Fundamentals Of Mechanical Design 6s136c

Fundamental Principles of Mechanical Engineering Technical Handbook

FLENDER Drives Answers for industry.

FLENDER Drives Fundamental Principles of Mechanical Engineering Technical Handbook

Technical Drawings

1

Standardization

2

Physics

3

Mathematics / Geometry

4

Mechanics / Strength of Materials

5

Hydraulics

6

Electrical Engineering

7

Materials

8

Lubricating Oils

9

Cylindrical Gear Units

10

Shaft Couplings

11

Vibrations

12

Bibliography

13

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Contents

Section 1 Technical Drawings Surface Texture Geometrical Tolerancing Sheet Sizes, Title Block, Non-standard Formats Type Sizes, Lines, Lettering Example

Page 5+6 7 – 21 22 23

Section 2 Standardization ISO Metric Screw Threads (Coarse Pitch Threads) ISO Metric Screw Threads (Coarse and Fine Pitch Threads) Cylindrical Shaft Ends ISO Tolerance Zones, Allowances, Fit Tolerances Parallel Keys and Taper Keys, Centre Holes

25 26 27 28 + 29 30

Section 3 Physics Internationally Determined Prefixes Basic SI Units Derived SI Units Legal Units Outside the SI Physical Quantities and Units of Lengths and Their Powers Physical Quantities and Units of Time Physical Quantities and Units of Mechanics Physical Quantities and Units of Thermodynamics and Heat Transfer Physical Quantities and Units of Electrical Engineering Physical Quantities and Units of Lighting Engineering Different Measuring Units of Temperature Measures of Length and Square Measures Cubic Measures and Weights Energy, Work, Quantity of Heat Power, Energy Flow, Heat Flow Pressure and Tension Velocity Equations for Linear Motion and Rotary Motion

32 32 33 33 34 35 35 – 37 37 + 38 38 39 39 40 41 41 42 42 42 43

Section 4 Mathematics / Geometry Calculation of Areas Calculation of Volumes

45 46

Section 5 Mechanics / Strength of Materials Axial Section Moduli and Axial Second Moments of Area (Moments of Inertia) of Different Profiles Deflections in Beams Values for Circular Sections Stresses on Structural and Fatigue Strength of Structures ·

48 49 50 51

1

Contents

Section 6

Page

Hydraulics Hydrostatics Hydrodynamics

53 54

Section 7 Electrical Engineering Basic Formulae Speed, Power Rating and Efficiency of Electric Motors Types of Construction and Mounting Arrangements of Rotating Electrical Machinery Types of Protection for Electrical Equipment (Protection Against and Foreign Bodies) Types of Protection for Electrical Equipment (Protection Against Water)

56 57 58 59 60

Section 8 Materials Conversion of Fatigue Strength Values of Miscellaneous Materials Mechanical Properties of Quenched and Tempered Steels Fatigue Strength Diagrams of Quenched and Tempered Steels General-Purpose Structural Steels Fatigue Strength Diagrams of General-Purpose Structural Steels Case Hardening Steels Fatigue Strength Diagrams of Case Hardening Steels Cold Rolled Steel Strips Cast Steels for General Engineering Purposes Round Steel Wire for Springs Lamellar Graphite Cast Iron Nodular Graphite Cast Iron Copper-Tin- and Copper-Zinc-Tin Casting Alloys Copper-Aluminium Casting Alloys Aluminium Casting Alloys Lead and Tin Casting Alloys for Babbit Sleeve Bearings Conversion of Hardness Values Values of Solids and Liquids Coefficient of Linear Expansion Iron-Carbon Diagram Pitting and Tooth Root Fatigue Strength Values of Steels Heat Treatment During Case Hardening of Case Hardening Steels

62 63 64 65 66 67 68 69 69 70 71 71 72 72 73 74 75 76 77 77 77 78

Section 9 Lubricating Oils Viscosity-Temperature-Diagram for Mineral Oils Viscosity-Temperature-Diagram for Synthetic Oils of Poly-α-Olefine Base Viscosity-Temperature-Diagram for Synthetic Oils of Polyglycole Base Kinematic Viscosity and Dynamic Viscosity Viscosity Table for Mineral Oils

2

80 81 82 83 84 ·

Contents

Section 10

Page

Cylindrical Gear Units Symbols and Units General Introduction Geometry of Involute Gears Load Carrying Capacity of Involute Gears

86 + 87 88 88 – 89 99 – 107

Gear Unit Types

107 – 110

Noise Emitted by Gear Units

111 – 114

Section 11 Shaft Couplings General Fundamental Principles

116

Torsionally Rigid Couplings, Flexible Pin Couplings Flexible Claw Couplings

117

Highly Flexible Ring Couplings, Highly Flexible Rubber Tyre Couplings Highly Flexible Rubber Disk Couplings, Flexible Pin and Bush Couplings

118

All-steel Couplings, Torque Limiters High-speed Couplings, Composite Couplings

119

Miniature Couplings, Gear Couplings Universal Gear Couplings, Multiple Disk Clutches

120

Fluid Couplings, Overrunning Clutches, Torque Limiters

121

Couplings for Pump Drives

122

Coupling Systems for Railway Vehicles

123

Coupling Systems for Wind Power Stations

124

Section 12 Vibrations Symbols and Units

126

General Fundamental Principles

127 – 129

Solution Proposal for Simple Torsional Vibrators

129 + 130

Solution of the Differential Equation of Motion

130 + 131

Formulae for the Calculation of Vibrations , Symbols and Units

131 132

Formulae for the Calculation of Vibrations

133 – 135

Evaluation of Vibrations

135 + 136

Section 13 Bibliography of Sections 10, 11, and 12

·

138 + 139

3

Table of Contents Section 1

1 Technical Drawings

Page

Surface Texture Method of Indicating Surface Texture on Drawings acc. to DIN EN ISO 1302 Surface Roughness Parameters

5 5+6

Geometrical Tolerancing General

7

Application; General Explanations

7

Tolerance Frame

7

Kinds of Tolerances; Symbols; Included Tolerances

8

Additional Symbols

8

Toleranced Features

9

Tolerance Zones

9

Datums and Datum Systems Theoretically Exact Dimensions Detailed Definitions of Tolerances

9 – 11 11 11 – 21

Sheet Sizes, Title Block, Non-standard Formats Sheet Sizes for Technical Drawings

22

Title Block for Technical Drawings

22

Non-standard Formats for Technical Drawings

22

Sizes of Type

23

Lines acc. to DIN ISO 128, Part 20 and Part 24

23

Lettering Example

23

4

·

Technical Drawings Surface Texture

1 1. Method of indicating surface texture on drawings acc. to DIN EN ISO 1302 1.1 Symbols for the surface texture Graphic symbols

Meaning Material removal by machining is required (without requirements).

preserved

Material removal by machining is required (with additional indication). Material removal is prohibited (without requirements).

non-porous Material removal is prohibited (with additional indication).

Material removal; surface roughness value Ra = 6.3 m. Material removal applies to the external contour of the view. Machining allowance specified by a numerical value in mm (e.g. 3 mm). lead-free 0.4 - 0.8

Material removal (by machining), surface roughness value Ra = 0.4 - 0.8 m. Requirement for the surface: “lead-free”.

1.2 Definition of the surface parameter Ra The centre line average height Ra of the assessed profile is defined in DIN EN ISO 4287 and

the evaluation length for assessing the roughness in DIN EN ISO 4288.

1.3 Indications added to the graphic symbols a = Requirements on the surface appearance b = Two or more requirements on the surface appearance c = Production method, treatment, coating, or other requirements concerning the manufacturing method, etc. d = Surface grooves and their direction e = Machining allowance (x) = No longer applicable (formerly: indication of Ra) 2. Surface roughness parameters 2.1 Peak-to-valley height Rt The peak-to-valley height Rt in m acc. to DIN 4762 Part 1 is the distance of the base profile to the reference profile (see figure 1). The base profile is the reference profile displaced to ·

such an extent perpendicular to the geometrical ideal profile within the roughness reference length, that s the point of the actual profile most distant from the reference profile (point T in figure 1).

5

Technical Drawings Surface Texture

1

Figure 1 2.2 Mean peak-to-valley height Rz The mean peak-to-valley height Rz in m acc. to DIN 4768 is the arithmetic average of the single irregularities of five consecutive sam-

pling lengths (see figure 2). Note: The definition given for Rz in DIN differs from its definition in ISO.

le = lm = lt = z1-z5 =

Start-up length

Run-out length

Figure 2

An exact conversion of the peak-to-valley height Rz into the centre line average height Ra and vice versa can neither be theoretically justified nor empirically proved. For surfaces which are generated by manufacturing methods of the group “metal cutting”, a diagram for the conversion from Ra into Rz and vice versa is shown in supplement 1 to DIN 4768, based on comparison measurements. The Ra values assigned to the Rz values are subject to scattering (see table).

Sampling length Evaluation length Traversed length Single irregularities

2.3 Maximum roughness height Rmax The maximum roughness height Rmax in m acc. to DIN 4768 is the largest of the single irregularities Z1 occurring over the evaluation length lm (see figure 2). Rmax is applied only in cases where the largest single irregularity (“runaway”) is to be recorded for reasons important for function. 2.4 Roughness grade numbers N In , it is not allowed to use roughness grade numbers (N grades), since they are given in inches.

3. Centre line average height Ra and roughness grade numbers in relation to the mean peak-to-valley height Rz Surface roughness value Ra

μm μin

50

12.5

6.3

3.2

1.6

0.8

0.4

0.2

0.1

500

250

125

63

32

16

8

4

2

1

N 11

N 10

N9

N8

N7

N6

N5

N4

N3

N2

N1

160

80

40

25

12.5

6.3

3.15

1.6

0.8

0.4

0.25

0.1

250

160

100

63

31.5

20

12.5

6.3

4

2.5

1.6

0.8

2000 1000

Roughness grade no. N 12

Surface rough- from ness value to Rz in μm

6

25

0.05 0.025

·

Technical Drawings Geometrical Tolerancing

1 4. General 4.1 The particulars given are in accordance with the international standard DIN ISO 1101, March 1985 edition. This standard gives the principles of symbolization and indication on technical drawings of tolerances of form, orientation, location and runout, and establishes the appropriate geometrical definition. The term “geometrical tolerances” is used in this standard as generic term for these tolerances. 4.2 Relationship between tolerances of size, form and position According to current standards there are two possibilities of making indications on technical drawings in accordance with: a) the principle of independence according to DIN ISO 8015 where tolerances of size, form and position must be adhered to independent of each other, i.e. there is no direct relation between them. In this case reference must be made on the drawing to DIN ISO 8015. b) the envelope requirements according to DIN 7167, according to which the tolerances of size and form are in direct relation with each other, i.e. that the size tolerances limit the form tolerances. 5. Application; general explanations 5.1 Geometrical tolerances shall be specified on drawings only if they are imperative for the functioning and/or economical manufacture of the respective workpiece. Otherwise, the general tolerances according to DIN ISO 2768 apply.

