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ОглавлениеPISTONS, CONNECTING RODS AND CRANKSHAFTS
In this chapter, I go into more depth on power-producing factors that may have only been touched on briefly in previous books, and you may not have seen the piston tech I provide.
You could easily buy pistons that needed to be modified for clearance somewhere on the crown. The biggest difference you are likely to find between one piston and another is in the plug positioning. The piston’s spark trough, or flame ditch, can be up to 1/2 inch off from where the plug is. This needs to be checked, and in some cases redone so the flame path has a clear run to the rest of the chamber. Make sure no part of the piston crown inhibits the flame front passage; cutting the spark ditch is always a move in the right direction. However, the passage of the flame through the charge is complex and understanding what goes on takes a lot of testing to find what is needed.
Fig. 2.1. Short of a supercharger or nitrous oxide, tapping in to the big-block Chevy’s displacement potential via a stroker crank and bigger bores is the best route to high-torque output. However, there are many issues to be dealt with. Although most are minor, those that are not have to be dealt with in an appropriate manner for best results.
An area where much work is currently being done is in the design of the cylinder heads’ quench pad and its interaction with the piston crown. For years, it was assumed that the quench pad should be just a flat surface that the piston closely approaches. Now, this is proving to be not the case. But it is not only the pistons’ quench area that interacts with the head; some less than obvious aspects of the piston design also affect the airflow.
Due to the fact that a performance Chevy big-block piston needs to generate a high compression and accommodate a high valve lift, especially on the intake, it presents some issues related to the breathing ability of the cylinder as a whole. The issue with the valve cutouts is easy to identify, but there is another flow-inhibiting issue that is rarely appreciated. What you read in the following paragraphs may be your first introduction to flow-inhibiting factors for pistons.
Extra Cubes: How Effective Are They?
Cube utilization is of such importance to the success of a big-inch build that it is worth a serious (if short) mention here. The point is this: Unless the rest of the engine spec is reevaluated to reflect the needs of the increased displacement, the time and effort involved in garnering those extra inches are largely wasted.
When displacement has been increased the biggest and most influential change in the engine’s optimal spec to utilize those extra cubes will be in the cam and valvetrain department. (See Chapter 9, Camshafts and Valvetrain Events, and Chapter 10, Valvetrain Optimization, which detail how increased displacement affects optimal cam event timing and lift.)
Fig. 2.2. At first glance, these appear to be trick pistons that are ready to be installed. In reality, these need some crown reworking. With this work, a 10-hp gain is achieved.
The edge of the intake valve pocket needs attention and it’s the easiest area to take care of. As can be seen in Figure 2.4 the air entering the cylinder around the short-side turn during the overlap period runs into the wall of the valve cutout. This needs to be rectified as shown in Figure 2.5. This “piston porting” exercise is easy enough to appreciate if you spend a little time studying these figures. And it definitely delivers performance benefits.
However, one aspect that is mostly peculiar to 24-degree heads is far from intuitive. A pressure/flow distribution plot (Figure 2.6) of the intake port reveals that the busiest exit area around the valve is not, as is normally the case with most other two valve heads, toward the cylinder’s center. Measuring seat exit velocities with a vented valve shows that the busiest section of the port is often toward the shrouded side. Once this becomes known it is easier to see why, when the block is chamfered in this area, the response is a sizable amount of extra output. The block chamfering, however, starts to aid flow after the valve is around 0.150 inch and more off the seat.
Fig. 2.3. The high dome and deep valve pocket design of a Chevy big-block piston can inhibit the flow of gases into and out of the cylinder. Air travels into the cylinder following the path of the blue arrows. During the overlap period, air runs into the ridge on the plug side (left) and the edge of the valve pocket (right). Exhaust gases exit in the direction of the red arrow, and the ridge (adjacent to the spark trough) impedes this exhaust gas flow. The black shaded areas are the first locations that need attention.
Fig. 2.4. This computational fluid dynamics illustration shows air entering the cylinder during the overlap period. At this stage of the intake event, the air that exits the short side turn runs into the wall of the valve cutout. This reduces the effectiveness of the overlap period and typically reduces the volumetric efficiency more than it might be supposed.
