosive environments or other special considerations, selecting from among the various steel alloys used for fasteners is a hard act to top.
2. In the drive to reduce space and weight, ever-higher strength alloys are considered. This is a good thing if not carried to extremes, but people need to remember that high strength is no good without toughness. Just because a fastener is strong doesn't mean it is tough - VERY high strength fasteners (250ksi - 270ksi yield, as found in some heat treats of some of the Nickel/Cobalt alloys) tend to be brittle and not respond well when subjected to high shock loads. Since they don't stretch much, they break, where a tougher (though less high strength) alloy will just stretch. Of course, the tougher alloy would usually have to be a larger diameter or use more fasteners in the joint; everything is a compromise in the fastener world.
3. Concerning toughness and resistance to high shock, there is an all to often ignored quality buried in fastener material chems and physicals called "Percent Elongation" which makes screening for toughness relatively easy. The Navy's big hurdle is passing tests for resistance to "Hi Shock and Undex" ("Undex" is short for UNDerwater EXplosion). The Navy flatly refuses to allow components (including the fasteners holding things together) in critical applications to use materials with a percent elongation less than 10%, and will only VERY rarely allow materials with a percent elongation less than 10% in a non-critical application by approval of an exemption specific to that component for a specific service on a specific class of ships (we're speaking combatants here; noncombatant ships are allowed a little more leeway in their design). In general, to meet hi shock and Undex requirements, materials with a percent elongation of 15% or more should be chosen, over 18% is better, and, as heat treating processes for metallic alloys in production quantities has improved over the last 20 -30 years, many fastener alloys that were in the 15% - 18% range can now be reliably produced in the 20% - 25% elongation range, which is great for toughness. Fortunately, by the nature of the beast and the general conservatism of mechanical joint designs used in shipbuilding, any fastener material with a percent elongation of over 10% will generally pass hi-shock/Undex in a properly designed bolted joint. To protect its flanks, the Department of Defense has some specifications that it piggybacks onto commercial specs (MIL-DTL-1222, for instance) to get what they need in this respect, so an ASTM A574 4340 socket head cap screw for a critical application would be ordered with a minimum percent elongation over 10%, which is higher than the minimum demanded for 4340 socket head cap screws in A574. And many engine building applications don't have the hi-shock loads present in Undex testing and can benefit from really high strength fasteners, even if their percent elongation is less than 10%, but you do have to match the alloy chosen to its shock environment carefully.
4. Once the user the user settles on the basic fastener configuration (hex head cap screw, socket head, 12 point, etc.) and the correct material, the most important things (more important than material and basic configuration, unless grave errors are made in selecting those two items) are the geometry of the fillet radius where the unthreaded shank joins the head and the transition from the threaded portion to the unthreaded shank of the fastener. Errors in design or execution in either of these two areas can result in fastener failure where it would not otherwise occur. To eliminate problems in the threaded to unthreaded shank portion, use of rolled threads is usually sufficient. For the shank to head transition, the proper fillet radius needs to be specified, and adherence to that specification needs to be checked religiously as part of the QA program.
5. When doing qualification testing or random sample receipt inspection of fastener lots, the most effective single physical test (after a good visual inspection and dimensional checks) to verify the quality of a finished fastener (including the items in 4. above) is a wedge tensile test per ASTM F606. The beauty of the wedge tensile test is that it is done on a finished fastener, not a machined tensile test specimen, so it picks up both material problems and manufacturing defects like an inadequate fillet radius. Generally, if you purchase fasteners from an outside source, they will state what specifications are used for manufacture and random sample visual inspection and dimensional checks are sufficient. Cash flow permitting, it's nice to send a few to a lab once in awhile for a wedge tensile test by an independent lab - I was spoiled; any Naval Shipyard had a lab with lots of neat machines, including a tensile test machine.
6. With regard specifically to rod bearing cap screws for ARC connecting rods, my inspection of a limited number of these fasteners (two rods from my engine builder's stock) left me impressed. Their high quality was obvious, as was correct design and execution of the fillet radius and the thread transition (to the extent that this can be determined by a visual inspection, but I'm willing to bet that they would pass a wedge tensile test without breathing too hard), but more interesting was an extra little design feature, a reduced diameter section in the unthreaded shank of each cap screw. That little feature actually improves resistance to shock and greatly improves performance in hi-shock situations. Basically, it makes the fastener a better spring, mitigating shock damage by increasing fastener toughness with a mechanical trick, rather than exotic metallurgy. As an example of what this feature can do, the Navy has subjected a group of fasteners (studs in this case) to a shock load that would make them fail; the average stretch was 1/16" before they broke. Repeating the test with studs identical except for a slightly reduced diameter in the unthreaded section, the fasteners actually stretched an average of 3/16" prior to failure. When not carried to extremes, that little feature is an excellent way to improve high shock performance, and it doesn't take much reduction to collect that benefit; obviously, too great a reduction in diameter in the unthreaded shank will reduce ultimate strength more than it benefits shock resistance, but it is usually possible to strike an effective balance without causing problems if the fastener design isn't right on the edge of failure due to the in service load profile to begin with.
7. Excerpts from a Navy training lecture - "What your mother and the professors didn't teach you"
A. Preload range of bolted joints assembled with a torque wrench:
Most people see a torque specification for the threaded fasteners in a joint assembly and think that (1), the desired preload is achieved with great precision, and (2) that the fastener-to-fastener preload variation within the joint is small. Neither of these impressions is correct. On the best of days, the fastener-to-fastener preload variation is about 35%, and often more. It is not unusual to see the largest preload measured in the bolt circle twice that of the smallest in joints assembled with a torque wrench by an inexperienced mechanic without benefit of a proper installation process. Although other standards have been (and when there is a specific reason, still are) used, most bolted joints found on submarines (that require use of a torque wrench for assembly) have a preload established at 2/3 of yield of the weakest element of the joint (150% of yield in the case of bearing stress), and, ideally, the limit will be reached as tensile stress in the bolt or stud.
So you have a torque from a drawing, Maintenance Standard, tech manual or whatever, based on, say, 2/3 of yield. When the mechanic is done assembling the joint, all the bolts in the bolt circle would have (theoretically) a tensile load of 2/3 of yield, because the torque was chosen to give that result. Unfortunately, there is a significant fastener-to-fastener variation in friction coefficient, AND a significant fastener-to-fastener variation in short term preload loss (the relaxation that occurs in the first 2 to 10 minutes after the wrench is removed from each fastener in the bolt circle for the last time). And a few other things (all told, about 76 different things, according to the Air Force, which did an excellent study on the subject, like prying loads and cross talk. The result is that all we know for sure is that each fastener in the bolt circle, having been tightened to a mean torque mathematically equivalent to a mean preload of 67% of yield, has an actual preload of somewhere between 40% and 90% of yield. It sounds crude, but it's close enough, even for hull integrity/high shock/UNDEX/hazardous fluid, etc, services. The distribution of this preload variation is more or less a bell shaped curve in a statistically valid sample.
B. The value of process instructions:
The real case is not quite as bad as the above makes it look, as the methods incorporated into Navy/Navy approved process instructions are designed to reduce this preload spread. The reality is that the multiple passes, check passes, etc, of Navy process instructions skew the curve. The top value (90% of yield) doesn't change, but the number of fasteners below 67% is significantly reduced, and the amount by which they are under 67% of yield is also reduced, while the number between 67% and 90% is increased. Since the best defense against long term preload loss is high initial preloads and minimization of fastener to fastener preload variation, the value of good processes and training in those processes is once again proven.
C. Good shop practice and precision:
As far as fastener to fastener preload variation is concerned, the variation for fasteners in a joint tightened without a torque wrench, but in stages and with check passes, is 5% to 10% more than the variation with a torque wrench, on the average. In reality, other than being cheap to use, the torque wrench only offers 3 advantages:
1. It assures adherence to a specified mean torque, which in many joints is not a particularly significant item, except for record purposes, or where minimizing the chance of exceeding the yield strength of the material or the threshold stress level of an H2 embrittlement prone material is important.
2. It results in slightly less preload variation than the use of "good shop practice" and a box end, open end or socket wrench, in the hands of an experienced technician.
3. It offers visual proof to witnesses that each fastener in the bolt circle has been properly tightened, needed for certification records for "critical joints".
For what it's worth, the most accurate method of tightening the fasteners in a bolted joint without resort to expensive ultrasonic measuring devices, strain gauge equipped bolts or other cost increasing approaches is angular turn of the nut. The fastener to fastener preload variation runs about 15%, much better than a torque wrench can ever hope for.
When fasteners are overtorqued severely during initial installation but survive the procedure without failure, the joint is generally OK for service without further action. Various embedment phenomena and other short-term preload losses will reduce the stresses to a high but acceptable level. The exception is where the fasteners are made of materials prone to hydrogen embrittlement. They may settle out, after short-term preload loss, at a level in excess of their threshold stress level, leaving a high probability of brittle failure in the future. Resolution where H2 embrittlement prone materials are involved should always favor loosening and re-torquing, one fastener at a time, to the correct value (Note to Tom: Many bolted joints on Navy ships require hydrostatic testing for verification. If the joint integrity is violated by loosening the fasteners, an expensive retest is required. By common sense and resulting executive fiat, the Navy does NOT regard the integrity of the joint as violated if the fasteners are loosened and re-tightened one at a time, saving the expense of retesting, hence the importance of "...one fastener at a time"). If discovery is after the joint is buried by interferences and cost is too great, use of Level I certs may allow acceptance based on stresses adjusted for the actual yield rather than the min spec numbers usually used in design calculations and by PC Bolts. Fasteners of materials such as Grade 5 steel and other materials not prone to H2 embrittlement may be left as is and torqued correctly at the next disassembly and re-assembly of the joint unless the activity is having a lot of such errors, in which case remedial retorquing is necessary to get production's attention.
