FAQ

Ring shear capacity ($P_r$) is the axial load required to physically shear the ring's cross-section, calculated as $P_r = \frac{D imes t imes ext{\pi} imes S_s}{K}$, where $S_s$ is the shear strength of the ring material ($S_s \approx 0.577 imes S_u$). Usually, $P_r$ is significantly higher than the groove deformation capacity ($P_g$). For a ring to fail in shear, the groove must be deep and the housing material must be extremely hard. In most engineering failures, the groove wall yields first, leading to an angular deflection of the ring (dishing), which causes it to climb out of the groove before shear stress reaches the ultimate limit.

The thrust capacity is often limited by the shear strength of the groove material rather than the ring itself. The allowable thrust load $P_g$ is given by $P_g = \frac{D imes d imes ext{\pi} imes S_y}{K}$, where $D$ is the shaft/bore diameter, $d$ is the groove depth, $S_y$ is the yield strength of the groove material, and $K$ is a safety factor (typically 2). If the groove material is soft (e.g., aluminum), the groove wall will yield and deform ('dish'), causing the ring to 'pop out' under load. In such cases, increasing the groove depth or using a harder housing material is necessary to meet the axial load requirements.

For external rings, centrifugal force causes the ring to expand radially, which can eventually lead to the ring lifting out of the groove. The maximum RPM ($V$) is calculated using the formula $V = \frac{715}{D_o} \sqrt{\frac{E imes I imes ext{cl}}{_x000d_ho imes A imes R_m^4}}$, where $D_o$ is the shaft diameter, $E$ is the modulus, $I$ is the moment of inertia, 'cl' is the centrifugal clearance (groove depth - ring radial wall), $\rho$ is the density, $A$ is the cross-sectional area, and $R_m$ is the mean radius. Engineers must ensure the operating RPM is at least 20% below this theoretical limit. If higher speeds are required, a 'self-locking' feature (a tab and slot) is integrated to mechanically prevent expansion.

Multiple fractures (fragmentation) usually indicate resonance or harmonic vibration. If the frequency of the external load matches the natural frequency ($f_n$) of the spring, the internal stresses can amplify until they exceed the ultimate tensile strength. The natural frequency is $f_n = \frac{1}{2} \sqrt{\frac{K}{m}}$, where $m$ is the mass. In high-speed machinery, 'surge' waves travel through the spring coils. If the spring is not properly damped or if the operating speed is too close to $f_n$, the resulting 'dynamic stress' can be 5-10 times the static stress. Solution: Increase the spring's stiffness $K$ to move $f_n$ out of the operating range or add mechanical damping.

Fretting fatigue is characterized by reddish-brown (for steel) or black (for stainless) debris found at the contact points between waves. It occurs due to micro-oscillations between the turns of the spring under load. This micro-motion wears away the protective oxide layer, leading to accelerated oxidation and crack initiation. In aerospace applications, this is common in components subject to high-frequency engine vibration. Mitigation strategies include applying high-pressure lubricants (e.g., Braycote), using a material with higher surface hardness, or increasing the spring's preload to reduce the relative motion between the turns.

Hydrogen embrittlement occurs during the acid pickling or electroplating (e.g., Zinc plating) of high-carbon steel springs. Atomic hydrogen ($H$) diffuses into the steel's grain boundaries. Under tensile stress, these hydrogen atoms migrate to the tips of micro-cracks, reducing the cohesive strength of the lattice and causing brittle fracture at stresses far below the yield strength. To mitigate this, springs must be 'baked' immediately after plating (typically at $375^{\circ}F$ for 4 to 24 hours depending on thickness) to drive out the trapped hydrogen. Failure usually presents as a clean, 'glass-like' fracture surface with no evidence of plastic deformation.

Permanent set is diagnosed by measuring the free height ($L_f$) before and after service. If $L_f$ has decreased, the spring was stressed beyond its proportional limit. Possible causes include: 1) The assembly was 'over-deflected' to its solid height during installation; 2) The operating temperature exceeded the material's relaxation limit; 3) The actual load was higher than the design load due to dynamic surging. Troubleshooting involves checking the 'solid height' stress. If $S_{solid} > 80\%$ of $S_y$, the design is 'set-prone'. A 'preset' operation (compressing the spring to solid during manufacturing) can help stabilize $L_f$ by inducing beneficial residual stresses.

