Racing pistons are manufactured from aluminum forgings. Certainly plenty of racing occurs with cast and hypereutectic pistons, but only where class rules forbid aluminum forgings. This book deals with high-level racing engines, so this discussion primarily pertains to forged racing pistons. With the exception of clearances and certain performance features, cast and hypereutectic pistons can be prepped the same way, so much of what is discussed here still applies.
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The primary aluminum alloy for piston forgings is 2618 although 4032 is often used when greater control of thermal expansion is required (see “Piston Alloys” on page 59). The downside is a reduction in strength due to higher silicon content. Choosing a piston for a racing engine requires careful evaluation of the final application in terms of engine speed and the sustained piston motion, velocity, and loading that occurs. Hence a brief discussion of piston speed is warranted.
Piston speed (velocity) refers to the average or mean speed of the piston as it moves up and down in the cylinder bore during each crankshaft revolution. Since the piston actually comes to a complete stop at the top of the stroke (TDC) and at the bottom of the stroke (BDC), its speed and acceleration at any given point is always changing. The piston is always accelerating from or decelerating to zero speed.
The official formula for mean piston speed yields an average speed based on two times the stroke (up and down for one revolution), times the engine speed (RPM), divided by 12 to convert to feet per minute (fpm). To simplify the formula, you can divide the numerator and the denominator by 2:
Piston Speedfpm = stroke x RPM ÷ 6
For example, piston speed for a 350 Chevy with a 3.48-inch stroke at 7,000 rpm is calculated like this:
Piston Speedfpm = 3.48 x 7,000 ÷ 6 P
iston Speedfpm = 4,060
Accepted mean velocity for most racing pistons is about 4,500 fpm, but many racing engines routinely exceed that with mean speeds of 5,500 fpm or higher. Interestingly, many Formula 1 engines with an engine speed of 18,000 rpm typically have a mean piston speed below 5,000 fpm primarily due to their very short stroke. In contrast a Cup engine exceeding 9,000 rpm may have a piston speed approaching 5,500 fpm.
Mean piston speed has long been used as a predictor of component durability under severe service. It is a good rule of thumb and it is even more instructive if you calculate maximum piston speed (MPS), since one of the axioms of engine performance dictates that power comes from engine speed. The more power strokes per minute, the more power available to do work.
You can get a very close approximation of MPS (ignoring rod center to center length and rod angularity) with the following formula:
Multiply the stroke times pi and divide by 12 to get feet per revolution. Then multiply by the maximum engine speed to get the maximum feet per minute.
MPSfpm = (stroke x π ÷ 12) x RPM
This speed occurs about mid stroke where the connecting rod is 90 degrees to the crankpin and the crank angle is approximately 75 degrees. Before that point the piston is accelerating; after it the piston is decelerating. When the piston is exactly at either TDC or BDC it is stopped and there is no acceleration.
Continuing our example of a 350 Chevy with a 3.48-inch stroke, let’s find its maximum piston speed at 7,000 rpm.
MPSfpm = (3.48 x 3.14 ÷ 12) x 7,000
MPSfpm = (10.92 ÷ 12) x 7,000
MPSfpm = .91 x 7,000 MPSfpm = 6,370 fpm
One of the most important considerations is the instantaneous piston acceleration and the staggering loads placed on the piston pin bore, piston pin, connecting rod, and rod bolts when the piston reverses at TDC. These are the most highly stressed components in the engine. Since an engine’s ability to make power is closely tied to the RPM it can turn, every effort is made to lighten valvetrain components to combat valve float. But the real limit often turns out to be piston mass and piston acceleration.
A typical 350 Chevy piston weighs 1.3 to 1.6 pounds. Special racing pistons weigh less, but imagine trying to accelerate one to over 6,800 fpm (350 Chevy at 7,500 rpm) maximum piston speed at mid-stroke and then slam it to a dead stop and reverse direction in about 1¾ inches (stroke ÷ 2).
