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理解轴瓦轴承(英文版)

理解轴瓦轴承(英文版)
理解轴瓦轴承(英文版)

Understanding Journal Bearings

Malcolm E. Leader, P.E.

Applied Machinery Dynamics Co.

Durango, Colorado

ABSTRACT

This paper covers the basic aspects of journal bearings including lubrication, design and application. Descriptions of various types of journal bearings are presented. Guidance is given for choosing the proper bearing type and keeping your bearings healthy. A section on do’s and don’ts gives practical information.

INTRODUCTION

Bearings are used to prevent friction between parts during relative movement. In machinery they fall into two primary categories: anti-friction or rolling element bearings and hydrodynamic journal bearings. The primary function of a bearing is to carry load between a rotor and the case with as little wear as possible. This bearing function exists in almost every occurrence of daily life from the watch on your wrist to the automobile you drive to the disk drive in your computer. In industry, the use of journal bearings is specialized for rotating machinery both low and high speed. This paper will present an introduction to journal bearings and lubrication. Lubrication technology goes hand-in-hand with understanding journal bearings and is integral to bearing design and application.

Since they have significant damping fluid film journal bearings have a strong impact on the vibration characteristics of machinery. The types of machinery we are concerned with range from small high speed spindles to motors, blowers, compressors, fans, and pumps to large turbines and generators to some paper mill rolls and other large slow speed rotors. Not covered here is the topic of bearings for reciprocating machinery. While some of the same principals apply, engine bearings have special needs and design considerations and deserve a more complete study. Reciprocating machinery bearings tend to be simpler in geometry and much more complicated in application than turbomachinery bearings. For example, the typical turbomachinery journal bearing consists of a thin layer of babbitt on steel while a connecting rod bearing may have numerous different layers of copper, steel, nickel, or other metals with a thin layer of babbitt on top. This layering is done for fatigue resistance to the pounding loads encountered in such machinery. Engine bearings are often required to withstand peak specific loads in excess of 3,000 PSI or about ten times a typical motor or turbine bearing. Reciprocating machines rely primarily on the squeezing of the oil film for load support.

WHEN TO USE FLUID FILM BEARINGS

There are applications where anti-friction bearings are the best choice. Commonly, smaller motors, pumps and blowers use rolling element bearings. Paper mill rolls often use large specialized spherical roller bearings. Clearly, anti-friction bearings are best for these applications. However, once the size of a pump (or fan or motor, etc.) gets large enough and fast enough, a gray area is entered. Here you will still find rolling element bearings used successfully but as speeds increase and temperatures rise, rotor dynamics often become a concern and critical speeds are encountered. This is when damping is required and fluid film bearings become increasingly necessary. My experience is that turbomachinery designers (and users) should consider using fluid film bearings if running above 3,000 RPM or the machine exceeds 500 HP. In my opinion, at 1,000 HP and up, all machines except very special cases should be on journal bearings specifically designed for that service. There are exceptions of course, and the decision where to apply what type of bearing is ultimately done for every machine individually based on good engineering practice and experience. Unfortunately, this decision is sometimes based on economics which keeps maintenance engineers and consultants employed. ADVANTAGES OF FLUID FILM BEARINGS

The primary advantage of a fluid film bearing is often thought of as the lack of contact between rotating parts and thus, infinite life. In a pure sense, this is true, but other complications make this a secondary reason for using these bearings. During startup there is momentary metal-to-metal contact and foreign material in the lubricant or excessive vibration can limit the life of a fluid film bearing. For these reasons, special care must be taken when selecting and implementing a lubrication system and special vibration monitoring techniques must be applied. The most important aspects of the health and longevity of a fluid film bearing are proper selection, proper installation, proper lubrication, and the alternating hydrodynamic loads imposed on the bearing surface by relative shaft-to-bearing vibration.

Some of the primary advantages of fluid film bearings are:

! Provide damping. Damping is required in order to pass through a critical speed. Damping is also required to suppress instabilities and subsynchronous vibration.

! Able to withstand shock loads and other abuse.

! Reduce noise.

! Reduce transmitted vibration.

! Provide electrical isolation of rotor to ground.

! Very long life under normal load conditions.

! Wide variety of bearing types for specific applications

The lubricant used provides these functions to all bearings:

! Remove heat generated in the bearing.

! Flush debris from load area.

Some disadvantages to fluid film bearings are:

! Higher friction (HP loss) than rolling element type.

! Susceptible to particulate contamination.

! Cannot run for any length of time if starved for lubricant such as a lube system failure.

! Radial positioning of rotor less precise.

Use of journal bearings is also an advantage in many applications when it comes to maintenance. Most fluid film bearings are split and rotor removal is not required to inspect and replace. While split rolling element bearings are also available they are costly and not common. Journal bearing fatigue damage is usually visible at an early stage and allows for better diagnostics of failure modes so that corrective action can be taken to prevent recurrence.

PLAIN BEARINGS AND BASIC CONCEPTS

In order to illustrate the basic nomenclature, geometry, and introduce the ideas of how fluid film bearings work, the simplest bearing called a plain journal bearing will be examined. Figure 1 is a photograph of a plain bearing. A steel base material is overlaid with a babbitt material and bored to a circular diameter equal to the shaft diameter plus the desired clearance. Scallops are cut at the splitline to admit oil. Figure 2 is a computer model of this same bearing.

Figure 1 - Typical Plain Journal Bearing Figure 2 - Computer Model of Plain Journal Bearing At zero speed, the shaft rests on the bearing at bottom dead center. As soon as shaft rotation begins the shaft “lifts off” on a layer of oil. In fluid film bearings, lubrication is required between a pair of surfaces with relative motion between them. There is always a convergent wedge developed that is formed due to the relative surface speeds and the lubricant viscosity to carry the applied load. An oil pressure film develops with equal and opposite force vectors to the applied load. One surface drags the lubricant, usually an oil, into a converging gap. As the space available in this gap decreases, the fluid develops a pressure gradient, or pressure hill. As the fluid leaves the gap, the high pressure helps expel it out the other side. A simple diagram of this is shown in figure 3.

Figure 3 - Basic Development of an Oil Wedge

Lubricants can be any fluid, including gasses. In early reference books (1,2,3) some of the lubricants discussed are tallow, lard (animal fat), vegetable oils, and whale and fish oils. Obviously, sometimes you used what was available! Even water can be used under some conditions. Mineral oils from petroleum have evolved from straight distillates to complex formulations with special additives today. Synthetic lubricants have also been developed, primarily poylalphaolefins and esters. Silicones, glycols and other fluids are also used in special applications. There is no ideal or universal lubricant, all are compromises to fit any given situation. Applications range from heavy low speed loads to light high speed loads. At one extreme solid lubricants may be necessary and at the other, gas bearings may be required. Obviously, most applications fall in the middle where grease and oil lubricants are used. In this discussion of journal bearings we will limit ourselves to light oil lubrication found in the majority of turbomachinery. Bearing Nomenclature

The shorthand that bearing analysts use with regards to journal bearings can be confusing and is certainly inconsistent from one analysis program to another. The terminology used in this paper is shown diagrammatically in figure 4. The symbolic notation and the definitions are as follows:

j R = Radius of Journal

b R = Radius of the Bearing

b b j C = Radial Clearance of the Bearing = R -R h = Radial clearance as a function of the angular position where the clearance is measured min h = Minimum oil film clearance

e = Eccentricity - the distance between the center o

f the bearin

g and the center of the shaft b ecb = e/C = Eccentricity Ratio - if zero, shaft is centered; if 1 then shaft touches bearing Line of Centers = Line connecting the center of the bearing and the center of the shaft ? = Attitude Angle = Angle from -Y axis to Line of Centers

ù or ù = Rotation Direction and Speed in RAD/SEC W = Gravity Load

Figure 4 - Bearing Nomenclature Figure 5 - Pressure Profile in a Journal Bearing

In our plain journal bearing example the load is supported by a high pressure oil region as shown in figure 5. Each line in the pressure profile represents an oil pressure vector at the centerline of the bearing. The sum of the vertical components add up to the applied load and the horizontal components cancel out for equilibrium. Oil inlet ports are placed in areas of minimum pressure. The pressure profile can also be examined in a three-dimensional format as shown

in figures 6 (low load) and 7 (for a highly loaded bearing).

