Journal Bearing

 KELOMPOK 2

Darma Adhi W. Galuh Intan P. Bhatara Putra M. Mulyani Muliyani Ali Akbar Tia Utari Eko Rusdiyanto

(11210009) (11210012) (11210014) (11210020) (11210023) (11210024) (11210028) (11210030)

Konsep Dasar Apa itu Bearing?

Bhatara Putra Mediriyanto

Tipe beban pada bearing 1. Bantalan radial (Journal radial (Journal Bearing)  Arah beban yang di tumpu bantalan ini adalah tegak lurus dengan sumbu poros 2. Bantalan aksial (Thrust Bearing)   Arah beban bantalan ini adalah sejajar dengan sumbu poros 3. Bantalan kombinasi (combination bearing)  Bantalan ini menumpu beban yang arahnya sejajar dan tegak lurus dengan sumbu poros

Introduction

• • •

•

Journal bearing termasuk salahsatu sliding bearing dan keterbalikkan dari ball bearing Journal bearing secara umum digunakan pada mesin piston kendaraan bermotor berbahan bakar bensin atau diesel Kelebihan : • Bearing type ini mampu menopang shaft yang berat. •  Awet dan tahan lama • Efek redaman dari film minyak membantu membuat mesin beroperasi dengan tenang dan halus. Kekurangan : • Membutuhkan suplai minyak pelumas yang besar • Hanya cocok untuk temperatur dan kecepatan rendah • Pembentukan lapisan minyak pelumas lambat

Bearing Diagram

Journal bearing  – Berfungsi sebagai bantalan poros engkol yang berputar Oil inlet  – Tempat  Tempat masuknya minyak pelumas Ketika oli pelumas masuk ke dalam bearing, oli akan memenuhi clearance/ gap antara shaft dan bearing sehinggga mengakibatkan tekanan fuida meningkat dan daya angkat hidrodinamis terhadap shaft

Type

Typical Loading

Application

(a)

Partial ar arc

Unidirectional load

Shaft guides, dampers

(a)

Circum fe ferential groove, Axial groove types

Variable load direction

Internal combustion engines

(a)

Cylindrical

Medium to heavy Unidirectional load

General machinery

(a)

Pressure dam

Light loads, unidirectional

High speed turbines, compressor

(a)

Overshot

Light loads, unidirectional

Steam turbines

(a)

Multilobe

Light loads, unidirectional

Gearing, compressor

(a)

Preloaded

Light loads, unidirectional

Minimize vibration

(a)

Tilting pad

Moderatic Variable loads

Minimize vibration

Movement of the bearing

Video

Infinitely Long Approximation (ILA)

Menentukan jari-jari shaft dan clearance

ILA Menentukan Menentukan Dimensionless pressure

Boundary Boun dary Condition

8.3 BOUNDARY CONDITIONS

  − Ṕ   

Assumsi :  = 0 θ = 0

Ṕ

Dimana : Ps = tekanan suplai C = radial clearence R = radius bearing = viskositas viskos itas pelumas = kecepatan putaran poros

 

8.4 FULL SOMMERFELD BOUNDARY CONDITIONS

Asumsi :  = 0  = 2π (360)

Ṕ θ

8.4 FULL SOMMERFELD BOUNDARY CONDITIONS

Substitusi Sommerfeld :

+θ cos 1 + ɛ θ Tekanan puncak terjadi ketika

Dimana :  = sudut angular pada tekanan maksimum = rasio eksentrisitas eksentrisitas

cos 



 =

−3 cos  (2+ )

 

 = eksentrisitas

C = radial clearence

8.4 FULL SOMMERFELD BOUNDARY CONDITIONS Besarnya tekanan puncak tak berdimensi pada distribusi tekanan adalah :

6  s i n  (2 ( 2 +      ) Ṕ  (2+ )(1 +  ) )

Dari persamaan 8.9 asumsi P = 0 pada

(8.9)

Ɵ  π besarnya tekanan ,

puncak tak berdimensi adalah :

)(4−5  +  ). 3(4− Ṕ   2(1− )(2+ ) Yang terjadi pada :



