History of BalancingThe 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 . students will be divided into small groups with those that have similar equipment and applications. • Industry standard tolerances vs. dynamic & couple).* Other topics include: • The different types of unbalance (static. drawing tolerances. tolerance specifications and machine setup. please specify balancer make. Service technicians and engineers will then supervise a number of exercises on machines and instruments that closely resemble the students' equipment. A mechanical aptitude with emphasis on rotating equipment and precision measurements (technician level) or machine shop experience is recommended. • Proper machine maintenance and troubleshooting.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. • Machine & rotor setup. Hands-on exercises will be used extensively to improve the operators technique on both balancing machine and instrumentation. Certification: Level I. • How to avoid interference of measurements due to drive and roller harmonics. *Since instruments and machines will be selected to closely represent the attendees' equipment. • Maximizing instrumentation features. • ISO tolerances and terminology. model and instrument with application. • Instrumentation functions & operation. After a brief overview of basic theory that includes the principles of machine operation. • Selecting the best balancing speed.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. Unbalance vs. Performance is decreased because of the absorption of energy by the supporting structure. Minimize power loss Unbalance in just one rotating component of an assembly may cause the entire assembly to vibrate. substantially reducing their service life. shafts. spindles. therefore. Minimize vibration c.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. which may eventually lead to their complete failure. etc. Vibrations set up highly undesirable alternating stresses in structural supports and housings. Increase bearing life g. Vibrations may be transmitted through the floor to adjacent machinery and seriously impair its accuracy or proper functioning. centrifugal force . Increase quality of product b. Minimize audible and signal noises d.” Why Balance? An unbalanced rotor will cause vibration and stress in the rotor itself and in its supporting structure. Minimize operator annoyance and fatigue f. Minimize structural stresses e. bushings.. gears. Balancing of the rotor is. necessary to accomplish one or more of the following: a. This induced vibration in turn may cause excessive wear in bearings. 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. however. impelling each particle outward and away from the axis of rotation in a radial direction.. Unbalance. therefore. if the speed is tripled. if an excess of mass exists on one side of a rotor. The higher the speed. Centrifugal force. When at rest. therefore. causes no vibration.e. If the mass of a rotating component is evenly distributed about its shaft axis. the centrifugal force quadruples. This figure shows the side view of a rotor having an excess mass m on one side. is independent of rotational speed and remains the same. i. the excess mass exerts no centrifugal force and. the centrifugal force is multiplied by nine. Yet the actual unbalance is still present. If the speed is doubled. will vibrate due to the excess centrifugal force exerted during rotation by the heavier side of the rotor.Centrifugal force acts upon the entire mass of a rotating component. Centrifugal force increases with the square of the speed A rotating element having an uneven mass distribution. whether the part is at rest or is rotating (provided the part does not deform during rotation). the entire rotor is being pulled in the direction of the arrow F. the part is "balanced" and rotates without vibration. However. Due to centrifugal force exerted by m during rotation. Causes of unbalance . 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. varies with speed. unbalance. permit radial displacement of parts of the assembly and thereby introduce unbalance. inclusions. and finishes. these masses can often be counterbalanced by designing for symmetry about the shaft axis. Unbalance may also occur due to lack of mass (such as a drill hole. Ideally. are sources of unbalance. Tolerances in fabrication. Symmetrical design and careful setting of tolerances and fits can often minimize balancing problems. which cannot be made concentric and symmetrical with respect to the shaft axis. necessary for economical manufacturing and assembly of several elements of a rotor. dimensional changes. including distortion. additional machining cost is involved and part strength may be affected. by the . including casting. Unmachined portions of castings or forgings. which permit any eccentricity or lack of squareness with respect to the shaft axis. Nonsymmetrical in use. If corrections are made by addition of material. grain. porous spot. including the following: a. The tolerances. often distort nonsymmetrically under service conditions. whether a balancing operation is to be performed or not. porosity. Variation within materials. and temperature changes. In the example illustrated. Either one may be caused by a variety of reasons. introduce substantial unbalance. In parts that require unbalanced masses for functional reasons. Large amounts of unbalance require large corrections. such as fans. aerodynamic forces. If such corrections are made by removal of material. machining. Fabricated parts. part shapes. Where low service speeds are involved and the effects of a reasonable amount of unbalance can be tolerated. and assembly. and shifting of parts due to rotational stresses. 