Pinion gear is one of the indispensable spare parts of ball mill equipment, according to the direction of tooth shape can be divided into straight teeth and oblique teeth, there are two kinds of key type and expansion type. Different specifications of the ball mill selection of pinion teeth and modulus are not the same. Mainly according to the speed of the mill and the bearing capacity of the gear itself. Regardless of the type of ball mill, the quality of the pinion is of great importance, which directly affects the operation and service life of the equipment.
The gears are forged with ZG45#, 35SiMn, 45MnB, 40Cr and other materials according to customer requirement and drawing, and the hardness can reach HRC40-55 through tempering. Mechanical properties of pinion or shaft gear materials should not be lower than the relevant provisions, tooth surface hardness should be higher than the hardness of big gear tooth surface 30HB above, machining accuracy should meet GB/T 10095.
General situation 1.5, 183, 2.2 meters ball mill is the national standard configuration, gear ring(gear rim) specifications are unified, the rest of the ball mill for non-standard products, manufacturers if need to replace pinion accessories, must know the gear modulus, tooth number and external diameter, internal diameter, stop, displacement coefficient and other data. It is better to have the drawing provided by the mill manufacturer. If there is no drawing, please find a professional to accurately measure the above data and draw a sketch. Provide customized check data to qualified manufacturers.
Module is the ratio of pitch t to PI (m = t/ PI) between the teeth of the same side of two adjacent gears, measured in millimeters. Modulus is one of the most basic parameters in gear manufacturing. The larger the modulus is, the higher the tooth is, the thicker the tooth is. National standard ball mill pinion module is not less than 14, commonly used gear module is 18 module, 20 module, 22 module, 24 module, 25 module, 26 module, 28 module, 30 module. The number of teeth and modulus of gears are set according to the requirements of mechanical and mechanical principles.
There is a minimum number of teeth, is to prevent in the hobbing root cutting phenomenon. Pinion generally must not be less than 17 teeth, pinion number is usually designed for singular, large gear is generally haff-type split structure of the number of teeth for even, and paired with the number of large gear teeth best prime, at least can not be integral multiples of each other, so as to avoid serious local tooth wear phenomenon. So its important to determine the transmission ratio. Balance the relation between the number of teeth and modulus according to center distance and transmission ratio.
Generally divided into two types, one is involute cylindrical spur gear, the other is helical cylindrical gear. Matching gear along the tooth width at the same time into meshing, easy to produce vibration noise, transmission is not smooth. Helical cylindrical gear is better than cylindrical spur gear in grinding machine, and can be used for high speed and heavy load. However, because the manufacturing cost of helical cylindrical gears is much higher than that of cylindrical spur gears, most pinion gears of ball mill on the market are cylindrical spur gears. Pinion and pinion shaft are usually connected in the form of expansion and internal keys.
There are many factors that determine the price of pinion gear. If the pinion gear of the same specification is made of different materials, different manufacturers and different regions, the final price will be different. Mainly is the pinion material and blank weight, followed by the specific requirements and the tedious degree of gear processing.The pinion gears of ball mill manufactured by our factory are all processed by hobbing teeth or grinding teeth, the accuarcy of teeth can up to ISO6, and the tooth surface is quenched with high frequency.
We can manufacture pinion for cement ball mill pinion pinion slag ball mill, ball mill pinion of fly ash, high fine grinding machine pinion, pinion bearing ball mill, sand ball mill pinion, wet ball mill pinion, gypsum ball mill pinion, the overflow type ball mill pinion, ceramic ball mill pinion, dressing ball mill pinion, the grid mill pinion, desulfurization ball mill pinion, the wind sweep gear grinding machine, drying coal power plant coal ball mill grinding pinion, rod mill pinion and so on.
There are two kinds of transmission forms of ball mill, i.e. center transmission and edge transmission. The two kinds of transmission have their own advantages and disadvantages. The central transmission adopts a high-speed planetary reducer, which is safe and reliable, with a service life of up to 10 years, low maintenance cost, but large one-time investment.
The girth gear(big ring gear) and pinion are the key of the edge transmission device. Their reliable operation is directly related to the stable production of the mill. In order to ensure their reliable operation, in addition to strengthening the daily maintenance and regular maintenance, the installation and alignment of the girth gear and pinion are also important.