Figure 6 - Hydrodynamic Pressure Profile - Low Load Figure 7 - Hydrodynamic Pressure Profile - High Load These figures show the bearing “unwrapped” with the top half on the left and the loaded bottom section on the right. It is important to note several things about the hydrodynamic pressure profiles. First, the peak pressure is significantly higher than the specific load (W/LD). Secondly, the pressure at the margins always returns to the boundary condition which is usually atmospheric pressure. The unloaded top half, even though it is cavitated is essentially at atmospheric

pressure which is why no cavitation damage ever occurs in this type of bearing.

Bearing Performance

While the stiffness and damping provided by a journal bearing are crucial, there are other design factors that must be considered in order to understand how bearings work. For example, if the eccentricity is too high there is a risk of metal-to-metal contact and higher dynamic loads being imparted to the babbitt causing premature fatigue. If the eccentricity is too low (journal is nearly centered) then the machine could more easily become unstable. Eccentricity is a function of both speed and load. Figure 8 indicates that, with a constant load, as speed increases, the eccentricity decreases.

Figure 8 - Plain Bearing Eccentricity versus Speed with a Constant Load

The attitude angle between the vertical axis in the load direction and the line of centers also changes with speed. A plot of this angle with eccentricity describes a plot of the locus of the shaft centerline as speed changes as seen in figure 9.

Figure 9 - Shaft Centerline Position in a Plain Bearing as a Function of Speed

If you combine the effect of speed and load on the bearing eccentricity, figure 10 tells the complete story. At low speeds b the eccentricity ratio e/C is high. At low loads, the eccentricity is high.

Figure 10 - Effect of Load and Speed on Plain Journal Eccentricity

There are additional parameters that must also be examined when analyzing fluid film bearings. One of the most important is the maximum temperature that will be generated in the fluid film. This is consistent with the horsepower losses for a given bearing and will be somewhat higher than the actual temperature measured with an RTD or thermocouple in the bearing shell. Figure 11 is a plot of the inlet oil temperature (kept at a constant 120E F in this case)and the predicted measured temperature and the predicted maximum oil film temperature for a constant load as speed is varied. While it is possible to operate journal bearings above 200E F, typically a bearing designer will seek to keep the maximum oil film temperature below that due to loss of babbitt fatigue strength. If possible, a good bearing design will have the maximum oil film temperature less than 175E F to allow for some margin for transient events. It is also min important to analyze the minimum oil film thickness as seen in figure 12. Here the ratio h /Cb is plotted as a function of speed with a constant load. This ratio is the minimum film thickness as a percentage of bearing clearance. If we knew the radial clearance was 5 mils then we can calculate the minimum film thickness at any speed. For 5 mils radial min clearance in this case the h would be about 1 mil at 100 RPM and more than 4 mils above 5,000 RPM.

Figure 11 - Temperatures in a Plain Bearing versus Speed at a Constant Load

Figure 12 - Minimum Oil Film Thickness in a Plain Bearing versus Speed at a Constant Load

LUBRICATION

For fluid film bearings viscosity is the most important factor. Unfortunately, there are two forms of viscosity terminology and numerous units associated with these measures. Absolute or dynamic viscosity is the ratio of the shear stress to the resultant shear rate as a fluid flows. The more a fluid resists shear, the thicker it is and the higher the absolute viscosity. This is measured in Poise (or Centipoise) or Reyns. Some reference books use different terminologies, so read carefully. Kinematic viscosity is the absolute viscosity divided by the specific gravity. The most common unit of measure is the Centistoke, abbreviated cSt. The following tables help define these two measures of viscosity. Figure 13 relates SUS to these measures.

Table 1 is a multiplying factor chart for converting between the various viscosity units. Note that since the specific gravity (?) is usually between 0.8 and 1.2, the absolute and kinematic viscosities are nearly the same for all practical purposes. Table 2 is a list of typical applications and the viscosity ranges found in these applications. This list is by no means complete nor meant to preclude other oils in similar services. Viscosity varies significantly with temperature and the variation is highly non-linear. Often blending of oil bases is done to reduce these effects.

Table 1 - Viscosity Conversion Factors

Table 2 - Typical Viscosity Ranges for Various Applications

Figure 13 shows how temperature affects three common ISO turbomachinery oils. At lower temperatures the differences are much more significant than at a typical operating temperature.

Figure 13 - Viscosity Variation with Temperature and Conversion Factors

The friction factor (and thus the horsepower loss) is a function of viscosity, load, and speed. This was quantified by Richard Stribeck, a German engineer who did extensive friction testing in 1902. There was a great deal of research at that time trying to find the best combinations of materials and lubricants that would give the lowest coefficient of friction. Figure 14 is the original Stribeck curve from Reference 2. The friction factor is plotted as a function of ZN/P where Z is the viscosity, N is the speed and P is the load. This is a non-dimensional equation.

Figure 14 - Stribeck Curve Relating Friction Factor to Viscosity, Speed, and Load

To define what viscosity is mathematically, we must go back to the early history of lubrication theory. As a shaft turns in a journal bearing the moving service shears the oil film between it and the stationary bearing surface. This shearing action creates a stress field in the lubricant film in PSI. The relative velocity between the surfaces defines the shear rate. The dynamic viscosity is defined as the ratio between the shear stress and the shear rate. Higher viscosities increase the shear stress while lower viscosities increase the shear rate due to a smaller minimum film thickness. Figure 15 shows this diagrammatically with a full unit derivation in English and SI units.

Or in SI units:

Figure 15 - Defining Dynamic Viscosity - English and Metric Units

The Reyn is named after Osbourne Reynolds famous for the “Reynolds number” that defines laminar and turbulent flows. However, a Reyn is much too large a unit and typical lubricants are in the micro Reyns region. Likewise, Poise is named after Jean Louis Marie Poiseuille a French physician who worked on fluid flow in the early 1800's. This unit is commonly used as centiPoise or cP which is 0.1 Pascal-second. The Stoke, after George Stokes, is a Poise divided by the fluid density. Usually used as centiStokes, or cSt. Water at room temperature has a viscosity of about 1 cSt.

Usually we want to minimize friction so we want to be somewhere in the center region of the Stribeck curve. Since speed and load are often fixed for a given machine, the bearing designer is left with choosing an appropriate lubricant viscosity. Other factors that affect this decision include the supply temperature and pressure, heat generation and removal, corrosion resistance and material compatibility.

Oil formulation is a complicated science and includes many additives to the base oils. These additives include high temperature breakdown resistance, extreme pressure, anti-foam, anti-sludge, corrosion resistance, anti-oxidation, and others. Often the formulations are proprietary depending on the supplier. Sometimes you have to tell your supplier what conditions and materials the oil will contact and ask them if their oil will react or degrade in such an environment. In some systems with high surface speeds, additive components have been known to centrifuge out of the oil and collect in areas like the inside of gear-type couplings. Here these materials build up and can restrict motion leading to increased vibration and wear.