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. c. location. it is the "heavy spot". Correction methods Corrections for rotor unbalance are made either by the addition of mass to the rotor. etc. rotating parts should always be designed for inherent balance. For example. this practice may eliminate the need for balancing. including motor windings. density. b. Manufacturing processes are the major source of unbalance. It is impractical to design a compensating unbalance into the armature. such as voids. d. Design and economic considerations prevent the adaptation of methods that might eliminate this distortion and thereby reduce the resulting unbalance. Nonsymmetrical design. and density of finishes. Limitations imposed by rotor design often introduce unbalance effects that cannot be corrected adequately by refinement of the design itself. Manufacturing tolerances and processes.) in which case it is called the "light spot”. cost is again a factor and space requirements for the added material may be a problem.The excess of mass on one side of the rotor in this figure is called unbalance. which is removed. In general. used where relatively large corrections are required. e. by relocating the shaft axis (“mass centering"). Welding provides a means of attaching a wide variety of correction masses at any desired angular locations. 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. This method is often used in balancing of wound armatures. provided the mass and its position are closely controlled. Addition of premanufactured weights. Milling. or Fly Cutting.. or in some cases. This is probably the most effective method of unbalance correction. 3. The ideal correction method permits a reduction of the maximum initial unbalance to less than balance tolerance in a single correction step. the mass of one washer may vary considerably from the mass of the next washer of the same type and size. thereby removing the intended amount of material with a high degree of accuracy. Mass Centering . 4. The selected correction method should ensure that there is sufficient space or material to allow correction of the maximum unbalance which may occur. Drilling. usually permit a single step reduction of 10:1 in unbalance if carried out carefully.. this is often difficult to achieve. Shaping. A typical application is addition of spring clips to the blades of automotive A/C blower wheels. This method is usually used only where the rotor design or material does not permit a more economical type of correction. Grinding.g. Material is removed from the rotor by a drill which penetrates the rotor to a measured depth. Then a second correction follows to reduce the remaining unbalance to its permissible value. by resistance welding the weights to the outside rotor surface. A depth gage or limit switch can be provided on the drill spindle to ensure that the hole is drilled to the desired depth. a preliminary correction is made. Removal of Mass 1. 2. i. However. The addition of mass may achieve a reduction ratio as large as 20:1 or higher. Addition of cut-to-size weights. Addition of Mass 1. The same limitations as in (2) apply. Care must be taken that welding heat does not distort the rotor. Addition of bolted or riveted standard washers. Variations in location introduce errors in correction. Addition of two-component epoxy. grinding must be considered a trial-and-error method of correction. drilling. Milling and grinding are less accurate. 2. The more common methods described below. and when means are provided for accurate measurement of cut with respect to those surfaces.e. 3. This method is often used in balancing of AC motor rotors.removal of material. This method is quick. It is difficult to evaluate the actual mass of the material. This is practiced on drive shafts. but somewhat limited in accuracy because the washers come in incremental sizes. This method permits accurate removal of mass when the rotor surfaces. from which the depth of cut is measured. It is difficult to apply the material so that its center-of-gravity is precisely at the desired correction location. for instance. unless carried out in automatic or semi-automatic balancing machines. which have integrated mass correction devices. are machined surfaces. Center drills. A final balancing operation is. all having a similar meaning. the machining operation will introduce some new unbalance. 