From the Rolls-Royce experimental archive: a quarter of a million communications from Rolls-Royce, 1906 to 1960's. Documents from the Sir Henry Royce Memorial Foundation (SHRMF).
Page discussing diesel crankshaft vibration, including analysis of harmonic orders, methods for increasing crankshaft frequency, and a stress graph.
| Identifier | ExFiles\Box 132\1\ scan0122 | |
| Date | 25th March 1939 | |
| 406 DIESEL CRANKSHAFT VIBRATION Table 7 HARMONIC ORDERS | FIRING ORDER 1-2-4-3 PHASE DIAGRAM | VECTOR DIAGRAM 1/2 - 2 1/2 - 4 1/2 - 6 1/2 | 1 - 3 1/2 - 5 1/2 - 7 1/2 | Σϵ=0.6875 1 - 3 - 5 - 7 | Σϵ=0.186 2 - 4 - 6 - 8 | Σϵ=2.846 torque increases the stress prohibitively. For example, the harmonic coefficient P for the third harmonic order will, when the components for gas pressure and inertia force are added, increase from ± 25 to ± 220, or multiply almost eight times, and the vibration stresses will be increased in the same proportion. The inertia torque has no influence on the sixth and higher harmonics and can be disregarded in connection with these. Therefore, the higher the frequency of the crankshaft assembly, the better, as it gives assurance against shaft breakage and tends toward smoother engine performance. Following are some means by which the frequency of the assembly can be increased. 1. Use of a shorter stroke, which decreases the moment of inertia of the crankpin and crank arms and the equivalent moment of inertia of the reciprocating part, and increases the torsional rigidity of the crankshaft. 2. Placing the cylinders as close together as possible, which reduces the length of the crankshaft. 3. Reducing weights of reciprocating and rotating parts to the minimum. There is much to be done by way of improving present-day piston material, and the heat conductivity of the material, with a view to reducing the weight. 4. The amplitude of vibration at the free end of the crankshaft should not exceed 0.5 deg., in order that the valve mechanism may function properly. If the amplitude of vibration is greater than 0.5 deg., the installation of a vibration damper at the forward end of the crankshaft should be considered. 5. Camshaft-driving gears and all other masses that are usually mounted on the forward end of the crankshaft, where the amplitude of torsional vibration is greatest, should be transferred to the rear, to a point near the node, where the amplitude is small. This will assure better valve-operating conditions. 6. In view of the fact that the camshaft drive is universally located at the forward end in present-day engines, a specially-flexible device for the driving gears is advisable, one that will absorb shocks due to crankshaft vibration. 7. The operating speed range of the engine should be kept below the critical speeds of harmonic orders up to the fourth, as at these critical speeds the inertia torque increases the stresses on the shaft considerably. 8. Hollow crankpins may be used. 9. Circular crank arms are preferable to rectangular ones, as the former have a smaller moment of inertia and greater width and therefore greater rigidity. 10. Counterweights on the crank arms, in conjunction with the flywheel, will facilitate starting, improve the smoothness of running, and result in uniform distribution of the masses along the length of the crankshaft. The counterweights at the same time will counteract the primary inertia force due to the moving masses of the individual cylinder. Any change in the frequency of the crankshaft assembly and of the values of the critical speeds will be shown immediately by changes in the shape of the normal elastic curve, and in the calculated stresses, and these two together will give a graphic picture of the possible performance from the standpoint of torsional vibration and smooth running of the engine. The calculations can later be checked by means of the torsiograph applied to the engine in running condition, and the performance factors established. Bibliography 1. FRAHM, H.{Arthur M. Hanbury - Head Complaints}—Recent Investigation Concerning Dynamical Action in Shafting of Marine Engines with Special Reference to Synchronous Torsional Vibration. Jl. Am. Soc. Naval Engrs., August, 1902. 2. GÜMBEL, L.—Torsional Vibration of Shafts. Jl. Am. Soc. Naval Engrs., May, 1902. 3. HOLZER, H.{Arthur M. Hanbury - Head Complaints}—Die Berechnung der Drehschwingungen. Berlin, 1921. 4. LEWIS, F.{Mr Friese} M.{Mr Moon / Mr Moore}—Torsional Vibration in Diesel Engines. Soc. Naval Arch. & Marine Engrs., November, 1925. 5. CARTER, B. C.—An Empirical Formula for Crankshaft Stiffness in Torsion. Engineering, July 13, 1928. (Turn to page 408, please) Figure 20 Graph Text: VIBRATION STRESS - LB. PER SQ. IN. REVOLUTIONS PER MIN. MAX. TENSILE STRESS OF STEEL 120,000 LB. PER SQ. IN. 4 CYL. ENGINE M.I.P 100 LB. PER SQ. IN. ω² = 7.5 x 10⁶ RAD.² -SEC.² f = 26,148 CYCLES PER MIN. FIRING ORDER 1-2-4-3 HARMONIC 6 ORDER SPEED RANGE OF ENGINE ENDURANCE LIMIT - REVERSE BENDING & TORSION SAFE STRESSES HARMONIC 12 ORDER Harmonic Orders shown: 12, 11 1/2, 10 1/2, 10, 9 1/2, 9, 8 1/2, 8, 7 1/2, 7, 6 1/2, 6, 5 1/2, 5, 4 1/2, 4 March 25, 1939 Automotive Industries | ||