5.4 Unless otherwise specified, the tolerance applies to the whole length or surface of the considered feature. 5.5 The datum feature is a real feature of a part, which is used to establish the location of a datum. 5.6 Geometrical tolerances which are assigned to features referred to a datum do not limit the form deviations of the datum feature itself. The form of a datum feature shall be sufficiently accurate for its purpose and it may therefore be necessary to specify tolerances of form for the datum features (see table on page 8). 5.7 Tolerance frame The tolerance requirements are shown in a rectangular frame which is divided into two or more compartments. These compartments contain, from top to bottom, in the following order (see figures 3, 4 and 5): - the symbol for the characteristic to be toleranced; - the tolerance value in the unit used for linear dimensions. This value is preceded by the symbol ∅ if the tolerance zone is circular or cylindrical; or by the symbol “S∅” if the tolerance zone is spherical; - if appropriate, the capital letter or letters identifying the datum feature or features (see figures 4, 5 and 6). Figure 3 Figure 4

5.2 Indicating geometrical tolerances does not necessarily imply the use of any particular method of production, measurement or gauging. 5.3 A geometrical tolerance applied to a feature defines the tolerance zone within which the feature (surface, axis, or median plane) is to be contained. According to the characteristic to be toleranced and the manner in which it is dimensioned, the tolerance zone is one of the following: - the area within a circle; - the area between two concentric circles; - the area between two equidistant lines or two parallel straight lines; - the space within a cylinder; - the space between two coaxial cylinders; - the space between two parallel planes; - the space within a parallelepiped or a sphere. The toleranced feature may be of any form or orientation within this tolerance zone, unless a more restrictive indication is given. ·

Figure 5 Figure 6 Remarks referred to the tolerance, for example “6 holes”, “4 surfaces”, or “6 x” shall be written above the frame (see figures 7 and 8). If it is necessary to specify more than one tolerance characteristic for a feature, the tolerance specifications are given in tolerance frames one under the other (see figure 9). 6 holes

6x Figure 7

Figure 8

Figure 9

7


1 Table 1: Kinds of tolerances; symbols; included tolerances Tolerances

Toleranced characteristics

Form tolerances

Circularity (Roundness)

–

Cylindricity

Straightness, Parallelism, Circularity

Profile any line

–

Profile any surface

–

Parallelism

Flatness

Perpendicularity

Flatness

Angularity

Flatness

Position

–

Concentricity (for centre points), Coaxiality (for axes)

–

Symmetry

Straightness, Flatness, Parallelism

Circular runout

Circularity, Coaxiality, Concentricity

Total runout

Concentricity, Coaxiality, Flatness, Parallelism, Perpendicularity

Orientation O i t ti tolerances Tol olerance ces of pos osition n

Included tolerances

– Straightness

Profile tolerances

Location tolerances

Symbols

Straightness Flatness

Runout tolerances

Table 2: Additional symbols Description

Symbols

direct Toleranced feature indications

by letter

Datum feature indication (by letter only) Datum target indication

∅2 = Dimension of the target area A1 = Datum feature and datum target number

Theoretically exact dimension Projected tolerance zone Maximum material requirement Dependent on dimensional, form, and position tolerances Least material requirement Dimension describing the least material state of a form feature Free state condition (non-rigid parts) Envelope requirement: The maximum material dimension must not breach a geometrically ideal envelope.

8

·


1 5.8 Toleranced features The tolerance frame is connected to the toleranced feature by a leader line terminating with an arrow in the following way: - on the outline of the feature or an extension of the outline (but clearly separated from the dimension line) when the tolerance refers to the line or surface itself (see figures 10 and 11).

5.9 Tolerance zones The tolerance zone is the zone within which all the points of a geometric feature (point, line, surface, median plane) must lie. The width of the tolerance zone is in the direction of the arrow of the leader line ing the tolerance frame to the feature which is toleranced, unless the tolerance value is preceded by the symbol ∅ (see figures 16 and 17).

Figure 11

Figure 10

- as an extension of a dimension line when the tolerance refers to the axis or median plane defined by the feature so dimensioned (see figures 12 to 14).

Figure 17

Figure 16

Where a common tolerance zone is applied to several separate features, the requirement is indicated by the words “common zone” above the tolerance frame (see figure 18). Common zone

Figure 12

Figure 13

Figure 14

- on the axis or the median plane when the tolerance refers to the common axis or median plane of two or more features (see figure 15).

Figure 18

5.10 Datums and datum systems Datum features are features according to which a workpiece is aligned for recording the toleranced deviations. 5.10.1 When a toleranced feature is referred to a datum, this is generally shown by datum letters. The same letter which defines the datum is repeated in the tolerance frame. To identify the datum, a capital letter enclosed in a frame is connected to a datum triangle (see figures 19 a and 19 b).

Figure 15 Note: Whether a tolerance should be applied to the contour of a cylindrical or symmetrical feature or to its axis or median plane, depends on the functional requirements. ·

Figure 19 a

Figure 19 b

9


1 The datum triangle with the datum letter is placed: - on the outline of the feature or an extension of the outline (but clearly separated from the dimension line), when the datum feature is the line or surface itself (see figure 20).

A single datum is identified by a capital letter (see figure 26). A common datum formed by two datum features is identified by two datum letters separated by a hyphen (see figures 27 and 29). In a datum system (see also 5.10.2) the sequence of two or more datum features is important. The datum letters are to be placed in different compartments, where the sequence from left to right shows the order of priority, and the datum letter placed first should refer to the directional datum feature (see figures 28, 30 and 31).

Figure 20 - as an extension of the dimension line when the datum feature is the axis or median plane (see figures 21 and 22). Note: If there is not enough space for two arrows, one of them may be replaced by the datum triangle (see figure 22).

Figure 26

Figure 27 Secondary datum

Primary datum

Tertiary datum

Figure 28

Figure 21

Figure 22

- on the axis or median plane when the datum is: a) the axis or median plane of a single feature (for example a cylinder); b) the common axis or median plane formed by two features (see figure 23).

5.10.2 Datum system A datum system is a group of two or more datums to which one toleranced feature refers in common. A datum system is frequently required because the direction of a short axis cannot be determined alone. Datum formed by two form features (common datum):

Figure 29 Figure 23 If the tolerance frame can be directly connected with the datum feature by a leader line, the datum letter may be omitted (see figures 24 and 25).

Figure 24

10

Figure 25

Datum system formed by two datums (directional datum “A” and short axis “B”).

Figure 30 ·


1 Datum system formed by one plane and one perpendicular axis of a cylinder: Datum “A” is the plane formed by the plane surface. Datum “B” is the axis of the largest inscribed cylinder, the axis being at right angles with datum “A” (see figure 31).

larity tolerance specified within the tolerance frame (see figures 32 and 33).

Figure 31

Figure 32

5.11 Theoretically exact dimensions If tolerances of position or angularity are prescribed for a feature, the dimensions determining the theoretically exact position or angle shall not be toleranced. These dimensions are enclosed, for example 30 . The corresponding actual dimensions of the part are subject only to the position tolerance or angu-

Figure 33

5.12 Definitions of tolerances Symbol

Definition of the tolerance zone

Indication and interpretation

5.12.1 Straightness tolerance The tolerance zone when projected in a Any line on the upper surface parallel to the plane is limited by two parallel straight plane of projection in which the indication lines a distance t apart. is shown shall be contained between two parallel straight lines 0.1 apart.

Figure 34

Figure 35 Any portion of length 200 of any generator of the cylindrical surface indicated by the arrow shall be contained between two parallel straight lines 0.1 apart in a plane containing the axis.

Figure 36 ·

11


1 Symbol



The tolerance zone is limited by a parallel- The axis of the bar shall be contained within epiped of section t1 · t2 if the tolerance is a parallelepipedic zone of width 0.1 in the specified in two directions perpendicular vertical and 0.2 in the horizontal direction. to each other.

Figure 37

Figure 38

The tolerance zone is limited by a cylinder The axis of the cylinder to which the tolerof diameter t if the tolerance value is ance frame is connected shall be contained preceded by the symbol ∅. in a cylindrical zone of diameter 0.08.

Figure 40

Figure 39 5.12.2 Flatness tolerance

The tolerance zone is limited by two paral- The surface shall be contained between two lel planes a distance t apart. parallel planes 0.08 apart.

Figure 41

Figure 42

5.12.3 Circularity tolerance The tolerance zone in the considered The circumference of each cross-section of plane is limited by two concentric circles the outside diameter shall be contained a distance t apart. between two co-planar concentric circles 0.03 apart.

Figure 43

Figure 44 The circumference of each cross-section shall be contained between two co-planar concentric circles 0.1 apart.

Figure 45

12

·


1 Symbol



5.12.4 Cylindricity tolerance The tolerance zone is limited by two The considered surface area shall be coaxial cylinders a distance t apart. contained between two coaxial cylinders 0.1 apart.

Figure 46

Figure 47

5.12.5 Parallelism tolerance Parallelism tolerance of a line with reference to a datum line The tolerance zone when projected in a plane is limited by two parallel straight lines a distance t apart and parallel to the datum line, if the tolerance zone is only specified in one direction.

The toleranced axis shall be contained between two straight lines 0.1 apart, which are parallel to the datum axis A and lie in the vertical direction (see figures 49 and 50).

Figure 49 Figure 48

The toleranced axis shall be contained between two straight lines 0.1 apart, which are parallel to the datum axis A and lie in the horizontal direction.

Figure 51 The tolerance zone is limited by a parallelepiped of section t1 · t2 and parallel to the datum line if the tolerance is specified in two planes perpendicular to each other.

Figure 53 ·

Figure 50

Figure 52 The toleranced axis shall be contained in a parallelepipedic tolerance zone having a width of 0.2 in the horizontal and 0.1 in the vertical direction and which is parallel to the datum axis A (see figures 54 and 55).

Figure 54

Figure 55

13


1 Symbol



Parallelism tolerance of a line with reference to a datum line The tolerance zone is limited by a cylinder The toleranced axis shall be contained in a of diameter t parallel to the datum line if cylindrical zone of diameter 0.03 parallel to the tolerance value is preceded by the the datum axis A (datum line). symbol ∅.

Figure 57

Figure 56 Parallelism tolerance of a line with reference to a datum surface

The tolerance zone is limited by two paral- The toleranced axis of the hole shall be conlel planes a distance t apart and parallel to tained between two planes 0.01 apart and the datum surface. parallel to the datum surface B.