G. Notes on choosing thread lubricants:
2. Once in awhile the subject of the range of friction coefficients that will be exhibited by a thread lubricant will come up. When discussing this subject, insist that the range be tied to a specific material (nut and bolt or stud) combination. The reason for this is that the same lubricant often has both a different mean friction coefficient and different extremes (and can have the same mean friction coefficient, but different extremes, the high and low values that establish the range) when you compare the mean and extremes for different combinations; alloy steel and alloy steel, CRES and CRES, Monel and Monel and KMonel with a monel nut, for instance. The extremes for each material combination establish the range FOR THAT COMBINATION. Taking the high for the combination that has the highest extreme and the low for the combination that has the lowest extreme, a method that will quite often be attempted by your Nuclear counterparts if you don't call them on it, doesn't give the range, it gives a number useless in calculations and of little interest in intelligent discussions.
8. Notes on tightening procedures (the following is a slightly edited excerpt from the Navy's Submarine Fastener Manual, from the sections that are the basis for development of local activity's process instructions and instruction packages for individual work packages, where such detail is necessary for a specific work package. As you can see, the basic approach is relatively simple and grounded in common sense. It also would pass as the generic basis for ARC's specific procedures):
Before tightening any fasteners, the following should be performed:
a. Examine fasteners for compliance with marking requirements.
b. Examine the internal and external threads for burrs, nicks, metallic slivers, etc., that could cause jamming or excessive resistance to tightening. Remove or correct as necessary.
c. Ensure the threads and mating bearing surfaces are clean and free of rust, chips, or other foreign matter.
d. Ensure the nut (or cap screw) seating surface is flat and contacts the mating surface all around.
e. Lightly lubricate the threads and bearing surface with the specified lubricant and remove excess lubricant to permit air to escape from under the nut (or head of the cap screw). Flange spot facing should also be lubricated.
f. If using torque measurement method, ensure the torque wrench has a current calibration sticker. Select a torque wrench such that the required torque is between 20% and 90% of the full-scale range of the torque wrench selected.
FASTENER TIGHTENING PROCEDURES. The following procedures are applicable to nuts, through bolts, studs, cap screws, and set-studs used on flat-face and raised-face flanges:
a. Prior to applying final torque, perform the prerequisites described in steps a through f above.
b. Assure proper alignment of the mating components.
c. Where the application requires O-rings or gaskets, ensure the O-ring or gasket is in its proper position. Make up the joint evenly by tightening diametrically opposite fasteners until the mating components contact each other. This will normally be accompanied by a noticeable increase in torque when metal-to-metal contact is made. Check all fasteners to ensure that no fasteners are loose. Continue to tighten fasteners sequentially. Apply approximately ten percent of the specified torque to ensure solid part contact. Finish torquing the joint in 25 percent increments of the specified torque.
d. For determining torque values used in this procedure, refer to paragraph 5-5, use Appendix E (PC Bolts), or seek guidance from your activity's Design Division.
e. When tightening nuts in set stud and nut type joints, check stud rotation by marking with a felt-tip marker on the nut end of each stud in a direction toward the center of the flange. Check the mark on each stud after tightening to ensure the stud did not rotate.
f. After completion of the last tightening pass, wait a minimum of 2 minutes, and execute a check pass or passes until the joint holds the specified torque setting. This minimizes the effects of short-term preload loss and helps minimize fastener-to-fastener preload variation.
9. Where hydrogen embrittlement is mentioned above, it shouldn't be a problem in the environment of a rod bearing cap screw, since you need high stress (above the material's threshold stress level) in the fastener, a source of free hydrogen and an electrolyte, and an electrical potential, like screwing a Kmonel stud into an HY-80 steel pressure hull, or a steel cap screw into an aluminum rod - the whole world is a battery waiting to happen once you introduce dissimilar metals. High strength steels (140 ksi yield and above for purposes of H2 embrittlement discussions) are subject to H2 embrittlement and can be assigned a threshold stress level of 80% of yield (threshold stress level is a bit of a moving target, but 80% of yield is a good working value for high strength steels), but for embrittlement to be a problem, you have to have ALL the factors, not unlikely in the marine environment, but HIGHLY unlikely inside an engine, even if the target mean preload is 90% - 100% of yield for the rod bearing cap screws.
10. Where PC Bolts is referred to above, it is a relatively simple to use computer program developed by the Navy for calculating fastener torques for bolted joints. I can send you a copy if you're interested in playing with it - it covers through bolted joints, cap screws threaded into blind holes, and set stud and nut type joints. The present version is a pleasant little calculating machine in 16 bit Dos code, so old it lacks mouse support, but it's lean, mean and gives you torque, preload and a large collection of joint component stresses. Or you can input a torque and get preload and the stresses, or input a desired preload and get the necessary mean torque and stresses. The Navy hands it out for free to anyone that wants it. My one mule consulting outfit heads the Beta test program for the new 32 bit Windows version, which we will have out soon and which also will be given away free to anyone that wants it, once it's ready to release into the wild. As a Navy employee I headed Beta testing for the three previous versions, so PC bolts is sort of the crown jewel of the contributions I've been able to make to solving fastener problems; that, and a whole bunch of friction coefficient testing that went into the lubricant library of PC Bolts, and which led me to take note of your rod installation instruction.
11. Oh yeah, one other little item; all of the above discussions assume that when the designer chose the fastener size and alloy and the preload we are going to apply to the fastener, he/she had a reasonably good idea of what the worst case in service load will be and chose the preload to be well above that load after short term preload loss and a reasonable profile for long term preload loss. If we blew any portion of that little set of choices, we have the potential for cyclic load reversal and good old fatigue failure. Short term preload loss can run anywhere from 5% to 40%, though proper tightening procedures will make it unlikely that you will ever see the latter figure. Our preferred approach is to take the max in service load and select a preload 4 times that amount, so that you can lose 50% of your initial assembly preload to any combination of short and long term preload loss and still have twice the preload needed to do the job - crude, but it works. Most bolted joints work well because their design is conservative (if nobody goofed), and hence they are very forgiving of minor errors, abuse and neglect. Of course, the more refined your design assumptions are from a standpoint of test results or other prior experience with a specific application on which to base your decisions, the closer you can cut it if space or weight considerations intrude, without putting the joint's integrity in danger. In combatant aircraft design, there are bolted joints that are subject to fatigue failure in order to reduce size to fit in the space available. With fatigue life to failure in a cyclic range that corresponds to 12 to 18 months of service, the aircraft maintenance plan calls for automatic replacement every 6 months. That's a valid solution for a fighter plane, but probably won't hack it for a passenger car.
12. I have repeatedly used the terms hi-shock and Undex in the discussions above. To give you an idea of the magnitude of the forces we're dealing with when Hi shock and explosion testing are invoked, civil engineering design for earthquake involves designing for accelerations primarily in the 9G to 11G range. Hi shock and Undex deal with G forces in the 450G range, give or take, and greater.
By way of background, I started my engineering career with a BS degree in Engineering, an Ensign's commission in the Naval reserve, and an Unlimited Horsepower US Coast Guard 3rd Engineer's License for both Steam and Diesel ships. After several years as a sea-going Marine Engineer, I came ashore in 1970 and eventually went to work for the Navy. I ended up as an engineer in the Submarine Fluid Systems Division, and spent the better part of the next thirty years in various Engineering and Naval Architect billets in the Naval Sea Systems Command, all in the submarine community. Got to ride/play with things most people only see in National Geographic specials and on the History Channel, and got paid for it. And got to retire when I was 55.
Submarines have thousands of bolted joints, many in critical applications, which, in the end, accidentally caused me to build a small empire within the Navy's engineering community as a fastener expert. Amazing how a collateral duty (that's all it ever was) gets to be a big deal.
In the early 1970s, the Navy's submarine community realized that they had a bunch of bolting problems; both with fastener selection and joint assembly, and by 1978 published a Submarine Fastener Manual. In 1981 I switched jobs and, as part of my new job's duties, inherited responsibility for that manual; after that, wherever I went in the submarine community, the job description for my new position would be rewritten to include the Submarine Fastener Manual and all submarine fastener problems as a collateral duty. As one who had been assembling motorcycle and automobile engines since he was 12, and new how to use a torque wrench, I thought I had this one wired. Within 3 days of inheriting the Fastener Manual I discovered that I about had the tip of the iceberg wired, and embarked on a very steep learning curve for the next two months. The learning hasn't stopped to this day, but the curve is rarely that steep any more. My specialty was "in service" engineering - how to select a fastener that would live for a specific application, installation procedures, including procedures for tightening with a torque wrench, use of angular turn, and use of ultrasonic direct stress measurement; review of failed fastener analysis reports and developing solutions for preventing those failures in the future. The less publicly visible (but probably more important) section handled fastener materials, manufacturing processes and QA/QC testing of materials and finished products, and contributed to my ongoing education on a regular basis. So that's how I got to be a fastener geek/goon.
Random comments, in no particular order of importance:
1. While there are many materials used for bolts, cap screws, machine screws and nuts, where there are not problems with corrosive environments or other special considerations, selecting from among the various steel alloys used for fasteners is a hard act to top.
2. In the drive to reduce space and weight, ever-higher strength alloys are considered. This is a good thing if not carried to extremes, but people need to remember that high strength is no good without toughness. Just because a fastener is strong doesn't mean it is tough - VERY high strength fasteners (250ksi - 270ksi yield, as found in some heat treats of some of the Nickel/Cobalt alloys) tend to be brittle and not respond well when subjected to high shock loads. Since they don't stretch much, they break, where a tougher (though less high strength) alloy will just stretch. Of course, the tougher alloy would usually have to be a larger diameter or use more fasteners in the joint; everything is a compromise in the fastener world.
