Torsional Vibration Analysis (Multi-DOF) Back
Vibration Analysis

Torsional Vibration Analysis — Multi-DOF Shaft System

Compute natural frequencies, mode shapes and Campbell diagrams for 2- to 4-disc torsional shaft systems in real time, and identify critical speeds for engines, turbines and gear trains at the design stage.

System Configuration
Number of discs n
Polar moment of inertia J (kg·m²)
J₁
kg·m²
J₂
kg·m²
J₃
kg·m²
Torsional stiffness k (N·m/rad)
k₁ (×10⁵)
kN·m/rad
k₂ (×10⁵)
kN·m/rad
Damping ratio ζ
%
Results
1st natural frequency f₁ (Hz)
2nd natural frequency f₂ (Hz)
ω₁ (rad/s)
Mode 1 node location
Mode Shapes (1st & 2nd)
Campbell Diagram (orders 1–6)
Natural Frequency Table
Modeω (rad/s)f (Hz)N_cr (rpm)
Theory & Key Formulas

$[K - \omega^2 M]\{\theta\}= \{0\}$

Stiffness matrix diagonal:
$K_{ii}= k_{i-1}+ k_i$
Off-diagonal: $K_{i,i+1} = -k_i$

What is Torsional Vibration Analysis?

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What exactly is "torsional vibration"? I get regular vibration, but what makes it torsional?
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Basically, it's the back-and-forth twisting of a rotating part. Imagine a car engine's crankshaft. It's not just spinning smoothly; it's constantly being "jerked" by the pistons, causing it to twist and untwist along its length. In this simulator, each disc you see represents a flywheel or gear, and the connecting lines are the elastic shafts that twist.
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Wait, really? So the whole system can vibrate in different patterns? How does that work?
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Exactly! Those patterns are called "mode shapes." For a 3-disc system, the first mode might have all discs twisting in the same direction. The second mode could have the middle disc still and the ends twisting opposite ways. Try it: use the "Number of Discs" selector above to switch from 2 to 4 discs and watch the mode shapes change. Each disc's inertia and shaft stiffness, which you control with the sliders, determines these patterns.
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You mentioned "critical speeds" in the description. Is that when the vibration gets dangerous?
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Yes, that's the key practical danger. A critical speed is when the engine's running speed matches a natural frequency of the shaft system. The vibration amplitude can become huge, leading to fatigue and failure. That's what the Campbell diagram in the simulator shows. Try increasing the "Engine Order" slider—you'll see lines crossing the natural frequency lines. Those intersections are the critical speeds you must avoid in design.

Physical Model & Key Equations

The core physics is described by an eigenvalue problem. For free vibration, we look for solutions where all discs oscillate at the same natural frequency ω. This leads to the classic equation balancing the system's stiffness (resistance to twist) and inertia (resistance to angular acceleration).

$$[K - \omega^2 M]\{\theta\}= \{0\}$$

Here, M is the mass (inertia) matrix (diagonal with values $J_1, J_2,...$), K is the stiffness matrix, ω is the natural frequency (rad/s), and {θ} is the mode shape vector (the relative twist angles of each disc). A non-zero solution only exists when the determinant $|K - \omega^2 M| = 0$, which is how we solve for ω.

The stiffness matrix K is assembled from the torsional stiffness $k_i$ of each shaft segment connecting the discs. For a disc $i$, its equation comes from the twist of the shafts on either side of it.

$$K_{ii}= k_{i-1}+ k_i, \quad K_{i,i+1}= -k_i$$

$k_i$ is the stiffness of shaft segment i ($k = \frac{GJ_p}{L}$, where $GJ_p$ is torsional rigidity). The diagonal term $K_{ii}$ is the sum of the stiffnesses connected to disc i. The off-diagonal term $K_{i,i+1}$ is the negative stiffness connecting disc i to its neighbor, coupling their motion. This structure is what you're changing when you adjust the "Shaft Stiffness" sliders in the simulator.