−3 2 + 

Load Carrying Based on Full Sommerfe Sommerfeld ld Condition

Load Carrying Based on Full Sommerfe Sommerfeld ld Condition Arah Radial

 Ѿ       

(8.10)

Arah Tangensial Tangensial

Ѿ     

(8.11)

Dari substitusi tekanan tak berdimensi pada persamaan 8.9 dengan persamaan 8.10 dan 8.11 maka didapatkan :

0

12 Ѿ    (1− ) 2 + 

(8.12)

Load Carrying Based on Full Sommerfe Sommerfeld ld Condition Dimana Beban tak berdimensi:

   Ѿ   

(8.13)

Dengan

 = beban yang diproyeksikan Ns = kecepatan poros dalam rev/s

Resultan dari

Ѿ dan Ѿ

Ѿ Ѿ + Ѿ  ( )()

(8.14)

  

Load Carrying Based on Full Sommerfe Sommerfeld ld Condition Attitude Angle

−Ѿ  ɸ  Ѿ  ∞  ɸ

(8.15)

. Dalam berbagai kasus, jika Ɛ =

0

Ɛ =

1

Ѿ  0 Ѿ∞

8.5 DEFINITION OF THE SOMMERFELD NUMBER 

8.5 DEFINITION OF THE SOMMERFELD NUMBER 

Bilangan Sommerfeld (S) merupakan bilangan tak berdimensi yang merupakan

parameter karakterisasi

performansi

sebuah bearing.

Bilangan ini menunjukkan karakteristik gesekan total dari bantalan.

      

(8.16)

substitusikan ke dalam persamaan Sommerfeld Number (8.14) maka :

1 Ѿ  

Dan penyelesaian S menjadi



  (1− ) (2+ )

12

(8.17)

8.6 HALF SOMMERFELD BOUNDARY CONDITION

Ѿ

6   1 −  ) (2+

(8.18)

 12 Ѿ 1 − )(2+

(8.19)

Total kapasitas beban dukung dan attitude angle adalah:

6 Ѿ 1 − )(2+  − ( − 4) . ∅    (−).

(8.21)

(8.20)

Eko Rusdiyanto

Contoh Soal 8.1

Fenomena Kavitasi Gaseous Cavitation

Kavitasi

Vapor Cavitation

Gaseous Cavit Cavitation ation Gaseous cavitation merupakan kavitasi yang disebabkan oleh adanya bagian dari minyak pelumas yang terlarut dengan udara pada kondisi jenuh (sekitar 10%), dan ketika tekanan sekitar menjadi turun bagian yang terlarut ini akan membentuk suatu kavitasi tetapi dibagian yang berbeda dari fluid film, hal ini yang menyebabkan kavitasi jenis gaseous tidak terlalu berbahaya.

Vapor Cavitation Vapor cavitation disebabkan oleh tingginya fluktuasi tekanan yang ada diantara film dari pelumas dan bearingnya itu sendiri, kavitasi  jenis ini cukup berbahaya berbahaya karena bisa menyebabkan kerusakan pada bearing (fatigue damage)

SWIFT-STEIBER (REYNOLD) BOUNDARY BOUNDARY CONDITION Perhitungan Perhitungan beban bearing dengan memperhatikan kavitasi kavitasi didalam perhitungannya   0     

INFINITEL INFINI TELY Y SHORT JOURNAL BEARING BEARI NG APPROXIMATION (ISA)

Ali Akbar

A. Infinitely Short Journal Bearing Approximation (ISA)

Integral 2 kali

Length-to-Diameter ratios up to L/D = ½ dengan trends rata-rata L/D = 1

Table Infinitely Long Journal Bearing Solutions with the Reynolds Boundary Condition

B. Full and Half Sommerfeld Solutions Solutions for Short Bearings (ISA)

Figure Short Bearing Eccentricity Eccentricity Ratio vs Sommerfeld Number

Figure Short Bearing Attitude Angle vs Sommerfeld Number

FINITE BEARING DESIGN & ANALYSIS This section focused on design and performance analysis based on the full solution of Reynold equation.