Cleaning of gear ring and mill flange positioning of gear ring coarse alignment of large gear ring positioning and coarse alignment of pinion installation of motor reducer primary grouting fine alignment of large gear ring fine alignment of pinion fine alignment of motor reducer.
In order to reduce the workload of installing gears on site, hobs with similar wear conditions should be selected during finishing of large and small gears, so as to make the meshing profile of large and small gears as close as possible and increase the contact area. Before the gear is installed, the tooth thickness vernier caliper can be used to check the tooth thickness of the gear ring indexing circle chord, or the tooth shape template can be made of sheet steel to retest the tooth shape. If necessary, it needs to be repaired on site.
After the ring gear is closed to the cylinder according to the installation mark, tighten the connecting bolts on the mating surface, and check whether there is a step on the side of the gear (the step is required to be as small as possible, otherwise it will affect the end runout of the ring gear). Use a feeler gauge of 0.04mm to check the matching condition. It is required that the insertion depth is not more than 50mm, and the cumulative length of the gap is not more than 1 / 4 of the tooth width.
When measuring the radial jump, a winch shall be used for disc grinding. Before disc grinding, the sliding shoes or hollow shaft oil station of the mill must be opened, and the floating capacity of the mill shall be checked with a dial indicator. Only when the floating capacity reaches 0.15mm can the disc grinding be carried out.
Most on-site dial gauges are installed in Figure 1 or figure 2, both of which have advantages and disadvantages. The position in Figure 1 can directly measure the radial jump at the top of the gear, but a worker is required to press and hold the dial indicator under the gear ring, and release the dial indicator when the tooth to be measured turns around, which is complex to operate and easy to cause injury to workers; the position in Figure 2 is simple to operate, but it is not the real radial jump of the gear. Therefore, we suggest to make a measuring tool (Figure 3) and install it on the head of the dial indicator, so that it is easier to measure the radial jump (Figure 4).
When measuring the end jump, a winch is still needed for disc grinding. Before disc grinding, the sliding shoes or hollow shaft oil station of the mill must be opened, and the floating capacity of the mill shall be checked with a dial indicator. Only when the floating capacity reaches 0.15mm, can the disc grinding be carried out. If the site personnel are sufficient, the end jump and the radial jump can be measured at the same time.
The end run out must be detected by double meter method to eliminate the error caused by the axial movement of the mill. When two dial indicators are installed on the same side of the ring gear, as shown in the left side of Figure 6, calculate the end jump according to line 4 of Table 1; when two dial indicators are installed on both sides of the ring gear, as shown in the right side of Figure 6, calculate the end jump according to line 5 of Table 1. The installation position of the two dial indicators must be 180 and 8 points need to be measured around the ring gear, as shown in Figure 5.
After the alignment of the big gear ring is completed, the pinion device shall be put in place according to the requirements of the drawings, the center height of the pinion shall be retested with a level gauge, and the sizing block shall be replaced if necessary to adjust the elevation. Use feeler gauge to roughly find tooth top clearance and tooth side clearance, and install diaphragm coupling, main reducer, auxiliary transmission device and main reducer lubricating oil station.
Grouting the anchor bolts of the pinion device once. After the primary grouting strength reaches 70% of the design strength, use the auxiliary drive plate to grind, and retest the runout of the big gear ring again, and then start the fine alignment of the pinion.
Most of the site is to use feeler gauge and red lead powder to detect the contact condition of the gear. There is no basis for the adjustment of this detection method, so the adjustment amount of the bearing seat can only be determined by experience, which is time-consuming and laborious. Compared with this method, the lead wire method can accurately calculate the amount of each adjustment, which is more accurate and saves installation time.
Before alignment, the lead wire shall be used to make a measuring tool as shown in Figure 7. In the figure, the left tool is used to measure the side gap and the right tool is used to measure the top gap. Set 12 sets of jackscrews at the 1-12 positions of the pinion bearing pedestal, and set 8 dial indicators at the 1-8 positions, as shown in Figure 8.
