FAQ

Dish-out is the axial deflection of the ring's inner or outer edge under load, transforming the flat ring into a conical shape. As the ring dishes, the effective contact area with the groove decreases, and the load vector shifts, creating a moment that promotes further deformation. Under dynamic impact (e.g., in pneumatic hammers), dish-out can lead to the ring 'walking' out of the groove. The amount of dish is proportional to $P \cdot R^2 / (E \cdot t^3)$. To counteract this, multi-turn spiral rings (2-turn or 3-turn) are used; they provide significantly higher moment of inertia ($I$) than a single-turn stamped circlip, thereby resisting dish-out and maintaining 360-degree contact even under shock loading.

The assembly thrust capacity is the lower of the Ring Shear Strength ($P_r$) and the Groove Yield Strength ($P_g$). $P_r = \frac{D \cdot t \cdot π \cdot S_s}{K}$, where $S_s$ is the shear strength and $K$ is a safety factor (usually 3). $P_g = \frac{D \cdot d \cdot π \cdot S_y}{K}$, where $d$ is the groove depth and $S_y$ is the yield strength of the housing/shaft material. In almost all cases involving soft materials like aluminum (6061-T6), the groove is the limiting factor. The groove will 'roll' or deform plastically, causing the ring to dish out and eventually fail. Engineers must calculate the 'groove deformation' limit using the formula $P_{def} = \frac{A_s \cdot S_y}{\phi}$, where $\phi$ is a factor accounting for the angle of the load.

The centrifugal force acting on a rotating ring tends to expand it. The critical speed is reached when the centrifugal force equals the ring's radial grip on the shaft. The formula is $N_{max} = \sqrt{\frac{4.48 \cdot 10^9 \cdot E \cdot I \cdot g}{\rho \cdot A \cdot R^4 \cdot (1+n)^2}}$ where $E$ is the modulus, $I$ is the moment of inertia, $\rho$ is the density, $A$ is the cross-sectional area, and $R$ is the mean radius. For high-RPM applications like electric vehicle motors, we use 'Self-Locking' spiral rings. These rings have a small tab that locks into a notch on the adjacent turn, mechanically preventing the ring from expanding due to centrifugal forces, thus allowing speeds $3-5$ times higher than standard rings.

SCC in 302SS wave springs is characterized by brittle-like transgranular or intergranular cracking, often with little to no visible macroscopic deformation. Under a Scanning Electron Microscope (SEM), the fracture surface will show 'branching' cracks—a hallmark of SCC. This occurs due to the combined effect of tensile stress (at the wave peaks) and a corrosive medium (chlorides from seawater). In marine environments, the 'crevices' formed where the waves touch are particularly vulnerable as they trap salt. To solve this, 316SS is often used for its molybdenum content which resists pitting, or for even higher performance, MP35N or Inconel 625 is utilized.

Wave shifting occurs when the peaks of one turn do not stay aligned with the valleys of the adjacent turn. In a Crest-to-Crest spring, if the turns shift, they may 'nest' (peak-to-valley), causing a sudden drop in the spring's effective height and an increase in the spring rate (since the number of active turns $n$ effectively decreases). This is usually caused by lack of radial constraint or high-frequency lateral vibrations. To prevent wave shifting, many engineers specify 'interlocking' waves or use a 'Wavo' spring (a round-wire wave spring) which has higher lateral stiffness. Alternatively, a tight-fitting pilot shaft can help maintain the axial alignment of the turns.

While often used interchangeably, they are distinct. Relaxation is the reduction in load $P$ over time while the spring is held at a constant deflection $f$. Creep is the increase in deflection $f$ over time while the spring is held at a constant load $P$. In most wave spring applications (e.g., preloading a bearing), the spring is held at a constant height, so relaxation is the primary concern. The rate of relaxation follows the Arrhenius equation: $Rate = A \cdot e^{-Q/RT}$. In critical aerospace applications, we must derate the initial load by 5-10% to account for the expected relaxation over the component's service life, especially if operating near the material's thermal limit.

Permanent set (plastic deformation) occurs when the stress at the wave peaks exceeds the material's yield strength $\sigma_y$. If this happens at high temperatures, it is likely due to 'stress relaxation'. Even if the calculated stress was below $\sigma_y$ at room temperature, the yield strength of materials like 17-7PH drops as temperature increases. Another possibility is 'over-compression' to the solid height where the actual stress $\sigma = \frac{48 \cdot E \cdot t \cdot f}{\u03C0^2 \cdot D_m^2}$ exceeds the elastic limit. To fix this, the spring should be 'preset' (compressed to solid height during manufacturing) to induce beneficial compressive residual stresses, or the material should be upgraded to a superalloy like Nimonic 90.