JOURNAL BEARING MATERIALS

Since the bearing is usually much less costly than the shaft, it is considered sacrificial if necessary. In addition, a low dry sliding friction coefficient between the shaft and bearing material is important for startup and shutdown conditions. Virtually every conceivable engineering material has been tried as a bearing material. Early man used wood or even

th

stone. Later iron, copper and leather were applied to relatively low speed shafts. In the 17 century Robert Hooke advocated steel shafts with “bell metal”, essentially bronze, bushes to replace the practice of using wood blocks on cast iron. There has been extensive research on the best bearing material to use.

In 1839 Isaac Babbitt patented several high tin and high lead alloys which are similar to modern formulations. Of course, his name is commonly used to describe many different bearing materials. Carbon, graphite, ceramics, plastics and composite materials are all used today in some applications. Lead alloys have been virtually eliminated in modern applications due to lower strength and environmental considerations. The most common bearing lining material used in turbomachinery today is high tin babbitt. The formulation and characteristics of the two most common babbitt alloys are given in table 3. There are other alloys with specialized properties and often proprietary composition.

These formulations are applied to a base material, typically steel, by either molding in a stationary fixture or spin casting which is a messy and exciting thing to experience. Spin casting is supposed to create a better bond with the substrate but it has been determined that impurities in the molten babbitt can migrate to the babbitt-base metal bond line during this process and weaken that bond. While babbitt won’t melt until the temperature is very high, it has lost almost half of its room temperature tensile strength at 212E F.

BEARING FAILURE MECHANISMS

There are many reasons a journal bearing might fail. A common mechanism is loss of lubricant and is not so much a bearing failure as a system failure. The next most common mechanisms is fatigue damage and one of the most important considerations for the material lining bearings is fatigue resistance. Thin babbitt has significantly more fatigue resistance than babbitt thickness greater than 15 mils as seen in figure 16. This is data from an extreme test conducted with about ten times the normal static load to speed up the test results.

Figure 16 - Bearing Fatigue Life as Babbitt Thickness Varies

Rotor imbalance and other conditions cause the journal to orbit in the bearing. This causes an oscillatory dynamic pressure to act on the babbitt surface. It is not uncommon for the peak static hydrodynamic pressure to approach 3 to 5 times the specific W/LD load. Oscillatory shaft motion adds an alternating pressure on top of this that impinges on the babbitt surface. Babbitt fatigues in a manner similar to the way potholes develop in a road surface. Surface cracks develop which propagate to the bond line below as shown in figure 17. The piece of babbitt detaches but is unable to leave the area because of the close clearance to the shaft. This piece of babbitt rattles around the pit, breaking into smaller bits which are carried away by the oil film. This action smooths the sides of the pit. Finally all the pieces of loose babbitt are gone leaving a rounded smooth hole. Figures 18 and 19 are micrographs of this process and figures 20 and 21 are photographs of a failing bearing. This type of damage is often mistaken as cavitation or erosion, which it is not. Virtually all journal bearings cavitate. Typically oil contains dissolved gasses and in the divergent unloaded area of the oil film it is normal for streamers to form. As the pressure wedge re-forms, the gasses dissolve back into the oil. This is a gradual process and no damage to the bearing occurs. There are cases where rapid rupture of the oil film is normal such as with high speed reciprocating machinery. Here vapor bubbles can form and collapse explosively. These impacts may locally fatigue the bearing surface. Entrained free air will aggravate this problem. The most common failure mechanisms are a result of operating conditions outside of the intended design. These include foreign material (dirt) or contamination (water, etc.) in the oil, overload, lube oil viscosity degradation, or, far too often, lubrication failure. Electrostatic and electromagnetic discharge across the oil film will erode the babbitt surface over time. Corrosion of the bearing babbitt is relatively rare especially if a regular oil analysis program is in effect. The steel backing can be attacked by water in the oil and copper alloys are more susceptible to corrosion from ammonia and other contaminates.

Figure 17 - Babbitt Failure due to Fatigue Pounding Damage

Figure 18 and 19 - Fatigue Cracks in Babbitt

Figures 20 and 21 - Early and Late Stage Babbitt Fatigue

JOURNAL BEARING DESIGN PARAMETERS

The design of a journal bearing for a particular application involves many considerations.

! Available Space

Many machinery manufacturers purchase bearings from specific bearing companies. It is advantageous to use standard bearing form factors for cost reduction. A machine that was designed for thin shell bearing liners is not usually a good candidate for conversion to tilting pad bearings simply because there is insufficient radial room in the bearing housing. Sometimes new housings can be made, but this involves additional cost.

! Friction and Heat Generation

The basic equation that bearing designers use to evaluate friction and heat generation is Petroff’s equation (9) which assumes a centered shaft in a plain bushing. Notice that journal load is not part of this calculation. Only the oil shear forces are used. The viscous shear force in this case is:

from this equation, an estimate of the temperature rise for a given oil flow rate is made using either empirical equations available in some of the references (10) or various computer programs. It is very important to understand the assumptions used in the empirical equations, particularly the heat capacity of the lubricant. This may vary significantly if using synthetic lubricants. As radial load on a bearing increases, the Petroff equation is not sufficient to predict the temperature rise and relatively complex computer programs are used. Getting sufficient oil into the bearing and subsequently draining it out are very detailed topics not covered here.

! Specific Steady Load

A very important concept is the specific load, P, is defined as:

or simply the journal load divided by the active bearing length times diameter. In English unit this is in PSI. The loading classification ranges (by this author) for this variable are:

0 to 50 PSI - Very Light

50 to 100 PSI - Light

100 to 200 PSI - Moderate

200 to 300 PSI - Heavy

more than 300 PSI - Special Design

There are problems associated with bearings that are too lightly loaded, specifically oil whirl and instability. For heavily loaded bearings heat generation and babbitt fatigue can become significant. One should always consider both the

minimum film thickness and the peak hydrodynamic load when evaluating a design. Special evaluation is required if the peak pressure exceeds 1,000 PSI or the minimum film thickness is less than 1.0 mils.

! Surface speed

The larger the journal diameter, the higher the surface speed for a given RPM. To calculate feet-per-minute (FPM) from RPM the relationship is: (use inches for D)

some typical examples are:

600 RPM with 8 Inch Journal = 1,257 FPM

3,600 RPM with 4 Inch Journal = 3,770 FPM

12,000 RPM with 6 Inch Journal = 18,850 FPM

At some point, which depends mainly on the lubricant and clearance, the oil film becomes turbulent. Turbulence increases the friction, reduces oil flow and generates more heat. For typical turbomachinery, surface speeds above 15,000 FPM should receive special scrutiny.

! Dynamic Load

Dynamic load is the result of shaft orbital motion in the oil film clearance space due to imbalance, misalignment, and other non-static forces. Thus, an alternating hydrodynamic pressure fluctuation is superimposed on the steady pressure being exerted on the babbitt. It is this alternating force which fatigues the babbitt. The easiest way to give guidelines for this is to limit maximum orbit vibration to less than 50% of the diametral clearance of the bearing. For example, if the bearing clearance is 6 mils, then the maximum acceptable displacement amplitude is 3 mils peak-to-peak measured on the orbit from two orthogonal proximity probes. Once the orbit exceeds 75% of the clearance, short-term damage is almost always happening and bearing life is likely limited to a few hours if the orbit size exceeds 90% of the available clearance. This author defines a dynamic load of 300 PSI (on top of the static load) as the upper limit of acceptability.