If there is no error in the processing of large and small gears, the graduation circle shall be tangent during installation, and its top clearance shall be 1 / 4 gear module. Considering the processing error caused by the hob wear and the floating amount of the mill when it is running relative to the static state and other factors, in order to avoid the gnawing back of the gear when it is running, it is recommended to increase the top clearance by 1.5-2.5mm during the installation, which can be adjusted according to the actual situation on site.
On the basis of meeting the above requirements, the larger the top clearance and side clearance is, the better, and the side clearance deviation on both sides of the gear is not more than 0.1mm, and the top clearance deviation on both sides of the gear is not more than 0.3mm. 8 points are measured in a week, and 5 points meet the requirements.
(3) Measuring points: 4, 7. If another point meets the requirements, the alignment is completed (i.e. at least 5 of the 8 points meet the requirements), otherwise, it is necessary to re measure after adjustment.
When adjusting the top clearance, refer to the reading of 5-8 position dial indicator and adjust 1-4 jacking screws; when adjusting the side clearance, refer to the reading of 1-8 position dial indicator and add or remove the adjusting gasket under the bearing pedestal.
After the lead wire method is used for alignment, the contact condition of the engagement surface of the large and small gears shall be retested with blue or red lead powder. The contact area shall not be less than 40% of the tooth height and 50% of the tooth width, and the contact on both sides of the gear shall be uniform, and only one side contact is not allowed, otherwise further adjustment is required.
In the daily production of theball mill, there will always be some sudden failures, such as severe vibration of the ball mill gear, cracking of the large gear ring of the ball mill, or the sudden increase in the temperature of the ball mill bearing. accident. Fodamon engineers shares the reasons and solutions of the severe vibration of the ball mill pinion.
The normal operation of the ball mill uses the motor to drive the ball mill to perform rotation operation. The motor drives the ball mill to directly connect the pinion through the elastic pin coupling and drive the big gear ring to realize the operation of the ball mill. In the process of operation, if the pinion and its base are found to vibrate violently, it is mainly because the distance between the pinion and the anchor bolt is far, and there is no fixed point in the middle, which makes the amplitude of the middle part of the pinion base larger, and the frequent vibration will cause serious wear of the pinion.
Solution: replace the pinion of the ball mill, add two anchor bolts on the inner side of the pinion and the meshing side of the big gear, and reinforce the base again. After the operation, the foundation shall be cleaned first, and the levelness, elevation and position and height of foundation bolts shall be checked.Drill the diameter hole (the size of the hole needs to be determined according to the size of the ball mill), install the anchor bolts, and grouting and pouring. After solidification, use the wedge iron to find the initial alignment. Then adjust the gap between the pinion gear and the large gear, correct the pinion gear by adjusting the height of the wedge, and then perform the second grouting, retain the motor-side half coupling, survey and map the pinion shaft, and redesign the small gear Gear half coupling.
Because the half coupling at the end of the pinion is in interference fit with the shaft, it needs to be hot installed. First, make a test bar that is larger than the online dimension of the shaft, fix the pinion shaft, and then heat the shaft hole of the half coupling. When the test bar can pass through the shaft hole of the half coupling completely, it can be assembled. Put the pinion shaft and bearing into the bearing housing, and adjust the tooth top clearance and tooth side clearance between the pinion and the big gear. Then the gap between the base and the ground shall be grouted twice, and the test run shall be carried out after solidification.It is suggested that if the vibration failure of the ball mill pinion is to be reduced, the ball mill can be installed in accordance with this method
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The typical IGC pinion shaft arrangement, having the impellers outboard of the seals and bearings, allows access for the application of VIGV mechanisms at the inlet of each stage (Fig. 4.5A), if desired, and similar accessibility exists for VDVs (Fig. 4.5B) with similar actuation mechanisms. In both cases, a single actuator rotates a control ring, which subsequently rotates each IGV or VDV in unison via identical cam features. Variable geometry can be used to increase the peak performance or operating range of a compressor, as is demonstrated in a later section. Because of the relative ease of access, VIGVs and VDVs are significantly more cost effective to implement on IGCs than with typical inline compressors. However, while this accessibility exists to varying degrees for every stage in an IGC, it is common to only implement VIGVs and/or VDVs for the first stage.