Fretting wear occurs at the wave peaks (the interfaces between turns) due to microscopic relative motion (oscillatory slip) during high-frequency vibration or cycling. This wear removes the protective oxide layer of the stainless steel, leading to localized pitting and stress concentrators that initiate fatigue cracks. Root causes include insufficient axial preload (allowing excessive movement) or resonance with system frequencies. Mitigation involves: 1) Increasing the initial preload to 'lock' the turns; 2) Applying a dry-film lubricant like $MoS_2$ or PTFE to reduce the coefficient of friction; or 3) Using a spring material with higher surface hardness, such as an aged Inconel X-750 or nitrided steel.

In mechanical seals, the wave spring must maintain a precise face pressure. The preload $P$ is sensitive to the installed height $L_i$. Due to the stack-up of tolerances in the seal housing, gland, and carbon face, $L_i$ can vary significantly ($ΔL$). Since $P = k \cdot (L_0 - L_i)$, any variation in $L_i$ results in a variation in $P$. If $k$ is high, even a small $ΔL$ can cause the seal to leak (low $P$) or wear prematurely (high $P$). Designers often select Crest-to-Crest springs with a 'flatter' spring rate (lower $k$) over a longer travel to minimize the sensitivity of the preload to these unavoidable manufacturing tolerances.

Installing nested wave springs into deep blind holes requires a dedicated insertion tool to prevent the turns from tangling or 'shingling' (where layers overlap incorrectly). The tool should be a plunger with a diameter slightly smaller than the spring $ID$. The spring is pre-compressed onto the tool, then inserted. It is vital to ensure that the individual layers are seated flush against the bottom of the hole. For high-volume automotive assembly, automated pick-and-place systems use a vacuum-assisted mandrel that holds the spring by its $ID$, ensuring that the multiple turns remain perfectly concentric during the high-speed insertion into the transmission housing.

For a bore-piloted wave spring, the outside diameter ($OD$) of the spring is the primary datum. The housing bore must be sized to accommodate the $OD$ at its maximum radial expansion (at solid height). If the bore is too tight ($D_{bore} < OD_{max}$), the spring will 'hoop' and seize, causing the spring rate to spike. Conversely, if the bore is too loose, the spring can shift off-center, leading to an uneven load $P$ across the wave peaks. In precision assemblies, a tolerance of H7/h6 is often targeted for the bore, and the spring $OD$ is specified with a tolerance that accounts for the manufacturing variation of the flat wire width $b$.

Springs with a high $L_0/D$ ratio (typically $> 1.5$) are prone to buckling during installation and operation. When compressed, the spring acts like a slender column. If not properly constrained by a bore or guided by a pilot (shaft), the spring will bow laterally, leading to non-axial loading and catastrophic failure of the waves. The critical buckling load is defined by $P_{crit} = π^2 \cdot E \cdot I / (K \cdot L)^2$. To prevent this, a housing bore should be used with a clearance of approximately 0.05mm to 0.15mm per side. Additionally, if the spring is used in a dynamic system, the guiding surface must be hardened to prevent the 'snaking' spring from wearing into the housing.

In high-speed bearing applications (e.g., turbochargers), the interface between the wave spring and the bearing race is critical. 'Plain ends' terminate at a wave peak, which can create localized 'point' loading, potentially leading to race distortion or uneven wear. 'Shim ends' (or flat ends) are integrated into the final 360 degrees of the spring, providing a flat parallel surface. This ensures that the spring force is distributed evenly over the entire circumference of the bearing race ($360^∘$ contact). This uniform preload is essential for preventing ball skidding and maintaining the stiffness of the spindle assembly at high RPMs, though it does slightly increase the spring's solid height.

Edge-winding (also known as 'No-Tooling-Cost' coiling) involves winding a pre-tempered flat wire on its edge. This process preserves the grain flow of the material along the circumference of the spring, which significantly enhances fatigue life and load consistency. In contrast, stamping involves punching the spring from a sheet, which results in 'cross-grain' orientation at various points around the circle, creating weak spots susceptible to premature failure. Furthermore, edge-winding eliminates the scrap associated with stamping and allows for the easy production of multi-turn Crest-to-Crest springs with shim ends, which provide a $360^∘$ contact surface for more uniform load distribution.

