Stroke length aside, connecting rods are one of the basic tunable components in a competition engine. As rod length (center to center) varies, it affects piston motion such that it can be used as a primary tuning ingredient. By influencing piston acceleration and velocity it dictates the rate at which a differential is created between atmospheric pressure (above the carburetor) and cylinder pressure during the intake stroke. Accordingly, it impacts major contributors to the VE equation intake and exhaust path cross sections, valve event timing, and optimum ignition point. Faster exposure to atmospheric pressure improves cylinder filling and thus VE, provided that intake tract dimensions and valve event timing are appropriately sized and synchronized.
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It is important to recognize that piston acceleration and velocity are both zero at TDC and BDC. At all points in between, acceleration and velocity are dictated by rod length. For any given rod length, the piston achieves maximum velocity at a precise point in the stroke relative to the crank angle where the rod axis is 90 degrees to the crank throw (typically, but not limited to about 70 to 75 degrees of crank angle). This point represents the highest rate of pressure drop exposure in the cylinder and is closely tied to intake valve timing for optimum cylinder filling.
Rod Length as a Tuning Component
To some degree, longer rods effectively slow the arrival and departure rate of the piston at both TDC and BDC. This is often referred to as piston dwell time and any number of magazine dyno tests have been conducted to prove that it does not significantly alter peak power. But they are missing the point. The real value of rod length tuning is realized in shaping the powerband to suit application-specific performance requirements. As a rule, connecting rod length can be employed to tighten or spread the RPM range between peak power and peak torque. This is an important function in matching engine performance to the vehicle and its specific racing requirements.
An example would be using a longer rod (and compatible inlet dimensions) in a superspeedway or Bonneville application to shift peak torque closer to peak power so that more effective torque is applied within the specific RPM range of operation. With slower piston departure from TDC, higher combustion pressure is applied to the piston over more crankshaft degrees.
Many builders agree that the rod length should be 1.7 to 1.9 times the given stroke length. Because the longer rod slightly increases piston dwell time, it also provides more time for combustion pressure to build against the piston before it is applied to the power stroke; hence, more net torque is applied in the most desirable range for the application. It is also generally conceded that longer rods tend to make a bit more power in most high-speed applications while shorter rods tend to boost lower-end torque due to faster piston acceleration and the associated higher port energy.
A piston with a shorter rod arrives at TDC more briskly and doesn’t hang around long before it departs swiftly. This is useful in some forms of racing. The piston achieves maximum velocity sooner and at less crank angle, which reduces cylinder volume exposure at the point of maximum pressure differential. Appropriate intake valve timing is required to ensure optimum efficiency under these conditions. Since the piston achieves maximum velocity sooner, the intake valve can be opened sooner to take advantage of the cylinder pressure differential. Less overall cylinder volume is exposed at this point, but the early initiation of flow chases the piston down the bore as volume exposure increases. This is commonly referred to as the piston tugging harder on the charge due to its increased acceleration.
Determining cam timing requirements for situations such as this one have become increasingly easier now that the average builder has access to powerful PC-based engine simulation software that illustrates any given crank angle to valve event timing. For example, the more aggressive action of a short rod combination implies the ability to consider slightly larger intake and exhaust dimensions (cross-sectional area) for the intake manifold and the headers without sacrificing vital port energy. PC modeling can help confirm this. In some cases you may be limited to a specific manifold and/or header combination whose fixed dimensions resonate about a particular rod length that can be identified through diligent PC simulation. Given the low cost of current simulation programs, there is no reasonable excuse for not modeling these concepts in advance on your PC.
Optimizing rod length further reveals tuning considerations that can improve combustion efficiency and reduce the amount of negative work performed on the piston prior to TDC. With a longer rod, the instant cylinder pressure rise approaching TDC is faster and greater and typically requires less total spark timing, depending on chamber efficiency. As spark timing is reduced, negative work (piston struggling against rising combustion pressure ahead of TDC) is diminished while increased dwell time provides greater post- TDC pressure rise against the piston (positive work).
Rod length also impacts piston selection to the extent that it dictates pin height (compression height), and in many cases the final location of the ring package. This is an important consideration since different racing applications require different ring pack placement to accommodate attending combustion- and heat-related issues (see Chapter 5). Longer rods tend to reduce pin height and often require an oil ring support rail because the pin bore encroaches on the oil ring groove. This adds weight to the ring package, but most builders feel that the benefit of the longer rod and attending piston configuration outweigh any mass penalty, at least in those applications where a longer rod positively impacts powerband positioning.