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. indicates that one side of the rotor has an . which in turn. namely to compensate for the opposed throws and crankpins of the crankshaft. namely a mass multiplied by its distance from the shaft axis. Units of unbalance Unbalance is measured in ounce•inches. i. to reduce initial unbalance in crankshaft castings or forgings.e. is rotated in a balancing machine. for instance. for example. Because material removal is uneven at different parts of the shaft. Furthermore. nor does it reduce the mass of the counterweights to a level where they no longer perform their proper function. still required. At this point the principal inertia axis of the shaft essentially coincides with the shaft axis of the balanced cage. An unbalance of 100 g•in. It is generally accomplished by drilling into the crankshaft counterweights. final correction (usually by drilling) does not exceed the material available for it. or gram•millimeters. then drill the shaft centers and thereby provide an axis in the crankshaft about which it is in balance. The shaft is mounted in a balanced cage or cradle.Such a procedure is used. guided along the axis of the cage. final unbalance corrections are small and balancing time is significantly shortened. gram•inches. therefore. However.. The subsequent machining of the crankshaft is carried out between these centers. its "radius". . the less energy is stored in the rotor). Balancing tolerances for various types of rotors will be discussed later in this book. gram•millimeters (abbreviated gmm). illustrating displacement of the principal axis of inertia from the shaft axis caused by the addition of certain unbalance masses in . 1925 on balancing terminology. View of Rotor With 100 g•in Unbalance In each case the mass.g. the smaller should be the residual unbalance. the required balancing accuracy. and safety concerns (i. simply multiply the mass by its radius. no more excessive vibration. regardless of rotational speed (provided the rotor does not change its shape over speed). therefore.g. the slower the rotational speed. e. For each of the four mutually exclusive cases an example is shown. Since a given excess mass at a given radius represents the same unbalance. ounce•inch (abbreviated oz•in) is too large for many balancing applications. gram•inch (abbreviated g•in). Types of unbalance The following paragraphs explain the four different types of unbalance as defined by the internationally accepted ISO Standard No. Generally. or 20 grams at a 5 inch radius.e. the speed at which the unbalance is measured is determined primarily by the type of balancing machine. While most countries use the metric system. e. A small residual unbalance will usually remain in the part. many branches of the industry use a combination of metric and English units. necessitating fractions or a subdivision into hundredths. its drive system. and subsequently use metric units of unbalance. To determine the unbalance. Once the unbalance has been corrected there will no longer be any significant disturbing centrifugal force and. just as there is a tolerance in any machining operation. namely 100 gram•inches.. when multiplied by its distance from the shaft axis.A. depending on its distance from the shaft axis. amounts to the same unbalance value. the higher the service speed.S. because it has proven to be the most practical. A true English unit. in the U. neither of which has become very popular. A given mass will create different unbalances.excess mass equivalent to 10 grams at a 10 inch radius. wherein the second length unit refers to the distance b between the two planes of unbalance. Couple unbalance is expressed in units of gram-millimeter2 (abbreviated gmm2). disk-shaped parts or for parts that are subsequently assembled into a larger rotor.certain distributions to a perfectly balanced rotor. and intersecting the CG. if large enough. Since this rotor will not rotate when placed on knife-edges. • The accuracy is limited by the friction between knife edge and journal. or similar. Static unbalance can be measured more accurately by centrifugal means on a balancing machine than by gravitational means on knife edges or rollers. This type of unbalance is found primarily in narrow. for instance. not the amount of unbalance. If the knife edges are level. At least two masses are required. gram•inch2 (abbreviated g•in2). a. b. the rotor will turn until the heavy spot reaches the lowest position. can be detected with gravity-type balancing devices. Static unbalance. a pair of precision ground knife edges. ounce•inch2 (abbreviated oz•in2). Static balancing by gravity is satisfactory only for relatively slowly revolving. 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. a dynamic method must be employed to detect couple unbalance. each placed in a different transverse plane (perpendicular to . 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. • The amount of unbalance can only be estimated and corrected by trial-and-error. Static Unbalance Static Unbalance Static unbalance exists when the principal axis of inertia is displaced parallel to the shaft axis. disk-shaped parts such as flywheels and turbine wheels. which is then balanced dynamically as an assembly. It can be corrected by a single mass correction placed opposite the center-of-gravity in a plane perpendicular to the shaft axis. This type of unbalance cannot be corrected by a single mass in a single correction plane. This condition arises when two equal unbalances are positioned at an axial distance on a rotor and spaced 180º from each other. Couple plus Static Unbalance results in Quasi-Static Unbalance. This is a special case of dynamic unbalance. 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. c. Dynamic Unbalance . 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. In the example for instance. In other words. 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. correction could be made by placing two masses at opposite angular positions on the main body of the rotor. a couple unbalance needs another couple to correct it. 