FAQ

Multi-turn wave springs, particularly those used in bearing preloads, require precise diametrical clearances to prevent binding. During compression, the mean diameter $D_m$ of a wave spring increases slightly due to the flattening of the waves. The expansion is governed by the relationship $\Delta D \approx \frac{0.051 \cdot (f^2 \cdot n^2)}{D_m}$. Consequently, if the spring is constrained by a housing (bore), the outer diameter $O_D$ must be sized such that $O_D + \Delta D < Bore_{min}$. Conversely, if the spring is located on a shaft, the inner diameter $I_D$ must remain larger than the shaft diameter throughout the full stroke. Failure to provide this 'expansion room' results in friction against the walls, leading to hysteresis in the load-deflection curve and potential galling of the mating surfaces.

17-7PH (UNS S17700) is a semi-austenitic precipitation-hardening stainless steel that provides high strength and corrosion resistance. In condition CH900, the material is cold-reduced to Condition C and then age-hardened at $900^{\circ}F$ ($482^{\circ}C$). This process transforms the austenite to martensite through cold work and then precipitates an aluminum-rich intermetallic phase. The resulting yield strength typically exceeds $200$ ksi. For wave springs, this high elastic limit allows for significant deflection without permanent set. However, designers must be cautious of the material's susceptibility to stress corrosion cracking (SCC) if exposed to chlorides while under high tensile stress. Furthermore, the fatigue limit in CH900 is optimized by the fine precipitate dispersion, making it superior to 302/304 stainless for high-cycle dynamic loading in flight control actuators.

For a Crest-to-Crest wave spring, the load $P$ for a given deflection $f$ is derived from the modified Moyer's formula: $P = \frac{E b t^3 n f}{K D_m^3} \cdot \frac{I_D}{O_D}$, where $E$ is the Modulus of Elasticity, $b$ is the radial wall, $t$ is the material thickness, $n$ is the number of active waves per turn, and $D_m$ is the mean diameter. The correction factor $K$ accounts for the curvature of the waves and the transition between crests. In a shim-end configuration, the load-deflection curve becomes more linear compared to plain ends because the shim provides a uniform 360-degree contact surface, reducing the localized stress concentrations and preventing 'wave-nesting' during the initial compression phase. If $K$ is not accurately modeled based on the $D_m/b$ ratio, the calculated rate can deviate by up to 15 percent as the spring approaches its solid height.

Stress Corrosion Cracking (SCC) occurs when a retaining ring is under high tensile stress (from its interference fit or installation) and is exposed to a corrosive medium. For high-carbon steel rings, even moisture can be enough. The failure is characterized by brittle cracking that occurs far below the material's yield strength. In failure analysis, SCC is identified by branched, transgranular or intergranular cracks. To prevent SCC, engineers must select materials like A286 or Inconel X-750 for corrosive environments and ensure that the 'installation stress' does not exceed a critical threshold, typically $30-50\%$ of the material's yield strength.

In turbine engines, high-frequency vibrations cause the retaining ring to rub against the groove walls and the retained component. This 'fretting' wears away the material, reducing the effective thickness $T$ of the ring and the depth $d$ of the groove. As these dimensions decrease, the thrust capacity drops according to $P_r \propto T$. The debris from fretting (metal oxides) also acts as an abrasive, accelerating the process. Failure is usually caught during overhaul by checking the 'axial play' of the assembly. Prevention involves using anti-fretting coatings like silver plating or using materials with higher surface hardness like 17-7PH.

If a groove wall is not square ($>0.5^{\circ}$ out of perpendicularity), the thrust load $P$ is not applied uniformly. This creates a component of force $P \cdot sin(\theta)$ acting radially, which encourages the ring to expand and exit the groove. In troubleshooting failures where the ring appears to have 'slipped' out, the groove geometry should be measured using a CMM (Coordinate Measuring Machine). A non-square groove significantly reduces the effective thrust capacity, often by as much as $50\%$, because the ring only contacts the 'high point' of the groove, leading to localized yielding and subsequent dishing.

Axial impact loading creates a dynamic stress wave that can exceed the static shear strength of the ring or the groove. In such cases, failure occurs via 'shear-off' of the groove wall or the ring itself. Unlike static failure, impact failure often leaves evidence of plastic flow or 'smearing' on the contact surfaces. To mitigate this, a safety factor of $4$ or $5$ should be used instead of the standard $3$. Additionally, using a wave spring in conjunction with a retaining ring can act as a shock absorber, damping the impact energy and protecting the retaining ring from peak force transients.

'Ring fly-out' is the premature exit of the ring from the groove during operation. In a failure analysis, the first step is to check for 'groove rounding' or deformation. If the groove wall is slanted, it indicates the thrust load exceeded the groove material's yield strength. Another cause is insufficient 'cling'—if the ring's ID/OD wasn't properly sized for the groove. If the ring itself is intact but found outside the groove, it likely failed due to centrifugal forces or high-frequency vibration that caused the ring to 'walk'. Examination of the ring's edges for wear patterns can indicate if it was fully seated prior to the event.

Multi-turn spiral rings (2-turn or 3-turn) provide a $360^{\circ}$ retaining surface with no gaps, which is superior for uniform load distribution compared to single-turn rings. They are often used to take up axial play in an assembly. By selecting a specific thickness $T$, the ring acts as a rigid shim. In assemblies with high axial tolerances, 'Laminar' rings (multiple rings stacked) can be used, although a single multi-turn ring is usually preferred for ease of assembly. The multi-turn design also increases the centrifugal capacity, as the turns support each other against radial expansion.

