Balancing Fundamentals Schenck [PDF]

Zitiervorschau

History of Balancing The first patent for balancing technology was filed by Henry Martinson of Canada in 1870, four years after the development of the dynamo by Siemens. Near the turn of the century, Akimoff (USA) and Stodola (Switzerland) attempted to develop Martinson's technology and apply it for industrial use. However, it was in 1907 when a modified version of the technology was patented by Dr. Franz Lawaczek, and offered to Carl Schenck, Darmstadt, Germany, for development. Schenck built the first industrial two-plane balancer, and subsequently bought exclusive world rights to the dynamic balancing machine in 1915.

Through the years, craftsmanship and quality have been the hallmarks of Schenck products. Technology advancements gave way to improved sensitivity, frequency selectivity and plane separation capability. The development of electronics and mechanical/electrical transducers, greatly reduced balancing time and paved the way for modern balancing technology.

Today Schenck balancing equipment is used with confidence for a wide range of applications - from the smallest rotors for dental drill instruments to the largest steam turbines in the world. Our precision balancing machines assure accurate, dependable rotor operation and are available in nearly any configuration for rotors weighing as much as 600,000 lbs.

Fundamental s of

Balancing

Fundamentals of Balancing is designed to give those less interested in theory and design the practical skills to increase balancing efficiency and streamline production. Hands-on exercises will be used extensively to improve the operators technique on both balancing machine and instrumentation. After a brief overview of basic theory that includes the principles of machine operation, tolerance specifications and machine setup, students will be divided into small groups with those that have similar equipment and applications. Service technicians and engineers will then supervise a number of exercises on machines and instruments that closely resemble the students' equipment.* Other topics include:

• The different types of unbalance (static, dynamic & couple). • ISO tolerances and terminology. • Industry standard tolerances vs. drawing tolerances. • Machine & rotor setup. • Instrumentation functions & operation. • Maximizing instrumentation features. • Selecting the best balancing speed. • How to avoid interference of measurements due to drive and roller harmonics. • Proper machine maintenance and troubleshooting. *Since instruments and machines will be selected to closely represent the attendees' equipment, please specify balancer make, model and instrument with application. Certification: Level I- Balancing Operator Who should attend: This course is intended for newly appointed balancing machine operators (up to 2-3 years) and other personnel directly related to the balancing process. A mechanical aptitude with emphasis on rotating equipment and precision measurements (technician level) or machine shop experience is recommended.

Balancing Fundamentals

Definition According to DIN/ISO 1925 Unbalance is “that condition which exists in a rotor when vibratory force or motion is imparted to its bearings as a result of centrifugal forces.” Why Balance? An unbalanced rotor will cause vibration and stress in the rotor itself and in its supporting structure. Balancing of the rotor is, therefore, necessary to accomplish one or more of the following: a. Increase quality of product b. Minimize vibration c. Minimize audible and signal noises d. Minimize structural stresses e. Minimize operator annoyance and fatigue f. Increase bearing life g. Minimize power loss Unbalance in just one rotating component of an assembly may cause the entire assembly to vibrate. This induced vibration in turn may cause excessive wear in bearings, bushings, shafts, spindles, gears, etc., substantially reducing their service life. Vibrations set up highly undesirable alternating stresses in structural supports and housings, which may eventually lead to their complete failure. Performance is decreased because of the absorption of energy by the supporting structure. Vibrations may be transmitted through the floor to adjacent machinery and seriously impair its accuracy or proper functioning.

Unbalance vs. centrifugal force

Centrifugal force acts upon the entire mass of a rotating component, impelling each particle outward and away from the axis of rotation in a radial direction. If the mass of a rotating component is evenly distributed about its shaft axis, the part is "balanced" and rotates without vibration. However, if an excess of mass exists on one side of a rotor, the centrifugal force acting upon this heavy side exceeds the centrifugal force exerted by the light side and pulls the entire rotor in the direction of the heavy side. This figure shows the side view of a rotor having an excess mass m on one side. Due to centrifugal force exerted by m during rotation, the entire rotor is being pulled in the direction of the arrow F.

