Search results

1 – 10 of over 6000
Article
Publication date: 29 June 2010

Zhimeng Luo, Jianzhong Zhou, Xiuqiao Xiang, Yaoyao He and Shan Peng

Shaft orbit is an important characteristic for vibration monitoring and diagnosing system of hydroelectric generating set. Because of the low accuracy and poor reliability of…

Abstract

Purpose

Shaft orbit is an important characteristic for vibration monitoring and diagnosing system of hydroelectric generating set. Because of the low accuracy and poor reliability of traditional methods in identifying the shaft orbit moving direction (MD), the purpose of this paper is to present a novel automatic identification method based on trigonometric function and polygon vector (TFPV).

Design/methodology/approach

First, some points on shaft orbit were selected with inter‐period acquisition method and joined together orderly to form a complex plane polygon. Second, by using the coordinate transformation and rotation theory, TFPV were applied comprehensively to judge the concavity or convexity of the polygon vertices. Finally, the shaft orbit MD is identified.

Findings

The simulation and experiment demonstrate that the method proposed can effectively identify the common shaft orbit MD.

Originality/value

In order to identity the shaft orbit MD effectively, a novel automatic identification method based on TFPV is proposed in this paper. The problem of identifying the shaft orbit MD is transformed into the problem about orientation of complex polygons, which are formed orderly by points on orbit shaft, and TFPV are applied comprehensively to judge the concavity or convexity of the polygon vertices.

Details

Sensor Review, vol. 30 no. 3
Type: Research Article
ISSN: 0260-2288

Keywords

Article
Publication date: 1 December 1951

J. Morris

1. Introduction THE effect of torque is usually neglected in the Bernoulli‐Euler treatment of the flexure of thin rods. In practical cases the torque will rarely have any…

Abstract

1. Introduction THE effect of torque is usually neglected in the Bernoulli‐Euler treatment of the flexure of thin rods. In practical cases the torque will rarely have any appreciable influence on static stability but from the dynamical aspect it may be the precursor of exponential instability in the absence of adequate damping forces. This phenomenon was suggested to me by Mr A. C. Hutchinson of Aliens, Bedford. Owing to torque there are cross deflexion coefficients which in particular cases are equal and opposite; and it is this negative reciprocity that gives rise to dynamic instability as we shall see. In the case, say, of a load on an overhung shaft it would appear that this instability is manifest at all speeds becoming more violent with the relative increase of torque. In practice it has frequently been found to be difficult if not impossible safely to run through particular ‘whirling’ speeds. In this connexion the problem of the stability of rotating shafts has reached a fresh peak of importance in consequence of the rapid development of the gas turbine as the prime mover in aircraft. The fundamental principles underlying the treatment of whirling phenomena are now well established and there is ample practical experience in support of the physical notion of shaft revolution propounded by Jeffcott in the Phil. Mag. for March 1919. Thus if a shaft is perfectly straight, completely balanced, and runs in bearings which are truly aligned; then, provided such masses as may be carried by the shaft have no appreciable moments of inertia, the normal ‘static’ flexural vibrations of the shaft will be independent of any imposed rotation, that is the path of any point on the neutral axis of the shaft will be uninfluenced by the rotation of the shaft section of which it is the centre. If the shaft is initially deflected by its weight, then the static vibrations will occur about the position of rest of the shaft and the imposed rotation will only apply to the cross‐sections of the shaft about their centres. If, however, the shaft is initially bent then each element may be considered as being ‘out of balance’ by the amount it is off the position it would occupy if the shaft were true and at rest. If we apply power to a bent shaft such as to maintain it in rotation with constant angular velocity then each element of shaft is regarded as being compelled to rotate with this constant angular velocity about the point on the effective elastic axis of the shaft, that is the axis of the deflected but corresponding unbent shaft. The unbalance will thus give rise to a forced vibration which will compel each element of the shaft to describe a circular path about its appropriate centre on the effective elastic axis; and this forced circular vibration will be independent of the static vibrations which are unforced and in consequence described as ‘free’. The radius of the path of any element in the forced circular vibration will depend inter alia on the speed of the imposed rotation, and the strain energy of the shaft so bent by flexure will be drawn or absorbed from the power which drives the shaft.