The output shaft is connected to the input pinion shaft of the gearbox by a high speed coupling shaft. This has flexible diaphragm elements and so allows some lateral misalignment between the GT and gearbox but, more importantly, allows axial tension (prestretch) to be incorporated when the unit is in the cold (ambient temperature) stationary condition. This axial prestretch is necessary to compensate for the thermal growth of the coupling shaft and the gearbox input pinion shaft when the unit reaches full speed operating temperature.
Figure 19.8 shows the coupling shaft arrangement. The axial thrust bearing, located on the gearbox input pinion shaft, carries the combined axial force (in a forward direction towards the GT) from the single helical gear train end thrust and that resulting from the cold prestretch of the high-speed coupling shaft. In the hot condition, at full operating power (steady state condition) an axial tension must be maintained in the drive train. This is because the output shaft bearing is designed to run with a tension loading if it is unloaded, or a compressive (aft) load is applied, the bearing will have a very short life.
During run-up of the GT the drive train only spends a very short time in the cold, full speed (7000 rpm) condition. The shafts quickly heat up to 120C. Because of centrifugal effects on the flexible diaphragm elements in the high-speed coupling shaft the tension force in the coupling varies with rotational speed. The relevant characteristics, based on type-test data for similar couplings, are shown in Fig. 19.8. The axial prestretch of the coupling can be adjusted by moving the GT forward or aft. The mounting system comprises a series of high-tensile linkages arranged as a three-dimensional spaceframe. During any axial adjustment of the GT position its lateral alignment, relative to the forward/aft axis of the vessel, must be maintained, otherwise the output bearing and gearbox input bearing may fail due to excessive lateral forces and vibration.
In passenger cars, trucks and tractors, a large proportion of closed-die hot forged steel components such as shafts, pinions, gears etc. are primarily used for strength, toughness, fatigue, along with dimensional accuracy. The paper provides an insight into the mechanism of grain growth in austenite during high temperature hot forging; stability of various precipitates in micro alloyed steels and its influence in restricting grain size; possible phase transformations during cooling; strengthening mechanisms and related aspects. Development of various forging grades over the years is analysed; it is shown that although the medium carbon, high manganese grades are predominantly used today in the hardened-and-tempered condition, several alternative grades are possible where the costly heat treatment and the accompanying risks of distortion / cracking could be avoided through air-cooling and yet obtain comparable or better properties. The requirement of through-thickness properties of large forgings and the role of acicular ferritic structures in very low carbon micro alloyed steel has been discussed. Demands such as machinabiity, weldability and the appropriate microstructures for the same have been looked into. Finally, it is shown that newer, non-steel materials (e.g. composites) as well as new versions of austempered ductile iron (ADI) are making their entry into the scenario which was so long considered a strong territory for steel forgings.
As in ball mills, the power draft of a rod mill is the product of capacity and work index, which is the energy required to break a mineral of a given size to the required size. The mill power is also increased by increasing the rod charge and the mill speed, while the mill power and capacity are both increased with increasing mill length.
For small ball mills, the power draw under dry batch grinding conditions was derived by Austin etal.  and the same considerations apply for rod mills. Equation (8.6) indicates that like any tubular mill the variation of mill power with speed in a rod mill is almost linear. This is true at the initial stages but breaks down when the critical speed is reached. At speeds in excess of the critical speed the power requirements decrease sharply. This is to be expected as rotation in excess of the critical speed results in the charge adhering to the inside liner and does not either cascade or cataract. Typical power requirements for two different mill loadings obtained in a laboratory size mill (0.6m 0.31m) with 20 lifters 25mm high are plotted in Figure8.8.
Figure8.8 shows the general characteristics of the change of mill power with mill speed for 17% and 40% mill loadings of a tumbling mill whose critical speed was 101rpm. It can be seen that at 40% loading the maximum mill power occurred at about 70% of the critical speed, while at a lower loading the maximum power drawn was nearly at the critical speed.
Equation (8.6) indicates that the power required is a function of the critical speed. Some manufacturers recommend an optimum speed of operation of their rod mills. For example, Marcy mills suggests that for their mills the peripheral speed should be governed by the relation
In industrial situations where conditions differ from Bonds set-up , Rowland and Kjos suggested in a series of papers [2,8] that Bonds Equation (3.25) can be used after correcting for different conditions encountered in industrial practice. Austin etal.  pointed out similar corrections required to Bonds equation to meet industrial conditions. These corrections are summarised below for specific conditions and are applicable to both rod and ball mills. More than one correction factor may be applicable. All factors are considered separately and the total correction is determined.
