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Balancing presentation
 

Balancing presentation

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From MUSTAFA KAMAL PASHA

From MUSTAFA KAMAL PASHA
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    Balancing presentation Balancing presentation Presentation Transcript

    • Dynamics of Rotating machinery with Emphasis on Balancing Technical Services
    • Dynamics of Rotating machinery with Case studies – Balancing fundamentals – Critical Speeds and vibratory Modes- How to identify and understand its significance. – Slow Roll and Bow shaft -Rotor dynamic perspective. – Damping – Bearings and support structures – Foundations – Case Studies
    • F=m*r*ω2 A disk with a mass M having an Unbalance weight m at a position r from its center. This unbalance causes an eccentric center of gravity e and results in a centrifugal force P when the disk is rotated at an angular speed ω Centrifugal Force F=mrω2 ω is angular velocity = 2πn/60 ; P is in Newtons The centrifugal force P changes its direction as the rotor rotates, which repeatedly acts on the bearing portion and so causes vibration of the whole machine.
    • Rotor Unbalance should not represented by Centrifugal force F since F changes as speed changes Unbalance U is represented by U=mr m : mass of unbalance r : radius of unbalance Dimension of unbalance g ・ mm Quality of Rotor Balance : ratio of unbalance U to rotor mass M e=U/M=mr/M Here e is a vector having a dimension of length which is given as μm where m is expressed in [g], r in [mm] and M in [kg] e is (eccentricity) of the center of gravity of the rotor. Expression of unbalance
    • Single Plane balancing applied to thin disc shaped rotors Static unbalance
    • Unbalances U1,U2 and U3 distributed on a rotor which is long in the axial direction can be substituted by two independent Unbalance vectors Ua and Ub on correction planes A and B respectively Dynamic Unbalance
    • Balancing -First order mode is carried out on three correction planes Balancing -Second order mode is carried out on four correction planes Rotors become flexible when speed is increased The boundary speed which separates the rigid rotor and the flexible rotor is called the critical speed. The number of additional correction planes necessary for eliminating deformation of a rotor is the same as the order of the critical speed. Three correction planes eliminating rotor deformation up to first-order critical speed Four correction planes eliminating deformation up to second-order critical speed. Multi Plane balancing of flexible rotors
    • Accuracy of balancing Balancing to the achievable limit is uneconomical Specific unbalance (e [μm]) expresses the unbalance state of a rotor independently of its mass and shape. Value of e is in inverse proportion to the maximum working revolution speed N [min-1] of the rotor, which means that eN is a constant value. (ISO) defines the product of specific unbalance and revolution speed as the balance quality. The balance quality has a dimension of [mm/s] because the dimensions of revolution speed and specific unbalance are [rad/s] and [mm] respectively. The grade of the balance quality is expressed by putting a letter G before a number which represents eN.
    • Procedure of determining allowable unbalance Rotor speed N , Mass of the rotor m Position of rotor bearings Position of correction planes Set the grade of balance quality according to the type of the rotor. Find allowable residual specific unbalance eper from rotor speed Use equation or from diagram Balance Quality = e*w Calculate the allowable residual unbalance from the allowable residual specific unbalance and mass of the rotor: Allowable residual unbalance Uper = E per * M(g ・ mm) Allocate the allowable residual unbalance to unbalances on each actual correction plane.