Details

Aircraft Engineering and Aerospace Technology, vol. 23 no. 12
Type: Research Article
ISSN: 0002-2667

Article
Publication date: 21 September 2012

R. Usubamatov, S.A. Adam and A. Harun

The purpose of this paper is to investigate the process of jamming of the hollow parts on the shaft and to derive a mathematical model for jamming in an assembly process.

Abstract

Purpose

The purpose of this paper is to investigate the process of jamming of the hollow parts on the shaft and to derive a mathematical model for jamming in an assembly process.

Design/methodology/approach

The mathematical model for jamming of parts on the shaft in an assembly process is based on the sizes, geometry, angular declination of part and shaft axes, and the frictional factor.

Findings

The equation for angular positional tolerance of coaxial parts and shafts, based on their geometry and sizes and leading to jamming, was derived.

Research limitations/implications

A mathematical model of parts jamming on the shaft is developed for assembly mechanisms. This research does not consider flexible deformations of components in assembly mechanisms, which results in the axis concentricity of part and shaft in the assembly process.

Practical implications

The results presented in the form of angular positional tolerance for coaxial parts and shafts based on their geometry and sizes make it possible to avoid the jamming of the parts. The results allow for formulating the angular positional tolerance of the assembly mechanisms that clamp the parts.

Originality/value

The proposed method for calculating the angular positional tolerance of coaxial parts and shafts for the assembly process should allow for increasing the reliability of the assembly process in the manufacturing industry.

Details

Assembly Automation, vol. 32 no. 4
Type: Research Article
ISSN: 0144-5154

Keywords

Article
Publication date: 13 November 2017

Jun Zha, Yaolong Chen and Penghai Zhang

The form error of shaft and hole parts is inevitable because of the machining error caused by rotation error of tool axis in machine tools where the elliptical form error is the…

Abstract

Purpose

The form error of shaft and hole parts is inevitable because of the machining error caused by rotation error of tool axis in machine tools where the elliptical form error is the most common in shaft and bearing bush. The purpose of this paper is to present the relationship between the elliptical form error and rotation accuracy for hydrostatic journal bearing in precision spindle and rotation table.

Design/methodology/approach

An error averaging effect model of hydrostatic journal bearing is established by using Reynolds equation, pressure boundary conditions, flux continuity equation of the land and kinetic equation of shaft in hydrostatic journal bearing. The effects of shaft and bearing bush on rotation accuracy were analyzed quantitatively.

Findings

The results reveal that the effect of shaft elliptical form error on rotation accuracy was six times larger than bearing bush. Therefore, to improve the rotation accuracy of hydrostatic journal bearing in spindle or rotation table, the machining error of shaft should be controlled carefully.

Originality/value

An error averaging model is proposed to evaluate the effect of an elliptical form error on rotation accuracy of hydrostatic journal bearings, which solves the Reynolds equation, the flux continuity equation and the kinetic equation. The determination of form error parameters of shaft and bearing bush can be yielded from finding results of this study for precision design of hydrostatic journal bearings.

Details

Industrial Lubrication and Tribology, vol. 69 no. 6
Type: Research Article
ISSN: 0036-8792

Keywords

Article
Publication date: 1 October 2002

Zbigniew Dżygadło and Witold Perkowski

The supercritical propulsion shaft equipped with a dry friction damper has been designed for a polish ultra light helicopter named IS‐2. Models of the shaft and the damper and…

Abstract

The supercritical propulsion shaft equipped with a dry friction damper has been designed for a polish ultra light helicopter named IS‐2. Models of the shaft and the damper and some results of analysis of the shaft flexural vibrations are presented.As it turned out the shaft vibrations strongly depend on parameters of the damper (especially on the damper gap) and can be regular or chaotic. There are two main cases: the damper with a small gap and the damper with a big gap, when compared to shaft eccentricity.