FINITE BEARING DESIGN & ANALYSIS

FINITE BEARING DESIGN & ANALYSIS

•

Minimum film thickness hmin = C ( 1 – Ɛ – Ɛ )

•

Friction force F=fW

•

Power loss E  p = F 2π R Ns

•

Temperature rise

 ΔT =

FINITE BEARING DESIGN & ANALYSIS

FINITE BEARING DESIGN & ANALYSIS •

Example:

A large pump has a horizontal rotor weighing 3200 lb supported on two plain 360o journal bearings , one on either side of the pump impeller. The specifications of the bearings are as follo follows: ws: R = 2 in L = 4 in C = 0.002 in N = 1800 rpm the lubricant viscosity µ = 1.3 x 10-6 reyns (SAE 10 at an inlet temperature of 166o F) Determine: a) Equil Equilibr ibrium ium posi positio tion n of the shaf shaftt cente centerr and loca locatio tion n of film film ruptur rupture e b) Minim inimum um fil film th thic ickn knes esss c) Loca Locati tion on an and d mag magni nitu tude de of maxi maximu mum m pre press ssur ure e d) Power loss e) Temperature rise

FINITE BEARING DESIGN & ANALYSIS

FINITE BEARING DESIGN & ANALYSIS

FINITE BEARING DESIGN & ANALYSIS

FINITE BEARING DESIGN & ANALYSIS

Muliyani

ATTITUDE ANGLE FOR OTHER BEARING CONFIGURATION

Where:

LUBRICANT SUPPLY ARRANGEMENT Supply Hole

Axial Groove

Circumferential Groove

SUPPLY HOLE

A common supply methode with small bearing and bushing is to place an inlet port at the bearing midplane opposite to the load line

AXIAL GROOVE

VARIOUS GROOVE POSITIONS AND INLET ARRANGEMENT

CIRCUMFERENTIALS GROOVE GROOVE

Alur yang melingkar ditempatk ditempatkan an pada centerlin centerline e

FLOW CONSIDERATION 

Axial flow due rot rotation ation

Where fL is a corretion factor for the film position as given below: 1. Oil hole hole or axial axial groove groove positio positioned ned in the the unloaded unloaded section section of the the bearing bearing opposite opposite to to the load line:

2. Oil hole or axial groove groove positioned positioned at the maximum film thickness

3. For double axial grooves running running parallel at ±90º angles to the load line:

4. For a full film starting from the maximum film thickness position

FLOW CONSIDERATION Pressure Induced Flow  Pressure

Inlet hole of diameter DH:

•

Qp : Pressure induced flow

•

Ps : Supply pressure

•

μi : Lubricant Viscosity

FLOW CONSIDERATION 1.

The film thickness parameter for an oil hole ar an axial groove positioned in the unloaded section of the bearing opposite the load line is

2. Positioned at the maximum film thickness

3. For double axial grooves grooves running parallel parallel ±90º angles to the load line

Total leakage flow rate

1. For an oil hole or an axial axial groov groove e position positioned ed in the unloade unloaded d section section of of the beari bearing ng opposite to the load line

2. For an an axial axial groove groove of of length length Lg (Lg/L= 0.3 0.3 to 0.8) 0.8) positioned positioned at the maxim maximum um film film thickness or two axial grooves grooves running parallel parallel at ±90 angles to the load line line

Mulyani

Example 8.4 •

•

Consider a journal bearing with the follo following wing specification that correspond to an actual bearing tested by Dowson et al , (1966): L/D = 0.75; R/C = 800; D = 0.102 m; W  =  = 11000 N, and operating speed is Ns = 25 rev/s. An axial groove was cut into the bearing surface in the unload portion of the bearing, opposite the load line. the groove width is ωg = 4.76 x 10- 3 m, and it is Lg = 0.067 m long. Lubricant Lubricant is supplied to the bearing at temperature temperature T i = 36.8 oC at a supply pressure of P  = 0.276x106 Pa (40 Psi) . The lubricant viscosity is a s function of temperature and varies according to µ =  µie-β(T-Ti) with µi = 0.03 Pa.s, and the temperature viscosity coefficient is estimated to be β = 0.0414. lubricant thermal conductivity conductivity k = 0.13 W/mK, and thermal diffusivity ɑt = 0.756 x 10-7 m2/s. Determine the flow rates, power loss, attitude angle, and maximum pressure. Parameterr Paramete