Figure 58

Figure 59

Parallelism tolerance of a surface with reference to a datum line The tolerance zone is limited by two paral- The toleranced surface shall be contained lel planes a distance t apart and parallel between two planes 0.1 apart and parallel to to the datum line. the datum axis C of the hole.

Figure 60

Figure 61

Parallelism tolerance of a surface with reference to a datum surface The tolerance zone is limited by two paral- The toleranced surface shall be contained lel planes a distance t apart and parallel between two parallel planes 0.01 apart and to the datum surface. parallel to the datum surface D (figure 63).

Figure 63 Figure 62

14

Figure 64

All the points of the toleranced surface in a length of 100, placed anywhere on this surface, shall be contained between two parallel planes 0.01 apart and parallel to the datum surface A (figure 64). ·


1 Symbol



5.12.6 Perpendicularity tolerance Perpendicularity tolerance of a line with reference to a datum line The tolerance zone when projected in a plane is limited by two parallel straight lines a distance t apart and perpendicular to the datum line.

The toleranced axis of the inclined hole shall be contained between two parallel planes 0.06 apart and perpendicular to the axis of the horizontal hole A (datum line).

Figure 65

Figure 66

Perpendicularity tolerance of a line with reference to a datum surface The tolerance zone when projected in a plane is limited by two parallel straight lines a distance t apart and perpendicular to the datum plane if the tolerance is specified only in one direction.

The toleranced axis of the cylinder, to which the tolerance frame is connected, shall be contained between two parallel planes 0.1 apart, perpendicular to the datum surface.

Figure 67

Figure 68

The tolerance zone is limited by a parallelepiped of section t1 · t2 and perpendicular to the datum surface if the tolerance is specified in two directions perpendicular to each other.

The toleranced axis of the cylinder shall be contained in a parallelepipedic tolerance zone of 0.1 · 0.2 which is perpendicular to the datum surface.

Figure 69

Figure 70

The tolerance zone is limited by a cylinder of diameter t perpendicular to the datum surface if the tolerance value is preceded by the symbol ∅.

The toleranced axis of the cylinder to which the tolerance frame is connected shall be contained in a cylindrical zone of diameter 0.01 perpendicular to the datum surface A.

Figure 71

Figure 72

·

15


1 Symbol



Perpendicularity tolerance of a surface with reference to a datum line The tolerance zone is limited by two The toleranced face of the workpiece shall parallel planes a distance t apart and be contained between two parallel planes perpendicular to the datum line. 0.08 apart and perpendicular to the axis A (datum line).

Figure 74

Figure 73

Perpendicularity tolerance of a surface with reference to a datum surface The tolerance zone is limited by two The toleranced surface shall be contained parallel planes a distance t apart and between two parallel planes 0.08 apart and perpendicular to the horizontal datum surperpendicular to the datum surface. face A.

Figure 76

Figure 75 5.12.7 Angularity tolerance Angularity tolerance of a line with reference to a datum line Line and datum line in the same plane. The tolerance zone when projected in a plane is limited by two parallel straight lines a distance t apart and inclined at the specified angle to the datum line.

The toleranced axis of the hole shall be contained between two parallel straight lines 0.08 apart which are inclined at 60° to the horizontal axis A - B (datum line).

Figure 77

Figure 78

Angularity tolerance of a surface with reference to a datum surface The tolerance zone is limited by two paral- The toleranced surface shall be contained lel planes a distance t apart and inclined between two parallel planes 0.08 apart which at the specified angle to the datum are inclined at 40° to the datum surface A. surface.

Figure 79

16

Figure 80 ·


1 Symbol



5.12.8 Positional tolerance Positional tolerance of a line The tolerance zone when projected in a plane is limited by two parallel straight lines a distance t apart and disposed symmetrically with respect to the theoretically exact position of the considered line if the tolerance is specified only in one direction.

Each of the toleranced lines shall be contained between two parallel straight lines 0.05 apart which are symmetrically disposed about the theoretically exact position of the considered line, with reference to the surface A (datum surface).

Figure 82 Figure 81

The axis of the hole shall be contained within a cylindrical zone of diameter 0.08 the axis of which is in the theoretically exact position of the considered line, with reference to the surfaces A and B (datum surfaces).

The tolerance zone is limited by a cylinder Figure 84 of diameter t the axis of which is in the theoretically exact position of the Each of the axes of the eight holes shall be considered line if the tolerance value is contained within a cylindrical zone of diameter 0.1 the axis of which is in the theoretically preceded by the symbol ∅. exact position of the considered hole, with reference to the surfaces A and B (datum surfaces).

Figure 83 Figure 85 Positional tolerance of a flat surface or a median plane The tolerance zone is limited by two parallel planes a distance t apart and disposed symmetrically with respect to the theoretically exact position of the considered surface.

Figure 86 ·

The inclined surface shall be contained between two parallel planes which are 0.05 apart and which are symmetrically disposed with respect to the theoretically exact position of the considered surface with reference to the datum surface A and the axis of the datum cylinder B (datum line).

Figure 87

17


1 Symbol



5.12.9 Concentricity and coaxiality tolerance Concentricity tolerance of a point The tolerance zone is limited by a circle of The centre of the circle, to which the tolerdiameter t the centre of which coincides ance frame is connected, shall be contained with the datum point. in a circle of diameter 0.01 concentric with the centre of the datum circle A.

Figure 88

Figure 89

Coaxiality tolerance of an axis The tolerance zone is limited by a cylinder of diameter t, the axis of which coincides with the datum axis if the tolerance value is preceded by the symbol ∅.

The axis of the cylinder, to which the tolerance frame is connected, shall be contained in a cylindrical zone of diameter 0.08 coaxial with the datum axis A - B.

Figure 90

Figure 91

5.12.10 Symmetry Symmetry tolerance of a median plane The tolerance zone is limited by two parallel planes a distance t apart and disposed symmetrically to the median plane with respect to the datum axis or datum plane.

Figure 92

The median plane of the slot shall be contained between two parallel planes, which are 0.08 apart and symmetrically disposed about the median plane with respect to the datum feature A.

Figure 93

Symmetry tolerance of a line or an axis The tolerance zone when projected in a plane is limited by two parallel straight lines a distance t apart and disposed symmetrically with respect to the datum axis (or datum plane) if the tolerance is specified only in one direction.

Figure 94

18

The axis of the hole shall be contained between two parallel planes which are 0.08 apart and symmetrically disposed with respect to the actual common median plane of the datum slots A and B.

Figure 95 ·


1 Symbol



Symmetry tolerance of a line or an axis The tolerance zone is limited by a parallelepiped of section t1 · t2, the axis of which coincides with the datum axis if the tolerance is specified in two directions perpendicular to each other.

The axis of the hole shall be contained in a parallelepipedic zone of width 0.1 in the horizontal and 0.05 in the vertical direction and the axis of which coincides with the datum axis formed by the intersection of the two median planes of the datum slots A - B and C - D.

Figure 96

Figure 97

5.12.11 Circular runout tolerance Circular runout tolerance - radial The tolerance zone is limited within any The radial runout shall not be greater than plane of measurement perpendicular to 0.1 in any plane of measurement during one the axis by two concentric circles a revolution about the datum axis A - B. distance t apart, the centre of which coincides with the datum axis.

Toleranced surface

Figure 99 The radial runout shall not be greater than 0.2 in any plane of measurement when measuring the toleranced part of a revolution about the centre line of hole A (datum axis).

Plane of measurement

Figure 98 Runout normally applies to complete revolutions about the axis but could be limited to apply to a part of a revolution. Figure 100

Figure 101

Circular runout tolerance - axial The tolerance zone is limited at any radial The axial runout shall not be greater than 0.1 position by two circles a distance t apart at any position of measurement during one lying in a cylinder of measurement, the revolution about the datum axis D. axis of which coincides with the datum axis. Cylinder of measurement

Figure 103 Figure 102 ·

19


1 Symbol



Circular runout tolerance in any direction The tolerance zone is limited within any cone of measurement, the axis of which coincides with the datum axis by two circles a distance t apart. Unless otherwise specified the measuring direction is normal to the surface.

The runout in the direction indicated by the arrow shall not be greater than 0.1 in any cone of measurement during one revolution about the datum axis C.

Cone of measurement

Figure 105 The runout in the direction perpendicular to the tangent of a curved surface shall not be greater than 0.1 in any cone of measurement during one revolution about the datum axis C.

Figure 104

Figure 106 Circular runout tolerance in a specified direction The tolerance zone is limited within any cone of measurement of the specified angle, the axis of which coincides with the datum axis by two circles a distance t apart.

The runout in the specified direction shall not be greater than 0.1 in any cone of measurement during one revolution about the datum axis C.

Figure 107 5.12.12 Total runout tolerance Total radial runout tolerance The tolerance zone is limited by two coaxial cylinders a distance t apart, the axes of which coincide with the datum axis.

The total radial runout shall not be greater than 0.1 at any point on the specified surface during several revolutions about the datum axis A-B, and with relative axial movement between part and measuring instrument. With relative movement the measuring instrument or the workpiece shall be guided along a line having the theoretically perfect form of the contour and being in correct position to the datum axis.

Figure 108 Figure 109

20

·


1 Symbol



Total axial runout tolerance The tolerance zone is limited by two par- The total axial runout shall not be greater allel planes a distance t apart and per- than 0.1 at any point on the specified surface pendicular to the datum axis. during several revolutions about the datum axis D and with relative radial movement between the measuring instrument and the part. With relative movement the measuring instrument or the workpiece shall be guided along a line having the theoretically perfect form of the contour and being in correct position to the datum axis.

Figure 110 Figure 111

·

21

Technical Drawings Sheet Sizes, Title Block, Non-standard Formats

1 Technical drawings, extract from DIN EN ISO 5457. 6. Sheet sizes The DIN EN ISO 5457 standard applies to the presentation of drawing forms even if they are

created by CAD. This standard may also be used for other technical documents. The sheet sizes listed below have been taken from DIN EN ISO 5457.

Table 3 Formats of trimmed and untrimmed sheets and of the drawing area 1)

Sheet sizes acc. to DIN EN ISO 5457, 5457 A series

Trimmed sheet a1 x b1 mm

mm

mm

A0

841 x 1189

821 x 1159

880 x 1230

A1

594 x 841

574 x 811

625 x 880

A2

420 x 594

400 x 564

450 x 625

A3

297 x 420

277 x 390

330 x 450

A4

210 x 297

180 x 277

240 x 330

1) The actually available drawing area is reduced by the title block, the filing margin, the possible sectioning margin, etc.