At TDC the piston is headed for the moon and the rod has to stop it and yank it back the other way. That’s enough to pull the piston pin right out of the piston, and it does on occasion. It also exerts similar loads on the rod bolts and rod cap. Acceleration (and thus g-force) is greatest just after TDC on the exhaust stroke (because there is no compression to cushion the piston).
The commonly used formula for calculating the maximum acceleration (MA) of a piston is:
MA = [(rpm2 x stroke ÷ 2,189) x (stroke ÷ 2 ÷ rod length) + 1]
For example, the maximum acceleration of a 350 Chevy piston at 7,500 rpm using the stock stroke of 3.48 inches and the stock rod length of 5.7 inches is calculated like this:
MA = [(7,5002 x 3.48) ÷ 2,189] x [(3.48 ÷ 2 ÷ 5.7) + 1]
MA = 7,500 x 3.48 x 1.3052
MA = 89,424.39 x 1.3052
MA = 116,716.71 ft/sec2
That’s insane acceleration for a 450- to 600-gram object that is not a cannon projectile. Because they are captured by the ring grooves, the piston rings are along for the ride, slamming up and down within the ring grooves trying desperately to maintain a seal with the cylinder wall. Is it any wonder that they experience ring flutter at very high engine speeds? Hence, the practice of using the tightest ring grooves possible without seizure and the thinnest and lightest rings that have minimal inertia.
If you apply this math to a Formula 1 engine you’ll find that the instantaneous acceleration is far beyond the normally accepted limit of 150,000 ft/sec2. How do they do it? The pistons are changing direction more than 150 times per second. It seems far beyond the physical limitation of the components involved, but then Formula 1 uses some very light and very strong exotic materials.
Piston Speed Issues
As discussed earlier, piston speed is an issue for component durability as it relates to engine speed and g-loading. But there is another factor of near equal importance. Note that even NASCAR and Formula 1 engines don’t like to exceed an MPS of 5,000 feet per minute. The reason is oil control.
In most engines, the pistons and rings are lubricated by oil splash being thrown off the spinning crank and rods. NASCAR and Formula 1 engines have an additional source provided by pin oilers, which spray a small jet of oil at the bottom of the piston to lubricate the piston pin and cool the piston crown. But most engines rely only on splash. This becomes a potential problem when you dry sump an engine to remove all oil from the pan. In many cases efforts are also made to isolate the cam tunnel to eliminate cam and lifter oil from dripping on the crankshaft. All of these things combine to reduce the amount of oil being splashed on the cylinder walls. Why is this important? At elevated engine RPM and the attending high piston speeds, reduced cylinder oiling makes it difficult to maintain a consistent hydrodynamic oil film between the piston and the cylinder wall. When MPS exceeds 5,500 rpm, inconsistent lubrication causes the piston to scuff and, worst-case scenario, friction-weld itself to the cylinder wall. This typically results in cylinder wall scoring and ultimately engine failure. The greater the piston speed, the bigger this problem can be, especially when other oiling system modifications combine to reduce cylinder wall oiling. When piston speed is too high, the minimal fog of lubricant in the crankcase has no time to attach itself to the cylinder wall.
Piston selection incorporates numerous factors including alloy type, expansion characteristics, skirt design, deck and dome configuration, pin type, and ring design and placement.
Virtually all pistons are made from either 2618 or 4032 aluminum. These alloys differ primarily in their material content and thermal and fatigue properties, which dictate their suitability for different applications. A 2618 piston has almost no silicon content. It expands approximately 15 percent more than a piston manufactured from 4032 alloy, but it yields higher strength at racing temperatures, particularly above 500 degrees F. Therefore, more race pistons are manufactured from 2618 than from 4032, which is more popular for street engines that utilize less cold clearance and require minimal startup noise. 2618 is the material of choice for Cup engines, Formula 1, and most high-end applications. Unless piston choice is limited to cast or hypereutectic pistons, the application will almost certainly be machined from one of the various proprietary 2618 forgings utilized by the major piston manufacturers.