In addition to creating more time for combustion pressure to rise and apply more torque to the crank, a longer rod speeds the burn rate due to enhanced charge density. Consequently the RPM separation between peak torque and peak power is reduced, effectively concentrating more torque in a narrower band, which benefits certain applications such as Bonneville, speedway oval, and drag racing engines that operate in a narrow RPM window with appropriately matched transmission and rear end gearing. Also, longer rods generally require appropriate intake manifold adjustments to accommodate slower piston motion around TDC. That frequently includes slightly smaller (cross-sectional area) intake runners to preserve port energy and in some cases advancing the cam to further enhance torque.
While retarding the cam for high-speed power is the traditionally accepted practice, it does not take into account the beneficial effects of specific rod length tuning in certain applications. Here again, PC modeling often illuminates unexpected paths to the most effective combination of valve timing and inlet- and exhaust-tract dimensions for a given rod length.
Finally, note that a longer rod and corresponding higher pin height is usually reflected in a shorter, lighter piston, which reduces reciprocating weight.
At the other end of the spectrum it’s often advantageous for road racers and some circle track applications to investigate shorter rod lengths and the associated higher port energy, which may prove to be useful for selected applications that are restricted in terms of manifolding or camshaft timing and are seeking more torque off the corners. While the shorter rod exposes less initial cylinder volume to the pressure drop, it promotes increased port velocity to aid cylinder filling efficiency. This calls for different valve timing than what is appropriate for a longer rod, particularly as it applies to the intake closing point. Because the instantaneous pressure rise is greater with a longer rod, it can effectively use a later closing intake to gain additional time for cylinder filling.
Conversely the shorter rod’s higher port energy offers superior filling efficiency, but calls for earlier intake closing due to slower cylinder pressure rise and reduced dwell time at TDC. This tends to build torque earlier in the RPM range and moves the peaks farther apart. As a bonus, higher port energy often contributes to enhanced throttle response. Of course with shorter rods, the piston tends to outrun the flame front after about 30 degrees from TDC so it is important to choose a faster burning combustion chamber and appropriate fuel to accommodate it.
Connecting Rod Materials
Primary factors influencing connecting rod design are extreme inertia forces and cylinder pressure as defined by maximum engine speed, rotating assembly geometry, and weight. Increased engine speed, displacement, component mass, and firing pressure all dictate design characteristics incorporated in racing rods. These are further influenced by dynamic drivetrain loading such as wheelspin or freewheeling propellers (out of the water) in the case of marine applications. Still, connecting rod failures are less prevalent today than in the past. Reasons include superior materials, precise preparation and assembly, and improved control of contributing factors such as spark timing, detonation suppression, and over-rev protection. Mechanical issues such as piston pin stiffness, lubrication, and bearing clearances are also credited with reducing overall connecting rod stress.
Two materials dominate modern connecting rod production: aluminum and forged steel. Aluminum rods predominately populate the professional drag racing ranks where load and life cycles are relatively short and the weight savings are beneficial to transient torque acceleration across narrow powerbands. Their utility has broadened in recent years, but outside of drag racing, most racing applications still rely on steel or titanium rods. For the purpose of this discussion I concentrate on these basic types with brief attention to other materials.
Lower classes frequently specify factory-style forged rods, although in some cases compressed powdered metal is employed along with an OEM-based cracking technique that physically breaks the big end of the rod in half along a pre-scored line. This creates a unique match of irregular fracture surfaces that only mate correctly between the original cracked halves. It is a useful technique that ensures absolutely accurate rod cap alignment and resistance to movement. It is used successfully in fairly powerful production engines, but is not yet widely employed in competition engines although aftermarket rod manufacturers now offer cracked rod technology.
Using compressed powdered metal rods of this kind is typically safe up to around 500 hp in production-based modified street engines, but most serious competition engines rely on aluminum or forged and billet steel aftermarket rods of varying configurations.