Drawing area

Drawing area a2 x b2

Untrimmed sheet a3 x b3

6.1 Title block Formats w A3 are produced in broadside. The title block area is in the bottom right corner of the trimmed sheet. For the A4 format the title block area is at the bottom of the short side (upright format).

Trimmed drawing sheet

Title block

6.2 Non-standard formats Non-standard formats should be avoided. When necessary they should be created using the

22

dimensions of the short side of an A-format with the long side of a greater A-format. ·

Technical Drawings Type Sizes, Lines Lettering Example

1 7. Type sizes Table 4: Type sizes for drawing formats (h = type height, b = line width) Paper sizes Application range g for lettering g Type, drawing no.

A 0 and A 1

A 2, A 3 and A 4

h

b

h

b

10

1

7

0.7

Texts and nominal dimensions

5

0.5

3.5

0.35

Tolerances, roughness values, symbols

3.5

0.35

2.5

0.25

7.1 The type sizes as assigned to the paper sizes in table 4 MUST be adhered to with regard to their application range. Larger type heights are

also permissible. Type heights smaller by approx. 20% will be accepted if this is required in a drawing because of restricted circumstances.

8. Lines acc. to DIN ISO 128, Part 20 and Part 24 Table 5: Line groups, line types and line widths Line group

0.5

Drawing format

0.7

A 4, A 3, A 2

Line type

A 1, A 0

Line width

Solid line (thick)

0.5

0.7

Solid line (thin)

0.25

0.35

Short dashes (thin)

0.25

0.35

Dot-dash line (thick)

0.5

0.7

Dot-dash line (thin)

0.25

0.35

Dash/double-dot line (thin)

0.25

0.35

Freehand (thin)

0.25

0.35

8.1 Line groups 0.5 and 0.7 with the pertaining line width according to table 5 may only be used. Assignment to the drawing formats A1 and

A 0 is prescribed. For the A4, A3 and A2 formats, line group 0.7 may be used as well.

9. Lettering example 9.1 Example for formats A 4 to A 2

DIN 332 - DS M24 DIN 509 E 2.5 x 0.4 DIN 509 F 2.5 x 0.4

·

23


Standardization

2

Page

ISO Metric Screw Threads (Coarse Pitch Threads)

25

ISO Metric Screw Threads (Coarse and Fine Pitch Threads)

26

Cylindrical Shaft Ends

27

ISO Tolerance Zones, Allowances, Fit Tolerances; Inside Dimensions (Holes)

28

ISO Tolerance Zones, Allowances, Fit Tolerances; Outside Dimensions (Shafts)

29

Parallel Keys, Taper Keys, and Centre Holes

30

24

·

Standardization ISO Metric Screw Threads (Coarse Pitch Threads)

ISO metric screw threads (coarse pitch threads) following DIN 13, Part 1 Nut

D1 + d * 2 H1 d 2 + D 2 + d * 0.64952 P d 3 + d * 1.22687 P H + 0.86603 P H 1 + 0.54127 P

Bolt

Nut thread diameter

h 3 + 0.61343 P R + H + 0.14434 P 6

Bolt thread diameter

Diameters of series 1 should be preferred to those of series 2, and these again to those of series 3. Pitch

Pitch diameter

d = D

P

d2 = D2

d3

D1

h3

H1

R

Tensile stress crosssection As 1)

Series 1 Series 2 Series 3

mm

mm

mm

mm

mm

mm

mm

mm2

0.5 0.6 0.7 0.75 0.8 1 1 1.25 1.25 1.5 1.5 1.75 2 2 2.5 2.5 2.5 3 3 3.5 3.5 4 4 4.5 4.5 5 5 5.5 5.5 6 6

2.675 3.110 3.545 4.013 4.480 5.350 6.350 7.188 8.188 9.026 10.026 10.863 12.701 14.701 16.376 18.376 20.376 22.051 25.051 27.727 30.727 33.402 36.402 39.077 42.077 44.752 48.752 52.428 56.428 60.103 64.103

2.387 2.764 3.141 3.580 4.019 4.773 5.773 6.466 7.466 8.160 9.160 9.853 11.546 13.546 14.933 16.933 18.933 20.319 23.319 25.706 28.706 31.093 34.093 36.479 39.479 41.866 45.866 49.252 53.252 56.639 60.639

2.459 2.850 3.242 3.688 4.134 4.917 5.917 6.647 7.647 8.376 9.376 10.106 11.835 13.835 15.294 17.294 19.294 20.752 23.752 26.211 29.211 31.670 34.670 37.129 40.129 42.587 46.587 50.046 54.046 57.505 61.505

0.307 0.368 0.429 0.460 0.491 0.613 0.613 0.767 0.767 0.920 0.920 1.074 1.227 1.227 1.534 1.534 1.534 1.840 1.840 2.147 2.147 2.454 2.454 2.760 2.760 3.067 3.067 3.374 3.374 3.681 3.681

0.271 0.325 0.379 0.406 0.433 0.541 0.541 0.677 0.677 0.812 0.812 0.947 1.083 1.083 1.353 1.353 1.353 1.624 1.624 1.894 1.894 2.165 2.165 2.436 2.436 2.706 2.706 2.977 2.977 3.248 3.248

0.072 0.087 0.101 0.108 0.115 0.144 0.144 0.180 0.180 0.217 0.217 0.253 0.289 0.289 0.361 0.361 0.361 0.433 0.433 0.505 0.505 0.577 0.577 0.650 0.650 0.722 0.722 0.794 0.794 0.866 0.866

5.03 6.78 8.78 11.3 14.2 20.1 28.9 36.6 48.1 58.0 72.3 84.3 115 157 193 245 303 353 459 561 694 817 976 1121 1306 1473 1758 2030 2362 2676 3055

Nominal thread diameter

3 3.5 4 4.5 5 6 7 8 9 10 11 12 14 16 18 20 22 24 27 30 33 36 39 42 45 48 52 56 60 64 68

Core diameter

1) The tensile stress cross-section is calculated acc. to DIN 13 Part 28 with formula ·

As =

Depth of thread

π 4

S

ǒd

2

+ d3 2

Ǔ

Round

2

25

2

Standardization ISO Metric Screw Threads (Coarse and Fine Pitch Threads) Selection of nominal thread diameters and pitches for coarse and fine pitch threads from 1 mm to 68 mm diameter, following DIN ISO 261

2

Nominal thread diameter d=D

Coarse itch pitch Series Series Series thread 1 2 3 1 1.2

Pitches P for fine pitch threads 4

3

2

0.25 0.25 0.3 0.35 0.35 0.4 0.45 0.45 0.5 0.6 0.7 0.8 1 1.25 1.5 1.75 2

1.4 1.6 1.8 2 2.2 2.5 3 3.5 4 5 6 8 10 12 14

1.5 1.5 1.5 1.5

15 16

1.5

2 17 18

20 22 24

2.5 2.5 2.5 3

2 2 2 2

3

2

3.5

2

3.5

2

25 26 27 28 30 32 33 35 36

4

3

2

4

3

2

4.5 4.5 5

3 3 3

2 2 2

5

3

2 2

38 39 40 42 45 48 50 52 55 56 60 64 68

26

58 65

5.5

4

3

2

5.5 6

4 4

3 3

6

4

3

2 2 2 2

1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5

1.25

1.25 1.25 1.25

1

1 1 1 1 1 1 1 1 1 1 1

0.75

0.75 0.75 0.75

0.5

0.5 0.5 0.5 0.5

1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5 1.5

·

Standardization Cylindrical Shaft Ends

Cylindrical shaft ends FLENDER works standard W 0470

Acc. to DIN 748/1 Diameter Series 1

2

Cylindrical shaft ends

ISO ISO Length tolertolerDiaL th ance Length ance meter zone Long Short zone

mm mm

mm

6

mm

mm

mm

16

FLENDER works standard W 0470

Acc. to DIN 748/1 Diameter Series 1

2

ISO ISO Length tolertolerDiaL th ance Length ance meter zone Long Short zone

mm mm

mm

mm

mm

100

210

165

100

mm m6 180

7

16

110

8

20

120

9

20

10

23

15

11

23

15

12

30

18

14 16

30 40

18 28

19 20 22

14 16

24 25

50 60

36 42

24 25

40

28 30

60 80

42 58

28 30

50

32 35 38

80 80 80

58 58 58

32 35 38

60

40 42

110 110

82 82

40 42

70

45 48 50

110 110 110

82 82 82

45 48 50

80

55

110

82

55

90

60 65

140 140

105 105

60 65

105

140 140

105 105

70 75

120

80 85

170 170

130 130

80 85

140

90 95

170 170

130 130

90 95

160

·

130

210 250

165 200

120 130

210

150

250 250

200 200

140 150

240

170

300 300

240 240

160 170

270

240 280 280

180 190 200

310

200

300 350 350

220

350

280

220

350

410 410 410

330 330 330

240 250 260

400

470

380

280

450

470 470

380 380

300 320

500

340

550

450

340

550

380

550 550

450 450

360 380

590

420

650 650

540 540

400 420

650

440

650

540

440

690

460

650 650

540 540

450 460

750

480

650 650

540 540

480 500

790

530

800 800 800 800

680 680 680 680

180

30

19 20 22

m6

110

160

28 36 36

70 75

165

140

40 50 50

k6

210

35

190

k6

240 250 260 280 300 320

360

400

m6 450

500

560 600 630

m6 6

n6

27

2

Standardization ISO Tolerance Zones, Allowances, Fit Tolerances Inside Dimensions (Holes) ISO tolerance zones, allowances, fit tolerances; Inside dimensions (holes) acc. to DIN 7157, DIN ISO 286 Part 2 μm + 500

2

Tolerance zones shown for nominal dimension 60 mm

+ 400 + 300 + 200 + 100 0 – 100 – 200 – 300 – 400 – 500

Nominal dimensions in mm

ISO Series 1 abbrev. Series 2 from to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to above to

1 3 3 6 6 10 10 14 14 18 18 24 24 30 30 40 40 50 50 65 65 80 80 100 100 120 120 140 140 160 160 180 180 200 200 225 225 250 250 280 280 315 315 355 355 400 400 450 450 500

