Post on 26-Dec-2014
transcript
Chapter 12
Lubrication & Journal bearings
(7 - 8 Lectures)(7 - 8 Lectures)
TOPICS1. Definitions and Objectives2. Types of Lubrication 3. Dynamic Viscosity 4. Bearing Characteristic Number5. Stable & Unstable Lubrication 6. Hydrodynamic Lubrication 7. Design Considerations 8. Heat Balance-Self-Contained Bearings9. Clearance10. Pressure-Fed Bearings11. Loads and Materials12 Boundary lubrication13. Types of Journal Bearings
Announcements
• HWK 3 Due Wed. 15/11
• Quiz 3 on Monday 13/11 Ch 14
Definitions and objectives• The role of a bearing is to provide relative positioning
and rotational freedom while transmitting a load between a shaft and a housing.
• There are two general types of bearings: • Rolling-contact bearings (anti-friction bearings, rolling
bearings). In the rolling-contact bearings the load is transmitted by rolling rather than by sliding.
• Journal Bearings (plain bearings, bushings, sleeve bearings). In journal bearings, the load is transmitted by sliding and the problem of this class of bearings is
essentially a lubrication problem.
Definitions and objectives
• Journal Bearings: cylindrical or semi-cylindrical bushing made of a suitable material.
– The Journal is the part of shaft or gear in bearing
• Among applications:1. High speed, high temperature, high varying loads:
• Automotive engines: connecting rod, crankshaft,…Metal alloys
• Turbo machinery: Metal alloys
2. Light loads, low speeds with little or no lubrication:• Nylon, Teflon, rubber
1. HydrostaticLow speed, light load
4. Elastohydrodynamic
For rolling contact
(gears, rolling bearings)
6. Solid Film
Extreme Temperatures
(Graphite or Molybdenum disulfide))
Types of Lubrication
Types of Lubrication • Hydrodynamic Lubrication
(HDL) (a) Full, thick Fluid film lubrication - surfaces separated by bulk lubricant film; Film conditions required for lubrication.
2. Boundary (Thin Film)Lubrication (b) partial lubrication (mixed) - both bulk lubricant and boundary film play a role; (c) boundary lubrication - performance depends essentially on boundary film
Viscosity
dy
du
A
F µτ ==
µ is absolute or dynamic viscosity (lbf.s/in2 or reyn. In ips system and
Pa.s in SI system)
du/dy is the rate of shear or velocity gradientIf rate of shear is constant: du/dy = U/h With h= c (clearance)
Fig. 12.1
Shear Stress
c
U
h
U
A
F µµτ ===
c=
Viscosity v.s. temperature
In general, Viscosity decreases with temperature increase.The increase in temperature comes from friction
Petroff’s Law
⇒=
===
==
==
==
= cr
PNfsf
f
s
s
TT
rl
WPrlPrffWrT
c
lNrlrT
rlArAT
c
rN
c
U
µπ
µπτπ
πτ
πµµτ
22
2;)2(
4)(2
2;)(
2
322
Bearing Characteristic Number (Sommerfeld Number)
Scrf
cr
PNS
22
2
π
µ
=
=
Petroff used a concentric shaft to define a group of dimensionless parametersThat allow the prediction of an acceptable coefficient of friction.
r/c = clearance ratio
Shear torque in lubricant
Friction torque
0.08-0.14For steel onBronze
a
bc
12
1’
2’f = 0.001-0.005Similar to precision BB
Stable and unstable lubricationThe McKee Brothers Plot
(a) Full, thick Fluid film lubrication - surfaces separated by bulk lubricant film; Film conditions required for lubrication.
(b) partial lubrication (mixed) - both bulk lubricant and boundary film play a role;
(c) boundary lubrication - performance depends essentially on boundary film
• Boundary lubrication should be expected for slow speeds:
U<10 ft/min (0.05 m/s)
Hydrodynamic Lubrication (HDL)
• For lubricated bearing the minimum film thickness h0 occurs to the left of load line because the shaft is pushed by the pressure build up on the right. The shaft is playing the role of a pump.
Fig. 12.6
• e: eccentricity
• h0 minimum film thickness
∀ ε = e/c = eccentricity ratio
• ß bearing angular length
Hydrodynamic Lubrication (HDL)Nomenclature
• Tower investigated bath-type lubrication in 157° partial bearing. He was able to determine the pressure distribution in oil film in axial and radial directions.
• Reynolds used Tower’s findings to propose a relationship between friction, pressure and velocity. His work is given under mathematical form in the following.
