Gear Noise and Vibration Episode 2 Part 6 ppt

240 Chapter 15 Root crack vibrations are submerged in similar ones generated by manufacturing errors and external disturbances. Very short shocks do not occur due to anything other than asperities or dirt and are of high frequency so they behave as shock fronts radiating out rather than as lumped mass vibrations. As mentioned previously, high frequency vibrations will not transmit satisfactorily through bearings, whether rolling or plain, because the pressure wave fronts reflect at the bearings instead of passing through. This means that it is not possible to detect these asperity (Smith) shocks using accelerometers mounted conventionally on the bearing housings. Experiments were carried out to check the link between Smith shock measurements on a rotor and on the bearing housing [3] and showed neligible coherence, with variations by a factor of 7 on the one giving only 20% variation on the other. This dictates that the detection system must be mounted on the pinion or wheel rotor to be effective. It is then remarkably sensitive. The system was used to monitor gearbox flank condition after lubrication failure in the form of removal of the oil system [4]. The instrumentation showed "failure" with local recorded acceleration levels exceeding 40 g about 125 minutes after oil removal. 40 g corresponds to saturation of the shock detection system so the levels would have exceeded this value. Fig. 15.4 shows some of the test traces obtained. The information for each of the 20 teeth is displayed separately for each four second batch of vibration and the levels have been staggered down 40 g for each tooth for clarity. The interesting observations were that there were indications of "failure" some 45 minutes before the final "failure" indication but more surprisingly that though the instrumentation showed high Smith shock levels the surface damage was barely visible and nowhere near the level which would have been noticed with a normal routine visual inspection. To see where the damage was it was necessary to refer to the traces to see which tooth was generating shocks. Needless to say the suggestion of mounting instrumentation on rotating shafts is not popular as slip rings or telemetry are required to transmit the information out. This factor will act as a major deterrent to the use of Smith shocks in a normal industrial setting as the instrumentation complications are only justified for critical applications. There is other information that can be deduced from the inspection of the Smith shock traces. Running in of gears involves asperity interactions which are (or should be) carefully controlled to give asperity removal rather than scuffing. There is in practice a fine dividing line between scuffing and running in of surfaces but both give shocks. The difference betwen them is that with scuffing the Smith shock intensities rise relatively rapidly with time whereas with running-in the shocks decay with time. They do not however Condition Monitoring 241 decay to zero but to a level dictated by the residual surface roughness after the running in process. Monitoring the shock levels during running in allows a direct check on the progress so that the next stage of running in can proceed as soon as there is stability. Most running in procedures are uncontrolled and rather inefficient with much time wasted on regimes which are doing nothing of use. Another slightly unwanted byproduct of monitoring Smith shocks is that, accidentally, they are the most sensitive and fast acting system ever encountered for detecting debris or dirt in the lubricating oil. Provided the debris is larger than about 3 um every single particle passing through the mesh will generate a shock which is easily detected. Initial attempts to view scuffing on a 80 mm centres test rig were subjected to high "noise" levels from dirt in the oil. Elimination of this background noise for easy viewing of scuffing involved using clean oil which was filtered down to about 2 \im or better. The major difference between debris shocks and scuffing shocks is that debris passage through the mesh only occurs once and so averaging over say 64 revs will virtually eliminate it whereas scuffing occurs in the same position each revolution for a small but finite number of revolutions. The final conclusion is the slightly surprising one that although pitting and root cracking are almost impossible to detect in a normal accuracy industrial gearbox it is relatively easy to detect scuffing or to control running- in under laboratory conditions. Whether the very high sensitivity of Smith shocks to dirt in the oil will justify their use for monitoring debris in critical installations remains to be seen. 15.6 Bearing signals Monitoring rolling bearing vibrations presents less practical problems than the corresponding gear vibrations. This is mainly because any vibration generated has a direct path to bearing housing accelerometers so there is no problem of lack of transmission of high frequencies. If we compare signals from pitting with those from a damaged ball bearing track then in both cases we have a contact running over a pit which is typically about 1 mm across. There is a big difference in frequencies as pitting may involve a pulse which is only about 0.1 millisec long and of the order of 1% of the mean load whereas a track pit may generate a pulse some 5 millisec long with an amplitude of the order of 10% of the mean load. In both cases there are characteristic frequencies or intervals between pulses involved which assist identification. Fig. 15.5 shows the typical trace obtained in one revolution but the next revolution, though it has the same time interval between pulses will have the pulses in a slightly different position as the cage speed is not synchronous with inner rotation speed. 