ISO Series 1 abbrev. Series 2

28

H7 P7

N7

N9

M7

K7

J6

– 6 –16 – 8 –20 – 9 –24

– 4 –14 – 4 –16 – 4 –19

– 4 –29 0 –30 0 –36

– 2 –12 0 –12 0 –15

0 –10 + 3 – 9 + 5 –10

+ – + – + –

–11 –29

– 5 –23

0 –43

0 –18

+ 6 –12

–14 –35

– 7 –28

0 –52

0 –21

–17 –42

– 8 –33

0 –62

–21 –51

– 9 –39

–24 –59

2 4 5 3 5 4

H8

J7 + – + – + –

F8 H11

G7

E9

D10

C11

A11

+ + + + + +

D9

4 6 6 6 8 7

+10 0 +12 0 +15 0

+14 0 +18 0 +22 0

+ 60 0 + 75 0 + 90 0

+12 + 2 +16 4 +20 + 5

+ + + + + +

20 6 28 10 35 13

+120 + 60 +145 + 70 +170 + 80

+330 +270 +345 +270 +370 +280

+ 6 – 5

+10 – 8

+18 0

+27 0

+110 0

+24 + 6

+ 43 + 75 + 93 +120 +205 + 16 + 32 + 50 + 50 + 95

+400 +290

+ 6 –15

+ 8 – 5

+12 – 9

+21 0

+33 0

+130 0

+28 + 7

+ 53 + 92 +117 +149 +240 + 20 + 40 + 65 + 65 +110

+430 +300

0 –25

+ 7 –18

+10 – 6

+14 –11

+25 0

+39 0

+160 0

+34 + 9

0 –74

0 –30

+ 9 –21

+13 – 6

+18 –12

+30 0

+46 0

+190 0

+40 +10

–10 –45

0 –87

0 –35

+10 –25

+16 – 6

+22 –13

+35 0

+54 0

+220 0

+47 +12

–28 –68 68

–12 –52 52

0 –100 100

0 –40 40

+12 –28 28

+18 – 7

+26 –14 14

+40 0

+63 0

+250 0

+54 +14

–33 79 –79

–14 60 –60

0 115 –115

0 46 –46

+13 33 –33

+22 – 7

+30 16 –16

+46 0

+72 0

+290 0

+61 +15

–36 –88

–14 –66

0 –130

0 –52

+16 –36

+25 – 7

+36 –16

+52 0

+81 0

+320 0

+69 +17

–41 –98

–16 –73

0 –140

0 –57

+17 –40

+29 – 7

+39 –18

+57 0

+89 0

+360 0

+75 +18

– 45 –108

–17 –80

0 –155

0 –63

+18 –45

+33 – 7

+43 –20

+63 0

+97 0

+400 0

+83 +20

+280 + 64 +112 +142 +180 +120 + 25 + 50 + 80 + 80 +290 +130 +330 + 76 +134 +174 +220 +140 + 30 + 60 +100 +100 +340 +150 +390 + 90 +159 +207 +260 +170 + 36 + 72 +120 +120 +400 +180 +450 +200 +106 +185 +245 +305 +460 + 43 + 85 +145 +145 +210 +480 +230 +530 +240 +122 +215 +285 +355 +550 + 50 +100 +170 +170 +260 +570 +280 +620 +137 +240 +320 +400 +300 + 56 +110 +190 +190 +650 +330 +720 +151 +265 +350 +440 +360 + 62 +125 +210 +210 +760 +400 +840 +165 +290 +385 +480 +440 + 68 +135 +230 +230 +880 +480

+470 +310 +480 +320 +530 +340 +550 +360 +600 +380 +630 +410 +710 +460 +770 +520 +830 +580 +950 +660 +1030 + 740 +1110 + 820 +1240 + 920 +1370 +1050 +1560 +1200 +1710 +1350 +1900 +1500 +2050 +1650

H7

H8

P7

N7

N9

M7

K7

J6

J7

H11

G7

F8

+ + + + + +

39 14 50 20 61 25

+ + + + + +

45 20 60 30 76 40

E9

60 20 78 30 98 40

D10 D9

C11 A11

·

Standardization ISO Tolerance Zones, Allowances, Fit Tolerances Outside Dimensions (Shafts) ISO tolerance zones, allowances, fit tolerances; Outside dimensions (shafts) acc. to DIN 7157, DIN ISO 286 Part 2 μm + 500

2

Tolerance zones shown for nominal dimension 60 mm

+ 400 + 300 + 200 + 100 0 – 100 – 200 – 300 – 400 – 500 ISO Series 1 x8/u8 abbrev. Series 2 1)

Nominal dimensions in mm

from 1 to 3 above 3 to 6 above 6 to 10 above 10 to 14 above 14 to 18 above 18 to 24 above 24 to 30 above 30 to 40 above 40 to 50 above 50 to 65 above 65 to 80 above 80 to 100 above 100 to 120 above 120 to 140 above 140 to 160 above 160 to 180 above 180 to 200 above 200 to 225 above 225 to 250 above 250 to 280 above 280 to 315 above 315 to 355 above 355 to 400 above 400 to 450 above 450 to 500 ISO Series 1 abbrev. Series 2

r6 s6 + + + + + +

20 14 27 19 32 23

n6

r5

+ 34 + 20 + 46 + 28 + 56 + 34 + 67 + 40 + 72 + 45 + 87 + 54 + 81 + 48 + 99 + 60 +109 + 70 +133 + 87 +148 +102 +178 +124 +198 +144 +233 +170 +253 +190 +273 +210 +308 +236 +330 +258 +356 +284 +396 +315 +431 +350 +479 +390 +524 +435 +587 +490 +637 +540

+ + + + + +

14 10 20 15 25 19

+ 72 + 53 + 78 + 59 + 93 + 71 +101 + 79 +117 + 92 +125 +100 +133 +108 +151 +122 +159 +130 +169 +140 +190 +158 +202 +170 +226 +190 +244 +208 +272 +232 +292 +252

+ 54 + 41 + 56 + 43 + 66 + 51 + 69 + 54 + 81 + 63 + 83 + 65 + 86 + 68 + 97 + 77 +100 + 80 +104 + 84 +117 + 94 +121 + 98 +133 +108 +139 +114 +153 +126 +159 +132

x8/u8 1)

s6

r5

h6 m5 m6 k5

+ + + + + +

16 10 23 15 28 19

+10 + 4 +16 + 8 +19 +10

+ 6 + 2 + 9 + 4 +12 + 6

+ 8 + 2 +12 + 4 +15 + 6

+ 4 0 + 6 + 1 + 7 + 1

k6 + 6 0 + 9 + 1 +10 + 1

j6 + – + – + –

4 2 6 2 7 2

js6

h9 h7 h8

f7 h11

+ 3 0 0 0 0 0 – 3 – 6 –10 –14 – 25 – 60 + 4 0 0 0 0 0 – 4 – 8 –12 –18 – 30 – 75 +4.5 0 0 0 0 0 –4.5 – 9 –15 –22 – 36 – 90

g6 – 2 – 8 – 4 –12 – 5 –14

e8 – – – – – –

6 16 10 22 13 28

– – – – – –

14 28 20 38 25 47

d9 – – – – – –

20 45 30 60 40 76

c11

a11

– 60 –120 – 70 –145 – 80 –170

–270 –330 –270 –345 –280 –370

+ 39 + 31 + 34 +23 +15 +18 + 9 +12 + 8 +5.5 0 0 0 0 0 – 6 – 16 – 32 – 50 – 95 –290 + 28 + 23 + 23 +12 + 7 + 7 + 1 + 1 – 3 –5.5 –11 –18 –27 – 43 –110 –17 – 34 – 59 – 93 –205 –400

+ 48 + 37 + 41 +28 +17 +21 +11 +15 + 9 +6.5 0 0 0 0 0 – 7 – 20 – 40 – 65 –110 –300 + 35 + 28 + 28 +15 + 8 + 8 + 2 + 2 – 4 –6.5 –13 –21 –33 – 52 –130 –20 – 41 – 73 –117 –240 –430

+ 59 + 45 + 50 +33 +20 +25 +13 +18 +11 + 43 + 34 + 34 +17 + 9 + 9 + 2 + 2 – 5 + 60 + 41 + 62 + 43 + 73 + 51 + 76 + 54 + 88 + 63 + 90 + 65 + 93 + 68 +106 + 77 +109 + 80 +113 + 84 +126 + 94 +130 + 98 +144 +108 +150 +114 +166 +126 +172 +132 r6

+8 –8

+39 +24 +30 +15 +21 +12 +9.5 +20 +11 +11 + 2 + 2 – 7 –9.5

+45 +28 +35 +18 +25 +13 +11 +23 +13 +13 + 3 + 3 – 9 –11

+52 +33 +40 +21 +28 +14 +12.5 +27 +15 +15 + 3 + 3 –11 11 –12.5 12.5

+60 +37 +46 +24 +33 +16 +14.5 +31 +17 +17 + 4 + 4 –13 13 –14.5 14.5

+66 +43 +52 +27 +36 +16 +16 +34 +20 +20 + 4 + 4 –16 –16

+73 +46 +57 +29 +40 +18 +18 +37 +21 +21 + 4 + 4 –18 –18

+80 +50 +63 +32 +45 +20 +20 +40 +23 +23 + 5 + 5 –20 –20 n6 m5 m6 k5

k6

j6

js6

–120 –310 0 0 0 0 0 – 9 – 25 – 50 – 80 –280 –470 –16 –25 –39 – 62 –160 –25 – 50 – 89 –142 –130 –320 –290 –480 –140 –340 0 –10 – 30 – 60 –100 –330 –530 0 0 0 0 –19 –30 –46 – 74 –190 –29 – 60 –106 –174 –150 –360 –340 –550 –170 –380 0 0 0 0 0 –12 – 36 – 72 –120 –390 –600 –22 –35 –54 – 87 –220 –34 – 71 –126 –207 –180 –410 –400 –630 –200 –460 –450 –710 0 0 0 0 0 –14 – 43 – 85 –145 –210 –520 –25 25 –40 40 –63 63 –100 100 –250 250 –39 39 – 83 –148 148 –245 245 –460 –770 –230 –580 –480 –830 –240 –660 –530 –950 0 0 0 0 0 –15 – 50 –100 –170 –260 – 740 29 –46 46 –72 72 –115 115 –290 290 –44 44 – 96 –172 172 –285 285 –550 –1030 –29 –280 – 820 –570 –1100 –300 – 920 0 –17 – 56 –110 –190 –620 –1240 0 0 0 0 –32 –52 –81 –130 –320 –49 –108 –191 –320 –330 –1050 –650 –1370 –360 –1200 0 0 0 0 0 –18 – 62 –125 –210 –720 –1560 –36 –57 –89 –140 –360 –54 –119 –214 –350 –400 –1350 –760 –1710 –440 –1500 0 –20 – 68 –135 –230 –840 –1900 0 0 0 0 –40 –63 –97 –155 –400 –60 –131 –232 –385 –480 –1650 –880 –2050 h6 h9 f7 h7 h8 h11 g6 e8 d9 c11 a11

1) Up to nominal dimension 24 mm: x8; above nominal dimension 24 mm: u8 ·

29

Standardization Parallel Keys, Taper Keys, and Centre Holes Dimensions of parallel keys and taper keys