Hydrodynamic Lubrication (HDL)-Theory B. Tower
Hydrodynamic Lubrication (HDL): theoryO. Reynolds
• Assuming pressure varies in x-direction only (no leakage)
• Assuming Velocity varies in x & y directions
)1(
0
ydxdp
dxdzdxy
dxdzpdydzdydzdxdxdppFx
∂∂=
=∂∂+−+−+=Σ
τ
τττ
• Assuming Newtonian viscous fluid + u = u(x,y)
• Assuming Constant viscosity and substituting Eq. (1) into (2):
• Integrating (3) twice (holding x constant):
• Assuming no slip at boundaries:
Hydrodynamic Lubrication (HDL): theory
)2(y
u
∂∂=µτ
)3(12
2
2
2
dxdp
y
uory
udxdp
µµ =
∂∂
∂∂=
)4(121
2
2CyCy
dx
dpu ++=µ
)5(2
@
00@0
1
2
−=⇓⇒==
=⇒==
dx
dph
h
UChyUu
Cyu
µ
• The velocity distribution in film is:
• Flow rate:
• Incompressible flow:
The above is the Reynolds Eq. For one-dimensional flow.
Considering Leakage (2-D):
Hydrodynamic Lubrication (HDL): theory
( ) )6(2
1 2 yh
Uhyyudxdp −−=
µ
)8(0 adx
Qd −=
)1112(633
−=+
∂∂
∂∂
∂∂
∂∂
dx
dhUz
ph
zx
ph
x µµ
)7(122
3
0 dxdphUhudyhQ
µ−==∫
)1012(63
−=
dx
dhUdx
dph
dx
d
µ
Hydrodynamic Lubrication (HDL):Theory
• There are no general analytical solutions to the 2-D Reynolds Equation.
• The Summerfeld Solution to Eq. 12-11
)1212(2
−
=
P
N
c
rfc
r µφ
Design ConsiderationsTwo groups of variables in the design of sliding bearings (eq:12.12)
A- The independent variables:• The viscosity µ,• The load per unit of projected bearing area, P The speed N• The bearing dimensions r, c, β and l
B- The dependent Variables or performance factors:• The coefficient of friction f• The temperature rise ∆T • The volume flow rate of oil Q • The minimum film thickness ho
The first (A) are somewhat under designer control and the second (B) are not.
Design Criteria for Journal Bearings (See Lab Manual for details)
• The value of the important parameter l/d is taken between 0.25 and 1.5. Values up to 2 were used in earlier designs. Nowadays the value of l/d is confined between 0.25 and 0.75. Short bearings are preferred when shaft deflections and misalignments are expected.
• The nominal value of clearance ratio r/c can be taken approximately as:
• 1000 for precision bearings when 25<d<150 mm• 500 for general machinery• 250 for rough machineryThe choice of the values of r/c depends on the tolerances and
surface roughness of shaft and bearing.
Design Criteria for Journal Bearings (See Lab Manual for details)
3. The minimum film thickness h0 can be estimated from one of these equations (Trumpler’s design criteria):
or
• The outlet temperature of the oil should be kept below 250°F (121°C). A value of 70°C (160°F)is usually specified as the average operating temperature
• Starting unit load Pst=Wst/ld is kept below 300 psi
• Design factor on starting load should be at least 2.
)(0004.0005.0
)(00004.00002.0
0
0
mmdh
indh
+≥
+≥
)(00025.00 indh =
Relationship between variables
Viscosity ChartsIn IPS units
Relationship between variables
Viscosity ChartsIn SI Units
Relationship between variables
Viscosity Charts
Relationship between variables
Minimum Film Thickness & Eccentricity ratio Chart
Optimal designZone
Relationship between variables
Minimum Film Thickness Angular position vs. S
Relationship between variables
Coefficient of friction variable vs. S
Relationship between variables
Flow variable vs. S
Relationship between variables
Maximum pressure ratio vs. S
Relationship between variablesTerminating Position of film pressure & maximum film pressure vs. S
Relationship between variablesLubricant Temperature rise ∆T
)(2
112
aQsQ
TQCTsQQCTQC ppsplossH
−∆=∆−+∆= ρρρ
Taking T1 as reference temperature:
)(Pr42 bc
fr
J
lNc
J
TNlossH ππ ==The heat loss due to friction
( ) ( )[ ] )(//5.01
/
4c
rcNlQQsQ
cfr
P
CJ Tp
−=
∆
πρ
Equating (a) to (b)
( ) ( )[ ] )1512(//5.01
/70.9 −−
=∆
rcNlQQsQ
cfr
Ppsi
FTWith ρ = 0.0311 lbm/in3 & Cp = 0.42 Btu/lbm.°Ffor petroleum lubricants and J=9336 lbf.in/Btu
Relationship between variablesLubricant Temperature rise vs. S
Sample problems on HDL
The analysis problems are of two general categories:• When the viscosity is specified as in example 12-
1 through 12-4 of 7th ed. The solution is straight forward.