Less obvious is that the flow pattern on the cylinder wall side of the port is spiraling past the edge of the bore shrouded intake valve. In effect, air is corkscrewing past the edge of the intake valve at about the 10 o’clock position, and during the overlap, the dome of a high-compression piston can block this flow. This suggests that you not only need to cut the top of the bore as discussed in Chapter 1, Displacement Decisions, but you should also find out if the piston dome can have any negative influence on the flow into the cylinder other than the effects of valve shrouding from the aspects indicated in Figure 2.3. From Figure 2.6, you can see that the flow into the cylinder does not follow a pattern that is by any means intuitive.
Fig. 2.5. Here are the two ways to relieve shrouding, which is caused by the valve pocket wall. The simple way is shown on the left and the more effective but tedious method is shown on the right.
Fig. 2.6. To appreciate what is going on here, first locate the ghosted image of the intake port and chamber and orientate that with the pressure differential contour lines around the valve. These contour lines show the pressure differential between the valve and valveseat on a 24-degree head as the valve progresses through its lift.
Although this appears to be a subject for Chapter 4, Cylinder Heads, the flow pattern developed has a strong influence on how the top of the bore and piston should be shaped. You need to recognize that the busiest area with the highest velocities occurs between the 9:00 and the 10:30 o’clock position. The edge of the bore and the piston dome can block flow in this region unless steps are taken to prevent it. The arrow indicates airflow through the port into the cylinder.
Fig. 2.7. Indicated here are the areas of a typical high-compression piston that need attention as far as valve pocket shrouding is concerned.
This flow test and the port probing just prior to the intake valve show an important flow pattern. As unlikely as it may seem, the flow corkscrews off the edge of the valve on the cylinder wall side of the port and then proceeds over the edge of the intake valve and into the cylinder. At least that is the way it would go if there were no obstructions. Because this flow pattern is generally unknown, piston domes rarely have a form that makes allowance for it. Depending on the height of the dome there is a potential 10 to 15 hp to be had by some subtle and some less than subtle reshaping.
The best piston crown shape to have is a flat one or one with a shallow dish in it. Unfortunately that usually results in a really undesirably low compression ratio unless the short-block has a lot of cubic inches. The first move is to address the edge of the piston’s intake valve pocket as per Figure 2.5, which shows the previously mentioned piston mod. From here on out the valve shrouding reduction moves are a little more subtle.
Fig. 2.8. This piston came out of a 900-hp bracket engine built by Throttle’s Performance. Although this engine ran very well, I knew there was more in it if the pistons were suitably reworked.
Fig. 2.9. A trough cut in the piston accommodates the spiral-flow pattern on the cylinder wall side of the port. The top edge of the trough needs to go under the valve head by about 0.100 inch and extend to the deck of the piston at the lower edge.
Fig. 2.10. This piston is nearing completion. The yellow arrow indicates the trough to accommodate the spiral flow seen at low and mid lift. The blue arrows to the right show the laid-back edges that inhibit flow during overlap. The blue arrows to the left show areas that have been lowered, so flow is improved to and from the area around the spark plug. The red arrow indicates the reworking location when bore chamfers on the block are used.
Fig. 2.11. These Mahle pistons have a crown shape that is on the way to emulating the recommended form. As such, they are a very effective piston right out of the box. In addition to a good crown and valve cutout form, these pistons come with 1.5-mm-wide compression rings, which typically have less bore drag than the 1/16-inch-wide rings.
Fig. 2.12. This JE piston is a classic example of why your big-block Chevy project should be focusing on as many cubes as possible. This piston, in a 572 with heads only minimally milled, delivered a 10:1 CR. The valve cutout illustrates just how little work there is to be done when a flat top or even a dished piston is used.
Fig. 2.13. For a drag-race-only application, gas porting through the piston crown is most often the preferred method to boost cylinder sealing.
The principle job of the top compression ring is to seal against the pressures experienced above it. To do this, it must have some radial load pressing it outward onto the cylinder walls. A relatively high-compressive preload can achieve this, but that means excess frictional losses on the induction or exhaust stroke. You need to increase the rings’ radial cylinder wall loading as the cylinder pressure increases. Gas porting the top ring groove is an attempt to do just that.
If maximum output for a given piston is the goal, gas porting is the way to go and is offered by most piston manufacturers. For drag racing, using a vertical style of gas port through the piston crown is the preferred method. However, in time, these can clog up, so for use other than drag racing, a horizontal gas port in the top ring groove is preferred.
Fig. 2.14. The two notable aspects of this performance JE piston are: the horizontal gas ports (arrows) and the 0.043/0.043/3-mm ring grooves.