Figure 22 - Dynamic Load - Vibration and Alternating Hydrodynamic Load

! L/D Ratio

The length to diameter ratio is one of the first things a bearing designer considers. Since the shaft diameter is often determined by other factors (torque and bending strength, for example) it is usually the active babbitt axial length that is controlled. This factor is “tuned” to give sufficient steady and alternating load capacity. However, it also significantly affects the stiffness and damping characteristics of the bearing. The longer the bearing, the lower the specific load and the lower the stiffness. A longer bearing also tends to have higher effective damping. Since damping is a frequency dependent stiffness, the speed of the machine must be considered as well. While, from a rotor dynamics standpoint, softer bearings are almost always a benefit, a too lightly loaded bearing may have stability problems. In the experience of this author, an L/D ratio less than 0.3 has poor damping and an L/D ratio greater than 0.75 usually shows little gain in effective damping. Heavily loaded bearings such as gears often exceed an L/D ratio of 1.0 and may need special care in aligning the bores to the shafting.

!Clearance

The basic guideline universally used for diametral journal bearing clearance is 1.5 mils-per-inch of journal diameter. That is, a 4 inch diameter shaft would need about 6 mils of diametral clearance. Always check if the specifications you see are for diametral clearance. Some manufacturers specify radial clearance - which is really difficult to measure! If the application has atypical loads and/or speeds, then this clearance rule may need to be adjusted. The bearing should be held in the housing with a zero to 1 mil interference fit. A loose bearing will have vibration problems and too much housing interference can reduce the clearance in the bearing. The issue of tolerances should also be considered. Clearance values are usually given as a range. The limits suggested here for minimum allowable clearance is 1.0 mil-per-inch shaft diameter plus one mil. The maximum allowable is 2.0 mils-per-inch. Remember tolerance costs money.

! Base Material

Normally steel is used as the backing material for babbitt bearings and is preferred for strength. Cuprous alloys are about half the strength of steel and have virtually no endurance limit. Bronze or even copper may be used where extra heat conduction is needed for a successful design. Caution should be used with new machinery. API 612 (Special Purpose Turbines) and API 617 (Centrifugal Compressors) both specifically disallow anything other than steel for a backing material. The reason for this is that it allows a relatively inexpensive and convenient path to fix a hot bearing or upgrade a design. If the original design includes copper pads, you are limited on the redesign.

!Grooving

On some bearings grooves are cut into the surface of the babbitt for a variety of reasons. The most often cited reasons for grooving are to direct lubricant to the loaded areas and for cooling. A well-designed bearing with steady loads (not reciprocating) does not need any grooves and, in fact, the sole effect is to increase stiffness, reduce load capacity, and damping. In some cases a circumferential groove is used (called a load groove) to increase the specific loading and increase stability.

JOURNAL BEARING DESIGN

The simplest form of journal bearing is a plain circular bushing with an ID that is slightly larger than the shaft OD. Early bearings were often made by pouring the babbitt or white metal directly into a sleeve containing the actual shaft or a same-size mandrel. Then, the surface of the babbitt was scraped to provide the desired clearance. This success of this procedure was highly dependent upon the skill of the mechanic. Scraping of bearings is strictly forbidden in most plants today and is not recommended. Many different bearing profile designs have been invented in the last century or so, but the basic geometries are primarily constrained by cost and ease of manufacturing. Special circumstances may require unique solutions.

Preload

Besides clearance the next most important factor of a bearing is preload. A preloaded bearing simply means that the center of curvature of the pad is not coincident with the geometric center of the bearing. Preload can be thought of as the percentage of clearance reduction. No preload means, with a centered shaft, even clearance all the way around and 50% preload would reduce the clearance by half at the closest point. Typical industrial bearings have preloads in the zero to 50% range. Preload is easily and cheaply added to a plain circular bearing. It is good to avoid getting too close to zero preload so that tolerances will not cause you to end up with negative preload. Negative preload pinches the inlet and outlet of the pads causing a scraper effect that will starve the bearing of lubrication.

Figure 23 - Pad Preload Definition

Offset

The other geometric parameter that is controlled is offset. When the closest clearance is half-way along the arc of a pad, this is the most common and is called 50% offset. Offset can be increased past 50% but offsets less than 50% are not normal and not recommended. Offset is easily manufactured and difficult to detect. An offset condition can be very useful both in controlling the stiffness and damping characteristics as well as reducing operating temperatures. The increase in offset opens the inlet side of the bearing pad admitting more fresh oil. Bearings with offset greater than 50% should not be run backwards since this essentially creates an offset less than 50%. In practice, offset bearings always include positive preload as well.

Figure 24 - Pad Offset Definition

JOURNAL BEARING TYPES

As mentioned above, the simplest bearing geometry is a circular bushing. It turns out that this is adequate for many machines and a poor bearing for many other applications. Some of the other common geometries of bearings are:

!Partial Arc Bearing

This is a bearing without a top half. Used with known unidirectional loads and to save energy. Has reduced damping and less ability to withstand vibration than a whole bearing. This type is not normally a good choice. Sometimes these are used to minimize friction and most often found in electric motors.

!Axial Groove Bearing

In its simplest form this is a plain circular bush with 2 or more axial grooves for oil distribution. Typically the 4-pad design is found but higher number of pads may be used. May be used when the load angle changes as operational variables change. This design has slightly better stability than a plain bearing without grooves.

!Elliptical (or lemon bore) bearing

Elliptical bearings are very common. In some applications bearing “crush” is specified which changes a circular bush into an elliptical arrangement. This bearing has better dynamic characteristics over a plain bearing and is more stable. Usually the horizontal clearance is 1.5 times to twice the vertical clearance (33 to 50% preload). The angle of the split between halves can be re-oriented for the best load capacity and damping. Dynamically this is a highly asymmetric bearing. The stiffness in the direction of the greater clearance is often an order of magnitude less than across the tight clearance direction. Damping is also significantly less in the direction of the maximum clearance.

Multi-lobe can refer to any number of lobes with the designer taking advantage of the ability to control the geometry of each pad. This usually means adjusting the preload and offset for each pad. The load angle can also be selected for optimum conditions. The three-lobe bearing is the most popular configuration but can be expensive to manufacture. Figure 25 shows a typical 3-lobe bearing and figure 26 shows the hydrodynamic pressure profile created by the individual lobes. The bearing designer can select the clearance, preload, and offset of each pad as well as the load orientation on-pad, between pad or somewhere in between. The optimum configuration will depend on the rotordynamics. Multi-lobe bearings like these are very useful in vertical machines.

Figure 25 - Multi-Lobe Bearing Figure 26 - 3-Lobe Bearing Pressure Profile

!Offset half bearing

This is a very simple and very effective bearing in some applications. It is a 2-pad multi-lobe with 100% offset. The pads are almost always preloaded as well. It is easy to make by boring a plain bushing off-center and the flipping the top half around. This bearing will not tolerate reverse rotation.

Figure 27 - Offset Half Bearing Figure 28 - Offset Half Bearing Hydrodynamic Pressure Profile

Pressure dam bearings, sometimes called pocket or step bearings, are plain bearings that may (rarely) have preload and even offset. Typically they are two-lobe bearings with a pocket cut out of the top half (or unloaded half) that creates an oil dam. The circumferential velocity head of the oil suddenly being impeded by the dam is converted to pressure head. This pressure head imposes a downward load on the shaft that will increase the eccentricity and may stabilize the system. This design also significantly increases the bearing stiffness and may cause the critical speed amplitude and amplification factor to increase. This is often an excellent first design modification if oil whirl is encountered with plain bearings. The height of the dam should be about the same as the diametral clearance of the bearing and up to 3 times that clearance.

A deeper pocket will have significantly reduced effect. If this bearing wears and clearance increases, the effectiveness of the dam will be diminished. Some designs now use multiple pockets, particularly in vertical machines where there is no gravity load. Computer programs are necessary to evaluate such cases. As always, a system design approach will be necessary to evaluate whether this bearing is appropriate. This bearing is also unidirectional in rotation.