    • G6.3 6.3 ●Machines for processing plants ● Turbine blades for main engines of merchant ships ●Drums for centrifugal separators ●Paper-making rolls and printing rolls ●Fans ●Completed gas turbine rotors for airplanes ●Flywheels ●Impellers of pumps ●Parts of machine tools and general machinery ●Medium- and large-sized armatures having no specific requirements for electric motors with axial center height of 80mm or more ●Small-sized armatures (mainly mass-production type) either for use being insensitive to vibration or for use with insulation against vibration ●Engine parts having specific requirements G2.5 2.5 ●Gas turbines, steam turbines and main engine turbines for merchant ships ●Rigid rotors for turbo generators ●Storage drums and disk turbo compressors for computers ●Main spindles for machine tools ●Medium- and large-sized armatures having specific requirements ●Small-sized armatures (excluding those defined in G6.3 and G1) ●Turbine-driven pumps,
    • Excessive Bearing Clearance Bent Shaft Misalignment or other Preload Electrical Influence Compliant Support or Foundation Soft Foot Mechanisms resulting in Syncronous 1X vibration other than unbalance
    • CROSS SECTIONAL ARRANGEMENT –TURBINE
    • VIBRATION MEASURING TRANSDUCERS SHAFT VIBRATION - PROXIMITY PROBE BEARING VIBRATION -VELOCITY PICK UP , ACCELEROMETER PHASE -OPTICAL PROBE , EDDY CURRENT PROBE
    • RECOMMENDED LOCATIONS OF VIBRATION MEASUREMENTS FOR PEDESTAL BEARINGS (AS PER ISO)
    • RECOMMENDED LOCATIONS OF VIBRATION MEASUREMENTS FOR HOUSING TYPE BEARINGS (AS PER ISO)
    • Measuring Amplifier 45O 45O Proximity Pick-up L RSHAFT RECOMMENDED LOCATIONS OF SHAFT VIBRATION MEASUREMENTS AS PER ISO
    • PROXIMITY PROBE & ACCELEROMETER
    • Natural Frequency The frequency of free vibration of a system. The frequency at which an undamped system with a single degree of freedom will oscillate upon momentary displacement from its rest position. Resonance Resonance is the condition which occurs when such forcing frequencies do in fact coincide with one or more natural frequencies. These may be a natural frequencies of the rotor, but often can be a natural frequency of the support frame, foundation . Forcing frequencies include those from sources such as unbalance, misalignment, looseness, bearing defects, gear defects, belt wear, etc. Critical speed Critical speeds are a special case of resonance in which the vibrating forces are caused by the rotation of the rotor
    • 0 10 20 30 40 50 60 70 80 90 0 300 600 900 1200 1500 1800 2100 2400 2700 3000 3300 RPM 0 60 120 180 240 300 360 0 300 600 900 1200 1500 1800 2100 2400 2700 3000 3300 I ,115 MW Generator Front Vertical Coast up , Before Balancing Micronspk-pkPhasedegrees
    • 0 10 20 30 40 50 60 70 80 0 300 600 900 1200 1500 1800 2100 2400 2700 3000 3300 RPM 0 60 120 180 240 300 360 0 300 600 900 1200 1500 1800 2100 2400 2700 3000 3300 115 MW Generator Rear Vertical Coast up , Before Balancing Micronspk-pkPhasedegrees
    • ROTOR AND BALANCE FORCE DETAILS GENERATOR ROTOR WEIGHT : 37000 KG GENERATOR ROTOR STATIC WEIGHT PER BEARING : 18500 KG BALANCING RADIUS FAN PLANE : 310 MM RETAINING RING PLANE : 460 MM DISTANCE BETWEEN RETAINING RING PLANE : 4850 MM DISTANCE BETWEEN FAN PLANE : 5740 MM APPROXIMATE WEIGHT OF TRIAL WEIGHT : 93 GRAM CENTRIFUGAL FORCE FOR 93 GRAMS AT BALANCE RADIUS AT 3000 RPM , FORCE UNITS : 430 KG RETAINING RING PLANE
    • 0 10 20 30 40 50 60 70 80 0 300 600 900 1200 1500 1800 2100 2400 2700 3000 3300 RPM 0 60 120 180 240 300 360 0 300 600 900 1200 1500 1800 2100 2400 2700 3000 3300 115 MW Generator Rear Vertical Coast up with 5x93 grams Couple correction weights Micronspk-pkPhasedegrees
    • GENERATOR FAN BLADES-BALANCING PLANE
    • Dynamics of Rotating machinery • Critical Speeds are dependent upon: – Rotor Flexibility - Mass and Stiffness ( D-dia of rotor, L- Bearing Span) – Support Stiffness which also includes the foundation stiffness. – The damping from the bearings dictates the amplification factor
    • To Summarize on critical speeds • It is always due to synchronous excitation. • Critical speeds in horizontal and vertical direction called as horizontal and vertical Modes depend on stiffness in those directions. • Horizontal mode is predominantly effects the vibration in horizontal direction and so in case of vertical mode. Can be measured by seismic in that direction only. • Since we also measure shaft vibrations at 45 deg so it is measuring both.