Aluminum racing rods are forged from heat-treated 7075 T6 aluminum alloy, which has a tensile strength of 83,000 psi. They are approximately 65 percent lighter than steel rods, but only have about half the strength of steel. Unfortunately, a significant portion of the weight advantage is sacrificed because aluminum rods have to employ increased bulk to maintain strength. Accordingly, block modifications are often required to accommodate the larger physical size of most aluminum connecting rods.
In many cases the oil pan rails must be ground for rod clearance and care must be taken to ensure that the larger bulk does not contact the camshaft. Increased rod size takes up more space in the crankcase and affects crankcase windage differently. Aluminum rods require a pinned lower rod bearing to ensure proper rod bearing alignment and to guard against spinning the bearing.
These rods are primarily used for high-RPM drag racing applications where their lighter weight provides a benefit in transient acceleration. They offer superb damping qualities that are useful in resisting detonation, but their lifespan is relatively short due to the increased tendency to work harder from repeated stretching and compression. If you are going to run aluminum rods take note of the following requirements:
- Accommodate the increased stretch factor
- Modify the block for clearance
- Check for camshaft clearance
- Use the correct pinned bearings
- Handle carefully to avoid nicking the surface
- Fit pins precisely and provide adequate pin oiling
- Replace rods after 50 to 60 runs to guard against breakage
- Consider how the bulkier rod shape affects windage
Steel rods are the most widely used due to their dimensional stability and exceptional long-term durability. Most steel connecting rods are manufactured from 4130 or 4340 alloy steel. Racing rods are commonly made from Mill Certified Aircraft Quality, vacuum carbon-arc deoxidized E4340 alloy, which has a tensile strength of up to 186,000 psi. Although normally calculated by dividing the maximum load by the original cross-sectional area of the component, tensile strength may be thought of as the amount of force required to pull a rod apart at the beam.
The onset of combustion pressure rise (negative work or torque) acts, to some degree, as a cushion for the piston as it approaches TDC, thus it helps to absorb the forces and g-loading that occurs when the piston reverses direction at TDC. This cushion prevents the piston from freewheeling through TDC and tends to soften the g-loading upon piston reversal because the piston is actually working against some amount of cylinder pressure at this point another delicate balance that affects net torque.
All engines incorporate a measure of piston-to-cylinder-head clearance to accommodate rod stretch at TDC. The degree of stretch or elasticity is governed by engine speed, piston weight, rod ratio, and the characteristic of connecting rod materials known as the modulus of elasticity. Simply put, “modulus” describes a load factor indicating a known range of deformation characteristic to a given material whereby the deformed part returns to its original shape when the load is removed. Part failure typically occurs when the modulus is exceeded under repetitive loading. Either the rod stretches permanently causing piston-to-cylinder-head contact and subsequent failure, or the piston-to-head clearance is insufficient to accommodate the known modulus of the rod’s parent material. In this case piston-to-head contact also ensues, usually with major carnage.
Steel rods stretch less than aluminum rods. Accordingly both types have a favored clearance range that accommodates known or (in some cases) anticipated stretch factors.
The commonly accepted minimum piston-to-head clearance for aluminum rods is about .055 inch depending on engine speed and associated component dynamics. The closer you shave it, the closer you flirt with component failure in the event of an over-rev or other unexpected abnormality. That doesn’t mean that many builders don’t fudge them close enough to produce witness marks on the piston tops where the pistons have been slightly kissing the head. If you see this you should regard it as a problem, not only because of inappropriate component contact, but because of the affect that is has on ring seal and the loss of cylinder pressure that may be occurring due to the ring unloading from the unexpected jarring.
Steel rods typically accept a piston-to-head clearance of about .035 inch, but this is often increased by .010 to .015 inch for high-RPM applications or those using lightweight cranks that are more prone to crank throw deflection both radially and along the direction of piston travel.
Characteristics of Rod Activity
In addition to rod stretch issues, builders also consider the material characteristics of rods in compression loading, particularly under conditions of detonation, which can fracture a rod that is too stiff or hammer the rod bearings repeatedly to the point of failure. While bearing failure typically occurs under compressive loading, most rod failures actually happen under tensile loading where the rod gets ripped apart upon piston reversal at TDC, particularly on the exhaust stroke where there is no opposing pressure to soften the transition. This is most often seen to occur about 1 to 1/2 inch down the beam from the pin. Or it fails at the hinge point where the beam expands to the big end. Rod bolt failures are rare unless improper torque and/or bolt stretch are applied.