Fig. 2.15. Many off-the-shelf budget-oriented pistons do not have gas porting to the top ring. This Goodson tool allows you to gas port your own pistons with little more than a drill.
Fig. 2.16. I typically use Total Seal rings because my dyno tells me I should. The use of thin cross-section rings also pays off. There can be as much as 8 to 11 hp difference between 0.062 and 0.043 compression rings.
Thinner rings are better than thick ones due to reduced friction; Total Seal rings are about as good as it gets because they seal tight. This is not just my opinion but the result of a lot of tests with rings at various gaps all the way down to the zero gap given with a Total Seal ring. Like the bores the piston rings need to have a low-friction prep. The first move is to use a very fine stone to remove the sharp corners from the edges of the rings’ outside diameter (OD). Then polish the rings with a Scotch-Brite pad until they feel really slick.
Connecting rod failure is rarely less than catastrophic. Fortunately, factory rods are fairly stout pieces but carry a lot of excess balance-pad mass. With work, they can be lightened considerably. Also, you can install a set of ARP bolts. But subsequently, your local machine shop needs to cut the caps and resize the big-end bore.
Piston and Rings: Final Specs
Many factors allow a good piston selection to become a great one. Here are the points you need to keep in mind:
• Select a piston and pin with the minimum weight to make an internal crank balance more easily.
• Maximize compression for the fuel octane to be used; first by minimizing cylinder head chamber volume and then by using a piston with a suitable crown.
• Use the thinnest rings your budget allows.
• Be sure the piston crown conforms to the porting specs.
• Deck the block to achieve a net quench clearance of 0.032 to 0.037 inch.
• Do not shroud the spark plug. (For further information see Chapter 11, Ignition Systems.)
• If the cam you intend to use has more than 280 degrees of off-the-seat duration, do a dummy assembly of just the number-1 cylinder with the cam installed to check the valve pocket clearance. (See Chapter 10, Valvetrain Optimization, for more on this subject.)
• If the budget allows for thermal barrier and anti-friction coating, have them applied.
Fig. 2.17. A stock rod can be converted to a fully floating pin by simply honing out the press-fit pin bore to give a 0.0006 to 0.001 clearance on the pin. Here, the pin has had a Tech-Line anti-friction coating applied; it is an option (not a necessity) if the right lubes are used.
Fig. 2.18. Swapping out stock rod bolts for ARP bolts is easy enough, but the big-end bore needs to be resized. The cost for that has to be added to the cost of a rod bolt upgrade.
Fig. 2.19. Before installing any stock-style bolt rods, be sure to put a sleeve over the bolt threads to protect the crank journal from damage during the installation.
Fig. 2.20. This Scat I-beam low-shoulder stroker clearance rod (PN 2-ICR6385) is made of 4340 steel. It’s strong, light, and affordable.
Fig. 2.21. I have used a number of these Callies stroker rods to good effect. The design seeks to minimize the material cut from the block necessary for clearance.
Fig. 2.22. These Scat rods (PN 2-454-6385-2200) look somewhat bulky, but they are, in fact, as light as most other quality rods. I have used them in nitrous engines up to 1,500 hp without failure.
Fig. 2.23. These K1 rods (PN CF6385APRB8) are also a cost-effective buy that is worth checking out. I have limited experience with these rods, but those I have used have held up in high-output builds without problem.
Fig. 2.24. Crower rods are certainly far from the cheapest available, but the quality is almost unbeatable. This billet rod (PN B93911PF-8) has just come out of a 900-hp bracket race engine after three busy seasons of racing. All of the rods look just as they did when they came out of the box.
Fig. 2.25. Manley has a great range of quality rods and is a good source if you are looking for a special lightweight rod for a specific application in which some race-rule-mandated aspect is limiting power, such as the use of a dual-plane intake.
When all of this is done, you will have about half the money into them that a set of good aftermarket rods cost. I use the word “good” here because I have seen failures with one or two brands of “off-shore”–sourced rods. Analysis has shown that the material spec was way off what it was supposed to be. If you stick with the rods I show here, you should be in about as good a shape as can be expected.
The rod I use most (because it has proven time after time to deal with the prolonged dyno sessions my mule engines go through) is Scat’s ICR6385 rod. In addition to being strong, it is typically lighter than a stock rod even though it is 1/4 inch longer between centers. Also, it is very affordable.
My gas dragster that has a touch of nitrous and produces a little more than 1,100 hp uses these rods. As of 2014, these rods are on their sixth season. In engines of up to 850 hp and 7,500 for a quarter-stroker, they have so far appeared bullet proof.