Centrifugal force increases with the square of the speed A rotating element having an uneven mass distribution, i.e., unbalance, will vibrate due to the excess centrifugal force exerted during rotation by the heavier side of the rotor. When at rest, the excess mass exerts no centrifugal force and, therefore, causes no vibration. Yet the actual unbalance is still present. Unbalance, therefore, is independent of rotational speed and remains the same, whether the part is at rest or is rotating (provided the part does not deform during rotation). Centrifugal force, however, varies with speed. The higher the speed, the greater the centrifugal force exerted by the unbalance and the more violent the vibration. Centrifugal force increases proportionately to the square of the increase in speed. If the speed is doubled, the centrifugal force quadruples; if the speed is tripled, the centrifugal force is multiplied by nine.

Causes of unbalance

The excess of mass on one side of the rotor in this figure is called unbalance. In the example illustrated, it is the "heavy spot". Unbalance may also occur due to lack of mass (such as a drill hole, porous spot, etc.) in which case it is called the "light spot”. Either one may be caused by a variety of reasons, including the following:

a. Tolerances in fabrication, including casting, machining, and assembly. b. Variation within materials, such as voids, porosity, inclusions, grain, density, and finishes. c. Nonsymmetrical design, including motor windings, part shapes, location, and density of finishes. d. Nonsymmetrical in use, including distortion, dimensional changes, and shifting of parts due to rotational stresses, aerodynamic forces, and temperature changes. Symmetrical design and careful setting of tolerances and fits can often minimize balancing problems. Large amounts of unbalance require large corrections. If such corrections are made by removal of material, additional machining cost is involved and part strength may be affected. If corrections are made by addition of material, cost is again a factor and space requirements for the added material may be a problem. Manufacturing processes are the major source of unbalance. Unmachined portions of castings or forgings, which cannot be made concentric and symmetrical with respect to the shaft axis, introduce substantial unbalance. Manufacturing tolerances and processes, which permit any eccentricity or lack of squareness with respect to the shaft axis, are sources of unbalance. The tolerances, necessary for economical manufacturing and assembly of several elements of a rotor, permit radial displacement of parts of the assembly and thereby introduce unbalance. Limitations imposed by rotor design often introduce unbalance effects that cannot be corrected adequately by refinement of the design itself. For example, electrical design considerations impose a requirement that one coil be at a greater radius than the others in a certain type of universal motor armature. It is impractical to design a compensating unbalance into the armature. Fabricated parts, such as fans, often distort nonsymmetrically under service conditions. Design and economic considerations prevent the adaptation of methods that might eliminate this distortion and thereby reduce the resulting unbalance. Ideally, rotating parts should always be designed for inherent balance, whether a balancing operation is to be performed or not. Where low service speeds are involved and the effects of a reasonable amount of unbalance can be tolerated, this practice may eliminate the need for balancing. In parts that require unbalanced masses for functional reasons, these masses can often be counterbalanced by designing for symmetry about the shaft axis. Correction methods Corrections for rotor unbalance are made either by the addition of mass to the rotor, by the