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.the shaft axis) and 180º opposite to each other. . 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. The data furnished by the machine permits changing the mass distribution of a rotor. perpendicular to the rotor shaft axis. when done accurately. A balancing machine is used to detect.ynamic unbalance. where the angular position of the static unbalance relative to the couple unbalance is neither 0º nor 180º. permitting the rotor to vibrate freely in at least one direction. never balance. 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. locate and measure unbalance. The balancing machine measures only unbalance. will balance the rotor. Dynamic unbalance is a combination of static unbalance and couple unbalance. Types of Balancing Machines The balancing machine is a measuring tool. Balance is a zero quantity. is that condition in which the central principal axis of inertia is neither parallel to. usually the horizontal. and therefore is detected by observing an absence of unbalance. nor intersects with the shaft axis. which. Resonance of rotor and bearing systems occurs at one half or less of the lowest balancing speed. rotor mass. Since the force that a given amount of unbalance exerts at a given speed is always the same. Balancing and Vibration Standards .though small . Hard-bearing Hard-bearing balancing machines are essentially of the same construction as soft-bearing balancing machines. the smaller will be the displacement of the bearings. Output from the pickups rises proportionately with the second or third power of the speed depending on the type of pickup used. The relationship between unbalance and bearing motion is very complex. Bearings (and the directly attached support components) vibrate in unison with the rotor. 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. thus adding to its mass. The greater the combined mass of the rotor and the bearings. 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. The output is not influenced by bearing mass. Suitable integrator circuitry then reduces the pickup signal inversely proportional to the square respectively cube of the balancing speed increase. 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. 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. Centrifugal force from a given unbalance rises with the square of the balancing speed. and the smaller will be the output of the devices which sense the unbalance. so that by the time balancing speed is reached. Restriction of vertical motion does not affect the amplitude of vibration in the horizontal plane. Unlike soft-bearing balancing machines. no matter whether the unbalance occurs in a small or large. so that a permanent relation between unbalance and sensing element output can be established. but the added mass of the bearings does. except that their bearing supports are significantly stiffer in transverse horizontal direction. the use of calibration masses or shakers is not required to calibrate the machine for a given rotor. and where the amplitude of vibration . resulting in a constant unbalance readout.is directly proportional to centrifugal forces produced by unbalance. or inertia. the angle of lag and the vibration amplitude have stabilized and can be measured with reasonable certainty. light or heavy rotor. 1800 r/min.Part 2: Land-based steam turbines and generators in excess of 50 MW with normal operating speeds of 1500 r/min.Balance quality requirements of rigid rotors -.Requirements for instruments for measuring vibration severity • ISO 3719:1994 Mechanical vibration -.Measurements on rotating shafts and evaluation criteria -.Balance quality requirements of rigid rotors -.Vocabulary • ISO 1940-1:1986 Mechanical vibration -.Mechanical mounting of accelerometers • ISO 5406:1980 The mechanical balancing of flexible rotors (withdrawn) • ISO 7475:2002 Mechanical vibration -.INTERNATIONAL STANDARDS • ISO 1925:2001 Mechanical vibration -.Characteristics to be specified for seismic pick-ups • ISO 8569:1996 Mechanical vibration and shock -.Measurements on rotating shafts and evaluation criteria -.Vibration of buildings -.Part 1: General guidelines • ISO 7919-2:2001 Mechanical vibration -.Balancing -.Part 3: Coupled industrial machines • ISO 7919-4:1996 Mechanical vibration of non-reciprocating machines -.Vocabulary • ISO 2371:1974 Field balancing equipment -.Balancing machines -. 3000 r/min and 3600 r/min • ISO 7919-3:1996 Mechanical vibration of non-reciprocating machines -.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 -.Symbols for balancing machines and associated instrumentation • ISO 4866:1990 Mechanical vibration and shock -.Part 4: Gas turbine sets • ISO 7919-5:1997 Mechanical vibration of non-reciprocating 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 -.Measurement and evaluation of shock and vibration effects on sensitive equipment in buildings .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 -.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 -.Evaluation of machine vibration by measurements on rotating shafts -.Description and evaluation (withdrawn) • ISO 2953:1999 Mechanical vibration -.Methods of describing equipment characteristics • ISO 5348:1998 Mechanical vibration and shock -.