Figure 29 - Pressure Dam Bearing Figure 30 - Pressure Dam Bearing Hydrodynamic Pressure Profile ! Tilting Pad Bearings

Tilting pad bearings are becoming the design of the turbomachinery industry. They are very robust, able to be designed to optimize the rotor dynamics and widely accepted as the very best bearing you can get. Tilting pad bearings also have no significant destabilizing cross-coupling. They can be costly and may not always be the best choice when all factors are considered. Figure 31 shows a typical modern tilting pad bearing. Figure 32 is a typical hydrodynamic pressure profile for two tilting pad bearings with different load orientations. Besides clearance, preload and offset, the bearing designer can choose the number of pads, different pivot arrangements, different types of end seals and load orientation. By choosing a 5 pad bearing with load directly over one pad the horizontal and vertical stiffness will be significantly different than a 4-pad design with load between pivots which will have symmetrical characteristics. Figure 33 shows this comparison. Note that Kxx is shorthand for the horizontal stiffness and Kyy means vertical stiffness. Thus, for a particular machine, the various factors may be manipulated to achieve the frequency of critical speeds, the lowest critical speed amplitude response, the lowest operating speed vibration, and the best rotor stability. Figure 33 indicates that the stiffness of the 5-Pad bearing with load between pivots will have a horizontal stiffness less than 1,000,000 LB/IN and a vertical stiffness in excess of 2,500,000 LB/IN. However, if a 4-Pad bearing is selected with load between pivots, the horizontal and vertical stiffnesses are equal at around 2,000,000 LB/IN throughout the speed range.

Figure 31 - Exploded View of Tilting Pad Bearing

Figure 32 - Hydrodynamic Pressure Profile for Two Tilting Pad Bearings

如何对滑动轴承轴瓦进行刮研

如何对轴瓦进行刮研 滑动轴承是靠平滑的面来支撑转动轴的,因而接触部位是一个面,它是在滑动摩擦下工作的轴承,工作平稳、可靠,无噪声,但起动时摩擦阻力较大(因轴瓦要承载转子自身重量,起动时转速低)。轴被轴承支撑的部位称为轴颈,与轴颈相配合起支撑轴颈作用的零件称为轴瓦,轴瓦由瓦体和轴承衬(表面合金)组成,滑动轴承工作时,轴瓦与轴之间要求有一层很薄的油膜起润滑作用,轴与轴瓦被润滑油分开而不发生直接接触,从而大大减小摩擦损失和轴瓦表面合金磨损,油膜还具有较强的吸振能力。如果存在润滑不良,则轴瓦与轴之间会直接摩擦,摩擦会产生很高的温度,虽然轴瓦是由特殊的有一定耐高温合金材料制成,但发生直接摩擦产生的高温仍然会将轴瓦烧坏,轴瓦还可能由于负荷过大、高温过高、润滑油存在杂物或黏度异常等因素造成轴瓦瓦衬变形直至烧瓦。 轴瓦刮研就是将精车后的弧面与所装配的轴进行刮合,通常是内孔面,轴瓦的瓦衬一般都要进行刮研,随着现代加工技术的不断改进,成品瓦侧隙可以达到要求,下瓦接触角、面就要根据现场轴颈经旋转后的情况来进行刮研,因滑动轴承工作时,轴瓦与轴之间要求有一层很薄的油膜起润滑作用。轴瓦刮研的目的是为了瓦衬形成圆的几何形状,使轴瓦与轴颈间存在楔形缝隙,以保证轴颈旋转时,摩擦面间能形成楔形油膜,使轴颈上升离开瓦衬,在油膜的浮力作用下运转,以减轻与瓦衬的摩擦,降底其磨损与动力的消耗,从而可以通过向轴瓦供油,带走轴瓦工作时产生的热量,以冷却轴承,并且在轴瓦与轴颈间隙形成稳定的有足够承载能力的油膜,以保证液态润滑。 轴瓦的光洁度及平整越好越有利于油膜形成(也包括侧隙油带的形成),下瓦接触角内一般是建立油膜产生油压的区域(也就是说下瓦接触角内是油膜压力区),现有的轴瓦刮研方法所加工出的轴瓦表面的光滑、平整度较低,在滑动轴承工作时难以形成稳定的油膜,从而轴与轴瓦之间的磨擦阻力大,运转时产生的热量大,磨损加剧,使用寿命短,维护频繁、成本高。 轴瓦的刮研是按轴瓦与轴承的配合来对轴瓦表面进行刮研加工,使其在接触角范围内贴合严密和有适当的间隙,从而达到设备中能形成良好的油膜。 1.开油槽:先用薄平錾进行初开,然后用细砂布和刮刀作用加工。

滑动轴承

第八章滑动轴承 8.1 重点、难点分析 本章的重点内容是滑动轴承轴瓦的材料及选用原则;非液体摩擦滑动轴承的设计准则及设计计算;液体动力润滑径向滑动轴承的设计计算。难点是液体动力润滑径向滑动轴承的设计计算及参数选择。 8.1.1 轴瓦材料及其应用 对轴瓦材料性能的要求:具有良好的减摩性、耐磨性和咬粘性;具有良好的摩擦顺应性、嵌入性和磨合性;具有足够的强度和抗腐蚀的能力和良好的导热性、工艺性、经济性等。 常用轴瓦材料:金属材料、多孔质金属材料和非金属材料。其中常用的金属材料为轴承合金、铜合金、铸铁等。 8.1.2 非液体摩擦滑动轴承的设计计算 对于工作要求不高、转速较低、载荷不大、难于维护等条件下的工作的滑动轴承,往往设计成非液体摩擦滑动轴承。这些轴承常采用润滑脂、油绳或滴油润滑,由于轴承得不到足够的润滑剂,故无法形成完全的承载油膜,工作状态为边界润滑或混合摩擦润滑。 非液体摩擦轴承的承载能力和使用寿命取决于轴承材料的减摩耐磨性、机械强度以及边界膜的强度。这种轴承的主要失效形式是磨料磨损和胶合;在变载荷作用下,轴承还可能发生疲劳破坏。 因此,非液体摩擦滑动轴承可靠工作的最低要求是确保边界润滑油膜不遭到破坏。为了保证这个条件,设计计算准则必须要求: p≤[p],pv≤[pv],v≤[v] 限制轴承的压强p,是为了保证润滑油不被过大的压力挤出,使轴瓦产生过度磨损;限制轴承的pv值,是为了限制轴承的温升,从而保证油膜不破裂,因为pv值是与摩擦功率损耗成正比的;在p及pv值经验算都符合要求的情况下,由于轴发生弯曲或不同心等引起轴承边缘局部压强相当高,当滑动速度高时,局部区域的pv值可能超出许用值,所以在p较小的情况下还应该限制轴颈的圆周速度v。 8.1.3液体动力润滑径向滑动轴承设计计算 液体动力润滑的基本方程和形成液体动力润滑(即形成动压油膜)的条件已在第一章给出,这里不再累述。 1.径向滑动轴承形成动压油膜的过程 径向滑动轴承形成动压油膜的过程可分为三个阶段: (1)起动前阶段,见图8-1a;