Coefficient of Thermal Expansion
2618 alloy expands roughly 15 percent more than 4032 alloy, hence its initial piston-to-wall clearance is 15 percent greater. Both alloys have approximately the same clearance at operating temperature with 2618 expanding approximately a half a thousandth (.0005-inch) more at 375 degrees F on a 4-inch-diameter piston. Pistons made from 4032 alloy are typically used for street/strip applications while most race applications use 2618 alloy for its superior tensile strength, fatigue resistance, and lower modulus of elasticity.
Piston Top Selection
Choice of piston top configuration is application specific and dependent on the compression ratio and combustion chamber configuration. High compression is still a staple of engine efficiency and big horsepower, but contemporary racing engines achieve it in different ways. In many cases large piston domes no longer prevail. Flat-top pistons and some gentle dome configurations are much more prevalent. They are designed specifically to work with modern cylinder heads that have shallower, highly efficient combustion chambers.
These combinations can still achieve high compression ratios and they are much more efficient due to improved mixture motion and combustion characteristics. Also, smaller combustion chambers help to build static compression and they generally provide shorter flame travel, which lessens the initial timing requirement and reduces negative work against the piston as it approaches TDC.
Piston Skirt Design
Piston skirts stabilize the piston in the bore. If the ring pack alone were capable of maintaining stability and optimum ring seal throughout all the various direction and pressure changes and the overall range of engine speed, you could remove the skirts altogether and eliminate a major friction source. Unfortunately that’s not possible, particularly with the current trend to lower tension rings.
Skirts are necessary to provide stability to the piston’s secondary motion, which is rocking in the bore caused by thrust loading, friction characteristics, piston center-of-gravity, pin location, and offset and temperature variations. The piston designer’s goal is to optimize stability and minimize frictional losses at the same time. This is accomplished by optimizing the skirt contact area based on the application. Shortening the skirt length reduces friction, but it may also tend to compromise stability. More effort is generally expended in shaping the surface contact area to minimize friction while maintaining stability.
Most pistons still incorporate a cam grind or elliptical shape to the skirt area as a means of controlling contact area. The same goes for the barrel profile, which also influences contact area. The barrel shape tapers inward, reducing the piston’s diameter as it approaches the bottom of the oil ring groove. This reduces the amount of skirt material exposed to cylinder wall contact. Critical factors are oil film viscosity, the thickness of the oil film, the degree of cylinder wall lubrication brought about by the engine’s particular crankcase and rotating assembly characteristics, and the presence or absence of pin oilers.
All of this is dependent on the engine speed range, stroke length, and rod angularity. Piston salesmen are well versed in the latest trends and procedures so don’t fool yourself into thinking you know more than they do. Quite likely, you don’t, but it’s a shared experience. Piston salesmen and designers are constantly evaluating feedback from racers to accumulate an enormous body of knowledge that covers a broad spectrum of applications.
Piston-to-wall clearance at the skirt is normally measured at the gauge point 1/2 inch up from the bottom of the skirt, but some manufacturers have very specific measurement points that you should use if they are provided.
To some degree the clearance needs to be greater on larger cylinder bores to allow for increased expansion on larger pistons. Because of their greater thermal expansion characteristics, 2618 pistons require more clearance than 4032 pistons. The rule of thumb is approximately .0012 inch per inch of bore diameter.
Many applications require additional clearance to accommodate extreme pressure and thermal loads and/or extreme engine speed. These are arbitrary guidelines and you should always determine the optimum clearance by consulting with a piston salesman and machinist.
Deck Clearance and Piston-to-Head Clearance
The deck clearance is defined as the distance between the piston top and the deck surface of the cylinder block when the piston is stopped at TDC. Deck clearance is one component of the piston-to-head clearance dimension that seeks to accomplish the minimal clearance possible without sustained piston-to-cylinder head contact. The purpose is to promote maximum quench effect (or squish) in a wedge-type chamber that physically forces the fuel mixture closer to the spark plug, where it is more prone to combust properly. The resulting turbulence enhances mixture quality and frequently permits a reduction in spark timing to reduce negative work against the piston prior to TDC. Power is typically increased and brake-specific fuel consumption is improved with less chance of detonation.