3. Concerning toughness and resistance to high shock, there is an all to often ignored quality buried in fastener material chems and physicals called "Percent Elongation" which makes screening for toughness relatively easy. The Navy's big hurdle is passing tests for resistance to "Hi Shock and Undex" ("Undex" is short for UNDerwater EXplosion). The Navy flatly refuses to allow components (including the fasteners holding things together) in critical applications to use materials with a percent elongation less than 10%, and will only VERY rarely allow materials with a percent elongation less than 10% in a non-critical application by approval of an exemption specific to that component for a specific service on a specific class of ships (we're speaking combatants here; noncombatant ships are allowed a little more leeway in their design). In general, to meet hi shock and Undex requirements, materials with a percent elongation of 15% or more should be chosen, over 18% is better, and, as heat treating processes for metallic alloys in production quantities has improved over the last 20 -30 years, many fastener alloys that were in the 15% - 18% range can now be reliably produced in the 20% - 25% elongation range, which is great for toughness. Fortunately, by the nature of the beast and the general conservatism of mechanical joint designs used in shipbuilding, any fastener material with a percent elongation of over 10% will generally pass hi-shock/Undex in a properly designed bolted joint. To protect its flanks, the Department of Defense has some specifications that it piggybacks onto commercial specs (MIL-DTL-1222, for instance) to get what they need in this respect, so an ASTM A574 4340 socket head cap screw for a critical application would be ordered with a minimum percent elongation over 10%, which is higher than the minimum demanded for 4340 socket head cap screws in A574. And many engine building applications don't have the hi-shock loads present in Undex testing and can benefit from really high strength fasteners, even if their percent elongation is less than 10%, but you do have to match the alloy chosen to its shock environment carefully.
4. Once the user the user settles on the basic fastener configuration (hex head cap screw, socket head, 12 point, etc.) and the correct material, the most important things (more important than material and basic configuration, unless grave errors are made in selecting those two items) are the geometry of the fillet radius where the unthreaded shank joins the head and the transition from the threaded portion to the unthreaded shank of the fastener. Errors in design or execution in either of these two areas can result in fastener failure where it would not otherwise occur. To eliminate problems in the threaded to unthreaded shank portion, use of rolled threads is usually sufficient. For the shank to head transition, the proper fillet radius needs to be specified, and adherence to that specification needs to be checked religiously as part of the QA program.
5. When doing qualification testing or random sample receipt inspection of fastener lots, the most effective single physical test (after a good visual inspection and dimensional checks) to verify the quality of a finished fastener (including the items in 4. above) is a wedge tensile test per ASTM F606. The beauty of the wedge tensile test is that it is done on a finished fastener, not a machined tensile test specimen, so it picks up both material problems and manufacturing defects like an inadequate fillet radius. Generally, if you purchase fasteners from an outside source, they will state what specifications are used for manufacture and random sample visual inspection and dimensional checks are sufficient. Cash flow permitting, it's nice to send a few to a lab once in awhile for a wedge tensile test by an independent lab - I was spoiled; any Naval Shipyard had a lab with lots of neat machines, including a tensile test machine.
6. With regard specifically to rod bearing cap screws for ARC connecting rods, my inspection of a limited number of these fasteners (two rods from my engine builder's stock) left me impressed. Their high quality was obvious, as was correct design and execution of the fillet radius and the thread transition (to the extent that this can be determined by a visual inspection, but I'm willing to bet that they would pass a wedge tensile test without breathing too hard), but more interesting was an extra little design feature, a reduced diameter section in the unthreaded shank of each cap screw. That little feature actually improves resistance to shock and greatly improves performance in hi-shock situations. Basically, it makes the fastener a better spring, mitigating shock damage by increasing fastener toughness with a mechanical trick, rather than exotic metallurgy. As an example of what this feature can do, the Navy has subjected a group of fasteners (studs in this case) to a shock load that would make them fail; the average stretch was 1/16" before they broke. Repeating the test with studs identical except for a slightly reduced diameter in the unthreaded section, the fasteners actually stretched an average of 3/16" prior to failure. When not carried to extremes, that little feature is an excellent way to improve high shock performance, and it doesn't take much reduction to collect that benefit; obviously, too great a reduction in diameter in the unthreaded shank will reduce ultimate strength more than it benefits shock resistance, but it is usually possible to strike an effective balance without causing problems if the fastener design isn't right on the edge of failure due to the in service load profile to begin with.
7. Excerpts from a Navy training lecture - "What your mother and the professors didn't teach you"
A. Preload range of bolted joints assembled with a torque wrench:
Most people see a torque specification for the threaded fasteners in a joint assembly and think that (1), the desired preload is achieved with great precision, and (2) that the fastener-to-fastener preload variation within the joint is small. Neither of these impressions is correct. On the best of days, the fastener-to-fastener preload variation is about 35%, and often more. It is not unusual to see the largest preload measured in the bolt circle twice that of the smallest in joints assembled with a torque wrench by an inexperienced mechanic without benefit of a proper installation process. Although other standards have been (and when there is a specific reason, still are) used, most bolted joints found on submarines (that require use of a torque wrench for assembly) have a preload established at 2/3 of yield of the weakest element of the joint (150% of yield in the case of bearing stress), and, ideally, the limit will be reached as tensile stress in the bolt or stud.
So you have a torque from a drawing, Maintenance Standard, tech manual or whatever, based on, say, 2/3 of yield. When the mechanic is done assembling the joint, all the bolts in the bolt circle would have (theoretically) a tensile load of 2/3 of yield, because the torque was chosen to give that result. Unfortunately, there is a significant fastener-to-fastener variation in friction coefficient, AND a significant fastener-to-fastener variation in short term preload loss (the relaxation that occurs in the first 2 to 10 minutes after the wrench is removed from each fastener in the bolt circle for the last time). And a few other things (all told, about 76 different things, according to the Air Force, which did an excellent study on the subject, like prying loads and cross talk. The result is that all we know for sure is that each fastener in the bolt circle, having been tightened to a mean torque mathematically equivalent to a mean preload of 67% of yield, has an actual preload of somewhere between 40% and 90% of yield. It sounds crude, but it's close enough, even for hull integrity/high shock/UNDEX/hazardous fluid, etc, services. The distribution of this preload variation is more or less a bell shaped curve in a statistically valid sample.
B. The value of process instructions:
The real case is not quite as bad as the above makes it look, as the methods incorporated into Navy/Navy approved process instructions are designed to reduce this preload spread. The reality is that the multiple passes, check passes, etc, of Navy process instructions skew the curve. The top value (90% of yield) doesn't change, but the number of fasteners below 67% is significantly reduced, and the amount by which they are under 67% of yield is also reduced, while the number between 67% and 90% is increased. Since the best defense against long term preload loss is high initial preloads and minimization of fastener to fastener preload variation, the value of good processes and training in those processes is once again proven.
C. Good shop practice and precision:
As far as fastener to fastener preload variation is concerned, the variation for fasteners in a joint tightened without a torque wrench, but in stages and with check passes, is 5% to 10% more than the variation with a torque wrench, on the average. In reality, other than being cheap to use, the torque wrench only offers 3 advantages:
1. It assures adherence to a specified mean torque, which in many joints is not a particularly significant item, except for record purposes, or where minimizing the chance of exceeding the yield strength of the material or the threshold stress level of an H2 embrittlement prone material is important.
2. It results in slightly less preload variation than the use of "good shop practice" and a box end, open end or socket wrench, in the hands of an experienced technician.
3. It offers visual proof to witnesses that each fastener in the bolt circle has been properly tightened, needed for certification records for "critical joints".
For what it's worth, the most accurate method of tightening the fasteners in a bolted joint without resort to expensive ultrasonic measuring devices, strain gauge equipped bolts or other cost increasing approaches is angular turn of the nut. The fastener to fastener preload variation runs about 15%, much better than a torque wrench can ever hope for.
When fasteners are overtorqued severely during initial installation but survive the procedure without failure, the joint is generally OK for service without further action. Various embedment phenomena and other short-term preload losses will reduce the stresses to a high but acceptable level. The exception is where the fasteners are made of materials prone to hydrogen embrittlement. They may settle out, after short-term preload loss, at a level in excess of their threshold stress level, leaving a high probability of brittle failure in the future. Resolution where H2 embrittlement prone materials are involved should always favor loosening and re-torquing, one fastener at a time, to the correct value (Note to Tom: Many bolted joints on Navy ships require hydrostatic testing for verification. If the joint integrity is violated by loosening the fasteners, an expensive retest is required. By common sense and resulting executive fiat, the Navy does NOT regard the integrity of the joint as violated if the fasteners are loosened and re-tightened one at a time, saving the expense of retesting, hence the importance of "...one fastener at a time"). If discovery is after the joint is buried by interferences and cost is too great, use of Level I certs may allow acceptance based on stresses adjusted for the actual yield rather than the min spec numbers usually used in design calculations and by PC Bolts. Fasteners of materials such as Grade 5 steel and other materials not prone to H2 embrittlement may be left as is and torqued correctly at the next disassembly and re-assembly of the joint unless the activity is having a lot of such errors, in which case remedial retorquing is necessary to get production's attention.
G. Notes on choosing thread lubricants:
2. Once in awhile the subject of the range of friction coefficients that will be exhibited by a thread lubricant will come up. When discussing this subject, insist that the range be tied to a specific material (nut and bolt or stud) combination. The reason for this is that the same lubricant often has both a different mean friction coefficient and different extremes (and can have the same mean friction coefficient, but different extremes, the high and low values that establish the range) when you compare the mean and extremes for different combinations; alloy steel and alloy steel, CRES and CRES, Monel and Monel and KMonel with a monel nut, for instance. The extremes for each material combination establish the range FOR THAT COMBINATION. Taking the high for the combination that has the highest extreme and the low for the combination that has the lowest extreme, a method that will quite often be attempted by your Nuclear counterparts if you don't call them on it, doesn't give the range, it gives a number useless in calculations and of little interest in intelligent discussions.