滑动轴承

滑动轴承 滑动轴承[huá dòng zhóu chéng] 滑动轴承(sliding bearing),在滑动摩擦下工作的轴承。滑动轴承工作平稳、可靠、无噪声。在液体润滑条件下,滑动表面被润滑油分开而不发生直接接触,还可以大大减小摩擦损失和表面磨损,油膜还具有一定的吸振能力。但起动摩擦阻力较大。轴被轴承支承的部分称为轴颈,与轴颈相配的零件称为轴瓦。为了改善轴瓦表面的摩擦性质而在其内表面上浇铸的减摩材料 层称为轴承衬。轴瓦和轴承衬的材料统称为滑动轴承材料。滑动轴承应用场合一般在低速重载工况条件下,或者是维护保养及加注润滑油困难的运转部位。常用的滑动轴承材料有轴承合金(又叫巴氏合金或白合金)、耐磨铸铁、铜基和铝基合金、粉末冶金材料、塑料、橡胶、硬木和碳-石墨,聚四氟乙烯(特氟龙、PTFE)、改性聚甲醛(POM)、等。滑动轴承(sliding bearing),在滑动摩擦下工作的轴承。滑动轴承工作平稳、可靠、无噪声。在液体润滑条件下,滑动表面被润滑油分开而不发生直接接触,还可以大大减小摩擦损失和表面磨损,油膜还具有一定的吸振能力。但起动摩擦阻力较大。轴被轴承支承的部分称为轴颈,与轴颈相配的零件称为轴瓦。为了改善轴瓦表面的摩擦性质而在其内表面上浇铸的减摩材料层称为轴承衬。轴瓦和轴承衬的材料统称为滑动

轴承材料。聚四氟乙烯(PTFE)、改性聚甲醛(POM)、等。滑动轴承应用场合一般在低速重载工况条件下,或者是维护保养及加注润滑油困难的运转部位。[1]滑动轴承种类很多。滑动轴承①按能承受载荷的方向可分为径向(向心)滑动轴承和推力(轴向)滑动轴承两类。②按润滑剂种类可分为油润滑轴承、脂润滑轴承、水润滑轴承、气体轴承、固体润滑轴承、磁流体轴承和电磁轴承7类。③按润滑膜厚度可分为薄膜润滑轴承和厚膜润滑轴承两类。④按轴瓦材料可分为青铜轴承、铸铁轴承、塑料轴承、宝石轴承、粉末冶金轴承、自润滑轴承和含油轴承等。⑤按轴瓦结构可分为圆轴承、椭圆轴承、三油叶轴承、阶梯面轴承、可倾瓦轴承和箔轴承等。滑动轴承轴瓦分为剖分式和整体式结构。为了改善轴瓦表面的摩擦性质,常在其内径面上浇铸一层或两层减摩材料,通常称为轴承衬,所以轴瓦又有双金属轴瓦和三金属轴瓦。轴瓦或轴承衬是滑动轴承的重要零件,轴瓦和轴承衬的材料统称为轴承材料。由于轴瓦或轴承衬与轴颈直接接触,一般轴颈部分比较耐磨,因此轴瓦的主要失效形式是磨损。轴瓦的磨损与轴颈的材料、轴瓦自身材料、润滑剂和润滑状态直接相关,选择轴瓦材料应综合考虑这些因素,以提高滑动轴承的使用寿命和工作性能。轴承的材料有1)金属材料,如轴承合金、青铜、铝基合金、锌基合金等轴承合金:轴承合金又称白合金,主要是锡、铅、锑或其它金属的合金,由于其

滑动轴承的装配工艺

滑动轴承的装配工艺 一滑动轴承的检修内容 1.检修油道是否畅通,润滑是否良好 2.检查滑动轴承的磨损情况,磨损超过标准时应更换 二滑动轴承的检修工艺 滑动轴承分整体式(轴套)和剖分式(轴瓦)两种 1.整体式滑动轴承拆卸与组装 滑动轴承的磨损超过标准时,应进行更换,先将要换下的轴套从机体上拆下,然后按下列程序进行装配。 ①清理机体内孔,疏通油道,检查尺寸。 ②压入轴套,根据轴套的尺寸和结合的过盈大小,可以用压入法、温差法或手锤加垫板将轴套敲入,压入时必须加油,以防轴套外圈拉毛或咬死等现象。 ③轴套定位,在压入之后,对负荷较重的滑动轴承,轴套还应固定,以防轴套在机体内转动。 ④轴套孔的修整,对于整体式的薄壁轴套在压入后,内孔易发生变形如内径缩小或成为椭圆形、圆锥形等,必须修轴套内孔的形状和尺寸,便于轴配合时符合要求,修整轴套孔可采用铰削、刮研、研磨等方法。 2.剖分式滑动轴承的拆卸与组装。

①拆卸 a拆除轴承盖螺栓,卸下轴承盖。 b将轴吊出。 c卸下上瓦盖与下瓦座内的轴瓦。 ②组装前 组装前应仔细检查各部尺寸是否合适,油路是否畅通,油槽是否合适。 ③轴瓦与轴颈的组装 a圆形孔,上、下轴瓦分别和轴瓦刮配,以达到规定间隙,要求轴瓦全长接触良好,剖分面上可装垫片以调整上面与轴颈的间隙。 b近似于圆形孔(其水平直径>垂直直径)轴承经加工后抽去剖分面上的垫片,以保证上瓦及两侧间隙,如不符合要求,可继续配刮直至符合要求为止。 c成形油楔面用加工保证,一般在组装时不宜修刮,组装时应注意油楔方向与主轴方向一致。 d薄壁轴瓦不宜修刮。 e主轴外伸长度较大时,考虑到主轴由于自身重量产生的变形,应把前轴承下瓦在主轴外伸端刮得低些,否则主轴可能会“咬死”。

滑动轴承作业

滑动轴承 学号 一 选择题 1. 宽径比d B /是设计滑动轴承时首先要确定的重要参数之一,通常取 d B / 。 A. 1~10 B.0.1~1 C. 0.3~1.5 D. 3~5 2. 下列材料中 不能作为滑动轴承轴瓦或轴承衬的材料。 A. ZSnSb11Cu6 B. HT200 C. GCr15 D. ZCuPb30 3. 在非液体润滑滑动轴承中,限制p 值的主要目的是 。 A. 防止出现过大的摩擦阻力矩 B. 防止轴承衬材料发生塑性变形 C. 防止轴承衬材料过度磨损 D. 防止轴承衬材料因压力过大而过度发热 4. 不是静压滑动轴承的特点。 A. 起动力矩小 B. 对轴承材料要求高 C. 供油系统复杂 D. 高、低速运转性能均好 5. 设计液体动压径向滑动轴承时,若通过热平衡计算发现轴承温升过高,下列改进措施中,有效的是 。 A. 增大轴承宽径比 B. 减小供油量 C. 增大相对间隙 D. 换用粘度较高的油 6. 含油轴承是采用 制成的。 A. 塑料 B. 石墨 C 铜合金 D. 多孔质金属 7. 液体摩擦动压径向轴承的偏心距e 随 而减小。 A. 轴颈转速n 的增加或载荷F 的增加 B. 轴颈转速n 的增加或载荷F 的减少 C. 轴颈转速n 的减少或载荷F 的减少 D. 轴颈转速n 的减少或载荷F 的增加 8. 径向滑动轴承的直径增大1倍,长径比不变,载荷不变,则轴承的压强p 变为原来的 倍。 A. 2 B. 1/2 C. 1/4 D. 4 9. 液体动压径向滑动轴承在正常工作时,轴心位置1O 、轴承孔中心位置O 及轴承中的油压分布应如图12-1的 所示。