F2 is known as the inefficiency factor for the wet closed circuit grinding. It is a function of the sieve size used to determine the value of work index, Wi, and the percentage passing this control size. This function has been determined for different percentages passing the controlling sieve size and is shown in Figure8.9.
The value of Wi is best taken from an impact test or a rod mill grindability test, whichever is the greater. For a ball mill, the value of the constant in Equation (8.12) equals 4000 according to Rowland and Kjos .
For efficient operation of rod mills the feed should preferably be uniform in top size. The manner of feeding either from conveyors or directly from bins and chutes affects the power consumption and mill performance. To correct for feed preparation is difficult. The rule of thumb suggested by Rowland and Kjos  is summarised in Table8.2.
The above considerations for determining the mill power draw serve as a guide for the selection of mills for a particular job. Examples 8.1 and 8.2 illustrate the method of calculating mill power draw and also to compute the size of mill required for specific purpose.
A uniform discharge from a closed circuit jaw crusher is 200t/h. The crusher feeds a wet rod mill such that 80% of the crusher product passes a 16mm screen. The rod mill feeds a wet ball mill at a feed size of 1.0mm (1000 m) and produces a product with 80% passing a 150 m screen. The rod mill is in an open grinding circuit. Determine:1.the shaft power of the rod mill,2.the size of the industrial mill.
1.Correction factor F1 is not applicable.2.Correction factor F2 does not apply to rod mills.3.Correction factor F3 has to be considered after L and D are determined (usually towards the end of the computation). Hence, F3 will be determined later.4.Since the feed size is 16,000 m, correction factor F4 has to be determined.Use Equation (8.12) to determine FOPT, but first determine reduction ratio R.Reduction ratio (R)=16,000/1000=16.0Optimum feed size (FOPT ) = 16,000/(13 1.1/Wi)0.5=16,000 (13 1.1/13.5)0.5= 17,952 mSince the feed is less than the optimum, no correction is necessary.5.Correction factor F5 is not applicable.6.Correction factor F6 is applicable when R is between R*+2 and R*2.R* is estimated after mill size determination.7.Correction factor F7 is not applicable for rod mills.8.Correction factor F8 is not applicable as the circuit is a rod-ball mill circuit and the rod mill is fed from closed circuit crushing.
Preliminary selection of a commercially available rod mill may now be made from the manufacturers catalogue. For example, Allis Chalmers catalogue shows that the nearest mill size would require 655kW . Such a mill would have the following tentative dimensions:Mill length=4.88mMill diameter=3.51m (inside diameter=3.31m)Rod length=4.72Rod load=40%Rod charge=90.7t
The diameter efficiency factor, F3 in step 2 can now be determined using Equation (8.10) (Figure8.10). As the ID of the mill has been provisionally established as 3.31m, thenF3=2.443.310.2=0.941 forD<3.81m
Referring again to Allis Chalmers rod mill performance in Appendix B-4, for a mill power draw of 618kW, the following mill will be finally suitable:Mill length=4.86mMill diameter=3.35m (inside diameter=3.15m)Rod length=4.72mRod load=45%Rod mass=93.5t
To determine the commercially available ball mill that would suit the conditions, refer to the manufacturers literature, for example, Allis Chalmers as published by Rowland and Kjos . From the tabulated data, the mill charge and other characteristics corresponding to a power draw of 1273kW areBall mill length=4.57mBall mill diameter=4.57m (inside liner diameter=4.39m)Ball mill load=35%Ball charge mass=113tBall size=64mm
Referring again to Allis Chalmers ball mill performance table, for a mill power draw of 1412kW, the same mill dimensions will be suitable if the ball charge is increased to 45% with a charge mass of 144t.
Since between-bearing helical gears are the most common type on site, only this type will be covered. However, the relations discussed, with minor modifications, will also apply to internal and external spur gears. Figure 4.4.5 shows the reaction forces that act on a helical pinion tooth.