8. Notes on tightening procedures (the following is a slightly edited excerpt from the Navy's Submarine Fastener Manual, from the sections that are the basis for development of local activity's process instructions and instruction packages for individual work packages, where such detail is necessary for a specific work package. As you can see, the basic approach is relatively simple and grounded in common sense. It also would pass as the generic basis for ARC's specific procedures):
Before tightening any fasteners, the following should be performed:
a. Examine fasteners for compliance with marking requirements.
b. Examine the internal and external threads for burrs, nicks, metallic slivers, etc., that could cause jamming or excessive resistance to tightening. Remove or correct as necessary.
c. Ensure the threads and mating bearing surfaces are clean and free of rust, chips, or other foreign matter.
d. Ensure the nut (or cap screw) seating surface is flat and contacts the mating surface all around.
e. Lightly lubricate the threads and bearing surface with the specified lubricant and remove excess lubricant to permit air to escape from under the nut (or head of the cap screw). Flange spot facing should also be lubricated.
f. If using torque measurement method, ensure the torque wrench has a current calibration sticker. Select a torque wrench such that the required torque is between 20% and 90% of the full-scale range of the torque wrench selected.
FASTENER TIGHTENING PROCEDURES. The following procedures are applicable to nuts, through bolts, studs, cap screws, and set-studs used on flat-face and raised-face flanges:
a. Prior to applying final torque, perform the prerequisites described in steps a through f above.
b. Assure proper alignment of the mating components.
c. Where the application requires O-rings or gaskets, ensure the O-ring or gasket is in its proper position. Make up the joint evenly by tightening diametrically opposite fasteners until the mating components contact each other. This will normally be accompanied by a noticeable increase in torque when metal-to-metal contact is made. Check all fasteners to ensure that no fasteners are loose. Continue to tighten fasteners sequentially. Apply approximately ten percent of the specified torque to ensure solid part contact. Finish torquing the joint in 25 percent increments of the specified torque.
d. For determining torque values used in this procedure, refer to paragraph 5-5, use Appendix E (PC Bolts), or seek guidance from your activity's Design Division.
e. When tightening nuts in set stud and nut type joints, check stud rotation by marking with a felt-tip marker on the nut end of each stud in a direction toward the center of the flange. Check the mark on each stud after tightening to ensure the stud did not rotate.
f. After completion of the last tightening pass, wait a minimum of 2 minutes, and execute a check pass or passes until the joint holds the specified torque setting. This minimizes the effects of short-term preload loss and helps minimize fastener-to-fastener preload variation.
9. Where hydrogen embrittlement is mentioned above, it shouldn't be a problem in the environment of a rod bearing cap screw, since you need high stress (above the material's threshold stress level) in the fastener, a source of free hydrogen and an electrolyte, and an electrical potential, like screwing a Kmonel stud into an HY-80 steel pressure hull, or a steel cap screw into an aluminum rod - the whole world is a battery waiting to happen once you introduce dissimilar metals. High strength steels (140 ksi yield and above for purposes of H2 embrittlement discussions) are subject to H2 embrittlement and can be assigned a threshold stress level of 80% of yield (threshold stress level is a bit of a moving target, but 80% of yield is a good working value for high strength steels), but for embrittlement to be a problem, you have to have ALL the factors, not unlikely in the marine environment, but HIGHLY unlikely inside an engine, even if the target mean preload is 90% - 100% of yield for the rod bearing cap screws.
10. Where PC Bolts is referred to above, it is a relatively simple to use computer program developed by the Navy for calculating fastener torques for bolted joints. I can send you a copy if you're interested in playing with it - it covers through bolted joints, cap screws threaded into blind holes, and set stud and nut type joints. The present version is a pleasant little calculating machine in 16 bit Dos code, so old it lacks mouse support, but it's lean, mean and gives you torque, preload and a large collection of joint component stresses. Or you can input a torque and get preload and the stresses, or input a desired preload and get the necessary mean torque and stresses. The Navy hands it out for free to anyone that wants it. My one mule consulting outfit heads the Beta test program for the new 32 bit Windows version, which we will have out soon and which also will be given away free to anyone that wants it, once it's ready to release into the wild. As a Navy employee I headed Beta testing for the three previous versions, so PC bolts is sort of the crown jewel of the contributions I've been able to make to solving fastener problems; that, and a whole bunch of friction coefficient testing that went into the lubricant library of PC Bolts, and which led me to take note of your rod installation instruction.
11. Oh yeah, one other little item; all of the above discussions assume that when the designer chose the fastener size and alloy and the preload we are going to apply to the fastener, he/she had a reasonably good idea of what the worst case in service load will be and chose the preload to be well above that load after short term preload loss and a reasonable profile for long term preload loss. If we blew any portion of that little set of choices, we have the potential for cyclic load reversal and good old fatigue failure. Short term preload loss can run anywhere from 5% to 40%, though proper tightening procedures will make it unlikely that you will ever see the latter figure. Our preferred approach is to take the max in service load and select a preload 4 times that amount, so that you can lose 50% of your initial assembly preload to any combination of short and long term preload loss and still have twice the preload needed to do the job - crude, but it works. Most bolted joints work well because their design is conservative (if nobody goofed), and hence they are very forgiving of minor errors, abuse and neglect. Of course, the more refined your design assumptions are from a standpoint of test results or other prior experience with a specific application on which to base your decisions, the closer you can cut it if space or weight considerations intrude, without putting the joint's integrity in danger. In combatant aircraft design, there are bolted joints that are subject to fatigue failure in order to reduce size to fit in the space available. With fatigue life to failure in a cyclic range that corresponds to 12 to 18 months of service, the aircraft maintenance plan calls for automatic replacement every 6 months. That's a valid solution for a fighter plane, but probably won't hack it for a passenger car.
12. I have repeatedly used the terms hi-shock and Undex in the discussions above. To give you an idea of the magnitude of the forces we're dealing with when Hi shock and explosion testing are invoked, civil engineering design for earthquake involves designing for accelerations primarily in the 9G to 11G range. Hi shock and Undex deal with G forces in the 450G range, give or take, and greater.
Wednesday, October 29, 2003
Are you Unbalanced?
By Tom Cole
Several years ago when we were developing our crankcase ventilation system for the Tecumseh Star engine, I got some seat-of -the-pants experience with the value and need of crankshaft balancing. I was accustomed to driving a 3” bore, 3” stroke test engine, which had one of our billet crankshafts in it. The 3x3 crank had been balanced, and ran very smooth considering it was producing about twice as much horsepower as the Star. The Star’s crank/rod/piston setup had not been balanced, and the difference took me by surprise. As I made 8000 rpm, my vision was so blurred from the engine’s vibration that I had to slow down to see the turn. My teeth and ribs felt like they were banging together and after my 15 lap stint at this ¼ mile asphalt oval, I was not interested in driving any more that day. My body was directly reporting to me the increased pain and fatigue that an unbalanced crankshaft can do to an engine and driver.
Clarence Clark is a friend of mine and he is what I would call an engine-building guru. For many years, Clarence’s company rebuilt the engines for the world’s largest fleet of racecars, the United Parcel Service. He has since traveled the country doing seminars for rebuilding supply companies like Goodson and Cobra Products. When I told Clarence about my experience, he went over to his tool box and produced a metal “H” which was made out of five, six inch long 3/8” metal pipes and two 3/8” pipe “T’s”. He handed me this contraption and told me to gently hold the cross pipe in one hand like an axle and spin it. Everything was fairly in balance and it spun easily making five or six revolutions. He then removed one of the four, six-inch uprights from the “H” and said “now it’s out of balance, spin it again.” I did and it only made one revolution! When I tried to spin it real hard, it only made two revolutions. This was an example of the need of both Force and Couple or “Dual Plane” balancing of a crankshaft. He told me one of the most interesting things about an out-of-balance crank or cam is that this tendency to stop (or inertia) increases with RPM so it requires more and more horsepower to obtain the same RPM as a balanced setup! It is hard to believe that we spend so much time and money on carburetors, valves, porting, flow and displacement and many of us ignore or are unaware of such a power robbing aspect of an engine.
Force and Couple balancing are technical terms used that really just mean top-to-bottom and side-to-side balance and are commonly referred to as “Dual Plane Balancing.” They are easily illustrated with a little history in tire balancing. Many years ago, when car tires were balanced, the rim and tire assembly was mounted on a shaft and then placed on a frame with the shaft resting in a ball bearing V fixture. The tire assembly would then rotate around until the heavy part came to rest at 6 o’clock. A wheel weight was then placed at 12 o’clock on one side of the tire. You kept adjusting this weight until the tire would not move regardless of how you repositioned it in the V fixture. This process is called, Static, or Force balancing and it is the method that is used on most kart and Jr. Dragster wheels today. Later, an improvement was made in this process by splitting the weight and putting half on the inside of the rim and half on the outside. This was the first form of Couple balancing in the tire industry. As years went by the advent of much wider tires came into being, so the need for more accurate Couple balancing increased because the part of the tire that was out of balance was often further from the centerline of the tire. And since one side of the tire could weigh more than the other, it became necessary to be more precise than to just split the weight in half to achieve optimum balance. We now have electronic tire balancers that spin the tire and wheel and calculate the amount of weight needed to Force balance. Then it calculates what amount goes on the inside of the rim and what goes on the outside. This is Dual Plane balancing of a tire.