2

Depth of keyDiameter Width Height way in shaft d b h t1

above to

1)

2)

Depth of keyway in hub t2 DIN 6885/1 6886/ 6887

Parallel keys and taper keys acc. to DIN 6885 Part 1, 6886 and 6887

Lengths, see below l1

Side fitting square and rectangular keys

l DIN

6885/1

6886

2)

mm mm mm 6 8 2 8 10 3 10 12 4 12 17 5 17 22 6 22 30 8 30 38 10 38 44 12 44 50 14 50 58 16 58 65 18 65 75 20 75 85 22 85 95 25 95 110 28 110 130 32 130 150 36 150 170 40 170 200 45 200 230 50 230 260 56 260 290 63 290 330 70 330 380 80 380 440 90 440 500 100 Lengths mm I1 or I

from to from to mm mm mm mm mm mm mm mm 2 1.2 1.0 0.5 6 20 6 20 3 1.8 1.4 0.9 6 36 8 36 Parallel key and keyway acc. to DIN 6885 Part 1 4 2.5 1.8 1.2 8 45 10 45 5 3 2.3 1.7 10 56 12 56 Square and rectangular taper keys 6 3.5 2.8 2.2 14 70 16 70 7 4 3.3 2.4 18 90 20 90 8 5 3.3 2.4 22 110 25 110 8 5 3.3 2.4 28 140 32 140 9 5.5 3.8 2.9 36 160 40 160 10 6 4.3 3.4 45 180 45 180 11 7 4.4 3.4 50 200 50 200 12 7.5 4.9 3.9 56 220 56 220 Taper and round-ended round ended sunk key and 14 9 5.4 4.4 63 250 63 250 keyway acc. to DIN 6886 14 9 5.4 4.4 70 280 70 280 16 10 6.4 5.4 80 320 80 320 1) The tolerance zone for hub keyway width b for 18 11 7.4 6.4 90 360 90 360 parallel keys with normal fit is ISO JS9 and 20 12 8.4 7.1 100 400 100 400 with close fit ISO P9. The tolerance zone for 22 13 9.4 8.1 110 400 110 400 shaft keyway width b with normal fit is ISO N9 25 15 10.4 9.1 125 400 125 400 and with close fit ISO P9. 28 17 11.4 10.1 140 400 140 400 2) Dimension h of the taper key names the 32 20 12.4 11.1 160 400 largest height of the key, and dimension tz the 32 20 12.4 11.1 180 400 largest depth of the hub keyway. The shaft Lengths 36 22 14.4 13.1 200 400 keyway and hub keyway dimensions not 40 25 15.4 14.1 220 400 deteraccording to DIN 6887 - ta taper er keys with gib 45 28 17.4 16.1 250 400 mined head - are equal to those of DIN 6886. 50 31 19.5 18.1 280 400 6 8 10 12 14 16 18 20 22 25 28 32 36 40 45 50 56 63 70 80 90 100 110 125 140 160 180 200 220 250 280 320 360 400

Dimensions of 60° centre holes in mm Recommended Bore diameters diameter d 2) d1 a 1) above to 6 10 1.6 5.5 10 25 2 6.6 2.5 8.3 3.15 10 25 63 4 12.7 5 15.6 63 100 6.3 20 Recommended diameters d6 2) d1 d2 d3 3) above to 7 10 M3 2.5 3.2 10 13 M4 3.3 4.3 13 16 M5 4.2 5.3 16 21 M6 5 6.4 21 24 M8 6.8 8.4 24 30 M10 8.5 10.5 30 38 M12 10.2 13 38 50 M16 14 17 50 85 M20 17.5 21 85 130 M24 21 25 130 225 M30* 26 31 225 320 M36* 31.5 37 320 500 M42* 37 43

30

b

d2

d3

Minimum dimensions t

0.5 0.6 0.8 0.9 1.2 1.6 1.4

3.35 4.25 5.3 6.7 8.5 10.6 13.2

5 6.3 8 10 12.5 16 18

3.4 4.3 5.4 6.8 8.6 10.8 12.9

Form B

Centre holes i shaft in h ft ends d (centerings) ( t i ) acc. tto DIN 332 Part P t1

Form B DIN 332/1

Form DS d4

d5

5.3 6.7 8.1 9.6 12.2 14.9 18.1 23 28.4 34.2 44 55 65

5.8 7.4 8.8 10.5 13.2 16.3 19.8 25.3 31.3 38 48 60 71

t1 t2 +2 min. 9 12 10 14 12.5 17 16 21 19 25 22 30 28 37 36 45 42 53 50 63 60 77 74 93 84 105

t3 +1 2.6 3.2 4 5 6 7.5 9.5 12 15 18 17 22 26

t4 ≈ 1.8 2.1 2.4 2.8 3.3 3.8 4.4 5.2 6.4 8 11 15 19

t5 ≈ 0.2 0.3 0.3 0.4 0.4 0.6 0.7 1.0 1.3 1.6 1.9 2.3 2.7

K Keyway

Form DS (with thread) DIN 332/2

1) 2) * 3)

Cutting-off dimension in case of no centering Diameter applies a lies to finished workpiece work iece Dimensions not acc. to DIN 332 Part 2 Drill diameter for tapping-size holes acc. to DIN 336 Part 1

·


Physics

Page

Internationally Determined Prefixes

32

Basic SI Units

32

Derived SI Units Having Special Names and Special Unit Symbols

33

Legal Units Outside the SI

33

Physical Quantities and Units of Lengths and Their Powers

34

Physical Quantities and Units of Time

35

Physical Quantities and Units of Mechanics

35 – 37

Physical Quantities and Units of Thermodynamics and Heat Transfer

37 + 38

Physical Quantities and Units of Electrical Engineering

38

Physical Quantities and Units of Lighting Engineering

39

Different Measuring Units of Temperature

39

Measures of Length

40

Square Measures

40

Cubic Measures

41

Weights

41

Energy, Work, Quantity of Heat

41

Power, Energy Flow, Heat Flow

42

Pressure and Tension

42

Velocity

42

Equations for Linear Motion and Rotary Motion

43

·

31

3

Physics Internationally Determined Prefixes Basic SI Units Internationally determined prefixes Decimal multiples and sub-multiples of units are represented with prefixes and symbols. Prefixes and symbols are used only in combination with unit names and unit symbols.

3

Factor by which the unit is multiplied

Prefix

Symbol

Factor by which the unit is multiplied

Prefix

Symbol

10-18

Atto

a

10 1

Deka

da

10-15

Femto

f

10 2

Hecto

h

10-12

Pico

p

10 3

Kilo

k

10-9

Nano

n

10 6

Mega

M

10-6

Micro

μ

10 9

Giga

G

10-3

Milli

m

10 12

Tera

T

10-2

Centi

c

10 15

Peta

P

10-1

Deci

d

10 18

Exa

E

– Prefix symbols and unit symbols are written – When giving sizes by using prefix symbols and without blanks and together they form the unit symbols, the prefixes should be chosen in symbol for a new unit. An exponent on the unit such a way that the numerical values are symbol also applies to the prefix symbol. between 0.1 and 1000. Example:

Example: 12 kN 3.94 mm 1.401 kPa 31 ns

1 cm3 = 1 . (10-2m)3 = 1 . 10-6m3 1 μs = 1 . 10-6s 106s-1 = 106Hz = 1 MHz

instead of instead of instead of instead of

1.2 . 104N 0.00394 m 1401 Pa 3.1 . 10-8s

– Prefixes are not used with the basic SI unit kilo- – Combinations of prefixes and the following units are not allowed: gram (kg) but with the unit gram (g). Units of angularity: degree, minute, second Example: Milligram (mg), NOT microkilogram (μkg). Units of time: minute, hour, year, day Unit of temperature: degree Celsius Basic SI units Physical quantity

Basic SI unit

Physical quantity

Name

Symbol

Metre

m

Mass

Kilogram

kg

Time

Second

s

Amount of substance

Electric current

Ampere

A

Luminous intensity

Length

32

Thermodynamic temperature

Basic SI unit Name

Symbol

Kelvin

K

Mol

mol

Candela

cd

·

Physics Derived SI Units Legal Units Outside the SI Derived SI units having special names and special unit symbols SI unit

Physical quantity

Relation

Name

Plane angle

Radian

rad

1 rad = 1 m / m

Solid angle

Steradian

sr

1 sr = 1 m2 / m2

Hertz

Hz

1 Hz = 1 s-1

Force

Newton

N

1 N = 1 kg . m / s2

Pressure, mechanical stress

Pascal

Pa

1 Pa = 1 N/m2 = 1 kg / (m . s2)

Energy; work; quantity of heat

Joule

J

1 J = 1 N . m = 1 W . s = 1 kg . m2 / s2

Power, heat flow

Watt

W

1 W = 1 J/s = 1 kg . m2 / s3

Coulomb

C

1C = 1A.s

Volt

V

1 V = 1 J/C = 1 (kg . m2) / (A . s3)

Electric capacitance

Farad

F

1 F = 1 C/V = 1 (A2 . s4) / (kg . m2)

Electric resistance

Ohm

Ω

1 Ω = 1 V/A = 1 (kg . m2) / A2 . s3)

Electric conductance

Siemens

S

1 S = 1 Ω-1 = 1 (A2 . s3) / (kg . m2)

Celsius temperature

degrees Celsius

°C

0 °C = 273.15 K Δ 1 °C = Δ 1 K

Henry

H

1 H = 1 V . s/A

Frequency, cycles per second

Electric charge Electric potential

Inductance

3

Legal units outside the SI Physical quantity

Plane angle

Volume

Time

Unit name Round angle Gon Degree Minute Second Litre Minute Hour Day Year

Unit symbol 1)

gon ° ’ ’’

2) 2) 2)

1 perigon = 2 π rad 1 gon = (π / 200) rad 1° = (π / 180) rad 1’ = (1/60)° 1’’ = (1/60)’ 1 l = 1 dm3 = (1/1000) m3

l min h d a

Definition

2) 2) 2) 2)

1 min = 60 s 1 h = 60 min = 3 600 s 1 d = 24 h = 86 400 s 1 a = 365 d = 8 760 h

Mass

Ton

t

1 t = 103 kg = 1 Mg

Pressure

Bar

bar

1 bar = 105 Pa

1) A symbol for the round angle has not been internationally determined 2) Do not use with prefixes ·

33

Physics Physical Quantities and Units of Lengths and Their Powers Physical quantities and units of lengths and their powers Physical quantity

SI unit Symbol Name

N.: Note L.U.: Further legal units N.A.: Units no longer allowed

l

Length

m (metre)