removal of material, or in some cases, by relocating the shaft axis (“mass centering"). The selected correction method should ensure that there is sufficient space or material to allow correction of the maximum unbalance which may occur. The ideal correction method permits a reduction of the maximum initial unbalance to less than balance tolerance in a single correction step. However, this is often difficult to achieve. The more common methods described below, e.g., drilling, usually permit a single step reduction of 10:1 in unbalance if carried out carefully. Milling and grinding are less accurate, unless carried out in automatic or semi-automatic balancing machines, which have integrated mass correction devices. The addition of mass may achieve a reduction ratio as large as 20:1 or higher, provided the mass and its position are closely controlled. If the method selected for reduction of maximum initial unbalance cannot be expected to bring the rotor within the permissible residual unbalance in a single correction step, a preliminary correction is made. Then a second correction follows to reduce the remaining unbalance to its permissible value. Addition of Mass 1. Addition of two-component epoxy. It is difficult to apply the material so that its center-of-gravity is precisely at the desired correction location. Variations in location introduce errors in correction. This method is often used in balancing of wound armatures. 2. Addition of bolted or riveted standard washers. This method is quick, but somewhat limited in accuracy because the washers come in incremental sizes, i.e., the mass of one washer may vary considerably from the mass of the next washer of the same type and size. This method is often used in balancing of AC motor rotors. 3. Addition of premanufactured weights. The same limitations as in (2) apply. A typical application is addition of spring clips to the blades of automotive A/C blower wheels. 4. Addition of cut-to-size weights. This is practiced on drive shafts, for instance, by resistance welding the weights to the outside rotor surface. Welding provides a means of attaching a wide variety of correction masses at any desired angular locations. Care must be taken that welding heat does not distort the rotor. Removal of Mass 1. Drilling. Material is removed from the rotor by a drill which penetrates the rotor to a measured depth, thereby removing the intended amount of material with a high degree of accuracy. A depth gage or limit switch can be provided on the drill spindle to ensure that the hole is drilled to the desired depth. This is probably the most effective method of unbalance correction. 2. Milling, Shaping, or Fly Cutting. This method permits accurate removal of mass when the rotor surfaces, from which the depth of cut is measured, are machined surfaces, and when means are provided for accurate measurement of cut with respect to those surfaces; used where relatively large corrections are required. 3. Grinding. In general, grinding must be considered a trial-and-error method of correction. It is difficult to evaluate the actual mass of the material, which is removed. This method is usually used only where the rotor design or material does not permit a more economical type of correction. Mass Centering

Such a procedure is used, for instance, to reduce initial unbalance in crankshaft castings or forgings. The shaft is mounted in a balanced cage or cradle, which in turn, is rotated in a balancing machine. The shaft is adjusted radially with respect to the cage until the unbalance indication for the combined shaft and cradle assembly is within a given tolerance. At this point the principal inertia axis of the shaft essentially coincides with the shaft axis of the balanced cage. Center drills, guided along the axis of the cage, then drill the shaft centers and thereby provide an axis in the crankshaft about which it is in balance. The subsequent machining of the crankshaft is carried out between these centers. Because material removal is uneven at different parts of the shaft, the machining operation will introduce some new unbalance. A final balancing operation is, therefore, still required. It is generally accomplished by drilling into the crankshaft counterweights. However, final unbalance corrections are small and balancing time is significantly shortened. Furthermore, final correction (usually by drilling) does not exceed the material available for it, nor does it reduce the mass of the counterweights to a level where they no longer perform their proper function, namely to compensate for the opposed throws and crankpins of the crankshaft.

Units of unbalance Unbalance is measured in ounce•inches, gram•inches, or gram•millimeters, all having a similar meaning, namely a mass multiplied by its distance from the shaft axis, i.e., its "radius". An unbalance of 100 g•in, for example, indicates that one side of the rotor has an

excess mass equivalent to 10 grams at a 10 inch radius, or 20 grams at a 5 inch radius.

View of Rotor With 100 g•in Unbalance In each case the mass, when multiplied by its distance from the shaft axis, amounts to the same unbalance value, namely 100 gram•inches. A given mass will create different unbalances, depending on its distance from the shaft axis. To determine the unbalance, simply multiply the mass by its radius. Since a given excess mass at a given radius represents the same unbalance, regardless of rotational speed (provided the rotor does not change its shape over speed), the speed at which the unbalance is measured is determined primarily by the type of balancing machine, its drive system, the required balancing accuracy, and safety concerns (i.e. the slower the rotational speed, the less energy is stored in the rotor). Once the unbalance has been corrected there will no longer be any significant disturbing centrifugal force and, therefore, no more excessive vibration. A small residual unbalance will usually remain in the part, just as there is a tolerance in any machining operation. Generally, the higher the service speed, the smaller should be the residual unbalance. Balancing tolerances for various types of rotors will be discussed later in this book. While most countries use the metric system, and subsequently use metric units of unbalance, e.g., gram•millimeters (abbreviated gmm), in the U.S.A. many branches of the industry use a combination of metric and English units, gram•inch (abbreviated g•in), because it has proven to be the most practical. A true English unit, e.g., ounce•inch (abbreviated oz•in) is too large for many balancing applications, necessitating fractions or a subdivision into hundredths, neither of which has become very popular.