If you want to spend a little more and get rods with even more strength that are 100-percent machined, you can do so without breaking the bank. Options to consider include those from Callies, Crower, K1, Manley, and Scat. When it comes time to spend money on connecting rods, remember that longer is always better. Usually, when building a short-deck block, a 6.385-inch-long rod is the best choice if you are looking for maximum inches. If you are building a tall-deck block (10.2 to 11.1 inches), rods are available up to 6.8 inches long.
There is more to crank selection than simply deciding what stroke and rear seal style is needed. All the bigger stock-displacement factory cranks are externally balanced, so it was either inconvenient to accommodate enough counterweight mass within the confines of the crankcase or there was simply not room. To balance the crank required additional counterbalance mass incorporated into the crank dampener and flywheel/flexplate. Although this Band-Aid fix is passably okay for the rear of the crank, it is undesirable for the snout because it leads to unnecessary bending moments about the number-1 main journal. Unless cost considerations require it, absolutely do not go with an externally balanced crank and dampener system; it needs to be internally balanced.
Fig. 2.26. The Scat 9000 1/4 stroker crank is by far the most popular aftermarket Chevy big-block crank. They are available in both internally and externally balanced form. To tell which one you have or may want, check the machining at the points indicated here. Internally balanced cranks have a large scallop machined from the big-end journal cheeks.
Fig. 2.27. I have had great success with top-of-the-line Callies forged cranks. However, what you see here is a little different. This budget Compstar crank is an offshore-sourced forging that is finished at the Callies plant in Fostoria, Ohio. As such, it is machined on the same equipment as their high-dollar cranks and undergoes the same stringent quality control.
Fig. 2.28. This is Scat’s entry-level forged 1/2-inch stroker crank (PN 4-454-4500-6535). Although counterweighted for a 6.535-inch-long rod, it can be paired with a 6.385 rod and installed in a stock short-deck block. With 0.060 oversize on the bore, this crank yields 525 ci.
Setting Bearing Clearances
It is very important to have bearing clearances within functionally acceptable limits. In a pro shop, this is typically done with expensive measuring equipment that is normally beyond the budget of a home engine builder. However, the use of Plastigauge can establish whether or not the clearance is acceptable. Your targets here are: mains, 0.0025 to 0.0033; rod journals, 0.0022 to 0.0027.
In my shop, we use this Fowler bore gauge to measure the bore diameters so we can establish bearing clearances. However, this gauge is expensive.
By assembling the bearing housing and bearing, then crushing the Plastigauge strip, the bearing clearance can be established within required accuracy.
A decent micrometer from one of the consumer tool houses, such as Harbor Freight, is affordable for the budget-minded home engine builder. Having this micrometer allows you to establish the crank journal sizes accurately.
When using Plastigauge with rod bearings, measure them in pairs. To prevent the rod skewing on the journal, insert feeler gauges (arrow) between the rods to take up the side clearance during nut/bolt tightening.
Crank end-float needs to be between 0.004 and 0.008; the target is 0.006. Use a fine emery cloth laid out on a machined flat surface to remove material from the bearing thrust face if it’s too tight.
Fig. 2.29. Manley has a good range of cranks available, with and without center counterweights.
Another aspect of crank design you may want to consider is whether or not to go with a design having center counterweights. The subject of center-counterbalanced weights may not have previously entered your thoughts. To understand what it’s about, see Figure 2.30.
Without center counterweights, a “couple” (a rocking effect caused by two forces) acts on the center main bearing; it is brought about because of the displacement of the two rod journals on either side of the main bearing. By fully counterweighting each throw, this couple is considerably reduced. But how important might this be in the grand scheme of things?
In terms of engineering finesse, a center-counterweighted crank is the way to go. It relieves the center main of some bearing loads and reduces the bending action caused by a lack of counterweight at this position. But it’s not an open-and-shut case. Having those extra counterweights usually means a slightly heavier crank, although not by as much as you might think. In a center-counterweighted crank, some of the mass for the center counterweights is taken from the next counterweights out. This means that, in part, some of the mass for the extra pair of weights is realized by the crank designer moving some of the mass from adjacent counterweights to the center counterweights.
Fig. 2.30. Shown here is the difference between a non–center-counterweighted crank (top) and one having center counterweights (bottom).