Deck clearance can be manipulated by altering the piston pin location relative to the fixed rod and stroke dimensions or by selective piston top machining in some cases. Many times, off-the-shelf pistons from major manufacturers can be used with considerable comfort as they incorporate common compression height dimensions that are usually compatible with available block deck height dimensions.
Many builders take a minimum cut off the block deck surface to square it with the crank centerline and then order custom pistons with pin placement specifically positioned to accommodate the piston-to-head clearance they intend to achieve with the proper combination of deck height and gasket thickness.
Wrist Pin Height
Choosing a pin location involves considerably more than simply fitting the whole reciprocating package into the available space. The position of the wrist pin in the piston must accommodate many critical dimensions. Most racing engines use longer than stock connecting rods, which help reduce piston weight while having positive effects on torque positioning and combustion efficiency. The higher the pin position, the shorter the piston and there is a resulting reduction in piston mass. This frequently requires the ring pack to be located higher on the piston.
In naturally aspirated applications, builders appreciate this because they like to move the ring pack to lighten the reciprocating assembly, improve piston stability, and minimize unburned gases in the crevice volume above the top ring. However, longer rods in a supercharged application are often more problematic because boosted applications need to move the ring pack down the piston to position it farther from excessive heat. Longer rods make this difficult to accomplish. In many cases a shorter rod can be specified for boosted applications because boost pressure reduces the need for the critical rod/stroke tuning relationships required for efficient naturally aspirated operation.
When long rods force the pin higher in the piston they require careful checking for interference between the top of the rod and the underside of the piston top. You need a minimum of .050-inch clearance between the rod and the piston at this point. In many cases the pin bore also interferes with the oil ring groove and a support rail must be added to stabilize the oil ring. There is a practical limit to how high the ring pack can be moved, typically no closer to the top than .200 inch, although many short-duration drag racing engines run it higher.
One important factor is the depth and location of the valve relief pockets relative to the proximity and depth of the top ring groove. This presents a potential structural weakness at the closest point and higher potential for burn-through or irreparable damage from detonation. All major piston manufacturers are very familiar with these requirements and salesmen normally keep customers out of trouble even though many may wish to go there. In many cases they can select an alternate forging and machine it to accommodate your requirements. If, in the case of naturally aspirated applications, you are using rod length to help position a torque curve, the sales people can get you as close as possible to your desired dimensions while making sure you don’t exceed limits that lead to early failures.
Piston coatings are now commonplace. Several types are used on pistons depending on the desired results. Various thermal barrier coatings are frequently applied to piston tops. They help hold heat in the combustion chamber to increase cylinder pressure while reducing the amount of heat transfered into the piston. Benefits also include reduced oil temperature and piston expansion.
Contact Reduction or Anti-Detonation Grooves
These grooves are designed to limit piston-to-cylinder-wall contact at elevated temperatures and high RPM. They are machined into the top ring land to protect the top ring by theoretically disrupting detonation pressure waves. Engine builders are split regarding their value, particularly since many pistons already incorporate a small degree of vertical taper on the top ring land to guard against contact between the piston and cylinder wall. The general consensus is that they don’t add power and that once detonation occurs, ring seal is lost and bigger problems usually ensue.
A V-shaped or U-shaped groove is machined into the second ring land to collect excess blow-by between the top and second ring. This groove collects excess residual combustion gases to help control top ring flutter while maintaining ring seal.
Cam and Barrel Shape
Every piston requires the correct cam and barrel skirt shape for the application and the anticipated engine speed. Different cam and barrel profiles are utilized for maximum performance. Piston manufacturers have a range of cam and barrel profiles that apply depending on the specific forging and the final application. When the skirt shape has been optimized it promotes:
- Tighter clearances
- Improved ring seal
- Increased power
- Improved durability
Piston skirt cam shapes are generally divided into two types: high-cam and low-cam profiles. Both are slightly elliptical but the high-cam shape is more pronounced. Both styles produce similar fictional losses, but the high-cam profile permits tighter cold-piston-to-wall clearances. Both cam profiles tend to resume their round shape once they come up to temperature. A piston sales rep can help you decide which is best for your application.