Types of unbalance The following paragraphs explain the four different types of unbalance as defined by the internationally accepted ISO Standard No. 1925 on balancing terminology. For each of the four mutually exclusive cases an example is shown, illustrating displacement of the principal axis of inertia from the shaft axis caused by the addition of certain unbalance masses in

certain distributions to a perfectly balanced rotor. a. Static Unbalance

Static Unbalance Static unbalance exists when the principal axis of inertia is displaced parallel to the shaft axis. This type of unbalance is found primarily in narrow, disk-shaped parts such as flywheels and turbine wheels. It can be corrected by a single mass correction placed opposite the center-of-gravity in a plane perpendicular to the shaft axis, and intersecting the CG. Static unbalance, if large enough, can be detected with gravity-type balancing devices, for instance, a pair of precision ground knife edges. If the knife edges are level, the rotor will turn until the heavy spot reaches the lowest position. The use of knife edges for the detection of unbalance is very limited because of the following: • The device can only indicate the angle of unbalance, not the amount of unbalance. • The amount of unbalance can only be estimated and corrected by trial-and-error. • The accuracy is limited by the friction between knife edge and journal. Static unbalance can be measured more accurately by centrifugal means on a balancing machine than by gravitational means on knife edges or rollers. Static balancing by gravity is satisfactory only for relatively slowly revolving, disk-shaped parts or for parts that are subsequently assembled into a larger rotor, which is then balanced dynamically as an assembly. b. Couple Unbalance

Couple Unbalance Couple unbalance is that condition for which the principal axis of inertia intersects the shaft axis at the center of gravity. This condition arises when two equal unbalances are positioned at an axial distance on a rotor and spaced 180º from each other. Since this rotor will not rotate when placed on knife-edges, a dynamic method must be employed to detect couple unbalance. Couple unbalance is expressed in units of gram-millimeter2 (abbreviated gmm2), gram•inch2 (abbreviated g•in2), ounce•inch2 (abbreviated oz•in2), or similar, wherein the second length unit refers to the distance b between the two planes of unbalance. This type of unbalance cannot be corrected by a single mass in a single correction plane. At least two masses are required, each placed in a different transverse plane (perpendicular to

the shaft axis) and 180º opposite to each other. In other words, a couple unbalance needs another couple to correct it. In the example for instance, correction could be made by placing two masses at opposite angular positions on the main body of the rotor. The axial location of the correction couple does not matter as long as its value is equal in magnitude but opposite in direction to the unbalance couple. c. Quasi-Static Unbalance

Quasi-Static Unbalance Quasi-static unbalance is that condition of unbalance for which the central principal axis of inertia intersects the shaft axis at a point other than the center of gravity. It represents the specific combination of static and couple unbalance where the angular position of one couple component coincides with the angular position of the static unbalance. This is a special case of dynamic unbalance.

Couple plus Static Unbalance results in Quasi-Static Unbalance, provided one Couple Mass has the same angular position as the Static Mass.

Note that the single unbalance mass in the first figure represents the same quasi-static unbalance as the 3 masses in the second!

d. Dynamic Unbalance

ynamic unbalance, is that condition in which the central principal axis of inertia is neither parallel to, nor intersects with the shaft axis. It is the most frequently occurring type of unbalance and can only be corrected (as is the case with couple unbalance) by mass correction in at least two planes perpendicular to the shaft axis. Dynamic unbalance is a combination of static unbalance and couple unbalance, where the angular position of the static unbalance relative to the couple unbalance is neither 0º nor 180º.