Building a street or street/strip engine with an internally balanced crank means you have eliminated an out-of-balance mass at either end of the crank. That’s such a big step in the right direction that it makes the issue of using a center counterweight or not far less important. The center main is big enough to take the added loads of the couple around it, so reliability is not likely to be an issue, at the level likely to be seen for engines up to approximately 7,500 rpm and 900 hp.
Fig. 2.31. This Scat non–center-counterweighted 4.5-inch-stroke crank weighed 63 pounds after balancing. This compares to about 75 pounds for a stock 4-inch-stroke forged crank.
If I’m building a cost-is-no-object engine and call the shots on a billet crank that needs to be as light as possible, I would go with the center counterweights for an endurance race engine. However, for anything short of Pro Stock or Pro Mod, the inherently lighter weight of a non-center-counterweighted crank is a good, although minor, option to consider.
The crank for one of my 598 builds (see Fig. 2.31) is a relatively high-dollar custom item, and I elected to go without center counterweights. If your application involves endurance and/or RPM at a relatively high level for extended periods, such as marine use, a center-counterweighted crank is your best option.
Fig. 2.32. This 1/4 stroker K1 crank is in the moderate price range and features center counterweights. It went into a GM-blocked street 540, 10.5:1 CR, build that netted 730 ft-lbs and 690 hp. All with a 550-rpm idle and plenty of vacuum.
Fig. 2.33. This Callies XL Magnum crank has a 4.750-inch stroke and is destined for a 632-ci build. It has three important features: hollow mains, center counterweights, and the shape of the counterweights, which are an effort to build a crank that is as light as possible while minimizing bearing loads.
The One-Piece Two-Piece Rear Seal
It would be really easy to skim right over the seemingly minor worth of the development of crank seal technology but it’s straight out of the world of Formula 1. It’s for that reason I show the rear seal type I am now using in all my builds.
The main asset of this seal is that it has extremely low friction on the crank and will run tens of thousands of race miles before leaving a visible line on the crank-seal diameter. It’s manufactured by GTS Racing Seals in Eastleigh, Hampshire, England, and available from TWPE. In addition, it seals in and out, so it makes for a better pan evacuation function. The low friction is immediately apparent as the crank spins with virtually the same ease as it did prior to installing the seal.
I have been testing these seals for about four years, and since then, nearly all of the cup car teams have switched over to them. To install it, put a single slit in a designated position, thread it into the oil seal grove in the block, and then position the cap.
Jacque builds her own race engines, and she is about to install this one-piece GTS seal on a 496 build employing a two-piece-seal crank.
This one-piece GTS seal is threaded into the block half of the rear main. The joint is positioned just below the split line. Although GTS seals call for the split to be vertical in the top half of the rear main, it seems to matter little where it is for an oil-tight seal, as long as it is not at the split line.
The inertial and gas pressure loads imparted to the crank at each rod station cause the crank to twist torsionally from end to end. When these forces are imparted at some multiple of the crank’s natural torsional frequency, they build up out of all proportion to the amount caused by the original excitation force.
A 4.75 stroke crank that was grossly under-dampened illustrates this condition. At about 4,000 rpm the crank starts to resonate and the rod journals vibrate back and forth during their rotation by almost 2 degrees. As a result, the position of the center of the rod journal on this seemingly very stiff crank is moving back and forth by an almost unbelievable 0.080 inch. This cyclic motion superimposed on the typical rotation of the crank plays havoc with the cam dynamics because the cam is coupled directly to the crank via the timing chain and gears.
Fig. 2.34. The Professional Products 7.6- or 8-inch dampener is my preference for a top-performing entry-level crank dampener. I have used them for many years and have seen good dyno results and zero problems.
Fig. 2.35. Torsional tests with dampeners from Innovators West have shown that this company knows how to build effective dampeners. The lesson is, buy direct from the company, not from a third party.
Fig. 2.36. The ATI Super Damper has a variety of elastomer rings and inertial mass rings so it can be tuned to your application. ATI offers a dampener torsional test service for which they go to the dyno being used and measure the existing torsionals. From this data, they optimize the dampener build to suit.
Fig. 2.37. This exploded view shows the internal components of an ATI dampener. These dampeners are tuned to the rotating assembly torsionals by sizing the mass and the stiffness of the rubber O-rings of the second element from the left.
In an extreme case, 50 to 75 hp is lost on a 750-hp engine, and a typical loss is 20 hp. Furthermore, the crank breaks at a fraction of its normal life. All this tells you that an effective crank dampener not only greatly extends crank life but also allows the generation of additional power.