The crankshaft of an engine has counterweights to dynamically offset the effect that the movement of the reciprocating mass has on the rotating mass of the crankshaft. The reciprocating mass is a percentage of the total mass of the top half of the connecting rod, the piston, wristpin, rings and circlips. The percentage used comes from a chart, which is calculated from expected RPM and stroke length. The rotating mass is the total mass of the bottom half of the connecting rod, the rod bolts, washers and bearings. To balance a single cylinder crankshaft, a bob weight is attached to the journal of the crankshaft that represents 100 % of the rotating mass and the percentage of the reciprocating mass from the chart. This assembly is then placed on the ball bearing V’s of our Stewart-Warner crankshaft-balancing machine and spun up to the expected operating RPM setting. The balancer can read the change in weight on both sides of the crankshaft and precisely tell the operator the amount of weight that needs to be added or removed from the left or right counterweights to achieve optimum dual plane balancing.
Balancing a crankshaft in itself does not produce more horsepower; it improves output potential by eliminating or greatly reducing a non-productive use of horsepower. It also helps prolong the life of an engine by reducing damaging vibration. In any situation where some grinding of the crankshaft or modification of the piston is permitted, it is a largely untapped source of performance gain.
Several years ago when we were developing our crankcase ventilation system for the Tecumseh Star engine, I got some seat-of -the-pants experience with the value and need of crankshaft balancing. I was accustomed to driving a 3” bore, 3” stroke test engine, which had one of our billet crankshafts in it. The 3x3 crank had been balanced, and ran very smooth considering it was producing about twice as much horsepower as the Star. The Star’s crank/rod/piston setup had not been balanced, and the difference took me by surprise. As I made 8000 rpm, my vision was so blurred from the engine’s vibration that I had to slow down to see the turn. My teeth and ribs felt like they were banging together and after my 15 lap stint at this ¼ mile asphalt oval, I was not interested in driving any more that day. My body was directly reporting to me the increased pain and fatigue that an unbalanced crankshaft can do to an engine and driver.
Clarence Clark is a friend of mine and he is what I would call an engine-building guru. For many years, Clarence’s company rebuilt the engines for the world’s largest fleet of racecars, the United Parcel Service. He has since traveled the country doing seminars for rebuilding supply companies like Goodson and Cobra Products. When I told Clarence about my experience, he went over to his tool box and produced a metal “H” which was made out of five, six inch long 3/8” metal pipes and two 3/8” pipe “T’s”. He handed me this contraption and told me to gently hold the cross pipe in one hand like an axle and spin it. Everything was fairly in balance and it spun easily making five or six revolutions. He then removed one of the four, six-inch uprights from the “H” and said “now it’s out of balance, spin it again.” I did and it only made one revolution! When I tried to spin it real hard, it only made two revolutions. This was an example of the need of both Force and Couple or “Dual Plane” balancing of a crankshaft. He told me one of the most interesting things about an out-of-balance crank or cam is that this tendency to stop (or inertia) increases with RPM so it requires more and more horsepower to obtain the same RPM as a balanced setup! It is hard to believe that we spend so much time and money on carburetors, valves, porting, flow and displacement and many of us ignore or are unaware of such a power robbing aspect of an engine.
Force and Couple balancing are technical terms used that really just mean top-to-bottom and side-to-side balance and are commonly referred to as “Dual Plane Balancing.” They are easily illustrated with a little history in tire balancing. Many years ago, when car tires were balanced, the rim and tire assembly was mounted on a shaft and then placed on a frame with the shaft resting in a ball bearing V fixture. The tire assembly would then rotate around until the heavy part came to rest at 6 o’clock. A wheel weight was then placed at 12 o’clock on one side of the tire. You kept adjusting this weight until the tire would not move regardless of how you repositioned it in the V fixture. This process is called, Static, or Force balancing and it is the method that is used on most kart and Jr. Dragster wheels today. Later, an improvement was made in this process by splitting the weight and putting half on the inside of the rim and half on the outside. This was the first form of Couple balancing in the tire industry. As years went by the advent of much wider tires came into being, so the need for more accurate Couple balancing increased because the part of the tire that was out of balance was often further from the centerline of the tire. And since one side of the tire could weigh more than the other, it became necessary to be more precise than to just split the weight in half to achieve optimum balance. We now have electronic tire balancers that spin the tire and wheel and calculate the amount of weight needed to Force balance. Then it calculates what amount goes on the inside of the rim and what goes on the outside. This is Dual Plane balancing of a tire.
The crankshaft of an engine has counterweights to dynamically offset the effect that the movement of the reciprocating mass has on the rotating mass of the crankshaft. The reciprocating mass is a percentage of the total mass of the top half of the connecting rod, the piston, wristpin, rings and circlips. The percentage used comes from a chart, which is calculated from expected RPM and stroke length. The rotating mass is the total mass of the bottom half of the connecting rod, the rod bolts, washers and bearings. To balance a single cylinder crankshaft, a bob weight is attached to the journal of the crankshaft that represents 100 % of the rotating mass and the percentage of the reciprocating mass from the chart. This assembly is then placed on the ball bearing V’s of our Stewart-Warner crankshaft-balancing machine and spun up to the expected operating RPM setting. The balancer can read the change in weight on both sides of the crankshaft and precisely tell the operator the amount of weight that needs to be added or removed from the left or right counterweights to achieve optimum dual plane balancing.
Balancing a crankshaft in itself does not produce more horsepower; it improves output potential by eliminating or greatly reducing a non-productive use of horsepower. It also helps prolong the life of an engine by reducing damaging vibration. In any situation where some grinding of the crankshaft or modification of the piston is permitted, it is a largely untapped source of performance gain.
Thursday, May 8, 2003
Setting Ignition Timing
By Tom Cole
Almost every day someone calls or emails us asking how to set the ignition timing on their engine. It is an important topic because as little as one degree can be the difference between an engine that runs up front and an engine that sputters and pops its way to last place. In this article, I am going to describe what I believe to be the most accurate and reliable method to set the timing on a Briggs and Stratton™ Engine. If you are using an ARC adjustable hub flywheel, begin by setting the hub index mark in the middle of the degree marks on the aluminum body. This will give you the maximum amount of adjustability after you set the timing based on the cam manufacturer’s specifications. The adjustable hub gives you an “at the track” advantage, because it allows you to easily advance or retard the ignition timing to tune for variable conditions.
The first thing you must do to set the timing is to identify the exact position of the flywheel’s magnet in relation to the coil when ignition occurs. To do this right, you need a plain old induction timing light and a car battery. Some folks will tell you about aligning the trailing edge of the magnet with the center of the little button just in front of the left leg of the coil, and for the most part, they are correct. But there is no such button on the Animal coil and factors such as coil gap make this method only a close approximation.
Below is the technique I use to find the exact trigger point using a timing light:
You are now ready to set the timing. Truthfully, it is more accurate to set the timing by fixing the piston at a measured distance before it reaches its highest point on the compression stroke. But, although everyone knows the “in the hole” distance for a pure stock setup, (30deg is .2115”) it is difficult to calculate the distance needed by different length rods since the distance traveled by the piston per degree of rotation varies with rod length and/or stroke. You can calculate it, but it just really isn’t worth the effort.
So, since the cam manufacturers generally provide you with a recommended ignition timing expressed in degrees before top dead center (BTDC) of the compression stroke, it is going to be best to set your timing using a degree wheel. (This is a good time to degree your cam too) On to the dreaded degree wheel…
Almost every day someone calls or emails us asking how to set the ignition timing on their engine. It is an important topic because as little as one degree can be the difference between an engine that runs up front and an engine that sputters and pops its way to last place. In this article, I am going to describe what I believe to be the most accurate and reliable method to set the timing on a Briggs and Stratton™ Engine. If you are using an ARC adjustable hub flywheel, begin by setting the hub index mark in the middle of the degree marks on the aluminum body. This will give you the maximum amount of adjustability after you set the timing based on the cam manufacturer’s specifications. The adjustable hub gives you an “at the track” advantage, because it allows you to easily advance or retard the ignition timing to tune for variable conditions.
The first thing you must do to set the timing is to identify the exact position of the flywheel’s magnet in relation to the coil when ignition occurs. To do this right, you need a plain old induction timing light and a car battery. Some folks will tell you about aligning the trailing edge of the magnet with the center of the little button just in front of the left leg of the coil, and for the most part, they are correct. But there is no such button on the Animal coil and factors such as coil gap make this method only a close approximation.
Below is the technique I use to find the exact trigger point using a timing light:
- Install only the crank, its bearings and its timing gear in the engine block and put on the side cover.
- Install the sheet metal guard that goes behind the flywheel.
- Install the flywheel with a standard key on the crank and snug it up with the starter nut. (No need to torque it, you are going to take it back off)
- Install the coil with the proper gap. (It will not be removed so tighten it up)
- Attach a spark plug to the plug wire and tape it to the block so as to ground it and create a spark.
- With the magnet at 12 o’clock under the coil put a white line on the outside rim of the flywheel at about 3 o’clock. This line needs to be plainly visible when looking at the block from the front.
- Attach the timing light to the battery and clamp the induction lead on the plug wire. Be careful that all wires are away from the flywheel.
- With the timing light pointing at the front of the block, turn the crankshaft clockwise with a drill or starter and you will see the timing strobe light up the white line. You need to spin it faster and more consistently than possible with a pull starter because a magneto has a retarding effect at higher rpm, and you want to compensate.
- While the strobe is flashing and the flywheel is spinning, make a white mark on the sheet metal guard or block that aligns with the white mark shown by the timing light on the flywheel. You can now place the magnet exactly where the spark is triggered when and if you ever remove the flywheel. BUT, if you move the coil or the metal guard, you have to start all over again.