In addition to a barrel shape or taper below the oil ring groove, pistons also taper inward above the oil ring to accommodate thermal expansion and prevent piston-to-wall contact upon piston rock. The difference in diameter of the top ring land compared to that of the oil ring land often is .025 to .040 inch.
Vertical Gas Ports
Vertical gas ports are a series (8 to 12) of vertical holes drilled around the perimeter of the piston deck at a radius that coincides with the back of the top piston ring groove. These holes direct combustion gas pressure to the back side of the top compression ring to force the ring out against the cylinder wall for improved sealing. These ports are typically used for short-duration drag racing applications and are designed to work with a tight ring groove side clearance of about 0.001 inch.
Lateral Gas Ports
Lateral gas ports perform the same function as vertical gas ports except they are half-round slots drilled horizontally into the top ring land at the top of the ring groove. Like vertical gas ports, they are essential for sealing low-tension compression rings. Lateral gas ports are more commonly used in longer-duration and endurance circle track and road racing applications. They offer similar ring sealing ability, but tend to acquire less carbon buildup than the vertical ports over time.
Piston-to-valve clearance is one of the most critical clearances in a racing engine. The intake valve comes closest to the piston at approximately 10 degrees ATDC and the exhaust valve is closest at about 10 degrees BTDC. Many racing pistons are now manufactured with valve reliefs that already incorporate the correct valve angle and accommodate the necessary depth, but you must check every single one of them personally, particularly if you are practicing individual cylinder tuning via different cam profiles and/or rocker ratio in alternate cylinders.
Minimum clearance should be .100 inch on the intake valve and .120 inch on the exhaust valve and this should be increased by .020 to .030 inch if you run aluminum rods. Engine builders frequently fudge this a bit, but it is a dangerous course.
The radial clearance on each valve should be at least .050 inch and the valve relief pocket must be machined at the proper angle and perfectly perpendicular to the valve axis.
JE Piston’s asymmetrical pistons have gained favor with many import engine builders. A recent development pioneered by JE, asymmetrical pistons, are manufactured with smaller, narrower skirts on the minor-thrust side. This configuration retains the standard design intent skirt for the major-thrust surface as required by the specific application.
The skirt on the minor-thrust side is reduced in size (width) to help minimize friction while still supporting piston stability within the bore. This also reduces overall piston weight. To date this piston design is only being applied in import engines with smaller, lighter pistons.
Some engine builders like to push the valve-to-piston clearance pretty tight, but there are guidelines that most builders use to keep themselves out of trouble. For steel and titanium rods with minimal stretch, .060 to .080 inch is the generally accepted minimum clearance. To some degree this also depends on the camshaft and valvetrain and the engine’s operational speed range. With lower engine speeds, some builders may fudge the intake clearance a little below .060 inch while maintaining the exhaust valve clearance at .100 to .120 inch. Two times the intake clearance is the generally accepted rule of thumb.
Higher-speed engines require even more precise fitment to ensure safe operation. This is particularly important in order to provide a safety factor in case of missed shifts or other conditions where an over-rev may occur. The intake valve comes closest to hitting the piston at approximately 10 degrees ATDC and the exhaust valve comes closest at about 10 degrees BTDC. Note that from this point the intake valve is actually chasing the piston down the bore as it opens, thus from the closest point, the clearance is always increasing.
Theoretically the intake valve should never hit the piston as long as you have adequate clearance at this point. While not good for power production, some small degree of pushrod and rocker arm flex at this point actually works in your favor in terms of valve-to-piston clearance.
On the exhaust side the piston is bearing down on the exhaust valve as it closes. It is imperative that contact be avoided at all cost. Additional exhaust valve clearance is required to accommodate thermal growth of the valve due to high exhaust temperatures. It is also important that the valve reliefs in the pistons are geometrically parallel with the faces of the valves. This can sometimes minimize damage if slight contact does occur.