图12-1 A. (a) B. (b) C. (c) D. (d) 10. 动压液体摩擦径向滑动轴承设计中,为了减小温升,应在保证承载能力的前提下适当 。 A. 增大相对间隙ψ,增大宽径比d B B. 减小ψ,减小d B C. 增大ψ,减小d B D. 减小ψ,增大d B 11. 动压滑动轴承能建立油压的条件中,不必要的条件是 。 A. 轴颈和轴承间构成楔形间隙 B. 充分供应润滑油 C. 轴径和轴承表面之间有相对滑动 D. 润滑油温度不超过50C ο 12. 在 情况下,滑动轴承润滑油的黏度不应选得较高。 A. 重载 B. 工作温度高 C. 高速 13. 与滚动轴承相比较,下述各点中, 不能作为滑动轴承的优点。 A. 径向尺寸小 B. 启动容易 C. 运转平稳,噪声低 D. 可用于高速情况下 14. 滑动轴承轴瓦上的油沟不应开在 。 A. 油膜承载区 B. 油膜非承载区 C. 轴瓦剖面上 15. 计算滑动轴承的最小油膜厚度m in h ,其目的是 。 A. 验算轴承是否获得液体摩擦 B. 汁算轴承的部摩擦力 C. 计算轴承的耗油量 D. 计算轴承的发热量 16. 设计动压径向滑动轴承时,若轴承宽径比取得较大,则 。 A. 端泄流量大,承载能力低,温升高 B. 端泄流量大,承载能力低,温升低 C. 端泄流量小,承载能力高,温升低 D. 端泄流量小,承载能力高,温升高 17. 双向运转的液体润滑推力轴承中,止推盘工作面应做成题图12-2 所示的形状。

滑动轴承的结构形式

滑动轴承的结构形式 食品包装机温馨提示:滑动轴承主要由滑动轴承座、轴瓦或轴套组成。 装有轴瓦或轴套的起支律轴与轴上零件作用壳体称为m 动轴承座:滑动轴承中与支承轴颈(以下简称轴颈)相配的圆筒形整休零件称为轴套.与轴颈相配的对开式零件就是轴瓦。 常用的径向滑动轴承有以下儿种结构形式。 (1)整体式汾动轴承.整体式滑动轴承是一种常见的整体式向心滑动轴承.滑动轴承座孔中压人用其有减摩擦特性的材料制成的轴弈.并用紧定螺钉fI 定。浴动轴承座顶部设有安装润淤装置螺纹孔。轴套卜开有油孔.并在内表而上开有油挤.以输送润滑油。减小摩擦;简单的轴套内孔则无油桐.淆动轴承磨祝后.只须更换轴弃即可。 整体式淆动轴承通常应川于轻载、低速或间歇工作的场合. (2)对开式附动轴承。对开式滑动轴承由轴承盖、轴承座、上轴瓦、下轴瓦和连接螺栓等组成,轴承座是轴承的革础部分.用螺栓固定于机架上:轴承盖与轴承座的结合面呈台阶形式.以保证两者定位可靠.并防iF 横向错动。轴承盖与轴承座采用探栓连接.并仄紧上、下轴瓦。通过轴承盖上连接的润滑装界.可将润浴油经油孔愉送到轴颈表面。 在轴承盖与轴承座之间一般留有5 mm 左右的间隙.并在上、下抽瓦的对开面处垫人适址的调整垫片,当轴瓦肺损后可根据其脾损程度.更换一些调整垫片.使轴颈与轴瓦之间仍能保持要求的间隙。 轴瓦的两端通常带有凸缘,以防止在轴承座中发生轴向移动;一般用销钉或紧定螺钉固定,以防止其周向转动。为了将润淤油引人和分布到轴承的整个作表面.轴瓦上加工有油孔,并在内表面上开有油抽。 (3)可调问w 式滑动轴承。滑动轴承的轴瓦在使川中难免磨损造成间隙增大.采用间隙可调橄的滑动轴承, 并延长轴瓦的使用寿命。可调式轴承采用带锥形表面的轴套.有内锥外柱和内柱外锥两种形式.通过轴颈与轴瓦问的轴向移动实现轴承径向间隙的调整。 (4)自位淞动轴承。自位汾动轴承是相对于轴颈表面可自行调恨轴线偏角的淤动轴承。 其特点是轴瓦与轴承盖、轴承座之间为球面接触.轴瓦在轴承中可随轴烦轴线转动. 以上资料由:食品包装机械 ,饮料灌装包装设备 ,封口机 锁口机 ,清洗灌装设备 ,开水器系列开水器系列,,米制品加工机械 ,夹层锅系列 ,烧烤小吃设备烧烤小吃设备提供提供提供!!

滑动轴承 ppt

第十二章滑动轴承 §12-1 滑动轴承概述 §12-2 滑动轴承的典型结构 §12-3 滑动轴承的失效形式及常用材料 §12-4 滑动轴承轴瓦结构 §12-5 滑动轴承润滑剂的选择 §12-6 不完全液体润滑滑动轴承的设计计算 §12-7 液体动力润滑径向滑动轴承的设计计算 §12-8 其它形式滑动轴承简介

滑动轴承概述1轴承的作用是支承轴。轴在工作时可以是旋转的,也可以是静止的。1.能承担一定的载荷,具有一定的强度和刚度。 2.具有小的摩擦力矩,使回转件转动灵活。 3.具有一定的支承精度,保证被支承零件的回转精度。根据轴承中摩擦的性质,可分为滑动轴承和滚动轴承。 一、轴承应满足如下基本要求: 二、轴承的分类 根据能承受载荷的方向,可分为向心轴承、推力轴承、向心推力轴承。 (或称为径向轴承、止推轴承、径向止推轴承)。 根据润滑状态,滑动轴承可分为:不完全液体润滑滑动轴承。 完全液体润滑滑动轴承。

滑动轴承概述2四、滑动轴承设计内容 三、滑动轴承的特点滚动轴承绝大多数都已标准化,故得到广泛的应用。但是在以下场合,则主要使用滑动轴承: 1.工作转速很高,如汽轮发电机。 2.要求对轴的支承位置特别精确,如精密磨床。 3.承受巨大的冲击与振动载荷,如轧钢机。 4.特重型的载荷,如水轮发电机。 5.根据装配要求必须制成剖分式的轴承,如曲轴轴承。 6.在特殊条件下工作的轴承,如军舰推进器的轴承。 7.径向尺寸受限制时,如多辊轧钢机。 轴承的型式和结构选择;轴瓦的结构和材料选择;轴承的结构参数设计;润滑剂及其供应量的确定;轴承工作能力及热平衡计算。

径向滑动轴承的典型结构1一、径向滑动轴承的结构 1.整体式径向滑动轴承 特点:结构简单,成本低廉。 应用:低速、轻载或间歇性工作的机器中。轴承座整体轴套 螺纹孔油杯孔 因磨损而造成的间隙无法调整。 只能从沿轴向装入或拆出。