N.: Basic unit L.U.: μm; mm; cm; dm; km; etc. N.A.: micron (μ): 1 μ = 1 μm Ångström (Å): 1 Å = 10-10 m

A

Area

m2 (square metre)

L.U.: mm2; cm2; dm2; km2 are (a): 1 a = 102 m2 hectare (ha): 1 ha = 104 m2

V

Volume

m3 (cubic metre)

H

Moment of area

m3

N.: moment of a force; moment of resistance L.U.: mm3; cm3

Ι

Second moment of area

m4

N.: formerly: geometrical moment of inertia L.U.: mm4; cm4

Symbol

3

L.U.: mm3; cm3; dm3 litre (l): 1 l = 1 dm3

1 m (arc) 1m = = 1m/m 1 m (radius) 1m

N.: 1 rad =

1 rad

1 degree + 1 o + rad 180 90 o + rad 2

α β γ

L.U. : rad, mrad Plane angle

Degree ( o) : 1 o + rad 180 o 1 Minute (Ȁ) : 1Ȁ + 60 Second (ȀȀ) : 1ȀȀ + 1Ȁ 60 Gon (gon) : 1 gon + rad 200

rad (radian)

N.A. : Right angle (L) : 1L + rad 2 Centesimal degree (g) : 1g + 1 gon Centesimal minute ( c) : 1 c + 1 gon 100 c cc cc Centesimal second ( ) : 1 + 1 100 Ω ω

34

Solid angle

sr (steradian)

N.: 1 sr =

1 m2 (spherical surface) 1

m2

(square of spherical radius)

= 1

m2 m2

·

Physics Physical Quantities and Units of Time and of Mechanics Physical quantities and units of time Symbol

Physical quantity

SI unit Symbol Name

N.: Note L.U.: Further legal units N.A.: Units no longer allowed N.: Basic unit L.U.: ns; μs; ms; ks Minute (min): 1 min = 60 s Hour (h): 1 h = 60 min Day (d): 1 d = 24 h Year (a): 1 a = 365 d (Do not use prefixes for decimal multiples and sub-multiples of min; h; d; a)

t

Time, Period, Duration

s (second)

f

Frequency, Periodic frequency

Hz (Hertz)

n

Rotational frequency (speed)

s-1

v

Velocity

m/s

L.U.: cm/s; m/h; km/s; km/h 1 kmńh + 1 mńs 3.6

a

Acceleration, linear

m/s2

N.: Time-related velocity L.U.: cm/s2

g

Gravity

m/s2

ω

Angular velocity

rad/s

L.U.: rad/min

α

Angular acceleration

rad/s2

L.U.: °/s2

Volume flow rate

m3/s

.

V

3

L.U.: kHz; MHz; GHz; THz Hertz (Hz): 1 Hz = 1/s N.:

Reciprocal value of the duration of one revolution L.U.: min-1 = 1/min

N.:

Gravity varies locally. Normal gravity (gn): gn = 9.80665 m/s2 ≈ 9.81 m/s2

L.U.: l/s; l/min; dm3/s; l/h; m3/h; etc.

Physical quantities and units of mechanics Symbol

Physical quantity

SI unit Symbol Name

m

Mass

kg (kilogram)

m’

Mass per unit length

kg/m

N.: m’ = m/l L.U.: mg/m; g/km In the textile industry: Tex (tex): 1 tex = 10-6 kg/m = 1 g/km

m’’

Mass in relation to the surface

kg/m2

N.: m’’ = m/A L.U.: g/mm2; g/m2; t/m2

·

N.: Note L.U.: Further legal units N.A.: Units no longer allowed N.: Basic unit L.U.: μg; mg; g; Mg ton (t): 1 t = 1000 kg

35

Physics Physical Quantities and Units of Mechanics Physical quantities and units of mechanics (continued) Physical quantity

SI unit Symbol Name

Density

kg/m3

J

Mass moment of inertia; second mass moment

.

Symbol

3

36

N.: Note L.U.: Further legal units N.A.: Units no longer allowed N.: = m/V L.U.: g/cm3; kg/dm3; Mg/m3; t/m3; kg/l 1g/cm3 = 1 kg/dm3 = 1 Mg/m3 = 1 t/m3 = 1 kg/l Instead of the former flywheel effect GD2 2 GD 2 in kpm 2 now : J + GD 4 L.U.: g . m2; t . m2

N.: kg

m2

m

.

Rate of mass flow

kg/s

F

Force

N (Newton)

L.U.: μN; mN; kN; MN; etc.; 1 N = 1 kg m/s2 N.A.: kp (1 kp = 9.80665 N)

G

Weight

N (Newton)

N.: Weight = mass acceleration due to gravity L.U.: kN; MN; GN; etc.

M, T

Torque

Nm

L.U.: μNm; mNm; kNm; MNm; etc. N.A.: kpm; pcm; pmm; etc.

Mb

Bending moment

Nm

L.U.: Nmm; Ncm; kNm; etc. N.A.: kpm; kpcm; kpmm; etc.

L.U.: kg/h; t/h

N.: 1 Pa = 1 N/m2 L.U.: Bar (bar): 1 bar = 100 000 Pa = 105 Pa μbar; mbar N.A.: kp/cm2; at; ata; atü; mmWS; mmHg; Torr 1kp/cm2 = 1 at = 0.980665 bar 1 atm = 101 325 Pa = 1.01325 bar 1 Torr + 101325 Pa + 133.322 Pa 760 1 mWS = 9806.65 Pa = 9806.65 N/m2 1 mmHg = 133.322 Pa = 133.322 N/m2

p

Pressure

Pa (Pascal)

pabs

Absolute pressure

Pa (Pascal)

pamb

Ambient atmospheric pressure

Pa (Pascal)

pe

Pressure above atmospheric

Pa (Pascal)

σ

Direct stress (tensile and compressive stress)

N/m2

L.U.: N/mm2 1 N/mm2 = 106 N/m2 = 1 MPa

τ

Shearing stress

N/m2

L.U.: N/mm2

ε

Extension

m/m

N.: Δl / l L.U.: μm/m; cm/m; mm/m

pe = pabs – pamb

·

Physics Physical Quantities and Units of Mechanics, Thermodynamics and Heat Transfer Physical quantities and units of mechanics (continued) Symbol

Physical quantity

W, A

Work

E, W

Energy

P

Power

.

SI unit Symbol Name


J (Joule)

N.: 1 J = 1 Nm = 1 Ws L.U.: mJ; kJ; MJ; GJ; TJ; kWh 1 kWh = 3 3.6 6 MJ N.A.: kpm; cal; kcal 1 cal = 4.1868 J; 860 kcal = 1 kWh

W (Watt)

N.: 1 W = 1 J/s = 1 Nm/s L.U.: μW; mW; kW; MW; etc. kJ/s; kJ/h; MJ/h; etc. PS; kkpm/s; N.A.: PS m/s kcal/h 1 PS = 735.49875 W 1 kpm/s = 9.81 W 1 kcal/h = 1.16 W 1 hp = 745.70 W

Q

Heat flow

η

Dynamic viscosity

Pa . s

N.: 1 Pa . s = 1 Ns/m2 L.U.: dPa . s; mPa . s N.A.: Poise (P): 1 P = 0.1 Pa . s

Kinematic viscosity

m2/s

L.U.: mm2/s; cm2/s N.A.: Stokes (St): 1 St = 1/10 000 m2/s 1cSt = 1 mm2/s

ν

3

Physical quantities and units of thermodynamics and heat transfer Symbol

Physical quantity

SI unit Symbol Name

T

Thermodynamic temperature

K (Kelvin)

t

Celsius temperature

°C

Q

Heat, Quantity of heat

J

N.: Note L.U.: Further legal units N.A.: Units no longer allowed N.:

Basic unit 273.15 K = 0 °C 373.15 K = 100 °C L.U.: mK

N.:

The degrees Celsius (°C) is a special name for the degrees Kelvin (K) when stating Celsius temperatures. The temperature interval of 1 K equals that of 1 °C.

1 J = 1 Nm = 1 Ws L.U.: mJ; kJ; MJ; GJ; TJ N.A.: cal; kcal a=

a

Temperature conductivity

m2/s

λ .

λ [ W / (m . K) ] = thermal conductivity [ kg / m3 ]

= density of the body

.

[ J / (kg K) ] = specific heat capacity at constant pressure ·

37

Physics Physical Quantities and Units of Thermodynamics, Heat Transfer and Electrical Engineering Physical quantities and units of thermodynamics and heat transfer (continued) Symbol

Physical quantity

SI unit Symbol Name


H

Enthalpy (Heat content)

J

s

Entropy

J/K

α h

Heat transfer coefficient

W / (m2 . K)

c

Specific heat capacity

J / (K . kg)

αl

Coefficient of linear thermal expansion

K-1

m / (m . K) = K-1 N.: Temperature unit/length unit ratio L.U.: μm / (m . K); cm / (m . K); mm / (m . K)

αv γ

Coefficient of volumetric expansion

K-1

m3 / (m3 . K) = K-1 N.: Temperature unit/volume ratio N.A.: m3 / (m3 . deg)

N.:

3

Quantity of heat absorbed under certain conditions L.U.: kJ; MJ; etc. N.A.: cal; Mcal; etc. 1 J/K = 1 Ws/K = 1 Nm/K L.U.: kJ/K N.A.: kcal/deg; kcal/°K L.U.: W / (cm2 . K); kJ / (m2 . h . K) N.A.: cal / (cm2 . s . grd) kcal / (m2 . h . grd) ≈ 4.2 kJ / (m2 . h . K) 1 J / (K . kg) = W . s / (kg . K) N.: Heat capacity referred to mass N.A.: cal / (g . grd); kcal / (kg . grd); etc.

Physical quantities and units of electrical engineering Symbol

Physical quantity

SI unit Symbol Name

I

Current strength

A (Ampere)

N.: Basic unit L.U.: pA; nA; μA; mA; kA; etc.

Q

Electric charge; Quantity of electricity

C (Coulomb)

1C = 1A.s 1 Ah = 3600 As L.U.: pC; nC; μC; kC

U

Electric voltage

V (Volt)

1 V = 1 W / A = 1 J / (s . A) = 1 A . Ω = 1 N . m / (s . A) L.U.: μV; mV; kV; MV; etc.

R

Electric resistance

Ω (Ohm)

1 Ω = 1 V / A = 1 W / A2 1 J / (s . A2) = 1 N . m / (s . A2) L.U.: μΩ; mΩ; kΩ; etc.