The radial valve relief clearance around each valve should be no less than .050 to .060 inch. Exhaust valve piston-to-valve contact is typically exacerbated by high engine speeds where the valvesprings have trouble keeping the lifters in contact with the cam lobes, causing some degree of lofting or float. Loss of lifter control frequently causes exhaust-valve-to-piston collisions due to thermal lengthening of the valve and the more critical dynamic relationship of the piston chasing the valve as it closes (see Chapter 11).
Race Piston Terminology
Understanding the many factors of race piston design is fundamental to making the correct choices to suit the intended racing application.
Contact Reduction or Anti-Detonation Grooves
These are grooves machined into the top ring land to protect the top ring by theoretically disrupting detonation pressure waves. They are designed to limit piston-to-cylinder-wall contact at elevated temperatures and high RPM. Engine builders are split regarding their value, particularly since many pistons already incorporated a small degree of vertical taper on the top ring land to guard against this contact. The general consensus is that they don’t add power and that once detonation occurs, ring seal is lost and bigger problems usually ensue.
This is a V- or U-shaped groove machined into the second ring land to collect excess blow-by between the top and second ring. This groove collects excess residual combustion gases to help control top ring flutter while maintaining ring seal.
Constant Pressure (CP) Groove
This is a channel or groove on the lower part of the top land that equalizes pressure to the back of the top ring groove. When used with lateral gas ports, the CP groove reduces carbon buildup in the gas ports and prevents the top land from pinching the top ring if the land contacts the cylinder bore. Cam and Barrel Shape
Cam and Barrel Shape
Every piston requires the correct cam and barrel skirt shape for the application and the anticipated engine speed. Different cam and barrel profiles are utilized for maximum performance. Piston manufacturers have a range of cam and barrel profiles that apply depending on the specific forging and the final application. When the skirt shape has been optimized it promotes tighter clearances, greater stability, improved ring seal, increased power, and improved durability.
Vertical Gas Ports
Vertical gas ports are a series (8 to 12) of vertical holes drilled around the perimeter of the piston deck at a radius that coincides with the back of the top piston ring groove. These holes direct combustion gas pressure to the back side of the top compression ring to force the ring out against the cylinder wall for improved sealing. These ports are typically used for short-duration drag racing applications and are designed to work with a tight ring groove side clearance of about .001 inch.
Lateral Gas Ports
These ports are meant to perform the same function as vertical gas ports except they are half-round slots drilled horizontally into the top ring land at the top of the ring groove. Like vertical gas ports, the intent is to enhance the sealing of low-tension rings. Lateral gas ports are more commonly used in longer duration and endurance circle track and road racing applications. They do not offer the same ring sealing ability as vertical ports and they tend to negatively affect top ring sealing unless everything is perfectly aligned.
Piston rings have one of the toughest jobs in a competition engine. They are subjected to various conflicting forces while constantly being slammed back and forth up to 150 times per second or more. Consider for a moment the primary goals of the piston rings in a modern racing engine.
- Seal combustion pressure within the combustion space
- Maintain compression ratio
- Prevent oil contamination of the fuel charge
- Transfer heat from the piston to the cylinder wall
- Resist ring flutter and chatter (vertical displacement) to maintain optimum seal
These are formidable goals for thin metal rings that are subjected to the massive forces of combustion. To better understand the ring’s function, let’s examine each ring in the cylinder kit separately.
The top ring’s sole function is to maintain the compression ratio and seal the combustion space against many thousands of pounds of combustion pressure upon ignition. This is a formidable task. The ring is required to maintain bi-directional sealing under all conditions even as it is subjected to the rapid high pressure rise of normal combustion, which then dissipates quickly as pressure is applied to the piston during the expansion cycle (power stroke).
Comparatively speaking, the top ring gets a bit of a breather during the exhaust and intake strokes, but it still has to maintain its seal while subjected to rapidly varying changes in pressure, direction, and velocity. It has to hold compression on the compression stroke and then gets slammed back the other direction under maximum combustion pressure on the power stroke.
Throughout all of this the piston is rocking and vibrating in the bore while crankshaft oscillations attempt to impart a jerking motion to ring travel to the point where the rings don’t know whether they are coming or going. And then the engine detonates and ring seal is totally lost.