    • Let us understand the vibratory modes. • The modes below the first flexural critical speed are called as rigid modes. • Rigid modes are bouncing or translatory have same phase on both bearings while in conical modes the phase is 180 deg. • In bending modes also the phase relationship in first and second modes is similar. • We need to study the phase angle vis a vis the design critical speed in overhung modes.
    • Rocking mode Conical mode First bending mode Second bending mode Overhungcantilever bending mode Rocking mode Conical mode First bending mode Rocking mode Conical mode Second bending mode First bending mode Rocking mode Conical mode Overhungcantilever bending mode Second bending mode First bending mode Rocking mode Conical mode
    • A typical shaft bow Bode’s Plot of a 120 MW generator.
    • Turbo machinery damping • Viscous damping – Proportional to velocity Bearings and Oil seals of large rotating machinery damping provided by lubricating oil Rotor system process fluids Pumps significant Gas turbines, Centrifugal compressors – insignificant
    • • Coulomb damping Sliding friction – rub Coulomb friction force is constant , depends on 1. Nature of sliding surfaces and 2. Perpendicular pressure between surfaces Turbo machinery damping
    • • Structural damping Internal friction in material due to vibratory stress and strain Proportional to maximum stress and therefore deflections Independent of frequency – vibratory stress Rotating machinery small compared to viscous damping Turbo machinery damping
    • Hydrodynamic bearingsHydrodynamic bearings •One of the basic purposes of a bearing is to provide a frictionless environment to support and guide a rotating shaft. •Industrial machinery with high horsepower and high loads, such as steam turbines, centrifugal compressors, pumps and motors, utilize journal bearings as rotor supports.
    • TO Develop Hydro Dynamic Pressures the following three parameters are required : 1) Load, 2) Speed and 3) Oil Wedge •Hydrodynamic principles, which are active as the shaft rotates, create an oil wedge that supports the shaft and relocates it within the bearing clearances. • Hydrodynamic bearings have relatively a low frictional resistance to turning but more importantly provide viscous damping to reduce lateral vibrations.
    • All heavy industrial turbo-machines use fluid film journal bearings of some type : • To support the shaft weight • To control the motions caused by I) unbalanced forces II) aerodynamic forces III) external excitations from seals and couplings.
    • • The damping is very important in many types of rotating machines where the fluid film bearings are often the primary source of the energy absorption needed to control vibrations. • Fluid film journal bearings also play a major role in determining rotor dynamic stability, making their careful selection and application a crucial step in the development of superior rotor-bearings systems.
    • Journal bearings have many differing designs to compensate for differing load requirements, machine speeds, cost, or dynamic properties. •Cylindrical Journal Bearings with & without oil rings . • Multi lobe Journal Bearings: 2 Lobe , 2 Lobe with loading arc, 2 Lobe Offset & 4 Lobe type • Tilting Pad Journal Bearings 4 Pad and 5 Pad type
    • CAPACITY OF HYDRODYNAMIC BEARINGSCAPACITY OF HYDRODYNAMIC BEARINGS Under operation, the capacity of hydrodynamic bearings is restricted by: • Minimum oil film thickness & • Babbitt temperature. • The critical limit for low-speed operation is minimum oil film thickness. In high-speed operation, babbitt temperature is usually the limiting criteria.
    • FLUID FILM JOURNAL BEARINGS SLOW SPEED HIGH SPEED RING LUBRICATED BEARINGS PRESSURE FED BEARINGS RADIAL LOADS RADIAL AND THRUST LOADS MULTI LOBE BEAINGS TILTING PAD BEARINGS CYLINDRICAL 2- LOBE 3- LOBE 4- LOBE 4- PAD 5-PAD VERTICAL ELLIPTICITY HORIZONTAL ELLIPTICITY SYMME- TRICAL 4- LOBE TILTED 4- LOBE
    • Fig.1. Limit for Satisfactory Bearing Operation under Hydrodynamic Condition.