Details

Aircraft Engineering and Aerospace Technology, vol. 74 no. 5
Type: Research Article
ISSN: 0002-2667

Keywords

Article
Publication date: 7 August 2018

Bhumi Ankit Shah and Dipak P. Vakharia

The purpose of this study is to identify the crack in the shaft at incipient stage. Transverse crack is the most common type of crack found on the periphery of the shaft. The…

Abstract

Purpose

The purpose of this study is to identify the crack in the shaft at incipient stage. Transverse crack is the most common type of crack found on the periphery of the shaft. The changes in dynamic behaviour of the rotor at high speed are enormous. The reliable operation of the machinery is paramount for the safety of individual and plant. Condition-based maintenance monitors the mechanical and operational condition of the machine. During such inspection, if any unhealthy symptoms are detected, then affected part is identified and taken out for the maintenance at most appropriate time.

Design/methodology/approach

Simulating the transverse crack of different depth and location is the most challenging part of the experimental analysis. To optimize the total experimental cost for simulation of crack in the shaft, inverted crack is proposed to be produced in shaft and investigation shall be carried out for of early crack detection in shaft using vibration analysis. The set of experiments has been conducted on healthy shaft, inverted cracked shaft and actual cracked shaft. Inverted crack methodology provides flexibility of simulating crack of any size and at any location, and it can be reconfigured for several times to obtain various set of results.

Findings

To derive objective of the study, steady state response analysis and transient response analysis are performed on the experiment test rig. Vibration signals are acquired from the bearing locations to detect the crack. The paper addresses the influence of the inverted crack on critical speed of the shaft and deviation of first and second harmonic component of the shaft because of introduction of inverted crack. The resultant Nyquist plots, orbit plots and frequency plots are compared with the baseline data (obtained with the healthy shaft) to identify the crack.

Originality/value

The present study focuses on methodology by which inverted crack is developed in the healthy shaft, which resembles the behaviour of actual crack, and it shall be used to study the changes in rotor stiffness caused by transverse crack. The experimental results obtained using the inverted crack shaft have same vibration characteristics but in reverse direction as it would have occurred with the cracked shaft.

Details

Industrial Lubrication and Tribology, vol. 70 no. 7
Type: Research Article
ISSN: 0036-8792

Keywords

Article
Publication date: 1 November 1950

A multi‐row radial air‐cooled aircraft engine, figs. 1 and 2, having in the embodiment shown four radial rows and seven helical banks, has each of its banks supplied by a separate…