Top rings are subjected to massive and near instantaneous forces just as they are attempting to change direction in a hostile environment of severe heat and pressure. Current trends toward thinner rings to reduce friction and resist ring flutter only exacerbate sealing problems; hence ring configuration and preparation are extremely critical. For years the traditional front-line application has been a high-strength ductile iron ring with a moly (molybdenum) inlaid face. Over time, this type of ring evolved to the current plasma moly facing that is sprayed onto the ring face at high temperature and velocity. This ring treatment is more resistant to the cracking and flaking that sometimes plagued inlaid rings under severe operating conditions.
Plasma-coated rings are preferred for many high-performance applications, but they are still susceptible to damage from high shock load conditions such as those found in nitrous-oxide and other power-adder applications. For these conditions, newer-style gas nitride ductile iron rings are usually a better choice. Gas nitriding is a specialized surface treatment used to harden the ring face for improved wear characteristics and resistance to detonation. Gas nitride rings are superior in every respect and have even found favor in many OEM applications. Standard steel rings are also a good choice for power-adder applications where severe duty is anticipated.
Racers in high-contamination environments such as dirt track or off-road racing often prefer chrome-faced top rings for their durability, but more recently many builders have switched to plasma-coated rings for their superior heat resistance under severe operating conditions. A new generation of thin vacuum-deposited chromium nitrite rings has reinvigorated chrome ring applications and many OEMs have now adopted them for their excellent wear characteristics. The general trend in racing rings points to thinner steel or stainless rings, many with exotic coatings such as tungsten or titanium nitride. These rings offer exceptional wear and sealing characteristics with enhanced friction reduction.
Coatings are applied using a positive vapor deposition process that has proved to be reliable and substantially beneficial in terms of power and durability although still quite expensive for all but well-funded applications. If any single ring type seems to be losing favor it is probably the L-shaped Dykes, or headland, ring with a 1/16-inch face and a .017- or .031-inch step on the back side to provide gas pressurization without requiring gas ports on the pistons.
Dykes rings require compatible piston ring grooves and they are generally harder to seat. They also exhibit higher wear characteristics due to gas pressurization behind the ring. They are in limited use, such as blown fuel drag racing applications where they are resistant to high pressures and excessive fuel wash in the cylinders.
Second compression rings are really misnamed since their primary function is to assist with oil control. Most ring manufacturers hold that the second ring is about 85 percent devoted to oil control and only 10 to 15 percent to compression sealing. Its main mission is to scrape oil missed by the oil ring to that it doesn’t find its way to the combustion chamber to contaminate the fuel mixture. Since heat is not an issue with the second ring, a conventional cast-iron ring with a reverse bevel and tapered face is still employed with the primary task of scraping oil.
A recent improvement is the Napier-style ring, which features a hook or claw shape on the back side that serves as a reservoir for oil being scraped off the cylinder walls. Napier-style rings have found favor in both OEM and racing applications because they provide exceptional oil control and reduced friction. They also permit builders to open the second ring gap wider to promote more effective pressure relief between the top and second ring, thus easing the top ring’s sealing effort.
Most oil ring packages employ a three-piece unit with a primary expander ring supported by upper and lower scraper rails. Many of the performance gains associated with ring technology have come from oil ring weight and tension reduction. Low-tension oil rings have long been known to aid performance via friction reduction, but recent improvements have realized additional gains. Ring tension is the primary focus of these efforts.
Ring tension is mainly a function of the ring’s radial depth in the ring groove. For decades oil rings have maintained the SAE standard of .199 inch, but recent development has reduced radial depth to .150 inch or less. With the piston ring groove machined to a corresponding depth and appropriate clearance, tension is reduced because the thinner ring is more flexible.