- Remove the flywheel, side cover and crank, and install the piston, rod, crank, cam, etc (leaving off the cylinder head) getting to the point where you are ready to install the flywheel and set the timing.
You are now ready to set the timing. Truthfully, it is more accurate to set the timing by fixing the piston at a measured distance before it reaches its highest point on the compression stroke. But, although everyone knows the “in the hole” distance for a pure stock setup, (30deg is .2115”) it is difficult to calculate the distance needed by different length rods since the distance traveled by the piston per degree of rotation varies with rod length and/or stroke. You can calculate it, but it just really isn’t worth the effort.
So, since the cam manufacturers generally provide you with a recommended ignition timing expressed in degrees before top dead center (BTDC) of the compression stroke, it is going to be best to set your timing using a degree wheel. (This is a good time to degree your cam too) On to the dreaded degree wheel…
- Using coarse grit sand paper, rough up the tapered part of the crank and then make sure it is clean. If you are un-willing to lick it, it isn’t clean enough!
- Similarly, rough up and clean the inside of the hole in the flywheel hub.
- Set the crank so, according to the degree wheel, you are at the cam manufacturers specified degrees of ignition timing before the piston reaches the highest point of its compression stroke BTDC.
- Put a few drops of Loctite™ on the tapered part of the crank and install the flywheel with the white timing lines aligned. DO NOT USE A KEY! Keys do very little to prevent a flywheel from spinning, and they will hinder your accuracy.
- Carefully tighten down the starter nut. Check and recheck the timing several times as you tighten to be sure that the white timing lines are still aligned and that the degree wheel is still where it is supposed to be. Then tighten the starter nut, A LOT. You are shooting for tight enough to hold it together, but not so tight as to split the flywheel.
- If you have an ARC adjustable hub flywheel, you can fine-tune your ignition timing using a dyno or a stopwatch and a Digatron CHT/Tach at the track.
Friday, April 12, 2002
Is 1° Costing You The Win?
By: Tom Cole
Do you REALLY know if advancing or retarding the ignition timing of your engine 1° will increase horsepower? I think most of us can agree that there is a good chance that we would find some improvement, if we checked. How easy is it to experiment with your motor? As I have said before, I am no expert. I am just a father like many of you who likes to build his son’s racing motors. I love the feeling when a motor I built takes the checkered flag, and I have felt the heat and disappointment from the driver (and the wife) when it is obvious that MY poor performing engine cost us the race.
One fact I have learned is that perfect timing is when your motor runs at the proper temperature, responds the way you want it to respond, and you win races. There are no magic numbers or exact calculations that can tell you where timing needs to be. There are only beginning reference points. Just like the carburetor, timing must be tuned to the conditions and variables specific to each motor, track, driver, and kart. Throw in the altitude and the weather conditions, and you can see that the ability to adjust timing quickly and accurately is something that can set you apart from the crowd.
Because the degree wheel infuriates me, and offset keys are a joke, I set ignition timing by aligning the trailing edge of the flywheel magnet with the middle of the coil trigger button while the top of the piston (before top dead center) is at a measured depth from the top of the cylinder. The depth of the piston in the hole varies with many factors, but your cam manufacturer will usually give you a good starting point. (30° BTDC = .2115” in the hole when popup is 0.000” on a stock engine) Once I have everything aligned, I carefully hand tighten the starter nut. Then I recheck my measurements, torque the starter nut down to somewhere around 1,000,000 in./lbs. (I give it all I’ve got with hand tools) and recheck my measurements again. Also note that I do not use an offset key. I lap the flywheel and crankshaft together and then apply Locktite™ to the taper (not the threads) on the crankshaft before installing the flywheel. Now I reassemble the motor, mount it up to the kart, and give it a test run. This whole process takes about 30 minutes.
Now how many of us are going to take that motor and go through that same process at least three more times to check and see if +or- 1° makes a noticeable difference? The method I use to set the timing has a margin of error of at least +or- 2° because I don’t know the exact firing point of the coil trigger button. I know I need to experiment, but what a pain! And how much stress am I putting on the stock flywheel each time I knock it off and torque it back down. Haven’t they been known to fly apart? And how well is that flimsy stock animal flywheel going to hold up? I have heard of engine builders breaking them when they are first installed!
It wasn’t until I faced all this trouble with that touchy little blue-plate slapper-cam motor that the value of ARC’s billet finned flywheel with its adjustable timing hub really came into view. I can literally be finished getting the motor’s ignition timing right with the ARC flywheel, before I can test the first timing change on a stock or fixed hub flywheel. How much time do you have at a race between practice and the first heat? I am too busy changing oil, cleaning and prepping tires, and making chassis adjustments to tear down a motor.
Do you REALLY know if advancing or retarding the ignition timing of your engine 1° will increase horsepower? I think most of us can agree that there is a good chance that we would find some improvement, if we checked. How easy is it to experiment with your motor? As I have said before, I am no expert. I am just a father like many of you who likes to build his son’s racing motors. I love the feeling when a motor I built takes the checkered flag, and I have felt the heat and disappointment from the driver (and the wife) when it is obvious that MY poor performing engine cost us the race.
One fact I have learned is that perfect timing is when your motor runs at the proper temperature, responds the way you want it to respond, and you win races. There are no magic numbers or exact calculations that can tell you where timing needs to be. There are only beginning reference points. Just like the carburetor, timing must be tuned to the conditions and variables specific to each motor, track, driver, and kart. Throw in the altitude and the weather conditions, and you can see that the ability to adjust timing quickly and accurately is something that can set you apart from the crowd.
Because the degree wheel infuriates me, and offset keys are a joke, I set ignition timing by aligning the trailing edge of the flywheel magnet with the middle of the coil trigger button while the top of the piston (before top dead center) is at a measured depth from the top of the cylinder. The depth of the piston in the hole varies with many factors, but your cam manufacturer will usually give you a good starting point. (30° BTDC = .2115” in the hole when popup is 0.000” on a stock engine) Once I have everything aligned, I carefully hand tighten the starter nut. Then I recheck my measurements, torque the starter nut down to somewhere around 1,000,000 in./lbs. (I give it all I’ve got with hand tools) and recheck my measurements again. Also note that I do not use an offset key. I lap the flywheel and crankshaft together and then apply Locktite™ to the taper (not the threads) on the crankshaft before installing the flywheel. Now I reassemble the motor, mount it up to the kart, and give it a test run. This whole process takes about 30 minutes.
Now how many of us are going to take that motor and go through that same process at least three more times to check and see if +or- 1° makes a noticeable difference? The method I use to set the timing has a margin of error of at least +or- 2° because I don’t know the exact firing point of the coil trigger button. I know I need to experiment, but what a pain! And how much stress am I putting on the stock flywheel each time I knock it off and torque it back down. Haven’t they been known to fly apart? And how well is that flimsy stock animal flywheel going to hold up? I have heard of engine builders breaking them when they are first installed!
It wasn’t until I faced all this trouble with that touchy little blue-plate slapper-cam motor that the value of ARC’s billet finned flywheel with its adjustable timing hub really came into view. I can literally be finished getting the motor’s ignition timing right with the ARC flywheel, before I can test the first timing change on a stock or fixed hub flywheel. How much time do you have at a race between practice and the first heat? I am too busy changing oil, cleaning and prepping tires, and making chassis adjustments to tear down a motor.
Thursday, January 17, 2002
Rolling Resistance
By: Tom Cole Date: January 17, 2002
Every rear wheel drive car or truck on the road or racetrack uses tapered roller bearings in their front hubs. The reasons for this are simple, less friction and combined radial (perpendicular to the axial) and thrust (parallel to the axial) load capacity. So why are kart racers using ball bearings which are designed only for radial loads on their karts? I can understand why Jr. Drag racers don’t care about thrust load. They just go fast in a straight line. But kart racers go fast and turn fast. Usually, the one who gets through the turns the fastest wins. So again, why are so many kart racers using ball bearings? Endplay! Somebody has actually sold people on the idea that karts with tapered roller bearing hubs have so much endplay that you can’t set the toe-in. What a pile! How is it that those NASCAR and F1 boys can set the alignment on their cars? Never mind that endplay tolerance for a wheel with tapered roller bearings is only .002” (that’s half the thickness of a piece of notebook paper). How precise are the widths of the inner and outer races of a .65-cent ball bearing? How precise is the relationship between the inner and outer races? Are they parallel? Are they precisely on the same plane? How precise is the bore of the hub? Is the center sleeve exactly long enough? Well guess what! None of this matters if you use a properly installed tapered roller bearing. The guy beating you doesn’t give a rip how the wheel rolls on the rack. How the wheels roll on the track, in the turn under thrust load and down the straightaway under radial load is what matters in wheel bearings. And he is not going to tell you why he’s beating you.
All ball bearings roll on a “point” of contact between the ball and the races. A tapered roller bearing distributes the load over the length of the roller in a “line” of contact. This greatly reduces the friction coefficient allowing tapered roller bearings of the same diameter as comparable ball bearings to carry a greater load and achieve a much greater fatigue life. Simply put, it will roll easier than a ball bearing. The angle of the races along with the taper of the bearing rods allows a tapered roller bearing hub to roll equally well in the turns or on the straight-aways. Standard ball bearings DO NOT roll as well through a turn as they do down the straightaway, and they do not handle the demands of a kart racer as well as a tapered roller bearing.
