Balancing the 2015-2016 Operating Budget Ward Forum Presentation.
Balancing presentation
-
Upload
pasha55 -
Category
Entertainment & Humor
-
view
2.136 -
download
15
description
Transcript of Balancing presentation
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 Fsince 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 mPosition 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 FrequencyThe 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 speedCritical 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
Mic
rons
pk-
pk
Phase
degre
es
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 M
icro
ns
pk-
pk
Phase
degre
es
ROTOR AND BALANCE FORCE DETAILS GENERATOR ROTOR WEIGHT : 37000 KGGENERATOR ROTOR STATIC WEIGHT PER BEARING : 18500 KG
BALANCING RADIUS FAN PLANE : 310 MMRETAINING RING PLANE : 460 MM DISTANCE BETWEEN RETAINING RING PLANE : 4850 MMDISTANCE BETWEEN FAN PLANE : 5740 MMAPPROXIMATE WEIGHT OF TRIAL WEIGHT : 93 GRAMCENTRIFUGAL 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 M
icro
ns
pk-
pk
Phase
degre
es
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
StationaryCasing
Thrust probe
Journalbearing
shaft NormalThrust
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 NormalThrust
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 HP = IP – 0.085/870 IP = 0.205 / 740 = h3/6310 Gen = 0.158 / 760 h2= hH1 + 405 * HP h3 = 485* IP h5=1100*Gen h1= h2+3350* HP hH1=(485+3495)*IP h6=(1100+7810)*Gen Alignment correction of case 1
h1=h1-(6310+3900+3350)* h3=h3-6310* = 0 h5=h5-1575*
h2=h2-(6310+3900)* h4=0 h6=h6-(1575+7810)*
hG
Gen
IP
h3
HP
h2 h1
hH1
V 0.085 D870
0.205 D740
h5
V 0.158 D760
= 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 HP = IP + 0.03 / 870 IP = 0.04 / 740 = h3/6310 Gen = 0.158 / 760 h2= hH1 + 405 * HP h3 = 485* IP h5=1100*Gen h1= h2+3350* HP hH1=(485+3495)*IP h6=(1100+7810)*Gen Alignment correction
h1=h1-(6310+3900+3350)* h3=h3-6310* = 0 h5=h5-1575*
h2=h2-(6310+3900)* h4=0 h6=h6-(1575+7810)*
hG
Gen
IP
h3
HP
h2 h1
hH1
0.03 D870
0.04 D740
h5
V 0.158 D760
= 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