2) The problem becomes more complex when only the lubricant inlet temperature is specified. To solve this type of problem an iterative procedure has to be followed. An example of the procedure is given in the following.
Problem # 12-12 (Modified)A 2-1/2 x2-1/2-in sleeve bearing uses grade 20 lubricant. The axial-
groove sump has an inlet temperature of 110° F. The shaft journal has a diameter of 2.500 in and the radial clearance is 0.002 in. lf journal speed is 1120 rev/min and the radial load is 1200 Ibf. Estimate
(a) The magnitude and location of the minimum oil-film thickness.(b) The eccentricity.(c) The coefficient of friction.(d) The power loss rate.(e) Both the total and side oil-flow rates.(f) The maximum oil-film pressure and its angular location.(g) The terminating position of the oil film.(h) The average temperature of the side flow.(i) The oil temperature at the terminating position of the oil film.
• Given: d = 2.5 in, b = 2.504 in, cmin = 0.002 in, W = 1200 lbf, SAE = 20, T1 = 110°F,N = 1120 rev/min, and l = 2.5 in.
• Required (see list)• Solution: to find any of these performance
factors we need to have the bearing characteristic number: S.
• To find average viscosity (From Fig. 12-11; 12) we need to have the average operating film temperature Tf (Eq. 12-14):
Procedure: (good for IPS and SI system)• For a first trial assume ∆T = (General) 20 – 80 °F
(10-50°C) For our case take ∆T = 40 °F
2. Tf =130 °F
Problem # 12-12
( ) avavavav
P
N
c
rS µµµµ 42 108.3
192
67.18625
5.25.2120067.182
002.
25.12×==
×
==
21T
f TT ∆+=
• Find µav = 3.8 µreyn (From Fig. 12-11; 12) using Tf = 130°F
• Calculate S = 3.8x104x3.8x10-6 = 0.144
• Calculate ∆TF or ∆TC using 12-18 or Fig 12-23; 24 with S=0.144 and l/d =1
• Recalculate Tfcal = 110+25.7/2≅122.85 °F
• Compare Tfcal to Tfassum if |difference| less than 6 °F or 3 °C Recalculate, For our case Tfassum -Tfcal = 130-122.85= 7.15 >6 °F
need to re-iterate:
1’ assume ∆T =30 °F
2’ Tf = 125°F 5’ ∆TF ≅ 27 °F
3’ µav ≅ 4.3 µreyn 6’ Tfcal = 110+27/2≅123.5°F
4’ S ≅ 0.163 7’ Tfassum -Tfcal = 125-
123.5=1.5<6 °F ACCEPT: Tf = 125°F or Tf = (125+123.5)/2= 124.25 °F
FT oFP
T
psi
F 7.2570.9/3.11923.170.9 =×=∆⇒=∆
Problem # 12-12
µav = 4.3 µreyn (From Fig. 12-11 for oF; 12 for oC) using Tf = 125°F yielding S=0.163
• Using Fig 12-16 with S=0.163 and l/d =1 h0/c = 0.49 ⇒ h0 = 0.0098 in
Using Fig 12-17 ⇒φ = 56 °• e= c- h0 =.002-.00098 = 0.001in. or using
Fig. 12-16 ⇒ ε =e/c = 0.5 ⇒ e = 0.001 in.• f :Fig 12-18 ⇒ (r/c)f= 4
f= 4/625=0.0064d) Power loss: H=(2πTN)/(778x12)= (2π
fWrN)/778x12=H = 0.121 Btu/s =436 Btu/hrH = 126 j/s=453 KJ/hr
Problem # 12-12
e) Using Fig 12-19 with S=0.163 and l/d =1
⇒ Q/rcNl = 4.15 ⇒ Q = 4.15x1.25x0.002x18.67x2.5=0.48 in3/s
Using Fig 12- 20 ⇒ Qs/Q=0.61⇒Qs = 0.29 in3/s
f) Using Fig 12-21 ⇒ P/Pmax = 0.44 ⇒ Pmax = 192/0.44=436 psig
• Using Fig 12-22 ⇒ θ Pmax = 18° & θp0 = 82°
• See part (a) Tav = 125°F
• T2= 110+30=140 °F
Problem # 12-12
NOTE: In cases where l/d curve is not available the interpolation equation (12-15; 16)
may be used when necessary.