Every rear wheel drive car or truck on the road or racetrack uses tapered roller bearings in their front hubs. The reasons for this are simple, less friction and combined radial (perpendicular to the axial) and thrust (parallel to the axial) load capacity. So why are kart racers using ball bearings which are designed only for radial loads on their karts? I can understand why Jr. Drag racers don’t care about thrust load. They just go fast in a straight line. But kart racers go fast and turn fast. Usually, the one who gets through the turns the fastest wins. So again, why are so many kart racers using ball bearings? Endplay! Somebody has actually sold people on the idea that karts with tapered roller bearing hubs have so much endplay that you can’t set the toe-in. What a pile! How is it that those NASCAR and F1 boys can set the alignment on their cars? Never mind that endplay tolerance for a wheel with tapered roller bearings is only .002” (that’s half the thickness of a piece of notebook paper). How precise are the widths of the inner and outer races of a .65-cent ball bearing? How precise is the relationship between the inner and outer races? Are they parallel? Are they precisely on the same plane? How precise is the bore of the hub? Is the center sleeve exactly long enough? Well guess what! None of this matters if you use a properly installed tapered roller bearing. The guy beating you doesn’t give a rip how the wheel rolls on the rack. How the wheels roll on the track, in the turn under thrust load and down the straightaway under radial load is what matters in wheel bearings. And he is not going to tell you why he’s beating you.
All ball bearings roll on a “point” of contact between the ball and the races. A tapered roller bearing distributes the load over the length of the roller in a “line” of contact. This greatly reduces the friction coefficient allowing tapered roller bearings of the same diameter as comparable ball bearings to carry a greater load and achieve a much greater fatigue life. Simply put, it will roll easier than a ball bearing. The angle of the races along with the taper of the bearing rods allows a tapered roller bearing hub to roll equally well in the turns or on the straight-aways. Standard ball bearings DO NOT roll as well through a turn as they do down the straightaway, and they do not handle the demands of a kart racer as well as a tapered roller bearing.
Monday, March 20, 2000
STROKER ROD DESIGN
By: Carl Amundsen Date: March 20, 2000
Evidently a lot of people have read and paid attention to our editorial " Oil Clearance is Not a Myth But a Calculation ". We appreciate all the calls and hope it has helped save some engines.
A constant question that keeps coming up is about the bore size on the stroker rod for the .875 (7/8) journal crankshaft.
Why are the rod bores on some brands of stroker rods as much as .004" out of round, right out of the box?
These rods all have angled or what is termed "splayed rod bolts". This is a necessary design because of the amount of stroke in the crankshaft and the need for clearance at the camshaft and inside the block.
The problem occurs when the parting surfaces at the rod and cap are not 90 degrees to the rod bolts. On the illustration below the right rod bolt enters at 15 degrees and the left rod bolt enters at 7 1/2 degrees. The parting surfaces of the rod and cap are at 75 degrees and 82 1/2 degrees respectively to the rod bolt.
It is easy to see from the illustration that the seat area under the rod bolt and the seat area of the cap will be at an angle to the parting surfaces of the rod assembly. This is where the problem begins.
During the manufacturing process the rod and cap are assembled together and the bolts are tightened to a specific torque value. The final machining of the bore now takes place, and comes out on size and round. The next thing that happens is that the rod bolts are removed and new ones are installed. Here is your problem. Because of the adverse angle at which the bolts enter the seam of the rod, they pull the bore out of round. It will always come out egg shaped, and the more you play with the torque values the worse it gets. This is not a one out of ten problem this will happen 100 times out of 100 times. It can be bad enough to lock the motor up. Bet on it. THE PROBLEM IS IN THE DESIGN.
To add insult to injury, some rods are designed with the rod bolts exposed to the bore. This just reduces the integrity of the bore and weakens the seam area. If that's not enough, you must grind some clearance into the back side of the bearings to accommodate the rod bolts.
THE FOLLOWING IS AN ILLUSTRATION OF THE ARC DESIGNED STROKER ROD, NOTICE THAT THE BOLTS AND THEIR PARTING EDGES ARE PERPENDICULAR TO ONE ANOTHER. THIS DESIGN ELIMINATES THE PROBLEMS CAUSED BY THE ANGLED BOLT TO PARTING EDGE SITUATION. OUR WAY STILL ISN'T PERFECT, BUT IT IS A HUGE STEP IN THE RIGHT DIRECTION.
Evidently a lot of people have read and paid attention to our editorial " Oil Clearance is Not a Myth But a Calculation ". We appreciate all the calls and hope it has helped save some engines.
A constant question that keeps coming up is about the bore size on the stroker rod for the .875 (7/8) journal crankshaft.
Why are the rod bores on some brands of stroker rods as much as .004" out of round, right out of the box?
These rods all have angled or what is termed "splayed rod bolts". This is a necessary design because of the amount of stroke in the crankshaft and the need for clearance at the camshaft and inside the block.
It is easy to see from the illustration that the seat area under the rod bolt and the seat area of the cap will be at an angle to the parting surfaces of the rod assembly. This is where the problem begins.
During the manufacturing process the rod and cap are assembled together and the bolts are tightened to a specific torque value. The final machining of the bore now takes place, and comes out on size and round. The next thing that happens is that the rod bolts are removed and new ones are installed. Here is your problem. Because of the adverse angle at which the bolts enter the seam of the rod, they pull the bore out of round. It will always come out egg shaped, and the more you play with the torque values the worse it gets. This is not a one out of ten problem this will happen 100 times out of 100 times. It can be bad enough to lock the motor up. Bet on it. THE PROBLEM IS IN THE DESIGN.
To add insult to injury, some rods are designed with the rod bolts exposed to the bore. This just reduces the integrity of the bore and weakens the seam area. If that's not enough, you must grind some clearance into the back side of the bearings to accommodate the rod bolts.
THE FOLLOWING IS AN ILLUSTRATION OF THE ARC DESIGNED STROKER ROD, NOTICE THAT THE BOLTS AND THEIR PARTING EDGES ARE PERPENDICULAR TO ONE ANOTHER. THIS DESIGN ELIMINATES THE PROBLEMS CAUSED BY THE ANGLED BOLT TO PARTING EDGE SITUATION. OUR WAY STILL ISN'T PERFECT, BUT IT IS A HUGE STEP IN THE RIGHT DIRECTION.
Tuesday, February 8, 2000
OIL CLEARANCE IS NOT A MYTH IT IS A CALCULATION
By: Carl Amundsen Date:2/8/2000
Over the years I have had many opportunities to discuss oil clearance with engine builders. Some are building for their own use, some are small shops while some are large. In any event when the question is posed "HOW DO YOU KNOW WHAT YOUR OIL CLEARANCE IS?" the responses vary. Some can give a good mathematical response, some don't have a clue. The common denominator here is that everyone knows they have oil clearance.
A common phenomenon that occurs when a motor fails is that no one looks in the mirror for answers. The vast majority of the time it is blamed on an engine part failure. The connecting rod too many times is the favorite fugitive.
A LESSON IN MEASUREMENTS
A very dangerous but commonly used method of measuring is a dial caliper. The average dial caliper is super for measuring the thickness, inside diameter, and outside of an empty toilet paper roll. A SHOCKING STATEMENT , NO DOUBT! If you look through some catalogs you will find a dial caliper for $40.00 that has an advertised accuracy of +/- .008 . That is an error factor of 16 thousandths. Pay $200.00 for a digital caliper with advertised accuracy of +/- .0015, the margin of error is now 3 thousandths. An error factor of 1 thousandths can cost you a motor. Calipers are great tools for some things but not for the task at hand. WE MUST BE ABLE TO READ IN TENTHS OF THOUSANDTHS(.0001)
ON WITH THE LESSON
If you were to take a good 0 to 1 inch micrometer and measure the thickness of a page in the ARC catalog, you will find it to be five thousandths (.005) thick. This is very important to remember as you read on. Now feel the paper with your fingers and imagine somehow splitting one of these pages 50 times. One single sheet would now be 1 tenth of 1 thousandth of an inch thick (.0001). I hope I have your attention, because this is getting down so that you can't even feel the thickness. Just incase I am going to be confusing you with decimals please remember the following:
1.0000 = 1 inch0.1000 = 100 thousandths 0.0100 = 10 thousandths0.0010 = 1 thousandth0.0001 = 1/10 thousandth
It makes no difference who you buy your parts from, nor does it matter what brand they are, ARC included,they need to be checked and double checked. NOBODY IS PERFECT.
For example let's build a stock stroke Briggs racing engine. The following is the measurement specifications on the typical parts to be used:
Stock Briggs crank rod journal size......................... .998 +/- .0015Aftermarket Connecting Rod................................... 1.150 +/- .005Aftermarket Rod Bearing thickness........................ .075 +/- .003
ALL THE PARTS WE ARE USING ARE WITHIN THE MANUFACTURERS TOLERANCES.
You are going to use a medium weight oil and you are shooting for .0025
The Connecting Rod is on the small side................. 1.1495The Rod Bearing is on the big side.......................... .0753The Crankshaft is on the big side.............................. .9985
Let's put this motor together and go racing.
The Connecting Rod is small by .0005...................... 1.1495The Thickness of the Rod Bearingare big by .003(2 bearings) =.1506............................. - .1506Net Rod Bore size with Bearingsinstalled................................................................... . 9989The Crank Rod Journal is big by .0005..................... - .9985The Calculated oil Clearance = ............................... .0004We were shooting for ............................................... .0025
With a little luck this motor will crank up and run, as long as the motor is running at no load and a low RPM it may be OK for a while. The minute you go racing the lack of oil flow between the bearing and the crank journal will cause heat build up. The bearing will seize on the crank journal, break the rod and just make a mess of everything. BAD PARTS RIGHT ? WRONG !