Real-World Applications

Marine Propulsion Systems: In large ship diesel engines connected to a long propeller shaft, torsional vibrations are a major design concern. Engineers use this exact multi-DOF analysis to place dampers and ensure no critical speed falls within the engine's operating range, preventing catastrophic shaft failure in the middle of the ocean.

Wind Turbine Drivetrains: The gearbox and generator in a wind turbine are subject to fluctuating torques from wind gusts. Analyzing the multi-disc system (blades, gearbox stages, generator) helps predict fatigue life and avoid resonance that could lead to expensive gearbox failures.

Automotive Crankshafts: In high-performance engines, the crankshaft, flywheel, and attached accessories (like a damper) form a torsional system. Identifying critical speeds ensures the engine can safely rev to high RPMs without inducing destructive vibrations that would break the crankshaft.

Aircraft Engine Turbine Shafts: The compressor and turbine discs in a jet engine are connected by a relatively slender shaft. Torsional analysis is crucial to ensure the system's natural frequencies are far from the excitation frequencies caused by blade passing, preventing high-cycle fatigue.

Common Misconceptions and Points to Note

First, the statement "A larger moment of inertia makes it harder to vibrate" is only half true and half false. While this holds for a simple single-degree-of-freedom system, the overall balance is critical in multi-inertia systems. For example, in a 4-inertia system, making only the end disks extremely heavy can cause a "localized vibration mode" where the lighter intermediate disks vibrate violently, potentially creating more problems. If you use the tool to increase only I1 and I4 and observe the mode shapes, you can see how I2 and I3 oscillate significantly.

Next, there's a pitfall: stiffness does not always mean "the stronger, the safer." Increasing the shaft diameter to raise torsional stiffness does increase the natural frequencies. However, when operating in high engine speed ranges, this can sometimes cause dangerous higher-order modes (e.g., the 4th mode) to descend into the operational speed range. Check the Campbell diagram to see how the lines for higher-order modes shift leftward when stiffness is increased.

Finally, consider interpreting results while ignoring damping. This simulator performs conservative (undamped) eigenvalue analysis, so it cannot determine the "severity" at resonance points. In reality, dampers are often present, and energy dissipates within the material. This means that even if a resonance point exists in the calculation, sufficient damping can suppress the amplitude, often preventing actual harm. Remember, this tool's role is to "identify potential problem areas"; final judgment requires transient response analysis that considers damping.

How to Use

  1. Enter polar moments of inertia (J1, J2, J3, J4) in mm⁴ for each disc; typical steel discs range 1e6–1e8 mm⁴
  2. Input shaft stiffness values (j1Val, j2Val, j3Val, j4Val) in N·mm/rad between adjacent discs; calculate as GIp/L where G=80 GPa for steel, Ip is polar second moment, L is segment length
  3. Click Solve to compute eigenvalues; simulator returns natural frequencies in Hz and corresponding mode shapes showing relative angular displacement
  4. Review Campbell diagram to identify critical speeds where running speed intersects natural frequency lines

Worked Example

Four-disc compressor shaft: J1=5e7 mm⁴, J2=8e7 mm⁴, J3=6e7 mm⁴, J4=4e7 mm⁴. Shaft segments: j1Val=1.2e6 N·mm/rad (steel, 80 mm diameter, 150 mm length), j2Val=9.5e5, j3Val=1.1e6, j4Val=8e5. Solver returns first natural frequency ω₁=187 rad/s (29.8 Hz), second mode ω₂=521 rad/s (82.9 Hz). Campbell diagram shows resonance avoidance band between 1800–2200 rpm required for safe operation below second critical speed.

Practical Notes

  1. Verify stiffness using FEA or analytical formulas; underestimated stiffness artificially lowers frequencies by 15–25%, causing missed resonances during commissioning
  2. Mode shape inspection reveals whether discs vibrate in-phase (first mode, lowest stress) or alternating (higher modes, stress concentration at bearings)
  3. For turbomachinery shafts, ensure operating speed stays 20% below lowest critical speed or above 120% of highest mode within speed range
  4. Damping ratio (~1–3% for steel shafts) not included in undamped solver; use results for preliminary design, validate with transient response analysis for damped behavior near resonance