Abstract

A multi‐row radial air‐cooled aircraft engine, figs. 1 and 2, having in the embodiment shown four radial rows and seven helical banks, has each of its banks supplied by a separate magneto 116 through the ignition harness 117 and a separate induction manifold 224 from the supercharger casing 36. The engine is made up of six main parts, namely the propeller shaft housing 30, magneto drive housing 32, crank‐case 34, supercharger 36, accessory drive housing 38 and auxiliary drive housing 39. The propeller shaft 40, fig. 3, is supported in the housing 30 by a bearing and thrust ball race 48, 50 and bearings 62 and 64 and is connected to the crank‐shaft 58 through sun‐and‐planet gearing 52, 54, 56 and 60. The gear 54 having its teeth cut helically so as to produce a rear‐wardly acting thrust component, has in its rear face a plurality of equi‐spaced pistons 78 which bear against the fixed ring 80 and gear 54 is splined to the casing 30 at 72, 74. Oil from a pump is forced into the cylinders behind the pistons 78 to balance the thrust component, and as the engine torque is directly proportional to the oil pressure in these cylinders, engine torque may be read directly from a pressure gauge connected to said cylinders. The magneto housing 32, no. 9, has seven magneto mountings 115 and seven magneto driving shafts 118 which are geared to the crank‐shaft 58. The crank‐case, figs. 9 and 11, is composed of a number of parts 124, 126, 128, 130 and 132 which are held together by the bolts 134. The valve cam rings 150 which are geared direct to the crank‐shaft 58 by gears 164, 168, 170 to run at one‐sixth engine speed, each carry a three‐lobe inlet and exhaust cam track 154 and 156 respectively which operate the rocker arms 155 and 157 through the tappets 146 and pushrods 81 and 83. The crank‐shaft (see Group XXIV) is carried by five bearings 208, the centre one of which is flanged to locate the shaft axially and transmit thrust to the crank‐case. The two outer bearings are pressed and pinned in the holes 131 of the partitions 163. The centre three bearings are carried in disks 138 which are axially split and have keys 139 inserted between their faces which engage the partitions 162. The disks 138 are made of a material with a higher coefficient of thermal expansion than the crank‐case so that for easy assembly they are made with a sliding fit which tightens to a force fit when the engine warms up. The cylinder liner 214 has a cooling muff 216 and a cylinder head of aluminium alloy shrunk on it and the rocker boxes 218 and 220 are diametrically opposed so that the cylinders may be reversed when the engine parts are to be used as a pusher instead of tractor assembly. As shown in fig. 1, the induction manifolds 224, which are attached to the flanges 75, fig. 9, on the cylinder heads, are located between the rocker boxes and thus keep the diameter of the engine to a minimum. The supercharger 36, fig. 12, with a vaned diffuser 230 has an inducer 274 and impeller 228 driven from the crank‐shaft through a spring drive and hydraulic damper unit 206 (see Group XXIV), a gear 240 attached to said unit, and through cither two low‐speed fluid couplings 234 (one only of which is shown) and gears 244, 250 and 326 or two high‐speed fluid couplings 247 (one only of which is shown) and gears 242, 248 and 326. A fuel feed valve 276 in web or boss 278 discharges fuel, as metered and proportioned to airflow by the carburettor, into the spinner cup 280 which is carried by and rotates with the impeller 228. The fuel passed through nozzles 282 into the airstream and valve 276 shuts when the fuel pressure exceeds a predetermined value. The accessory drive housing 38 contains seven radial shafts and mountings for engine accessories such as oil pumps, tachometers, etc. All the shafts, one of which 346 is shown, are driven from the bevel gear 294 which is splined at 313 to the shaft 308 which runs through the hollow impeller shaft and is splined at 310 to the crank‐shaft 58. The cooling fan 364 which has turbine blades 365 between the ring shrouds 367 and 368 is driven by low‐ or high‐speed fluid couplings 372 or 375 and a series of gears 398, 400 and 382 which drive the gear 380 carried by the spider 378 splined to the shaft 308 at 313. The power take‐off shaft which may be used to drive an auxiliary supercharger, a second propeller or a second fan, etc., is keyed to the shaft 336 which is driven by the sun‐and‐planet gearing 406, 408 and 410.

Details

Aircraft Engineering and Aerospace Technology, vol. 22 no. 11
Type: Research Article
ISSN: 0002-2667

Article
Publication date: 1 December 1997

Terry Ford

Reports on the development of the Super CMV mainshaft for the Trent class jet engine from Rolls‐Royce. Describes in depth how the new mainshaft was designed to outperform the…

Abstract

Reports on the development of the Super CMV mainshaft for the Trent class jet engine from Rolls‐Royce. Describes in depth how the new mainshaft was designed to outperform the earlier generations. In comparison with the previous generation, the new mainshaft, through the use of improved materials, is able to cope with approximately twice the amount of torque transmitted to the fan, has a shaft weight per unit of transmitted torque ratio of 25 per cent less, a diameter that is similar in size, and yet was still able to be manufactured at a similar cost. Also reports on the increase in fatigue lives for the shaft oil holes and splines that was achieved through design improvements.