242 Chapter 15 S3 1 1 1 1 1 1 1 1 1 1 time 1 revolution Fig 15.5 Expected trace from ball bearing with single inner track pit. Pulses shown as dashes are for the next revolution. In practice the main problem with rolling bearings arises if the damage is not detected in the initial stages when there is only a single pit. Once damage has spread over a significant arc of the track the vibration signal generated is roughly continuous and the characteristic pulses disappear into a general background noise. In ball bearings any ball surface damage gives a rather intermittent signal as the ball tracks over different parts of its surface. The standard techniques of frequency analysis and monitoring the amplitudes of the ball rotation and passing frequencies work well and will usually give clear warning of trouble. Roller bearings tend to present more problems as individual pits generate small pulses (as each pit carries a small fraction of the total load) and so generates a smaller pulse. One unusual case occurs with fluid coupling drives which may be fitted between electric motors and gearboxes to cushion startup as, although these are running at normal speeds of 1450 or 1750 rpm, the internal bearings are only running at slip speeds of the order of 20 rpm. The relative speeds are so low that track or ball damage does not generate significant vibrations so it is almost impossible to monitor these bearings. In many such cases the use of a fluid coupling is redundant so it is preferable to remove the coupling and to rely on the protection systems for the motor to protect the gearbox as well as preventing motor overheating. The motor systems need to have the normal thermal (slow acting) cutout but also to have a current overload cutout which comes into action after the motor is up to speed. Soft start controllers can achieve the same result. Condition Monitoring 243 References 1. Ray, A.G., 'Monitoring Rolling Contact Bearings under Adverse Conditions.' Conference on Vibrations in Rotating Machinery, I. Mech. E., Sept. 1980, p 187. 2. McFadden, P. D., 'Detection of gear faults by decomposition of matched differences of vibration signals. 1 2000 Mechanical Systems and Signal Processing. 14(5) pp 805-817 3. Smith, J.D., Transmission of Smith shocks through rolling bearings.' Journal of Sound and Vibration, Jan. 1995. 181 pp 1-6 4. Smith, J.D., 'Continuous monitoring of Smith shocks after lubrication failure.' Proc. Inst. Mech. Engrs., Vol 209C, 1995, pp 17-27. [...]... displacement at position 1 is due mainly to the force F acting on the wheel at 1 but is also due to the force F acting on the pinion at position 2 So 8, = Fpi, + Fp 12 and 52 = Fp 12 + Fp 22 As the transmission error is the sum of 81 and S2 T.E = F ( p n + 2 p , 2 + p 22) giving F Once F is known, the displacement 83 at the support foot can be found as 8 3 =F(p 1 3 etc ... "free" on soft test supports, the vibration is of amplitude V and we wish to predict the hull vibration 8 and the force P transmitted through the mount when the gearbox is installed main gearcase forces and deflections hull H Fig 16. 9 Idealised conditions at gearbox foot 25 8 Chapter 16 The first move is to decide what the internal forcing force is and we use superposition to find out how much force would... gearbox where we can measure vibration when running on a test rig We can measure the dynamic response at the foot to an external vibrator and we can measure the structural (hull) dynamic response For simplicity, assume that the test rig has extremely soft antivibration mounts so that the gearbox feet are not restrained on the test rig and the foot vibration is then the "free" vibration level Fig 16. 9... deform under the load which is normally well over 1000 N 0.03 100 20 0 300 Frequency in Hz 400 500 Fig 16. 5 Frequency distribution of input for 4 millisec half sine pulse Chapter 16 25 2 force transducer lightweight tip accelerometer light hollow handle Fig 16. 6 Alternative approaches to impact testing It is sometimes necessary to hit the gearcase using the hammer with a rubber block interposed to give... (vectorially) but it does not work the other way round 16. 5 Sweep, impulse, noise or chirp When vibration testing to get a transfer function or response of a system there are four basic choices: (a) the traditional very slow sweep with a vibrator, 25 6 Chapter 16 (b) noise (band limited) with a vibrator (it is not necessary to have pure white noise but all frequencies in the range should be present),... covering the range of response we can have any shape of pulse We then take a frequency analysis of the input and of the output and the ratio (complete with phase) is the frequency transfer function required 14 12 50 100 150 20 0 25 0 300 350 Fig 16. 4 Build up of harmonics to give a pulse 400 25 1 Vibration Testing Having too short an impulse is not helpful as too much of the input energy goes into high... the contact tip material and the geometry The idea that an impulse will give all harmonics at roughly equal amplitudes is rather strange to some students as they are accustomed to the idea that all frequencies and equal amplitudes in a vibration give random "white noise" The only difference between the two is that the phases of the components in white noise are completely random whereas for an impulse.. .25 0 Chapter 16 the system and open loop operation may be needed For this it is advisable to have a combined system where the average force is controlled (at very low frequency) by the slow servo loop while the vibrating force component is set by hand 16. 3 Hammer measurements Electromagnetic vibrators are clumsy, delicate and expensive and hydraulics is cumbersome so... foot F and hull H and, to complicate matters, an intermediate elastic isolator E All vibrations and forces are in the same (vertically downward) direction We can then call the various responses (or receptances) P with suffixes denoting where we excite and where we measure All the receptances are complex to allow for phase effects When the gear drive is running "free" on soft test supports, the vibration. .. the vibrator and though hydraulic vibrators are relatively small and compact for their power they do not work satisfactorily at 1 kHz As electromagnetic vibrators are large, heavy and delicate (as well as being expensive) they cannot be mounted inside a gearbox and sometimes are too large to fit near bearing housings but the resulting vibration near an accessible foot may be wanted Fig 16. 8 shows a . the pinion at position 2. So 8, = Fpi, + Fp 12 and 5 2 = Fp 12 + Fp 22 . As the transmission error is the sum of 81 and S 2 T.E. = F(p n +2p, 2 + p 22 ) giving F. Once F. in Fig. 16. 3 and then add the results and the estimates of the tooth effects as discussed in section 16. 6. 16 .2 Hydraulic vibrators As electromagnetic vibrators are large and very. V and we wish to predict the hull vibration 8 and the force P transmitted through the mount when the gearbox is installed. main gearcase forces and deflections hull H Fig 16. 9