Types of Balancing Machines The balancing machine is a measuring tool. A balancing machine is used to detect, locate and measure unbalance. The data furnished by the machine permits changing the mass distribution of a rotor, which, when done accurately, will balance the rotor. Balance is a zero quantity, and therefore is detected by observing an absence of unbalance. The balancing machine measures only unbalance, never balance. Soft-bearing The soft-bearing balancing machine derives its name from the fact that it supports the rotor to be balanced on bearings which are very flexibly suspended, permitting the rotor to vibrate freely in at least one direction, usually the horizontal, perpendicular to the rotor shaft axis.

Resonance of rotor and bearing systems occurs at one half or less of the lowest balancing speed, so that by the time balancing speed is reached, the angle of lag and the vibration amplitude have stabilized and can be measured with reasonable certainty. Bearings (and the directly attached support components) vibrate in unison with the rotor, thus adding to its mass. Restriction of vertical motion does not affect the amplitude of vibration in the horizontal plane, but the added mass of the bearings does. The greater the combined mass of the rotor and the bearings, the smaller will be the displacement of the bearings, and the smaller will be the output of the devices which sense the unbalance. The relationship between unbalance and bearing motion is very complex. A direct indication of unbalance can be obtained only after calibrating the indicating system for a given rotor by making several calibration runs with calibration weights of known value attached to the rotor in the chosen correction planes. Calibrating a soft-bearing machine by shaking the rotor (without spinning it) has been attempted by several manufacturers but proven inaccurate because the polar moment of inertia is ignored. Hard-bearing Hard-bearing balancing machines are essentially of the same construction as soft-bearing balancing machines, except that their bearing supports are significantly stiffer in transverse horizontal direction. This results in a horizontal resonance for the rotor and bearing support system which occurs at a frequency several orders of magnitude higher than that for a comparable soft-bearing balancing machine. The hard-bearing balancing machine is designed to operate at speeds well below this resonance in an area where the phase angle lag is constant and practically zero, and where the amplitude of vibration - though small - is directly proportional to centrifugal forces produced by unbalance. Since the force that a given amount of unbalance exerts at a given speed is always the same, no matter whether the unbalance occurs in a small or large, light or heavy rotor, the output from the sensing elements attached to the balancing machine bearing supports remains proportional to the centrifugal force resulting from unbalance in the rotor. The output is not influenced by bearing mass, rotor mass, or inertia, so that a permanent relation between unbalance and sensing element output can be established. Centrifugal force from a given unbalance rises with the square of the balancing speed. Output from the pickups rises proportionately with the second or third power of the speed depending on the type of pickup used. Suitable integrator circuitry then reduces the pickup signal inversely proportional to the square respectively cube of the balancing speed increase, resulting in a constant unbalance readout. Unlike soft-bearing balancing machines, the use of calibration masses or shakers is not required to calibrate the machine for a given rotor. Balancing and Vibration Standards