More often than not the big end is still attached to the crankshaft with a tensile failure at the beam. Consider also that paired rods on a common crank throw have the opportunity to transfer the effects of detonation (via load sharing) in one cylinder to the adjacent rod, which may cause extreme stress or temporary lubrication problems that may spin a bearing or seize a rod.
There is also little doubt that rods absorb crankshaft radial deflection and some degree of vibratory distress under cyclic loading. It may not reveal itself on the dyno, but at the end of a long straightaway, after 250 laps, or just before the 5-mile marker on a long pull at Bonneville, rod stress may show itself. This belies the importance of proper crankshaft dampening and accurate balancing or (in many cases) overbalancing to smooth engine operation within the effective operating range (RPM) of the individual engine.
Another area of concern is the relationship between the rod and piston pin. Even the best rods can be undone by careless pin fitting or poor pin selection. While tighter pin clearances are desirable to prevent piston rattling on the pin, bending forces introduce pin distress that can destroy a piston or break a rod. Piston pin modulus must be sufficient to resist bending along the pin axis as well as radial distortion in the pin bore where the pin temporarily becomes egg shaped and bites the pin bore or the rod bushing, frequently with unpleasant consequences.
Excessive bending along the pin axis typically causes the pin to seize in the small end of the rod. Even if the condition is not severe enough to cause pin seizure, it introduces additional frictional drag and heat rise that may lead to pin failure over time. Numerous steps can be taken to address piston pin issues, particularly in high-speed, high-horsepower engines where pin distortion is more prevalent.
When selecting pistons, it is important to work closely with the piston supplier to secure the best pin configuration possible (see Chapter 5). The lightest possible pin is desirable, but not at the expense of pin distortion. Increased wall thickness is often used along with a tapered inside diameter that thickens toward the center of the pin and rides on inboard-style pin bosses that allow a shorter overall pin.
Piston manufacturers deal with this regularly and usually have a good solution if you can provide accurate information regarding stroke length, rod type and length, crankshaft and rod material, pin bushing material, and anticipated engine speed. In particular, manufacturers now offer diamond-like coatings and other coating procedures that significantly reduce pin bore galling. These coatings have a very low coefficient of friction (combined with proper pin bore honing) that can virtually eliminate wrist pin problems. This is particularly true in applications that employ block-mounted pin oilers to provide additional pin lubrication while also cooling the piston.
One dimension that may promote engine failure is the clearance between the top of the connecting rod and the underside of the piston top. This is particularly notable on racing pistons with a low compression height dictated by longer rod length. In this case the rod rides high in the piston and has a greater chance of interference with the bottom of the deck surface whose shape often mirrors the piston top to achieve a uniform thickness. The shape of the rod’s small end and its thickness above the pin bore also influence this relationship.
To ensure adequate operating clearance it is necessary to check each rod and piston assembly carefully. Fortunately, most racing pistons use floating pins that make it easy to check this clearance prior to assembly. There are two places where the rod can interfere with the bottom of the piston. The first is the top of the rod interfering with the underside of the piston deck. The second is the upper radius on either side of the rod pin bore where interference may occur with the radius of the piston pin boss to the piston deck.
To check these clearances, apply machinist’s bluing to the bottom of the piston and the inner radius of the pin bosses.
Assemble the rod to the piston with the correct orientation and swing the rod to both sides until it contacts the inside of the skirt. Slide the piston to the right and left and repeat in closer proximity to each pin boss. Remove the pin and rod and check the bluing for witness marks that indicate contact. While performing this operation observe carefully with a bright light, as the clearance may be inadequate even though it does not physically indicate contact with the bluing. You need at least .050-inch minimum clearance, which can be verified by passing a large bent paper clip between the piston and rod. If you observe witness marks and/or not enough clearance with the paper clip, you may be able to gain clearance by grinding the bottom of the piston slightly.
Be sure to check piston deck thickness prior to performing this operation. If you observe interference at the pin bosses, you can often chamfer the small end of the rod to gain clearance. Many connecting rod manufacturers offer tapering of the small end to accommodate these issues. Be sure to recheck the clearance after performing any of this work. If you modify the rod, make certain you recheck the small end weight for correct balance.
Written by John Baechtel and Posted with Permission of CarTechBooks
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