Nilai

Parameter Parameter

Nilai

L/D

0.75

Ti 

36.8 oC

R/C

800

Ps 

0.276x106 Pa (40 Psi)

D

0.102 m

µi

0.03 Pa.s

W

11000 N

β 

0.0414

Ns

25 rev/s

k

0.13 W/mK

ωg

4.76 x 10-3 m

ɑt

0.756 x 10-7 m2/s

L

0.067 m

•

Menggunakan rumus Sommerfeld number :

     <   0. 0 3  25. 0  0. 0 762  0. 1 02  0.338    ∆  (Ǫ   778 12 0.48 0.0315 (20.37) 33.4: , )

Mean outlet temperature is T0 = 120 + 33.4 = 153oF

BEARING STIFFNESS, ROTOR VIBRATION, AND OIL O IL WHIRL INSTABILIT INSTABILITY Y

•

Spring mass ma ss system system

•

The bearing horizontal horizontal natural frequency

Tia Utari

Example 8.6

Determine whirl stability for a horizontal rotor and its bearings with the following characteristics characteristics : D = 2R = 5 in L = 2.5 in C = 0.005 in N = 90 rad/s rad/s (5400 rpm) Ks = 5 x 10^6 lb/in rotor stiffness W = 5000 lb rotor weight (m = W/g W/g = 5000/386 = 13 lb

μ = 2 x 1

0 lbs/ (reyns) viscosity

/in rotor mass

Example 8.6 Con’t

Analysis Using the graph

Unit bearing load P= =



Sommerfeld number

 (/) S = μN   (.  ∶.   )   = 2*10 / ∗90/  

(DL)

 /   ∗. 

= 200 psi

= 0.225

Characteristic bearing number = S(L/D) = 0.225 (2.5 in/ 5 in

)

Stability on rotor stiffness

 = (0.005 in/ 5000 lb)*(5*10 lb/in)

(C/W)

= 10

= 0.05625 next

Example 8.6 Con’t

Stability on case (C/W)m = (0.005 in/ 5000 lb)*(13 lb = 8.28

ω

Rotor will be free of oil whip instability instability

/in*(2π*90/))

Effective Temperature Misalignment and Shaft Deflection

Operating Clearance

Maximum Bearing Temperature

General Design Guides

Eccentricity and Minimum Film Thickness

Turbulent and Parasitic Loss Effect

Flooded versus Starved Condition Bearing Load Dimensions

Effective Eff ective Temperature

Temperature rata-rata pada viskositas tertentu

Global effective temperature

For small bearing

Dimana : J  panas mekanik ρ densitas oil  leakage flowrate  temperatur awal  kapasitas panas  conduction & radiation  power loss

   Φ 

Maximum Bearing Bea ring Temperature Temperature

Temperature

Minimum film thickness

Turbulent and Parasitic Loss Effect

Turbulent : •

Bearing diameter

•

Large film thicknesses (clearance)

•

High surface speed

•

Low fluid viscosities

•

High Reynolds numbers Parasitic Parasitic loss : •

Putaran Putaran dan turbulensi pada oil grooves dan clearence

•

Losses pada percepatan feed oil terhadap surface speed yang tinggi

•

Vortex pada feed dan oil grooves

•

Surface drag yang terjadi antara oil dengan high surface speed

Flooded versus Starved Condotion

Leakage flow rate Flooded

Kerja bearing jelek

Condition

Good Cooling

Starved Condition

•

 < 1 

Bearing Load and Dimensions

Projected Loading PL= W/(L*D)

Rentan vibrasi Power loss tinggi High oil flow

overheating

Eccentriciry and Minimum Film Thickness

 = C – e

= C(1 – Ɛ)

• Excessive bearing temperature • Susceptibility to wear

too small

too large • Poorer vibration • Higher power loss

Operating Clearance

Clearance is 0.002 per inch of diameter