Removing a spiral ring (especially a multi-turn version) requires a 'removal notch' or 'removal end' which is designed into the ring. A small screwdriver or specialized tool is used to pry the end out of the groove, and the ring is kemudian uncoiled. The primary challenge is avoiding scratches or gouges in the groove material, which can become stress concentrators. In high-strength shafts, a scratch can lead to fatigue failure of the shaft itself. Using plastic or soft-metal prying tools is recommended for aluminum or titanium housings. For 'internal' rings, ensuring the ring doesn't snap back and strike the bore wall is critical.

The groove radius $R$ must be kept minimal (typically $R \le 0.005$ inches) to ensure the ring's flat surface makes maximum contact with the groove wall. If the radius is too large, it creates a 'ramp' that facilitates the ring dishing and eventually popping out under axial load. During assembly, the presence of a radius or a chamfer on the mating part also affects the load path. The calculation for the allowable load $P$ is derated based on the ratio of the radius to the ring thickness. Designers must specify 'square' grooves to maximize the efficiency of the spiral ring's multi-turn design.

Verification of ring seating in automated systems is often performed using laser displacement sensors or vision systems that measure the ring's 'protrusion' height from the shaft or bore. Since a spiral ring has no ears or lugs (unlike stamped rings), its profile is uniform. A correctly seated ring will have a consistent diameter. Another method involves a 'push-off' test where a calibrated axial force is applied to ensure the ring is locked in the groove. In high-volume automotive production, electronic sensors on the installation tool monitor the torque/force curve to detect if the ring 'snapped' into the groove correctly.

Spiral retaining rings are installed by winding them into the groove, which involves expanding the ring (for shafts) or contracting it (for bores). To prevent permanent set, the ring should not be expanded more than $1\%$ beyond its yield point. The maximum installation diameter $D_{max}$ is limited by the material's elastic strain limit $\epsilon = \sigma_y / E$. Using a tapered mandrel or a sleeve helps distribute the expansion force evenly around the circumference, preventing localized yielding at the ring's gap. If a ring is over-expanded, it will not seat tightly in the groove, leading to axial play and potential failure.

Elgiloy (a Co-Cr-Ni alloy) is selected for spiral retaining rings in subsea oil and gas tools due to its exceptional resistance to Sour Gas ($H_2S$) and its ability to maintain high strength at cryogenic and elevated temperatures. It is NACE MR0175 compliant, meaning it is resistant to sulfide stress cracking. The material's high modulus of elasticity ($E \approx 30 \times 10^6$ psi) and excellent fatigue endurance make it superior to 17-7PH in environments where cyclic loading and extreme corrosion coexist. Processing involves cold working followed by age hardening to achieve tensiles exceeding 250 ksi.

During the spiral coiling process, significant residual stresses are induced in the material as it is bent into its circular shape. Stress relieving involves heating the rings to a temperature below the critical range (e.g., $750^{\circ}F$ for carbon steel) for a specific duration. This stabilizes the ring's dimensions, prevents 'spring-back' or 'out-of-roundness', and improves the fatigue life by reducing the peak internal stress. For high-precision spiral rings, this process is essential to ensure that the ring stays flat and maintains its specified diameter over its shelf life and operational life.

316 Stainless Steel (UNS S31600) has slightly lower tensile strength than 302 (UNS S30200) due to its higher nickel and the addition of molybdenum. Consequently, a ring made of 316 will have a lower thrust capacity (approximately $10-15\%$ less) than a geometrically identical 302 ring. However, 316 is significantly more resistant to pitting and crevice corrosion in chloride-rich environments (e.g., seawater). Engineers must account for the lower yield strength ($ \sigma_y \approx 30$ ksi annealed, though higher when cold-worked for rings) when calculating the safety factor for the assembly.

Black Oxide (MIL-DTL-13924) is a conversion coating that provides a dark appearance and a small degree of corrosion resistance, primarily intended for indoor applications or where the ring will be submerged in oil. Zinc Phosphate (heavy) provides better corrosion protection and acts as a substrate for lubricants. For spiral rings, the coating must be thin enough to not interfere with the ring's ability to flex during installation. Carbon steel rings (SAE 1070-1090) must be carefully processed to avoid hydrogen embrittlement if these coatings involve an acid pickling step.

Beryllium Copper (Alloy 25, UNS C17200) is specified for its unique combination of high strength (comparable to steel), excellent electrical conductivity, and non-sparking properties. It is also non-magnetic, making it ideal for MRI medical equipment or sensitive electronic sensors. BeCu rings are typically age-hardened at $600^{\circ}F$ to achieve their full spring properties. Additionally, BeCu offers excellent corrosion resistance in marine environments. However, due to the toxicity of beryllium dust during manufacturing and the high material cost, its use is restricted to applications where these specific properties are mandatory.

The edge margin $Y$ is the distance from the groove to the end of the shaft or bore. It must be sufficient to prevent the material from shearing off under axial load. The formula is $Y = \frac{3P}{\pi \cdot D \cdot \sigma_y}$. If $Y$ is too small, the material behind the groove will experience shear failure or 'blowout'. In aerospace applications where weight is critical, $Y$ is optimized but generally kept to at least $3 \times$ the groove depth $d$ to maintain a safety factor of 2.0 or higher against the yield strength of the host material.

Dishing occurs when the thrust load causes the ring to bend into a conical shape. This is usually due to the groove's edge margin being too small or the groove wall deforming. As the ring dishes, it begins to 'walk' out of the groove. The relationship for the moment leading to dishing involves the load $P$ and the lever arm between the load application point and the groove support. If the groove radius is too large ($R > 0.1 \cdot T$), the ring will dish prematurely. The safety factor must be derated by a factor of $K \approx 1 - (R/T)$ to account for this geometric instability.