The axial load, WA, is calculated directly as shown in Figure 4.4.5, and can be applied to either the gear shaft, pinion shaft, or divided between the gear and pinion shaft. Most gear designs absorb all thrust on the gear shaft (low speed), since this usually results in the lowest thrust bearing losses.
Shaft centerline position is monitored by two proximity probes that are mounted at each radial bearing. They will record the position of the shaft and therefore the load angle of the shaft in the bearing. The load angle in gear applications changes with transmitted torque (power/speed).
Critical integral gear compressors have required emergency shutdown due to excessive radial bearing wear that did not cause high levels of vibration prior to shutdown (high vibration occurred when Babbitt material was excessively worn).
Heavily loaded radial bearings do not exhibit high vibration, and can go undetected if shaft position is not monitored, and bearing pad temperature probes are not located at the load point of the bearing.
This best practice has been used since the 1980s to optimize highly loaded gear radial bearing life, by predicting and recommending machine shutdowns at convenient times, thus eliminating costly emergency shutdowns.
Tumbling mills are most commonly rotated by a pinion meshing with a girth ring bolted to one end of the machine (Figure 7.10). The pinion shaft is either coupled directly or via a clutch to the output shaft of a slow-speed synchronous motor, or to the output shaft of a motor-driven helical or double helical gear reducer. In some mills, electrical thyristors and/or DC motors are used to give variable speed control. Very large mills driven by girth gears require two motors, each driving separate pinions, with a complex load sharing system balancing the torque generated by the two motors. (See also Knecht, 1990.)
The larger the mill, the greater are the stresses between the shells and heads and the trunnions and heads. In the early 1970s, maintenance problems related to the application of gear and pinion and large speed reducer drives on dry grinding cement mills of long length drove operators to seek an alternative drive design. As a result, a number of gearless drive (ring motor) cement mills were installed and the technology became relatively common in the European cement industry.
The gearless drive design features motor rotor elements bolted to a mill shell, a stationary stator assembly surrounding the rotor elements, and electronics converting the incoming current from 50/60Hz to about 1Hz. The mill shell actually becomes the rotating element of a large low speed synchronous motor. Mill speed is varied by changing the frequency of the current to the motor, allowing adjustments to the mill power draw as ore grindability changes.
The gearless drive design was not applied to the mills in the mineral industry until 1981 when the then-worlds largest ball mill, 6.5m diameter and 9.65m long driven by a 8.1MW motor, was installed at Sydvaranger in Norway (Meintrup and Kleiner, 1982). A gearless drive SAG mill, 12m diameter and 6.1m length (belly inside liners) with a motor power of more than 20MW, went into operation at Newcrest Minings Cadia Hill gold and copper mine in Australia, with a throughput of over 2,000t h1 (Dunne et al., 2001). Motor designs capable of 35MW have been reported (van de Vijfeijken, 2010).
The major advantages of the gearless drive include: variable speed capacity, removal of limits of design power, high drive efficiency, low maintenance requirements, and less floor space for installation.
Packaged Centrifugal Compressors are integrally geared, skid-mounted centrifugal units consisting of a driver (motor or turbine), a bull gear, and up to four pinions. Each pinion shaft terminates in an impeller. These machines are often described as packaged air machines, since the most common service is general plant compressed air supply. Generally, the driver and bull gear speed is 3,600 rpm or less, and the pinion speeds can be as high as 60,000 rpm. These machines are produced as a package with the entire machine mounted on a common foundation which also includes a panel with control and monitoring instrumentation. Because of the large number of these machines manufactured, proper monitoring locations for proximity probes have been established by the various machine manufacturers. Nearly all of these machines are supplied by the OEM with one proximity probe per impeller (one or two per pinion) and sometimes one probe on the bull gear shaft. In some instances, manufacturers have been able to respond to user specifications by supplying X-Y proximity probes and Keyphasor probes on the various shafts.
Process Centrifugal Compressors are usually larger than packaged units and have radial-flow or axial-flow stages usually mounted in the center span of the rotor between two fluid film bearings. These machines should be monitored with XY radial proximity probes at each journal bearing and at least one axial probe at the thrust bearing. If thrust position measurement is connected into automatic shutdown, then two axial probes should be installed in a voting logic configuration to reduce the possibility of false trips. If radial vibration channels are also wired into automatic shutdown, then extra false trip protection should be incorporated in the radial vibration monitors.