    • Pressure Distribution in a Journal Bearing
    • Oil Ring Bearing
    • Different Oil Ring Designs
    • Cross Sectional View of Ring Lubricated Journal Bearing RING LUBRICATED BEARINGS
    • Cylindrical and Multi-Lobe Journal Bearings Pressure Fed Bearings
    • Fluid film Thrust bearings 1. Supports Axial Forces Constant thrust loads Differential pressure across wheels (Turbines and Compressors) Gears – Axial force components Dynamic axial loads Bent rotors , Misaligned shafts 2. Maintains rotor in fixed axial position with respect to Casing Axial clearances between Blade rows determine Turbine efficiency Wheels and diaphragms in Compressors
    • Thrust bearing assembly should fulfill requirement for Axial position Axial float Axial location – axial position shims behind active thrust shoes Axial float – Total thrust float shims behind inactive thrust shoes Motors and Generators no thrust bearing Magnetic forces across air gap center the rotor within the stator Fluid film Thrust bearings Dynamics of Rotating machinery
    • Active pads Inactive pads Shims for axial position Shims for thrust float Thrust Float Stationary Casing Thrust probe Journal bearing shaft Normal Thrust Thrust bearing ---Centrifugal compressors, small turbines
    • Active pads Inactive pads Shims for axial position Shims for thrust float Thrust Float Thrust probe Thrust cum Journal bearing shaft Normal Thrust Thrust bearing ---Large Steam and Gas turbines Active thrust collar Inactive thrust collar
    • Steam Turbine Design Philosophy KWU Design Russian Design
    • Run-out Diagram , Rotation angle of Shaft alignment and Bearing Height Correction (Before Initial Correction for BKTPP unit 2 ) 405 3495 485 5825 475 1100 3350 3900 6310 1575 7810 θ HP = θ IP – 0.085/870 θ IP = 0.205 / 740 ψ = h3/6310 θ Gen = 0.158 / 760 h2= hH1 + 405 * θHP h3 = 485* θIP h5=1100*θGen h1= h2+3350* θHP hH1=(485+3495)*θIP h6=(1100+7810)*θGen Alignment correction of case 1 h1=h1-(6310+3900+3350)*ψ h3=h3-6310* ψ = 0 h5=h5-1575*ψ h2=h2-(6310+3900)*ψ h4=0 h6=h6-(1575+7810)*ψ hG θ Gen θIP h3 θHP h2 h1 hH1 V 0.085 D870 Λ 0.205 D740 h5 V 0.158 D760 ψ = Rotation angle of Shaft Alignment h1 h2 h5 h6
    • Run-out Diagram , Rotation angle of Shaft alignment and Bearing Height Correction (Before Secondary Correction for BKTPP unit 2 ) 405 3495 485 5825 475 1100 3350 3900 6310 1575 7810 θ HP = θ IP + 0.03 / 870 θ IP = 0.04 / 740 ψ = h3/6310 θ Gen = 0.158 / 760 h2= hH1 + 405 * θHP h3 = 485* θIP h5=1100*θGen h1= h2+3350* θHP hH1=(485+3495)*θIP h6=(1100+7810)*θGen Alignment correction h1=h1-(6310+3900+3350)*ψ h3=h3-6310* ψ = 0 h5=h5-1575*ψ h2=h2-(6310+3900)*ψ h4=0 h6=h6-(1575+7810)*ψ hG θ Gen θIP h3 θHP h2 h1 hH1 Λ 0.03 D870 Λ 0.04 D740 h5 V 0.158 D760 ψ = Rotation angle of Shaft Alignment h1 h2 h5 h6
    • SCHEMATIC FOR GERB SPRING TIE ROD SHIM TG DECK TG COLUMN NOTE: 1. THESE READINGS ARE IN ADDITION TO READING TAKEN BY GERB ON THE PROTOCOL DOCUMENT. 2 TURBINE ENGINEER ALONG WITH CIVIL ENGINEER TO ASSOCIATE. A. STICK MICRO METER READING AT FOUR LOCATIONS BETWEEN DECK AND COLUMN. MARK THE LOCATION OF READING (USE METAL MARKER). B. STICK MICROMETER READING AT FOUR LOCATION OF EACH SPRING ASSEMBLY. C RECORD TOTAL THICKNESS OF SHIM HEIGHT AND NUMBER OF SHIMS. A AB B C
    • TIE ROD SCHEMATIC FOR M/S GERB’S CONDENSER SPRING ASSEMBLY SHIM JACK BOLTS CONDENSER FOUNDATION CONDENSER BOTTOM PLATE BACK
    • THANK YOU