Cryogenic treatment of 302 stainless steel wave springs, involving immersion in liquid nitrogen ($-196^∘C$), promotes the transformation of retained austenite to martensite. This transformation increases the hardness and wear resistance of the wave peaks. More importantly for precision optics or cryo-valves, it improves dimensional stability by relieving internal stresses caused by the edge-winding process. Without cryo-processing, the spring might undergo subtle 'walking' or diameter changes when cycled between ambient and cryogenic temperatures, which would compromise the axial preload on sensitive lens assemblies.

Elgiloy is the gold standard for implantable medical wave springs due to its biocompatibility, extreme fatigue resistance, and non-magnetic properties. It exhibits a high UTS of up to 2600 MPa after cold work and aging. Its modulus of elasticity ($E = 190$ GPa) is stable across a wide temperature range. In medical pumps, Elgiloy's resistance to pitting and crevice corrosion in saline environments (body fluids) outperforms 316L stainless steel. The processing involves a cold-reduction of approx 85% followed by aging at $500^∘C$ for 5 hours. This creates a dense dislocation network that prevents the initiation of micro-cracks during the millions of cycles required for heart-assist devices.

High-carbon steel like SAE 1070 is highly susceptible to hydrogen embrittlement (HE) during acid pickling and electroplating processes (e.g., zinc or nickel plating). Atomic hydrogen diffuses into the grain boundaries, particularly in the high-stress areas at the wave peaks, leading to brittle fracture under load. To mitigate this, a 'bake-out' process is mandatory. The springs must be placed in an oven at $190^∘C$ to $220^∘C$ within 1 to 4 hours of plating. The baking duration (typically 8 to 24 hours) allows the hydrogen to diffuse out of the steel matrix. For critical aerospace components, mechanical plating or organic coatings are often substituted to eliminate the risk of HE entirely.

At operating temperatures of $350^∘C$, 17-7PH stainless steel begins to suffer from over-aging and a significant reduction in yield strength, leading to stress relaxation (creep). Inconel X-750 (AMS 5699) is the superior choice for subsea O&G due to its exceptional resistance to chloride-induced stress corrosion cracking (SCC) and high-temperature stability. X-750 undergoes precipitation hardening through the formation of $\gamma'$ phase ($Ni_3(Al, Ti)$), which maintains its mechanical properties up to $700^∘C$. The spring must be heat treated to the No.1 temper (HT) to optimize for relaxation resistance. Furthermore, X-750's lower modulus ($E \approx 213$ GPa) compared to 17-7PH ($E \approx 200$ GPa at temp) requires a slight adjustment in the number of turns to maintain identical load characteristics.

The number of waves $N$ is inversely proportional to the deflection and directly proportional to the cube of the load capacity. Specifically, the load $P$ is proportional to $N^4$. However, increasing $N$ reduces the arc length between peaks, which increases the stiffness but can lead to manufacturing difficulties and higher stress concentrations at the peaks. For stability, $N$ should be an integer or half-integer to ensure uniform load distribution. In applications with rotating shafts, an odd number of waves is often preferred to avoid harmonic resonance with the shaft's rotational frequency, especially when the operating speed approaches the spring's natural frequency $f_n = \frac{1}{2π}\sqrt{\frac{k}{m}}$.

Fatigue life estimation for wave springs involves calculating the alternating stress $\sigma_a$ and the mean stress $\sigma_m$. For a wave spring, the maximum stress occurs at the wave peaks and is given by $\sigma = \frac{3 \cdot π \cdot E \cdot t \cdot N^2 \cdot f}{4 \cdot D_m^2 \cdot n}$. Using the Goodman relation $\frac{\sigma_a}{\sigma_e} + \frac{\sigma_m}{\sigma_{ut}} = 1$, where $\sigma_e$ is the endurance limit and $\sigma_{ut}$ is the ultimate tensile strength (approx. 1650 MPa for 17-7PH CH900), engineers can determine if the spring will survive $10^6$ cycles. In high-cycle applications like medical pumps, we aim for a safety factor $S_f > 1.2$. Residual stresses from the coiling process must be relieved via heat treatment at $480^∘C$ to ensure the mean stress does not shift during service.