It is not necessary to mount a Keyphasor probe on the compressor if the coupled driver operates at the same speed and is equipped with a Keyphasor. However, if a gearbox is part of the system, then the compressor(s) on the opposite side of the gearbox from the driver should have a Keyphasor. Temperatures of all bearings, oil supply, and ambient conditions should also be monitored. On axial compressors, an accelerometer can be a useful auxiliary measurement for determining blade disturbances.
Reciprocating Compressors of the horizontal type can usually be monitored by using X-Y proximity probes to observe the piston rod or plunger of the compressor in addition to X-Y probes at accessible main crankshaft bearings. Monitoring the average position of the plunger can determine rod misalignment, packing wear, and wear on the sliding element and cylinder liner. Monitoring the dynamic motion signal of the probe can determine rod vibration or flexure (deflection). This is most necessary on hypercompressors. A Keyphasor could be installed on one of the drive rods to provide speed and timing information. An alternative location for the Keyphasor may be found on the crankshaft or motor driver. The Keyphasor should be installed so that the voltage pulse occurs when one cylinder is at top dead center to observe the relationship of all plunger positions as a function of stroke.
Screw Compressors have two rotors with interlocking lobes and act as positive displacement compressors. If the machine is equipped with fluid film bearings, the optimum installation should include X-Y proximity probes at each radial bearing of each rotor as well as at least one axial probe for thrust position measurement of each rotor. Thrust position monitoring of each rotor is very important because a thrust bearing failure means an axial rub will occur between the rotating elements. If rolling element bearings are used, monitoring with casing velocity transducers and Dual Path Monitors is acceptable; however, due to the very close axial clearances between rotors, axial shaft position monitoring is still of importance. The case-mounted transducer should be a velocity pickup with a Dual Path Monitor measuring both velocity and displacement vibration. Temperatures of machine components should also be monitored. Two Keyphasor probes, one for each rotor, would be desirable.
The location of the gear-box thrust bearing will determine whether thrust transmission will occur for a given machinery train arrangement. For the train configuration shown in Figure 3-77, the thrust bearing is generally part of the high-speed pinion shaft assembly. Using a coupling friction coefficient of 0.3, the pinion thrust bearing could be exposed to a maximum axial thrust of
However, the maximum possible axial force imposed on the gear mesh on Figure 3-77 is not related to F1, but rather, as will be seen later, to the tangential driving load FT and the magnetic centering force MA of the motor.
If the thrust bearing is mounted on the low-speed gear shaft, as shown in Figure 3-78, it could be exposed to the same maximum thrust as the bearing shown in Figure 3-77. However, the thrust force is now transmitted across the gear mesh. The maximum possible axial force imposed on the mesh is now not only a function of FT but of F1 as well.
The analysis can, of course, be extended to drivers other than motors. Unlike motors, which are generally furnished with axially free-floating rotors, these other drivers will probably incorporate thrust bearings.
There are several designs available, operated by a single acting cylinder with spring return, a double acting cylinder or twin cylinders. The basic principal is the same in each: the piston rod of a pneumatic cylinder becomes a rack which rotates the pinion shaft. In a twin cylinder arrangement it is possible to have a three- or four-position actuator according to which ports are pressurised, as shown in Figure 10. The angular rotation is limited in principal only by the stroke length of the cylinder, but in practice, standard units rarely exceed about 360. Up to 400 Nm is possible with this type of design.
The turbine shaft serves the function of a pinion shaft of the gearbox, thereby increasing the compactness and sturdiness of the design. This is especially suited as a cost-efficient solution in ORC-based power plant applications where a single turbine alone cannot transfer the entire power provided by the process. The same applies to processes with multiple pressure levels.
The conclusion of this is that the inward radial flow turbine design form Atlas Copco is extremely reliable and requires little maintenance, even if this is a high performance design with advanced features.
This high reliability and low maintenance has been verified in hundreds of turbines which have been made by Atlas Copco and delivered not only to ORC plants, but also to applications with similar (and not quite so similar) fluids such as natural gas condensate removal, liquid natural and petroleum gas, air separation, and so on.