Conventional low-tension rings provide about 12 to 14 pounds of tension and low-tension rings in OEM performance engines like the Modular Ford or the GM LS series now operate with less than 10 pounds of radial tension. They offer exceptional oil control in a production environment and superior frictional qualities that dramatically impact fuel economy. By way of comparison, the best high-end Sprint Cup ring sets use a 1.5- to 2.0- mm oil ring with only 2.5 to 4 pounds of radial tension. These professional racing rings, such as Perfect Circle’s U-Flex oil ring, are made in limited sizes for Cup requirements and require very precise CNC ring groove machining and dry sump lubrication supported by additional vacuum pump crankcase pressure evacuation to limit pressure below the rings.
Piston Ring Terminology
The following design elements apply to piston ring selection.
The clearance between the ring axial height and the height of the piston ring groove.
The vertical thickness or height of the piston ring in the axial direction.
Clearance between the back or inside diameter of the ring and the back of the ring groove when the ring face is flush with the ring land.
The clearance between the ends of the ring when compressed within the cylinder bore.
The end gap clearance when the ring is not compressed.
The inside diameter of the ring when compressed to the bore diameter.
The outside diameter of the ring when compressed to the bore diameter.
The width of the ring in the radial direction.
Ring Axial Sides
The top and bottom surfaces of the ring.
The front of the ring that contacts the cylinder wall.
The installed position of the ring, which imparts cross sealing due to a chamfered area on either side of the ring. This configuration makes the top ring cone slightly upward and the second ring come slightly downward.
Now let’s examine the currently favored techniques for piston ring preparation and installation. Savvy engine builders recognize the importance of proper cylinder wall preparation in terms of surface finish and efforts to control cylinder shape under load at operating temperature (see Chapter 3). Assuming optimum cylinder wall preparation for the purpose of this discussion, we can examine the role and preparation of rings as it applies to a racing environment.
Successful piston ring application begins with proper ring groove preparation. Because the ring seals on its face and against the ring land, precision-fit ring grooves are essential for proper piston ring operation. The tightest possible clearance between the ring and the ring groove is desirable for optimum ring control, but it is absolutely critical that the ring not stick in the groove. Ring clearance is necessary to allow for component expansion and to allow gas pressure to migrate behind the ring.
Many high-end builders prefer to purchase piston blanks and cut their own ring grooves to exacting specifications. This is found at the highest level of competition, but it is primarily a function of habit and does not constitute a slam against ring grooves cut by any of the major piston manufacturers, which also go to great lengths to cut very precise ring grooves.
Piston ring micro-welding occurs when aluminum from the bottom of the top compression ring groove transfers to the bottom of the top compression ring, causing it to stick in the groove. This causes ring rotation to cease and typically results in increased blow-by as the ring loses bore conformity and torsional twist. The primary cause is usually poor ring groove finish or the ring not being exactly perpendicular to the piston. It is more frequently seen near the ring gap since this is a potentially greater source of hot gas leakage, which contributes to the problem.
The condition is often exacerbated by improper ring groove clearance (too tight) and may be precipitated by excessive temperature, which affects the hardness of the piston material. Micro-welding typically begins during engine break-in, particularly with engines that run a vacuum in the crankcase.
One remedy is to perform the break-in with zero vacuum so the rings receive proper lubrication during the initial seating process. If done correctly this often avoids the problem. Many builders fail to even recognize that micro-welding has occurred even though the engine may lose power and blow-by increases dramatically with a significant loss in crankcase vacuum. They often attribute it to something else and never really solve the problem because they don’t examine the rings carefully upon disassembly. Hard anodizing the top ring groove helps prevent it unless ring groove machining is poorly done. Builders have hard anodized ring grooves for years and more recently they have employed coated rings and ring lands, typically with phosphate dry lubricant.
Most manufacturers now offer top ring groove anodizing and specially coated rings specifically designed to combat micro-welding. All racing engines are susceptible to this problem and race piston manufacturers have done their part by specifically targeting the quality of their ring groove machining to achieve the smoothest possible surface finish with minimal peaks that can attach themselves to the rings. Engine builders look for it specifically upon teardown in order to take the necessary steps during assembly and initial fire-up to help prevent it. In addition to coatings, they often hand lap the rings to smooth the surface and they pay very close attention to ring clearance in the groove.
Written by John Baechtel and Posted with Permission of CarTechBooks