INTERNATIONAL STANDARDS • ISO 1925:2001 Mechanical vibration -- Balancing -- Vocabulary • ISO 1940-1:1986 Mechanical vibration -- Balance quality requirements of rigid rotors -- Part 1: Determination of permissible residual unbalance • ISO 1940-2:1997 Mechanical vibration -- Balance quality requirements of rigid rotors -- Part 2: Balance errors • ISO 2041:1990 Vibration and shock -- Vocabulary • ISO 2371:1974 Field balancing equipment -- Description and evaluation (withdrawn) • ISO 2953:1999 Mechanical vibration -- Balancing machines -- Description and evaluation (available in English only) • ISO 2954:1975 Mechanical vibration of rotating and reciprocating machinery -- Requirements for instruments for measuring vibration severity • ISO 3719:1994 Mechanical vibration -- Symbols for balancing machines and associated instrumentation • ISO 4866:1990 Mechanical vibration and shock -- Vibration of buildings -- Guidelines for the measurement of vibrations and evaluation of their effects on buildings • ISO 5343:1983 Criteria for evaluating flexible rotor balance (withdrawn) • ISO 5344:1980 Electrodynamic test equipment for generating vibration -- Methods of describing equipment characteristics • ISO 5348:1998 Mechanical vibration and shock -- Mechanical mounting of accelerometers • ISO 5406:1980 The mechanical balancing of flexible rotors (withdrawn) • ISO 7475:2002 Mechanical vibration -- Balancing machines -- Enclosures and other protective measures for the measuring station (available in English only) • ISO 7626-1:1986 Vibration and shock -- Experimental determination of mechanical mobility -Part 1: Basic definitions and transducers • ISO 7626-2:1990 Vibration and shock -- Experimental determination of mechanical mobility -Part 2: Measurements using single-point translation excitation with an attached vibration exciter • ISO 7626-5:1994 Vibration and shock -- Experimental determination of mechanical mobility -Part 5: Measurements using impact excitation with an exciter which is not attached to the structure • ISO 7919-1:1996 Mechanical vibration of non-reciprocating machines -- Measurements on rotating shafts and evaluation criteria -- Part 1: General guidelines • ISO 7919-2:2001 Mechanical vibration -- Evaluation of machine vibration by measurements on rotating shafts -- Part 2: Land-based steam turbines and generators in excess of 50 MW with normal operating speeds of 1500 r/min, 1800 r/min, 3000 r/min and 3600 r/min • ISO 7919-3:1996 Mechanical vibration of non-reciprocating machines -- Measurements on rotating shafts and evaluation criteria -- Part 3: Coupled industrial machines • ISO 7919-4:1996 Mechanical vibration of non-reciprocating machines -- Measurements on rotating shafts and evaluation criteria -- Part 4: Gas turbine sets • ISO 7919-5:1997 Mechanical vibration of non-reciprocating machines -- Measurements on rotating shafts and evaluation criteria -- Part 5: Machine sets in hydraulic power generating and pumping plants • ISO 8042:1988 Shock and vibration measurements -- Characteristics to be specified for seismic pick-ups • ISO 8569:1996 Mechanical vibration and shock -- Measurement and evaluation of shock and vibration effects on sensitive equipment in buildings

• ISO 8821:1989 Mechanical vibration -- Balancing -- Shaft and fitment key convention • ISO 9688:1990 Mechanical vibration and shock -- Analytical methods of assessing shock resistance of mechanical systems -- Information exchange between suppliers and users of analyses • ISO 10055:1996 Mechanical vibration -- Vibration testing requirements for shipboard equipment and machinery components • ISO 10137:1992 Bases for design of structures -- Serviceability of buildings against vibration (available in English only) • ISO/TS 10811-1:2000 Mechanical vibration and shock -- Vibration and shock in buildings with sensitive equipment -- Part 1: Measurement and evaluation • ISO/TS 10811-2:2000 Mechanical vibration and shock -- Vibration and shock in buildings with sensitive equipment -- Part 2: Classification • ISO 10814:1996 Mechanical vibration -- Susceptibility and sensitivity of machines to unbalance • ISO 10816-1:1995 Mechanical vibration -- Evaluation of machine vibration by measurements on non-rotating parts -- Part 1: General guidelines • ISO 10816-2:2001 Mechanical vibration -- Evaluation of machine vibration by measurements on non-rotating parts -- Part 2: Land-based steam turbines and generators in excess of 50 MW with normal operating speeds of 1500 r/min, 1800 r/min, 3000 r/min and 3600 r/min • ISO 10816-3:1998 Mechanical vibration -- Evaluation of machine vibration by measurements on non-rotating parts -- Part 3: Industrial machines with nominal power above 15 kW and nominal speeds between 120 r/min and 15 000 r/min when measured in situ • ISO 10816-4:1998 Mechanical vibration -- Evaluation of machine vibration by measurements on non-rotating parts -- Part 4: Gas turbine driven sets excluding aircraft derivatives • ISO 10816-5:2000 Mechanical vibration -- Evaluation of machine vibration by measurements on non-rotating parts -- Part 5: Machine sets in hydraulic power generating and pumping plants (available in English only) • ISO 10816-6:1995 Mechanical vibration -- Evaluation of machine vibration by measurements on non-rotating parts -- Part 6: Reciprocating machines with power ratings above 100 kW • ISO 10817-1:1998 Rotating shaft vibration measuring systems -- Part 1: Relative and absolute sensing of radial vibration • ISO 10819:1996 Mechanical vibration and shock -- Hand-arm vibration -- Method for the measurement and evaluation of the vibration transmissibility of gloves at the palm of the hand • ISO 11342:1998 Mechanical vibration -- Methods and criteria for the mechanical balancing of flexible rotors (available in English only) • ISO 11342/Cor1:2000 Mechanical vibration -- Methods and criteria for the mechanical balancing of flexible rotors (Technical Corrigendum 1) • ISO 13373-1:2002 Condition monitoring and diagnostics of machines -- Vibration condition monitoring -- Part 1: General procedures • ISO 14694:2003 Industrial fans -- Specifications for balance quality and vibration levels • ISO 14695:2003 Industrial fans -- Method of measurement of fan vibration • ISO 14839-1:2002 Mechanical vibration -- Vibration of rotating machinery equipped with active magnetic bearings -- Part 1: Vocabulary • ISO 16063-1:1998 Methods for the calibration of vibration and shock transducers -- Part 1: Basic concepts • ISO 16063-11:1999 Methods for the calibration of vibration and shock transducers -- Part 11: Primary vibration calibration by laser interferometry (available in English only) • ISO 16063-12:2002 Methods for the calibration of vibration and shock transducers -- Part 12: Primary vibration calibration by the reciprocity method (available in English only)