Now let's build another motor and go the opposite way, still trying to achieve an oil clearance of .0025
The Connecting Rod is on the big side........................ 1.1505The Rod Bearing is on the small side.......................... .0747The Crankshaft is on the small side.............................. .9965
The Connecting Rod is big by .0005............................ 1.1505The Thickness of the Rod Bearingare small by .003(2 bearings) =.1494............................. - .1494Net Rod Bore size with Bearingsinstalled...................................................................... 1.0011The Crank Rod Journal is small by .0015..................... - .9965The Calculated oil Clearance = .................................. .0046We were shooting for .................................................. .0025
This motor is going to run, but what is going to happen here is: The bearing is going to get pounded at the top and bottom of the rod bore, because there is an air gap between the two surfaces. The oil is not thick enough to prevent this from happening. This is going to convert the rod bore into the shape of an egg standing on end. The crankshaft will now start to loose it's round shape and wear.
EVERY MINUTE IT RUNS, THESE PARTS WILL INCREASE THE OIL CLEARANCE UNTIL IT EXPLODES. HOW LONG WILL IT LAST? I CAN'T SAY EXACTLY, BUT NOT TO LONG!
The crying shame here is that everyone will point their fingers at the connecting rod, bearings and/or crankshaft as the culprit.
THE MOST OVERLOOKED AREA IN ENGINE BUILDING IS AS FOLLOWS:You have a motor that has run for many months , but you notice you're getting a little blow-by and it could probably use a set of rings. It needs freshening up so we tear it down. Every thing looks great so we touch the bore with a hone and put a new set of rings in. Back to the races. If it ain't broke don't fix it. STOP! Under the most ideal conditions engine parts will change with use. Bearings can look good but will wear, connecting rod bores will change shapes and crank journals will wear and rod bolts will stretch and fatigue. It is just as important now as ever to check the dimensions of the parts. Don't go to sleep.
Some may think there is some deep mystery behind the term oil clearance. The reason is, everyone has a different opinion as to what it should be. As a rule of thumb, it can be anywhere between .0015 and .0035 and be in the ball park. If the clearance is on the low side, use thin oil, heavy oil will not work. If clearance is on the high side , use a heavier oil, thin oil will not work. How can you make a judgment on the oil to use unless you know what your oil clearance is? THERE IS NO MAGIC, YOU CANNOT KNOW WHAT YOUR OIL CLEARANCE IS UNLESS YOU MEASURE THE PARTS THAT AFFECT IT.
DEFINITION OF WHAT OIL CLEARANCE IS:
"The distance between two very smooth moving surfaces that will allow oil to be present at all times, coating both parts with a film of oil so that metal on metal contact never happens. It must be small enough to retain enough oil and large enough to allow a fresh cool supply to move through every microsecond."
If you do not check your parts and build 4 motors and 3 of them seem to have a pretty good life span you are lucky. If one fails out of the gate or shortly thereafter, shame on you. It takes less than 15 minutes to check and measure all the parts in a motor and the benefits are fantastic. It is possible to build an engine that will run until the cows come home, or something like that. There are many engine builders out there that do a fantastic job in this area, but on the other hand there are more that don't.
ASSUME NOTHING, BELIEVE NOTHING AND LAST BUT NOT LEAST, CHECK EVERYTHING.
I hope we have given you an insight to one reason for engine failures. If you need us, give us a call, it's all free. 1-800-521-3560
Over the years I have had many opportunities to discuss oil clearance with engine builders. Some are building for their own use, some are small shops while some are large. In any event when the question is posed "HOW DO YOU KNOW WHAT YOUR OIL CLEARANCE IS?" the responses vary. Some can give a good mathematical response, some don't have a clue. The common denominator here is that everyone knows they have oil clearance.
A common phenomenon that occurs when a motor fails is that no one looks in the mirror for answers. The vast majority of the time it is blamed on an engine part failure. The connecting rod too many times is the favorite fugitive.
A LESSON IN MEASUREMENTS
A very dangerous but commonly used method of measuring is a dial caliper. The average dial caliper is super for measuring the thickness, inside diameter, and outside of an empty toilet paper roll. A SHOCKING STATEMENT , NO DOUBT! If you look through some catalogs you will find a dial caliper for $40.00 that has an advertised accuracy of +/- .008 . That is an error factor of 16 thousandths. Pay $200.00 for a digital caliper with advertised accuracy of +/- .0015, the margin of error is now 3 thousandths. An error factor of 1 thousandths can cost you a motor. Calipers are great tools for some things but not for the task at hand. WE MUST BE ABLE TO READ IN TENTHS OF THOUSANDTHS(.0001)
ON WITH THE LESSON
If you were to take a good 0 to 1 inch micrometer and measure the thickness of a page in the ARC catalog, you will find it to be five thousandths (.005) thick. This is very important to remember as you read on. Now feel the paper with your fingers and imagine somehow splitting one of these pages 50 times. One single sheet would now be 1 tenth of 1 thousandth of an inch thick (.0001). I hope I have your attention, because this is getting down so that you can't even feel the thickness. Just incase I am going to be confusing you with decimals please remember the following:
1.0000 = 1 inch0.1000 = 100 thousandths 0.0100 = 10 thousandths0.0010 = 1 thousandth0.0001 = 1/10 thousandth
It makes no difference who you buy your parts from, nor does it matter what brand they are, ARC included,they need to be checked and double checked. NOBODY IS PERFECT.
For example let's build a stock stroke Briggs racing engine. The following is the measurement specifications on the typical parts to be used:
Stock Briggs crank rod journal size......................... .998 +/- .0015Aftermarket Connecting Rod................................... 1.150 +/- .005Aftermarket Rod Bearing thickness........................ .075 +/- .003
ALL THE PARTS WE ARE USING ARE WITHIN THE MANUFACTURERS TOLERANCES.
You are going to use a medium weight oil and you are shooting for .0025
The Connecting Rod is on the small side................. 1.1495The Rod Bearing is on the big side.......................... .0753The Crankshaft is on the big side.............................. .9985
Let's put this motor together and go racing.
The Connecting Rod is small by .0005...................... 1.1495The Thickness of the Rod Bearingare big by .003(2 bearings) =.1506............................. - .1506Net Rod Bore size with Bearingsinstalled................................................................... . 9989The Crank Rod Journal is big by .0005..................... - .9985The Calculated oil Clearance = ............................... .0004We were shooting for ............................................... .0025
With a little luck this motor will crank up and run, as long as the motor is running at no load and a low RPM it may be OK for a while. The minute you go racing the lack of oil flow between the bearing and the crank journal will cause heat build up. The bearing will seize on the crank journal, break the rod and just make a mess of everything. BAD PARTS RIGHT ? WRONG !
Now let's build another motor and go the opposite way, still trying to achieve an oil clearance of .0025
The Connecting Rod is on the big side........................ 1.1505The Rod Bearing is on the small side.......................... .0747The Crankshaft is on the small side.............................. .9965
The Connecting Rod is big by .0005............................ 1.1505The Thickness of the Rod Bearingare small by .003(2 bearings) =.1494............................. - .1494Net Rod Bore size with Bearingsinstalled...................................................................... 1.0011The Crank Rod Journal is small by .0015..................... - .9965The Calculated oil Clearance = .................................. .0046We were shooting for .................................................. .0025
This motor is going to run, but what is going to happen here is: The bearing is going to get pounded at the top and bottom of the rod bore, because there is an air gap between the two surfaces. The oil is not thick enough to prevent this from happening. This is going to convert the rod bore into the shape of an egg standing on end. The crankshaft will now start to loose it's round shape and wear.
EVERY MINUTE IT RUNS, THESE PARTS WILL INCREASE THE OIL CLEARANCE UNTIL IT EXPLODES. HOW LONG WILL IT LAST? I CAN'T SAY EXACTLY, BUT NOT TO LONG!
The crying shame here is that everyone will point their fingers at the connecting rod, bearings and/or crankshaft as the culprit.
THE MOST OVERLOOKED AREA IN ENGINE BUILDING IS AS FOLLOWS:You have a motor that has run for many months , but you notice you're getting a little blow-by and it could probably use a set of rings. It needs freshening up so we tear it down. Every thing looks great so we touch the bore with a hone and put a new set of rings in. Back to the races. If it ain't broke don't fix it. STOP! Under the most ideal conditions engine parts will change with use. Bearings can look good but will wear, connecting rod bores will change shapes and crank journals will wear and rod bolts will stretch and fatigue. It is just as important now as ever to check the dimensions of the parts. Don't go to sleep.
Some may think there is some deep mystery behind the term oil clearance. The reason is, everyone has a different opinion as to what it should be. As a rule of thumb, it can be anywhere between .0015 and .0035 and be in the ball park. If the clearance is on the low side, use thin oil, heavy oil will not work. If clearance is on the high side , use a heavier oil, thin oil will not work. How can you make a judgment on the oil to use unless you know what your oil clearance is? THERE IS NO MAGIC, YOU CANNOT KNOW WHAT YOUR OIL CLEARANCE IS UNLESS YOU MEASURE THE PARTS THAT AFFECT IT.
DEFINITION OF WHAT OIL CLEARANCE IS:
"The distance between two very smooth moving surfaces that will allow oil to be present at all times, coating both parts with a film of oil so that metal on metal contact never happens. It must be small enough to retain enough oil and large enough to allow a fresh cool supply to move through every microsecond."
If you do not check your parts and build 4 motors and 3 of them seem to have a pretty good life span you are lucky. If one fails out of the gate or shortly thereafter, shame on you. It takes less than 15 minutes to check and measure all the parts in a motor and the benefits are fantastic. It is possible to build an engine that will run until the cows come home, or something like that. There are many engine builders out there that do a fantastic job in this area, but on the other hand there are more that don't.
ASSUME NOTHING, BELIEVE NOTHING AND LAST BUT NOT LEAST, CHECK EVERYTHING.
I hope we have given you an insight to one reason for engine failures. If you need us, give us a call, it's all free. 1-800-521-3560
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