Sample problem on Design of HDL Journal Bearings (to be solved during help session)
• Design a journal bearing to carry a radial load of 1500 lb while the shaft rotates at 850 rpm. The shaft stress analysis determines that the minimum acceptable diameter at the journal is 2.10 in.
• The shaft is part of a machine requiring good
precision.• Power loss in the bearing should not exceed 1%
of the 15 hp driving power.
Procedures for design of oil lubricated journal bearings
• A- Full-film (Hydrodynamic) Lubrication• Step1: Often, the shaft diameter at the bearing is
determined by strength and deflection analyses. If the shaft diameter is not known Table 12-5 or Table 28-8 of the Standard Handbook of Machine Design can be utilized to get a rough estimate of the unit load P=W/ld (with W being the applied load). This value is combined with the value of l/d (ratio of bearing length to bearing diameter), determined in the next step, to find the dimensions of the bearing.
Procedures for design of oil lubricated journal bearings
• Step2: The value of the important parameter l/d is taken between 0.25 and 1.5. Values up to 2 were used in earlier designs. Nowadays the value of l/d is confined between 0.25 and 0.75. Short bearings are preferred when shaft deflections and misalignments are expected.
• Step3: The minimum film thickness h0 can be estimated from one of these equations: )(00025.00 indh =
)(0004.0005.0
)(00004.00002.0
0
0
mmdh
indh
+≥
+≥
Procedures for design of oil lubricated journal bearings
Step4: The nominal value of clearance ratio r/c (r = bearing radius and c = clearance) can be taken approximately as:1000 for precision bearings when 25<d<150 mm500 for general machinery250 for rough machineryThe choice of the values of r/c depends on the tolerances and surface roughness of shaft and bearing. This guideline when combined with the results of steps 1 and 2 will allow you to get the nominal value of c.
Procedures for design of oil lubricated journal bearings
• Step5: Now the bearing characteristic number (S = Sommerfeld number) can be determined from the chart of Fig. 12.164 .
• Step6: Next, the viscosity µ of the oil is determined using:
Where:P (unit load) = W/ld, with W being the applied load. N = speed in revolutions per second.
N
P
r
cS
2
=µ
Procedures for design of oil lubricated journal bearings
• Step7: The outlet temperature of the oil should be kept between 200°F (93°C) and 250°F (121°C). A value of 70°C (160°F) is usually specified as the average operating temperature [2, 9, 18-20]. The chart of Fig 12-11 or 12-12 [3,4]) can be entered to select an oil grade. If the selected lubricant has a viscosity higher than the value computed in step 6, recalculate S and find the new h0.
• Step8: Now, find the friction coefficient from Fig. 12-17. The friction coefficient should be kept as low as possible consistent with h0 (i.e. in the optimum zone between the minimum friction line and the maximum load line in Fig. 12.14 [3,4]. As a general rule friction coefficients below 0.01 are acceptable (see Table 28-1 of the Standard Handbook of Machine Design [5]).
Procedures for design of oil lubricated journal bearings
• Step9: Power loss due to friction can be calculated from:
•• Its value can be compared to the input power to take a
decision concerning f and h0. • Step10: Select a suitable bearing material from Table
12-5 [3,4] or from Tables 28-2 to 28-4 of the Handbook [5]. Unit load, maximum operating temperature and conditions should be used as criteria for material selection.
• Step11: Write a summary of your design results.
)(1050
hpfWrN
H =
Self-Contained Bearings
Examples of Pillow-blocks withPolymer Bearings
Ring oiled bearing
Pillow-blocks or pedestal bearings are used for:•Fans,•Blowers•Pumps and small motors
Self-Contained Bearings
)(Pr42
bc
fr
J
lNc
J
TNgenH ππ ==
Two general types of lubrication: 1) Oil-Ring and 2) Oil BathSince the warm lubricant stays within the bearing housing; it shouldbe designed such that the heat generated by friction is dissipated.