• ISO 16063-13:2001 Methods for the calibration of vibration and shock transducers -- Part 13: Primary shock calibration using laser interferometry NATIONAL STANDARDS ANSI S2.7-1982 (R1997)

Balancing Terminology

(identical to ISO 1925)

ANSI S2.60-1987 (R1997)

Balancing Machines - Enclosures and Other Safety Measures

(identical to

ANSI S2.42-1982 (R1997)

Procedures for Balancing Flexible Rotors

ANSI S2.38-1982 (R1997)

Field Balancing Equipment - Description and Evaluation

ANSI S2.19-1989 (R1997)

Mechanical Vibration - Balance Quality Requirements of Rigid Rotors - Part 1, Determination of Permissible Residual Unbalance (identical to ISO 1940)

ISO 7475) (identical to ISO 5406) (identical to ISO

2371)

SAE Documents ARP587B : Balancing Machines - Description and Evaluation Horizontal, Two-Plane, SoftBearing Type for Gas Turbine Rotors ARP588B : Balancing Machines - Description and Evaluation Vertical, Single-Plane, SoftBearing Type for Gas Turbine Rotors ARP1134 : Adapter Interface - Turbine Engine Blade Moment Weighing Scale ARP1202 : Balancing Machines, Dynamic, Ball Type Slave Bearings for Rotor Support ARP1382 : Design Criteria for Balancing Machine Tooling ARP4048 : Balancing Machines - Description and Evaluation Horizontal, Two-Plane, HardBearing Type for Gas Turbine Rotors ARP4050 : Balancing Machines - Description and Evaluation Vertical, Two-Plane, HardBearing Type for Gas Turbine Rotors ARP4162A Balancing Machine Proving Rotors : ARP4163 : Balancing Machines, Tooling Design Criteria (as of 7-2003 being worked on, will supersede ARP 1382) ARP5323 : Balancing Machines - Description and Evaluation Vertical, Single-Plane, HardBearing Type for Gas Turbine Rotors ARP510A : Moment Weight of Turbine and Compressor Rotor Blades AIR1839 : A Guide to Aircraft Turbine Engine Vibration Monitoring Systems ANSI and ISO Documents may be ordered through www.ansi.org API Standards may be ordered through a distributor, Global Engineering Documents at http://global.ihs.com

SAE Standards may be ordered through www.sae.org