The primary failure mode is Stress Corrosion Cracking (SCC) or simple oxidation leading to pitting. Carbon steel (SAE 1070-1090) is highly susceptible to rust in the presence of moisture. Pitting acts as a severe stress concentrator ($K_t$). Because wave springs operate at high stress levels (often 30-50% of tensile strength), a pit can quickly transition into a fatigue crack. Once a crack initiates, the reduced cross-sectional area leads to rapid propagation until the spring snaps. To prevent this, carbon steel springs are typically finished with Zinc Phosphate (oil dipped) or Vapor Degreased and coated with an anti-corrosion oil.

An OD-piloted wave spring expands radially as it is compressed axially. If the bore surface is rough ($>63 \mu in$ Ra), the friction between the spring's OD and the bore increases significantly. This 'frictional locking' effectively increases the observed spring rate during compression because the friction opposes the radial expansion. During decompression, the spring may 'hang up' or return slowly. For precision applications like optical lens positioning, the bore should be hard-anodized or coated with a dry-film lubricant (e.g., MoS2) to ensure the spring rate remains consistent and repeatable within $\pm 1\%$ of the design value.

In nested wave springs (where two or more turns are coiled in parallel), the gap of the wire must be oriented correctly to ensure load balance. If the gaps are aligned, there is a localized 'weak spot' in the spring's circumference. In high-vibration environments, such as automotive transmissions, this can lead to 'fretting' at the gap edges and uneven wear of the mating surfaces. For optimal performance, the gaps of consecutive springs (if using multiple single-turn nested springs) should be staggered by $180 / N$ degrees, where $N$ is the number of waves, to ensure the most consistent radial load distribution and prevent harmonic resonance.

Snaking is a buckling failure where a multi-turn spring moves radially out of alignment under axial load. This occurs when the free height to mean diameter ratio ($L_f / D_m$) exceeds approximately $1.5$. To prevent snaking, the spring should be guided by a bore or a shaft. If the application does not allow for a pilot, the spring must be designed with a lower $L_f / D_m$ ratio or use a nested design which is inherently more stable. For high-deflection applications, internal guides or 'sleeves' are often integrated into the assembly to ensure the spring compresses axially and maintains a uniform load distribution across its circumference.

Shims are used to calibrate the Work Height $H_w$ and compensate for tolerances in the housing and shaft. Since $P = K(L_{free} - H_w)$, a small error in $H_w$ leads to a significant error in $P$. By using a precision-ground shim, the assembly's stack height can be controlled to within $\pm 0.001$ inch. In multi-turn springs, the shim also provides a flat, parallel bearing surface. Without a flat surface, the 'wave' of the spring may not seat properly, leading to an 'off-center' load or 'tipping' moment, which induces non-uniform stresses and potentially causes the spring to 'snake' or buckle sideways within the bore.

When a wave spring is installed on a shaft (ID piloted), the shaft diameter $D_s$ must be sized to account for the spring's ID contraction during compression. Although wave springs primarily expand at the OD, there is a minor geometric change at the ID. The recommended clearance is typically $0.005$ to $0.020$ inches depending on the spring size. If the shaft is too large, the spring will bind, causing a localized stress increase and preventing the spring from reaching its intended work height. The shaft should be polished to at least $32 \mu in$ Ra to minimize wear on the spring's inner edges during cycling.

MP35N is a multiphase alloy (Ni-Co-Cr-Mo) that offers a unique combination of ultra-high strength (up to 300 ksi tensile) and excellent biocompatibility. Compared to 316L, MP35N has a much higher fatigue limit and modulus of elasticity. In medical devices like heart valves or orthopedic implants where the spring must be tiny yet powerful, MP35N allows for extreme miniaturization. 316L is often too soft ($S_y \approx 40$ ksi annealed) to function as a high-performance spring without excessive bulk. Furthermore, MP35N is highly resistant to crevice corrosion and pitting in saline body fluids, where 316L might fail over long durations.