As seen above the heat generated (in Btu/s) by friction can be estimated:
Where J= 9336 in.lbf/Btu
The heat to be dissipated & surface temperature of housing are respectively:
)19;1712(1
)19;1712()(1
bbTT
T
aaTTA
fb
fCR
lossH
−++=
−∞−+
=
∞
ααα
)(1050
hpinfWrN
genH =
Tf is the average film temperature which is unknown and found by trial and error to satisfy Hgen=Hloss as in the following example.See also (Eq. 12-20) for Tf
Or in (hp)
See Eq. 12-18 for ħCR and Table 12-2 for α
Example on self-contained bearings
Example on self-contained bearings
Example on self-contained bearings
ClearanceAmong the independent variables under designer’s control, clearance is the most difficult to hold accurate during manufacture and It may also increase during service because of wear.
When selecting a clearance for a JB a number of performance variables and expected in service wear should be taken into account.
Bearing Noisy+ h0 decreases
ClearanceTable 12-3: Max., Min. & Average Clearances for 1.5 in. dia. JB based on fit
Clearance
Temperature limits for mineral oils
Oils with antioxidants + O2 supply unlimited
O2 insignificant
Pressure-Fed Bearings
•At high bearing loads and high temperature: turbo machinery, car engines, ESP, ….•Lubricant is supplied at supply pressure Ps through supply hole drilled opposite to load bearing area side.
Pressure-Fed Bearings
w
Unit load
( )
−=
−==
2
''
2312'4'2
2/
wll
rl
W
rl
WP
Velocity Profile
)2112(4'8
22 −−=
ycl
pu s
µ
Pressure-Fed Bearings
Example of pressure-fedGrooved bearings
Centrally located full annular groove
Circumferential groove axial pressure distribution
Pressure-Fed Bearings
Use charts with l/d
)6(2
1 2 yh
Uhyyu
dxdp −−=
µ
'8
)2112(4'8
2
max
22
l
cpu
ycl
pu
s
s
µ
µ
=
−−=
( ) )2212(5.11'3
23
−+= εµ
πl
rcpsQ s
( )2312'4'2
2/ −==rl
W
rl
WPrl
WP
2=
Qs from Fig. 12-19; 20
Pressure-Fed lubricantNatural circulation of oil
Velocity
Side-Flow
Unit load
Use charts with l’/d
( )( )
( )( ) )2912(
5.11
/)10(978
)2812(5.11
/0123.0
42
26
42
2
−+
=∆
−+
=∆
rp
WScfrT
rp
WScfrT
s
C
s
F
ε
εTemperature rise
∆T from Fig. 12-23; 24
• An eight-cylinder diesel engine has a front main bearing with diameter 3.5 in. and length 2 in. The bearing has a central annular oil groove 0.250 in. wide. It is pressure-lubricated with SAE 30 oil at an inlet temperature of 180°F and at a supply pressure of 50 psi. Corresponding to a radial clearance of 0.0025 in, a speed of 2800 rev/min, and a radial load of 4600 lb, find the temperature rise and the minimum oil-film thickness.
Example on Pressure-Fed BearingsProblem 12-34; 16 (modified)
• Given: d = 3.5 in, l = 2.0 in, Ps = 50 psi, w = 0.25 in; cmin = 0.0025 in, W = 4600 lbf, SAE = 20, T1 = 180°F; N = 2800 rpm
• Required: ∆TF, h0, Pmax, θ Pmax & θp0
• Solution: Use Eq. 12-28 to compute ∆TF
Problem # 12-34; 16 (modified)
( ) avavP
N
c
r avS µµµ
410045.375160
28002700
2×===
psirl
WP
c
r
inwl
l
751875.075.14
4600
'4
7000025.0
75.1
875.02
'
=××
==
==
=−=
( )( ) )2812(
5.11
/0123.042
2
−+
=∆rp
WScfrT
s
F ε
• For a first trial assume ∆T = 30 °F 2. Tf = 180+30/2 = 195 °F • Find µav = 1.4 µreyn (From Fig. 12-11; 12) using Tf =
195°F• Calculate S = 0.0426• Use S = 0.0426 and l’/d = ¼ to find ε = 0.93 from Fig. 12-
15; 16 & (r/c)f = 2.2 from Fig. 12-17; 18
• Calculate ∆TF
• Recalculate Tfcal = 180+22.64/2≅191.3 °F
• Compare Tfcal to Tfassum if |difference| >6 °F Recalculate, For our case
Tfassum -Tfcal = 195-191.3= 3.7 <6 °F
ACCEPT: ∆T = 30 °F
Problem # 12-34; 16
( )( ) FTF °=
××+×=∆ 64.22
75.15093.05.11
246000426.02.20123.042
• Using Fig 12-14; 16 with S=0.0426 and l/d =1/4 h0/c = 0.07 ⇒ h0 = 0.000175 in
Trumpler’s Criteria satisfied?Trumpler’s Criteria satisfied?