滑动轴瓦

轴瓦(bush) 汽轮机中,轴瓦是轴承的重要构件之一,轴瓦是滑动轴承和轴接触的部分,非常光滑,一般用青铜、减摩合金等耐磨材料制成,在特殊情况下,可以用木材、塑料或橡皮制成。也叫“轴衬”,形状为瓦状的半圆柱面。 其主要作用是:承载轴颈所施加的作用力、保持油膜稳定、使轴承平稳地工作并较少轴承的摩擦损失。 分为轴向推力瓦和径向瓦,径向瓦起到支撑转子和转动部分的作用,推力瓦承担轴向定位和轴向推力的作用,是重要的静止部件。 滑动轴承工作时,轴瓦与转轴之间要求有一层很薄的油膜起润滑作用。如果由于润滑不良,轴瓦与转轴之间就存在直接的摩擦,摩擦会产生很高的温度,虽然轴瓦是由于特殊的耐高温合金材料制成,但发生直接摩擦产生的高温仍然足于将器烧坏。轴瓦还可能由于负荷过大、温度过高、润滑油存在杂质或黏度异常等因素造成烧瓦。烧瓦后滑动轴承就损坏了。滑动轴承(sliding bearing),在滑动摩擦下工作的轴承。滑动轴承工作平稳、可靠、无噪声。在液体润滑条件下,滑动表面被润滑油分开而不发生直接接触,还可以大大减小摩擦损失和表面磨损,油膜还具有一定的吸振能力。但起动摩擦阻力较大。轴被轴承支承的部分称为轴颈,与轴颈相配的零件称为轴瓦。为了改善轴瓦表面的摩擦性质而在其内表面上浇铸的减摩材料层称为轴承衬。轴瓦和轴承衬的材料统称为滑动轴承材料。常用的滑动轴承材料有轴承合金(又叫巴氏合金或白合金)、耐磨铸铁、铜基和铝基合金、粉末冶金材料、塑料、橡胶、硬木和碳-石墨,聚四氟乙烯(PTFE)、改性聚甲醛(POM)、等。 滑动轴承应用场合一般在低速重载工况条件下,或者是维护保养及加注润滑油困难的运转部位 滑动轴承的分类 滑动轴承种类很多。 ①按能承受载荷的方向可分为径向(向心)滑动轴承和推力(轴向)滑动轴承两类。 ②按润滑剂种类可分为油润滑轴承、脂润滑轴承、水润滑轴承、气体轴承、固体润滑轴承、磁流体轴承和电磁轴承7类。 ③按润滑膜厚度可分为薄膜润滑轴承和厚膜润滑轴承两类。 ④按轴瓦材料可分为青铜轴承、铸铁轴承、塑料轴承、宝石轴承、粉末冶金轴承、自润滑轴承和含油轴承等。 ⑤按轴瓦结构可分为圆轴承、椭圆轴承、三油叶轴承、阶梯面轴承、可倾瓦轴承和箔轴承等。 滑动轴承的轴瓦结构和轴承材料 轴瓦分为剖分式和整体式结构。为了改善轴瓦表面的摩擦性质,常在其内径面上浇铸一层或两层减摩材料,通常称为轴承衬,所以轴瓦又有双金属轴瓦和三金属轴瓦。 轴瓦在轴承上,直接与轴相接触的部分,承受载荷并且与轴具有相对运动。为减少摩擦,磨损对轴瓦材料提出各种要求,除要求摩擦副间摩擦系数小,耐磨外还应满足以下几点:1.应具有足够的抗压强度、抗疲劳强度和承受冲击的能力。轴承合金厚度为(0.013-0.13) mm。 2.抗粘着性好。当载荷大、转速高,轴承间隙过小,表面光洁度不高,润滑不良时,要注意精心选择材料的匹配,防止摩擦副的粘着磨损一胶合。 3.具有适应性和容纳异物的能力。硬度低,塑性好和弹性系数低的材料具有良好的适应性和容纳异物的能力。

滑动轴承轴瓦结构及其定位概要

滑动轴承的轴瓦结构及其定位 轴瓦在轴承中直接与轴颈接触,其结构是否合理对轴承的承载能力及使用寿命影响很大。有时为了节省贵重材料或结构需要,常在轴瓦的内表面上浇注或轧制一层轴承合金,称为轴承衬。轴瓦应具有一定的强度和刚度,在轴承中定位可靠,便于输入润滑剂,容易散热,并且装拆、调整方便。 一、轴瓦的类型及结构 常用的轴瓦有整体式和对开式两种结构。整体式轴瓦按材料及制法不同,分为整体轴套和单层、双层或多层材料的卷制轴套。非金属整体式轴瓦既可以是整体非金属轴套,也可以是在钢套上镶衬非金属材料。 整体式轴瓦需从轴端安装和拆卸,可修复性差。 对开式轴瓦有厚壁轴瓦和薄壁轴瓦之分。厚壁轴瓦用铸造方法制造,内表面可附有轴承衬,常将轴承合金用离心铸造法浇注在铸铁、钢或青铜轴瓦的内表面上。为使轴承合金和轴瓦贴附得好,常在轴瓦内表面上制出各种形式的榫头、凹沟或螺纹。 图1 轴瓦的结构 薄壁轴瓦由于能用双金属板连续轧制等新工艺进行大量生产,故质量稳定,成本低,但轴瓦刚性小,装配时不再修刮轴瓦内圆表面,轴瓦受力后,其形状完全取决于轴承座的形状,因此,轴瓦和轴承座均需精密加工。薄壁轴瓦在汽车发动机、柴油机上得到广泛的应用。

二、轴瓦的定位 轴瓦和轴承座不允许有相对移动。为防止轴瓦与轴承座之间产生轴向和周向的相对移动,要对轴瓦进行定位。 为了防止轴瓦移动,可将其两端做出凸缘来作轴向定位,如图所示。 图2 凸缘定位 也可以用紧定螺钉或销钉将其固定在轴承座上,或在轴瓦剖分面上冲出定位唇以供定位用。 图3 紧定螺钉和销钉定位 三、油孔和油槽 开油槽目的:把润滑油导入轴颈和轴承所构成的运动副表面。 开油槽原则:尽量开在非承载区,尽量不要降低或少降低承载区油膜的承载能力;轴向油槽不能开通至轴承端部,应留有适当的油封面。 轴向油槽分为单轴向油槽和双轴向油槽。对于整体式径向轴承,轴颈单向旋转时,荷载方向变化不大,单轴向油槽最好开设在最大油膜厚度位置,以保证润滑油从压力最小的地方输入轴承。对开式径向轴承,常把轴向油槽开设轴承剖分面处(剖分面与荷载作用线成90°),如果轴颈双向旋转,可在轴承剖分面上开设双轴向油槽,通常轴向油槽应较轴承宽度稍短,以便在轴承两端留出封油面,防止润滑油从端部大量流失。周向油槽适用于载荷方向变动范围超出180°的场合,它常设在轴承宽度中部,把轴承分为两个独立部分;当宽度相同时,设

机械设计考研练习题-滑动轴承

滑动轴承 一 选择题 (1) 宽径比d B /是设计滑动轴承时首先要确定的重要参数之一,通常取 d B / C 。 A. 1~10 B.0.1~1 C. 0.3~1.5 D. 3~5 (2) 下列材料中 C 不能作为滑动轴承轴瓦或轴承衬的材料。 A. ZSnSb11Cu6 B. HT200 C. GCr15 D. ZCuPb30 (3) 在非液体润滑滑动轴承中,限制p 值的主要目的是 C 。 A. 防止出现过大的摩擦阻力矩 B. 防止轴承衬材料发生塑性变形 C. 防止轴承衬材料过度磨损 D. 防止轴承衬材料因压力过大而过度发热 (4) 在滑动轴承材料中, B 通常只用于作为双金属或三金属轴瓦的表层材料。 A. 铸铁 B. 轴承合金 C. 铸造锡磷青铜 D. 铸造黄铜 (5) 在滑动轴承轴瓦材料中,最易用于润滑充分的低速重载轴承的是 C 。 A. 铅青铜 B. 巴氏合金 C. 铝青铜 D. 锡青铜 (6) 滑动轴承的润滑方法,可以根据 A C 来选择。 A. 平均压强p B. 3pv C. 轴颈圆周速度v D. pv 值 (7) B 不是静压滑动轴承的特点。 A. 起动力矩小 B. 对轴承材料要求高 C. 供油系统复杂 D. 高、低速运转性能均好 (8) 设计液体动压径向滑动轴承时,若通过热平衡计算发现轴承温升过高,下列改进措施中,有效的是 C 。 A. 增大轴承宽径比 B. 减小供油量 C. 增大相对间隙 D. 换用粘度较高的油 (9) 巴氏合金用于制造 B 。 A. 单层金属轴瓦 B. 双层及多层金属轴瓦 C. 含油轴承轴瓦 D. 非金属轴瓦 (10) 含油轴承是采用 D 制成的。 A. 塑料 B. 石墨 C 铜合金 D. 多孔质金属 (11) 下述材料中, C 是轴承合金(巴氏合金)。 A. 20CrMnTi B. 38CrMnMo C. ZSnSb11Cu6 D. ZCuSnl0Pbl (12) 液体摩擦动压径向轴承的偏心距e 随 B 而减小。

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