Details

Aircraft Engineering and Aerospace Technology, vol. 69 no. 6
Type: Research Article
ISSN: 0002-2667

Keywords

Article
Publication date: 6 October 2023

Fugang Zhai, Shengnan Li and Yangtao Xing

This paper aims to study the motion trajectory of the oil seal for shaft in eccentric state and derive equation of lip motion trajectory.

Abstract

Purpose

This paper aims to study the motion trajectory of the oil seal for shaft in eccentric state and derive equation of lip motion trajectory.

Design/methodology/approach

This paper analyzes the force during the motion of the eccentric lip by considering the material viscoelasticity, and a cam-plate mechanism is established as an equivalent model for the motion between the shaft and the lip; according to this, the equation of lip motion trajectory is derived.

Findings

The trajectory of the lip lags that of the shaft in the eccentric state because the viscoelasticity-affected lip recovery velocity is lower than the shaft recovery speed. The lip trajectory enters the lag phase earlier and the lag phase’s duration is longer with the increase of the eccentricity and rotational speed, because the deviation of the recovery velocities between the lip and the shaft will be exacerbated.

Originality/value

Innovatively, by considering the viscoelasticity of the material, the cam-plate mechanism is used to equivalent the motion of the shaft-lip to derive the equation for the radial motion trajectory of the eccentric lip. The regularity of lip motion is the key to determining the performance of oil seals, and the eccentric lip trajectory research method revealed in this paper provides a research basis for the performance research and optimization of eccentric oil seals.

Peer review

The peer review history for this article is available at: https://publons.com/publon/10.1108/ILT-06-2023-0161/

Details

Industrial Lubrication and Tribology, vol. 75 no. 9
Type: Research Article
ISSN: 0036-8792

Keywords

Article
Publication date: 30 June 2023

Abdul Kareem Abdul Jawwad and Mofid Mahdi

This article aims to investigate and model the effects of welding-generated thermal cycle on the resulting residual stress distribution and its role in the initiation and…

Abstract

Purpose

This article aims to investigate and model the effects of welding-generated thermal cycle on the resulting residual stress distribution and its role in the initiation and propagation of fatigue failure in thick shaft sections.

Design/methodology/approach

Experimental and numerical techniques were applied in the present study to explore the relationship(s) between welding residual-stress distribution and fatigue failure characteristics in a hydropower generator shaft. Experimental techniques included stereomicroscopy, optical and scanning electron microscopy (SEM), chemical analysis and mechanical testing. Finite element modelling (FEM) was used to model the shaft welding cycle in terms of thermal (temperature) history and the associated development of residual stresses within the weld joint.

Findings

Experimental analyses have confirmed the suitability of the used material for the intended application and confirmed the failure mode to be low cycle fatigue. The observed failure characteristics, however, did not match with the applied loading in terms of design stress levels, directionality and expected crack imitation site(s). FEM results have revealed the presence of a sharp stress peak in excess of 630 MPa (about 74% of material's yield strength) around weld start point and a non-uniform residual stress distribution in both the circumferential and through-thickness directions. The present results have shown very close matching between FEM results and observed failure characteristics.

Practical implications

The present article considers an actual industrial case of a hydropower generator shaft failure. Present results are valuable in providing insight information regarding such failures as well as some preventive design and fabrication measures for the hydropower and other power generation and transmission sector.

Originality/value

The presence of the aforementioned stress peak around welding start/end location and the non-uniform distribution of residual-stress field are in contrast to almost all published results based on some uniformity assumptions. The present FEM results were, however, the only stress distribution scenario capable of explaining the failure considered in the present research.

Details

Multidiscipline Modeling in Materials and Structures, vol. 19 no. 5
Type: Research Article
ISSN: 1573-6105

Keywords

1 – 10 of over 6000