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 -.Description and evaluation (available in English only) • ISO 2954:1975 Mechanical vibration of rotating and reciprocating machinery -.Part 1: Determination of permissible residual unbalance • ISO 1940-2:1997 Mechanical vibration -.Part 2: Balance errors • ISO 2041:1990 Vibration and shock -.Balancing machines -. Part 4: Gas turbine driven sets excluding aircraft derivatives • ISO 10816-5:2000 Mechanical vibration -.Part 1: Measurement and evaluation • ISO/TS 10811-2:2000 Mechanical vibration and shock -.• ISO 8821:1989 Mechanical vibration -.Vibration and shock in buildings with sensitive equipment -.Evaluation of machine vibration by measurements on non-rotating parts -.Methods and criteria for the mechanical balancing of flexible rotors (available in English only) • ISO 11342/Cor1:2000 Mechanical vibration -.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 -. 3000 r/min and 3600 r/min • ISO 10816-3:1998 Mechanical vibration -.Part 12: Primary vibration calibration by the reciprocity method (available in English only) .Evaluation of machine vibration by measurements on non-rotating parts -.Vibration of rotating machinery equipped with active magnetic bearings -.Method for the measurement and evaluation of the vibration transmissibility of gloves at the palm of the hand • ISO 11342:1998 Mechanical vibration -.Method of measurement of fan vibration • ISO 14839-1:2002 Mechanical vibration -.Serviceability of buildings against vibration (available in English only) • ISO/TS 10811-1:2000 Mechanical vibration and shock -.Part 6: Reciprocating machines with power ratings above 100 kW • ISO 10817-1:1998 Rotating shaft vibration measuring systems -.Part 2: Classification • ISO 10814:1996 Mechanical vibration -.Part 1: Relative and absolute sensing of radial vibration • ISO 10819:1996 Mechanical vibration and shock -.Evaluation of machine vibration by measurements on non-rotating parts -. 1800 r/min.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 -.Balancing -.Part 5: Machine sets in hydraulic power generating and pumping plants (available in English only) • ISO 10816-6:1995 Mechanical vibration -.Vibration and shock in buildings with sensitive equipment -.Part 1: General procedures • ISO 14694:2003 Industrial fans -.Methods and criteria for the mechanical balancing of flexible rotors (Technical Corrigendum 1) • ISO 13373-1:2002 Condition monitoring and diagnostics of machines -.Hand-arm vibration -.Analytical methods of assessing shock resistance of mechanical systems -.Evaluation of machine vibration by measurements on non-rotating parts -.Part 1: Vocabulary • ISO 16063-1:1998 Methods for the calibration of vibration and shock transducers -.Vibration testing requirements for shipboard equipment and machinery components • ISO 10137:1992 Bases for design of structures -.Shaft and fitment key convention • ISO 9688:1990 Mechanical vibration and shock -.Vibration condition monitoring -.Part 2: Land-based steam turbines and generators in excess of 50 MW with normal operating speeds of 1500 r/min.Evaluation of machine vibration by measurements on non-rotating parts -.Specifications for balance quality and vibration levels • ISO 14695:2003 Industrial fans -.Information exchange between suppliers and users of analyses • ISO 10055:1996 Mechanical vibration -.Evaluation of machine vibration by measurements on non-rotating parts -.Part 1: General guidelines • ISO 10816-2:2001 Mechanical vibration -.Susceptibility and sensitivity of machines to unbalance • ISO 10816-1:1995 Mechanical vibration -.Part 1: Basic concepts • ISO 16063-11:1999 Methods for the calibration of vibration and shock transducers -. Part 1. Two-Plane.org API Standards may be ordered through a distributor. HardBearing Type for Gas Turbine Rotors ARP4162A Balancing Machine Proving Rotors : ARP4163 : Balancing Machines.Turbine Engine Blade Moment Weighing Scale ARP1202 : Balancing Machines. HardBearing Type for Gas Turbine Rotors ARP4050 : Balancing Machines .19-1989 (R1997) (identical to ISO 1925) (identical to Balancing Terminology Balancing Machines . SoftBearing Type for Gas Turbine Rotors ARP1134 : Adapter Interface .Description and Evaluation Horizontal.42-1982 (R1997) ANSI S2. Two-Plane.ihs. will supersede ARP 1382) ARP5323 : Balancing Machines .38-1982 (R1997) ANSI S2.ansi. SoftBearing Type for Gas Turbine Rotors ARP588B : Balancing Machines .Description and Evaluation Vertical. Ball Type Slave Bearings for Rotor Support ARP1382 : Design Criteria for Balancing Machine Tooling ARP4048 : Balancing Machines .Enclosures and Other Safety Measures ISO 7475) Procedures for Balancing Flexible Rotors (identical to ISO 5406) (identical to ISO Field Balancing Equipment .com SAE Standards may be ordered through www.Balance Quality Requirements of Rigid Rotors . Global Engineering Documents at http://global.Description and Evaluation Horizontal.sae. Determination of Permissible Residual Unbalance (identical to ISO 1940) SAE Documents ARP587B : Balancing Machines . Dynamic.org .60-1987 (R1997) ANSI S2.• 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.Description and Evaluation 2371) Mechanical Vibration . Tooling Design Criteria (as of 7-2003 being worked on.Description and Evaluation Vertical.7-1982 (R1997) ANSI S2.Description and Evaluation Vertical. Single-Plane. Single-Plane. Two-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.