Wave springs are typically manufactured from flat wire produced by rolling round wire. This process creates a natural round edge or 'deburred' edge. Any micro-fissures or sharp 'burrs' on the edge of the wire act as stress risers (K_t). In a cyclic application, these risers facilitate crack initiation. For high-cycle fatigue ($>10^6$ cycles), wire must be inspected for surface decarburization (in carbon steels) and edge quality. Polishing or vibratory finishing the springs after coiling and heat treatment can increase fatigue life by 20-30% by removing surface defects and inducing a slight compressive residual stress on the material's exterior.

Elgiloy (complying with NACE MR0175) is the gold standard for sour gas environments due to its extreme resistance to Stress Corrosion Cracking (SCC) and Hydrogen Embrittlement. Unlike carbon steels or even some stainless steels, Elgiloy's cobalt-base chemistry prevents the formation of brittle hydrides. Its fatigue resistance is also superior in corrosive media. When processing Elgiloy wave springs, the material is typically cold-worked and then age-hardened at $900^{\circ}F$ to $1000^{\circ}F$. This results in a material with high elastic modulus ($E \approx 28.5 \times 10^6$ psi) and high corrosion fatigue limits, essential for the 20-year service life required in subsea BOP (Blowout Preventer) stacks.

The CH900 condition (Cold Reduced and Aged at $900^{\circ}F$) provides the highest possible tensile and yield strength for 17-7PH stainless steel. In aerospace actuators, weight and space are constrained, necessitating high power density. By using CH900, engineers can design springs with thinner material $t$ while maintaining the same load $P$, because $P \propto t^3$ but stress $S \propto t^{-2}$. The aging process at $900^{\circ}F$ precipitates aluminum-rich intermetallic compounds that pin dislocations, increasing the yield strength to approximately 260 ksi. This allows for higher operating stresses and greater deflection ranges without permanent set compared to the RH950 or TH1050 conditions.

Inconel X-750 (AMS 5699) is a nickel-chromium alloy that remains ductile and maintains its elastic modulus at cryogenic temperatures, whereas many martensitic steels become brittle. 17-7PH, while excellent at room temperature, can suffer from reduced fracture toughness at temperatures below $-100^{\circ}F$. Inconel X-750's precipitate phase (gamma prime) provides stability down to $-400^{\circ}F$. For cryogenic seals, Inconel X-750 is preferred due to its lower coefficient of thermal expansion and resistance to hydrogen-assisted cracking in specialized subsea or aerospace propellant environments. The heat treatment for X-750 typically involves a solution anneal followed by age hardening ($1350^{\circ}F$ for 20 hours) to maximize yield strength.

The Work Height ($H_w$) is the axial space the spring occupies under a specific load. To avoid plastic deformation, the stress at $H_w$ must not exceed the yield strength ($S_y$) of the material. For nested springs, the stress is $S = \frac{3 \pi \cdot P \cdot D_m}{4 \cdot b \cdot t^2 imes N^2}$. The safety factor is $SF = S_y / S$. If $H_w$ is too close to the solid height ($H_s = n \cdot t$), the spring enters the non-linear range where the actual load exceeds the calculated linear load due to contact between turns. Designers must ensure that $H_w > H_s + (0.1 \times f_{total})$ to maintain linearity and prevent early fatigue failure caused by local stress concentrations at the solid height.

Friction in multi-turn wave springs occurs at the contact points between waves (crests) and at the interface between the spring and the bore or shaft. This leads to a hysteresis loop in the load-deflection curve where the loading force $P_{load} = P_{theoretical} + F_{friction}$ and the unloading force $P_{unload} = P_{theoretical} - F_{friction}$. The magnitude of $F_{friction}$ is a function of the coefficient of friction $\mu$ and the normal force at the contact points. In high-frequency applications, this friction generates heat, which can lead to localized thermal expansion and a reduction in the fatigue life. For precision applications, lubrication or PEEK coating is applied to reduce $\mu$ and flatten the hysteresis loop.