Many wire-rope manufacturers and machine designers are under the impression that the significant stress in a wire rope is the tensile stress, or possibly the stress due to tension and bending. This paper proves by mathematical analysis that by far the greatest stress in a wire rope results from Hertz contact stresses at points of contact of wire-on-wire, and asserts that the usual mode of failure of a wire rope is fretting-fatigue initiated at such points of contact. Design relationships based on these concepts should be of great value to designers who use wire rope.
DISCLAIMERSThe findings in this report are not to be construed as an official Uepartment of the Army position unless so designated by other authorized documents.When Government drawings, specifications, or other data are used for any purpose other than in connection with a definitely related Government procurement operation, the U.S. Government thereby incurs no responsibility nor any obligation whatsoever; and the fact that the Government may have formulated, furnished, or in any way supplied the said drawings, specifications, or other data is not to be regarded by implication or otherwise as in any manner licensing the holder or any other person or corporation, or conveying any rights or permission, to manufacture, use, or sell any patented invention that may in any way be related thereto.Trade names cited in this report do not constitute an official endorsement or approval of the use of such commercial hardware or software. DISPOSITION INSTRUCTIONSDestroy this report when no longer needed. Do not return it to the originator. / -madUnclassified Security CUt>ific»tton DOCUMENT CONTROL DATA R&D nmciNATiNC ACTIVITV n ntf.nrmf •uthortThe objectives of this program were to analyze technology applicable to tension members at it relates to the functional requirements of heavy outslzed loads externally suspended from helicopters, and to develop a comprehensive design theory and conceptual designs for tension members which will provide a basis for future detail design, fabrication, and test programs.The tension-member concepts selected for study Included wire rope, wire-rope belt, synthetic rope, synthetic tape, steel tape, roller chain, and jointed links. A weightedparameter technique was used to begin evaluation of these candidate concepts, followed by an analysis of practical considerations with reference specifically to the 1972, 1975, jj and 1980 time frames.The results of this study Indicate that, on a long-term basis, synthetic tape and wirerope belt are the most promising concepts. Only wire rope Is acceptable In the short term (1-2 years). Problems that remain to be solved for synthetic tape Include high aerodynamic drag and high stored elastic energy. Wire-rope belt Is an untried concept for this high-load application. Since high-strength synthetic material Is being studied In other programs, wire-rope belt was chosen for the preliminary design phase of this program.
IntroductionThis paper presents a numerical technique for obtaining" the rotational critical frequencies of multiple-span flexible shafts of nonuniform cross section with modern automatic computing equipment. This type of calculation is important in fields such as steam and gas turbines, windtunnel compressors, and propeller shafts. To guarantee smooth running of the mechanism, the normal operating speeds must be adequately separated from the critical speeds. The latter must, therefore, be calculated with reasonable accuracy.The advent of high-speed automatic computers has reduced time required to compute critical speeds in complex mechanisms, since computers are capable of performing extensive arithmetic calculations rapidly and accurately. In order to realize this advantage, however, it is necessary to specify fully the method of calculation and eliminate all engineering judgment decisions. This paper describes such a development. A previously published technique by M. A. Prohl [1] for the calculation of critical frequencies has been modified and advanced to render it suitable for use on automatic computers, and a program has been written and used on an IBM Type 650 calculator. The theoretical foundations of Prohl's calculation are presented in the appendix. Section 2 presents the modifications required to adapt the method to the computer. It will be seen that the modifications employed could also have been applied to other published methods, thereby rendering them equally amenable to computer solution.Section 3 contains a description of the computer program as written. In section 4, some results of our experience with the use of this program are presented. Modifications of Basic TheoryIn the appendix, the differential equation d'[ d y]dx--~ EI Tx2. j = t~o~2y(1) * Present address: Technical Operations, Incorporated. t Numbering of equations agrees with the derivation in the appendix. which represents the deflection y of a shaft whirling at a critical speed w is transformed into a set of recursion relations by breaking the shaft into n sections of length (hx)~. The mass of the shaft is distributed by averaging the mass of each two adjoining sections and assuming this average mass lumped at the junction between the sections. The recursion relations are 2 V,, = Vn-1 + mnwyn-1, -Mn = Mn-1 + V,~(Ax),,, = -o -+ O,,_~, Y,,=Y,,-i+ f~n (2M,_~fi-M,~) (Ax)" where V= M= 0--y= mn = Ax= (6)t + o._,(~)., speed, radians/sec shear, lb moment, inch-lb slope, radians deflection of point x, inches concentrated mass at junction point length of shaft section, inches flexibility constant of the section = Ax/EI, 1/inch-lb E = modulus of elasticity, psi I = moment of inertia of section, in 4.The critical speeds for any particular shaft can be found by the method of excess moments. The excess moment, M~x¢, is found by assuming appropriate initial values of V, M, 0, and y, choosing a value for w, and integrating to the terminal end of the beam by repeated use of equations (6). This process is repeated for various values of until those w...
To obtain information concerning the ability of conventional parallel-shaft gears to transmit motion smoothly at low velocities under high load, experiments were conducted on a single commercial gear mesh about the same size as the last stage proposed for a radio telescope drive. Two torque levels of 130,000 and 197,000 lb-in. were used. The largest amplitude of angular disturbance which was measured was less than 10 arc sec, while the average disturbance was about one arc sec at the low speed shaft. The smooth performance of the industrial-quality gear trains eliminated the expense of special designing. In similar applications where smooth motion is a requirement, components of good reliability and comparatively ready availability can be specified with confidence.
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