• hh00 ≥≥ 0.0002+0.00004(3.5)=0.00034 in 0.0002+0.00004(3.5)=0.00034 in not not satisfied?satisfied?
• TTmaxmax = T = Tss+ + ∆∆T= 180+22.64=202.64 °F T= 180+22.64=202.64 °F <250<250
°F °F OKOK
• PPstst = 751 psig = 751 psig <350 psi <350 psi not satisfied?not satisfied?
• Using Fig 12-20; 21 ⇒ P/Pmax = 0.16 ⇒ Pmax =
751/0.16=4694 psig • Using Fig 12-21; 22 ⇒ θ Pmax = 8° & θp0 = 24°
Problem # 12-34; 16
A- Loads: Typical values of unit load P
JOURNAL BEARING LOADS & MATERIALS
JOURNAL BEARING LOADS & MATERIALS
B- Materials: To minimize wear of journal bearings, Metallic Materials (Table 12-5 for Hydrodynamic Lubrication and 12-6 for Boundary Lubrication) are selected for:
1. Mechanical Properties
• Conformability: to compensate for small shaft misalignments and deflections (i.e. Low E and yield: Lead base Babbit=90% Pb + 10% Cu)
∀ • Embeddability: to allow foreign particles to become embedded into the bearing which prevents scratching of shaft and sleeve (Tin base and Lead base Babbit)
∀ • High Fatigue Strength: to support the compressive cyclic loading (Trimetal, Silver, Steel base, Solid Brass…)
JOURNAL BEARING MATERIALS 2. Thermal Properties • High Thermal Conductivity: to remove heat rapidly from the
bearing (Ag, Cu, Pb). • Thermal Coefficient of Expansion not too different from that
of casing and shaft. 3. Metallurgical Properties • Compatibility: to avoid fusing under heat and contact
dissimilar materials (Mainly not same melting point) for shaft and bearing are more compatible than similar materials.
4. Chemical Properties • Corrosion Resistant: to resist corrosion by lubricant
improvement additives (Sn, Al, Ag...).
Non-Metallic Materials (Table 12-6) such as Wood, Rubber, Carbon Graphite, Derlin, Teflon, Nylon… Most have low thermal conductivity.
JOURNAL BEARING MATERIALS
Boundary (thin-film)-Lubrication • In certain applications boundary lubrication should be
designed for (see your lab manual for the procedure of boundary lubrication design).
• Boundary lubrication should be expected for slow speeds (start ups and shut downs) : U<10 ft/min (0.05 m/s)..
• In boundary lubrication the bearing performance depends essentially on boundary film.
• The coefficient of friction is reduced by using animal and vegetable oils containing fatty acids that stick to metal surfaces.
Materials for Boundary (thin-film, boundary friction, oilite, oiles and bushed pins)-Lubrication
To minimize metal-to-metal contact in boundary lubrication:•Mix animal or vegetable oils with lubricant•Use porous metallic materials (Table)•Use non-metallic materials•Use indented bearings
Table 12-8
Sample problem on Design of Boundary-Lubricated Journal Bearings
• Design a boundary lubricated plain-surface bearing to carry a radial load of 2.5 kN from a shaft rotating at 1150 rpm. The nominal minimum diameter of journal is 75 mm.
Given: Boundary lubricated JB. W=2.5 kN; n= 1150 rpm; d = 75 mm
• Solution: (see class work)
Types of Radial Journal Bearings
Radial
(Plain Bearings, sleeves)
Thrust Journal Bearing
Thrust
Journal Bearings
Types of bearings
Plain Bearings
Journal Bearings
Self-lubricated Journal Bearings
Bushes Polymer Bearings
Types of Bearings
Radial Journal Bearings for Pinion Shaft in Gear Box for GE Turbine
• Housing for Gear Box showing Radial Journal Bearing Supports
Types of Bearings
Types of Bearings
Radial Journal Bearings for Pinion Shaft in Gear Box for GE Turbine
Types of Radial Journal Bearings
Types of Radial Journal Bearings
Types of Radial Journal Bearings
Typical Groove Patterns
Thrust Journal Bearing
• Thrust Bearing for GE
Turbine Shaft
Thrust Journal Bearing