Crankshaft Dampener installation
In all probability at least half of all conventional elastomer dampener failures are due to incorrect fitting or removal. For the dampener to be fully functional, it must be a very tight fit on the crankshaft. This tight fit plus the insurance of a top-quality crank-securing bolt are virtually mandatory for a performance power plant.
Do not hammer the dampener on or use the dampener-securing bolt to pull the dampener into place. Using a big gear puller around the dampener OD almost certainly damages or destroys the elastomer ring between the outer inertia ring and the dampener hub. Rent the right tool from one of the big auto parts chains, such as O’Reilly’s, Pep Boys, Advanced Auto Parts, etc., or purchase a really good tool, such as the one from Moroso shown here.
This Moroso crank dampener installer/remover is in my toolbox because it’s a true pro-quality tool. If you intend to build engines on a regular basis, it’s a must-have item for the toolbox.
This ARP crank bolt is a highly recommended addition to the fit and function of the dampener.
It is neither practical nor possible to rid the rotating assembly of all torsionals. But if those torsionals are reduced to some practical lower limit, you essentially eliminate any problems. Assuming the cam and valvetrain have inherently good mechanical dynamics, an effective dampener (compared to an indifferent one) can reduce the peak-to-peak torsional vibrations. On a 4.25-inch-stroke crank, this means a reduction from about 1.2 degrees to about 0.375 degree. Anything under about 0.4 degree is acceptable, but it is possible to make worthwhile reductions even on this figure.
Bear in mind that the longer the stroke, the lower the natural resonant torsional frequency. In this book, I typically deal with strokes between 4.25 and 4.75 inches. A really effective dampener can, with such stroke lengths, reduce torsionals that occur in the important 4,000 to 8,000 range to a peak-to-peak torsional vibration of as little as 0.2 degree. But to get that you must be very selective in your choice of a dampener.
Fig. 2.38. The ATI dampener is sold in several pieces. To minimize the effect of crank nose tolerances, the hub may require honing to achieve the desired fit to the particular crank being used.
Size and Weight
Engine builders have a strong tendency to focus unduly on dampener diameter and weight. The thinking here is that a lighter dampener, having less moment of inertia, allows the engine to accelerate faster so it must be better for performance. In 99 percent of instances, however, that proves to be not the case. The dampener attribute you should rank as number one, if optimizing performance on the track is the goal, is its ability to effectively dampen those unwanted torsionals. Everything else comes in far behind that requirement.
Although there are going to be instances to the contrary, my experience is that a good 8-inch dampener marginally edges out a 7-inch dampener in terms of engine output by the odd couple of horsepower for an engine in the 800-hp range, even though it’s typically 2 to 2½ pounds heavier (about 30 percent greater moment of inertia). My advice here is that for endurance applications go with a good 8-inch dampener.
The important weight/mass in a dampener is the inertial mass on the outside of the dampener. If you want to save weight and a little inertia, an aluminum hub is an option. However, my thoughts here are that the aluminum hub doesn’t take as many removals before the all-important fit on the crank becomes questionable. Remember, unless that fit is super tight the dampener does not work as it should.
Also, because I am on the topic of dampener retention, be sure to use an ARP crank bolt to hold the dampener in place: no exceptions here.
Dyno shops with torsional vibration test gear are virtually nonexistent. That being the case, how do you test the effectiveness of a dampener? Fortunately that is very simple. The more effectively the torsionals are dampened, the more horsepower the engine makes. I have spent quite a few hours doing torsional measurement but at the end of the day the best dampener is the one that delivers the best output.
Fig. 2.39. Both ATI and BHJ offer an aluminum hub option. This saves about 1½ pounds.
Fig. 2.40. Shown side by side, the difference between a 7-inch (left) and an 8-inch (right) BHJ dampener looks much greater than 1 inch. With longer-stroke builds (4.5 inches and over), the 8-inch dampener appears to be the best choice.
Balance
If you have made the decision to go with internal balance, you have addressed the number-one issue in the right manner. Also be aware that a little underbalance or overbalance does not affect how smoothly the engine runs. The biggest factor affecting just how smooth the engine runs is how close to uniform the weights of the rods and pistons are.
Fig. 2.41. BHJ manufactures professional-quality high-tech dampeners that are similar in design to conventional factory elastomer dampeners. These American-made custom-tailored dampeners have a very long service life capability and should be considered a top choice for a fit-and-forget application.