G

Electric conductance

S (Siemens)

C

Electric capacitance

F (Farad)

38


N.:

Reciprocal of electric resistance 1 S = 1 Ω-1 = 1 / Ω; G = 1 / R L.U.: μS; mS; kS 1 F = 1 C/V = 1 A . s/V = 1 A2 . s / W = 1 A2 . s2 / J = 1 A2 . s2 / (N . m) L.U.: pF; μF; etc. ·

Physics Physical Quantities and Units of Lighting Engineering, Different Measuring Units of Temperature Physical quantities and units of lighting engineering Symbol

Physical quantity

SI unit Symbol Name

Ι

Luminous intensity

cd (Candela)

L

Luminous density; Luminance

Φ

Luminous flux

lm (Lumen)

E

Illuminance

lx (Lux)

cd /

N.: Note L.U.: Further legal units N.A.: Units no longer allowed N.:

Basic unit 1 cd = 1 lm (lumen) / sr (Steradian) L.U.: mcd; kcd

3

L.U.: cd / cm2; mcd / m2; etc. 1 cdńm 2 1 asb + N.A.: Apostilb (asb):

m2

Nit (nt): Stilb (sb):

1 nt = 1 cd / m2 1 sb = 104 cd / m2

1 Im = 1 cd . sr L.U.: klm 1 lx = 1 lm / m2

Different measuring units of temperature Kelvin K TK

Degrees Celsius °C tC

Degrees Fahrenheit °F tF

Degrees Rankine °R TR

T K + 273.15 ) t C

t C + T K * 273.15

t F + 9 @ T K * 459.67 5

TR + 9 @ TK 5

T K + 255.38 ) 5 @ t F 9

t C + 5 ǒt F * 32Ǔ 9

t F + 32 ) 9 @ t C 5

T R + 9 ǒt c ) 273.15 Ǔ 5

TK + 5 @ TR 9

t C + 5 T R * 273.15 9

t F + T R * 459.67

T R + 459.67 ) t F

Comparison of some temperatures 0.00 + 255.37 + 273.15 + 273.16 + 373.15

1)

– 273.15 – 17.78 0.00 + 0.01 + 100.00

1)

– 459.67 0.00 + 32.00 + 32.02 + 212.00

0.00 + 459.67 + 491.67 + 491.69 + 671.67

1) The triple point of water is +0.01 °C. The triple point of pure water is the equilibrium point between pure ice, air-free water and water vapour (at 1013.25 hPa). Temperature comparison of °F with °C

·

39

Physics Measures of Length and Square Measures Measures of length Unit

3

Inch in

Foot ft

Yard yd

Stat mile

Naut mile

mm

m

km

1 in 1 ft 1 yd 1 stat mile 1 naut mile

= = = = =

1 12 36 63 360 72 960

0.08333 1 3 5280 6080

0.02778 0.3333 1 1760 2027

– – – 1 1.152

– – – 0.8684 1

25.4 304.8 914.4 – –

0.0254 0.3048 0.9144 1609.3 1853.2

– – – 1.609 1.853

1 mm 1m 1 km

= = =

0.03937 39.37 39 370

3.281 . 10-3 3.281 3281

1.094 . 10-3 1.094 1094

– – 0.6214

– – 0.5396

1 1000 106

0.001 1 1000

10–6 0.001 1

1 German statute mile = 7500 m 1 geograph. mile = 7420.4 m = 4 arc minutes at the equator (1° at the equator = 111.307 km) 1 internat. nautical mile 1 German nautical mile (sm) 1 mille marin (French)

}

= 1852 m = 1 arc minute at the degree of longitude (1° at the meridian = 111.121 km)

Other measures of length of the Imperial system 1 micro-in = 10-6 in = 0.0254 μm 1 mil = 1 thou = 0.001 in = 0.0254 mm 1 line = 0.1 in = 2.54 mm 1 fathom = 2 yd = 1.829 m 1 engineer’s chain = 100 eng link = 100 ft = 30.48 m 1 rod = 1 perch = 1 pole = 25 surv link = 5.029 m 1 surveyor’s chain = 100 surv link = 20.12 m 1 furlong = 1000 surv link = 201.2 m 1 stat league = 3 stat miles = 4.828 km

Astronomical units of measure 1 light-second = 300 000 km 1 l.y. (light-year) = 9.46 . 1012 km second distances to the stars) = 1 parsec (parallax second, 3.26 l.y. 1 astronomical unit (mean distance of the earth from the sun) = 1.496 . 108 km Typographical unit of measure: 1 point (p) = 0.376 mm Other measures of length of the metric system : 1 toise = 1.949 m 1 myriametre = 10 000 m Russia: 1 werschok = 44.45 mm 1 saschen = 2.1336 m 1 arschin = 0.7112 m 1 werst = 1.0668 km Japan: 1 shaku = 0.3030 m 1 ken = 1.818 m 1 ri = 3.927 km

Square measures Unit

sq in

sq ft

sq yd

sq mile

cm2

dm2

m2

a

ha

km2

1 square inch 1 square foot 1 square yard 1 square mile

= = = =

1 144 1296 –

– 1 9 –

– 0.1111 1 –

– – – 1

6.452 929 8361 –

0.06452 9.29 83.61 –

– 0.0929 0.8361 –

– – – –

– – – 259

– – – 2.59

1 cm2 1 dm2 1 m2 1a 1 ha 1 km2

= = = = = =

0.155 15.5 1550 – – –

– 0.1076 10.76 1076 – –

– 0.01196 1.196 119.6 – –

– – – – – 0.3861

1 100 10000 – – –

0.01 1 100 10000 – –

– 0.01 1 100 10000 –

– – 0.01 1 100 10000

– – – 0.01 1 100

– – – – 0.01 1

Other square measures of the Imperial system 1 sq mil = 1 . 10-6 sq in = 0.0006452 mm2 1 sq line = 0.01 sq in = 6.452 mm2 1 sq surveyor’s link = 0.04047 m2 1 sq rod = 1 sq perch = 1 sq pole = 625 sq surv link = 25.29 m2 1 sq chain = 16 sq rod = 4.047 a 1 acre = 4 rood = 40.47 a 1 township (US) = 36 sq miles = 3.24 km2 1 circular in + sq in + 5.067cm 2 (circular area with 1 in dia.) 4 1 circular mil + sq mil + 0.0005067mm 2 4 (circular area with 1 mil dia.)

40

Other square measures of the metric system

Russia: 1 kwadr. archin 1 kwadr. saschen 1 dessjatine 1 kwadr. werst

= = = =

0.5058 m2 4.5522 m2 1.0925 ha 1.138 km2

Japan: 1 tsubo 1 se 1 ho-ri

= 3.306 m2 = 0.9917a = 15.42 km2

·

Physics Cubic Measures and Weights; Energy, Work, Quantity of Heat Cubic measures US gallon – 7.481 202

0.01442 24.92 672.8

Imp gallon – 6.229 168.2

– 0.02832 0.7646

= = =

1 US liquid quart 1 US gallon

= =

57.75 231

0.03342 0.1337

1 4

0.25 1

0.8326 3.331

0.2082 0.8326

946.4 3785

0.9464 3.785

– –

1 Imp quart 1 Imp gallon

= =

69.36 277.4

0.04014 0.1605

1.201 4.804

0.3002 1.201

1 4

0.25 1

1136 4546

1.136 4.546

– –

= =

0.06102 61.02 61023

– 0.03531 35.31

– 1.057 1057

– 0.2642 264.2

– 0.88 880

– 0.22 220

1 1000 106

0.001 1 1000

106 0.001 1

1 cm3 1 dm3 (l) 1 m3

US liquid quart 0.01732 29.92 807.9

16.39 – –

1 cu in 1 cu ft 1 cu yd

Unit

cu ft – 1 27

dm3 (l) 0.01639 28.32 764.6

cu in 1 1728 46656

Imp quart

1 US minim = 0.0616 cm3 (USA) 1 US fl dram = 60 minims = 3.696 cm3 1 US fl oz = 8 fl drams = 0.02957 l 1 US gill = 4 fl oz = 0.1183 l 1 US liquid pint = 4 gills = 0.4732 l 1 US liquid quart = 2 liquid pints = 0.9464 l 1 US gallon = 4 liquid quarts = 3.785 l 1 US dry pint = 0.5506 l 1 US dry quart = 2 dry pints = 1.101 l 1 US peck = 8 dry quarts = 8.811 l 1 US bushel = 4 pecks = 35.24 l 1 US liquid barrel = 31.5 gallons = 119.2 l 1 US barrel = 42 gallons = 158.8 l (for crude oil) 1 US cord = 128 cu ft = 3.625 m3

cm3

m3

3

1 Imp minim = 0.0592 cm3 (GB) 1 Imp fl drachm = 60 minims = 3.552 cm3 1 Imp fl oz = 8 fl drachm = 0.02841 l 1 Imp gill = 5 fl oz = 0.142 l 1 Imp pint = 4 gills = 0.5682 l 1 Imp quart = 2 pints = 1.1365 l 1 lmp gallon = 4 quarts = 4.5461 l 1 lmp pottle = 2 quarts = 2.273 l 1 Imp peck = 4 pottles = 9.092 l 1 Imp bushel = 4 pecks = 36.37 l 1 Imp quarter = 8 bushels = 64 gallons = 290.94 l

Weights Unit

dram

1 dram 1 oz (ounce) 1 lb (pound)

= = =

1 16 256

oz

lb

0.0625 0.003906 1 0.0625 16 1

short cwt

long cwt

short ton

long ton

g

kg

t

– – 0.01

– – 0.008929

– – –

– – –

1.772 28.35 453.6

0.00177 0.02835 0.4536

– – –

1 short cwt (US) = 25600 1 long cwt (GB/US) = 28672

1600 1792

100 112

1 1.12

0.8929 1

0.05 0.04464 45359 0.056 0.05 50802

45.36 50.8

0.04536 0.0508

1 short ton (US) = 1 long ton (GB/US) =

32000 35840

2000 2240

20 22.4

17.87 20

1 1.12

– –

907.2 1016

0.9072 1.016

1 1000 106

0.001 1 1000

10-6 0.001 1

1g 1kg 1t

– –

= 0.5643 0.03527 0.002205 – – = 564.3 35.27 2.205 0.02205 0.01968 = – 35270 2205 22.05 19.68

1 grain = 1 / 7000 lb = 0.0648 g 1 stone = 14 lb = 6.35 kg 1 short quarter = 1/4 short cwt = 11.34 kg 1 long quarter = 1/4 long cwt = 12.7 kg 1 quintal or 1 cental = 100 lb = 45.36 kg 1 quintal = 100 livres = 48.95 kg 1 kilopound = 1kp = 1000 lb = 453.6 kg

(GB) (GB) (USA) (GB / USA) (USA) (F) (USA)

0.8929 1

– – – – 1.102 0.9842

1 solotnik = 96 dol = 4.2659 g (CIS) (CIS) 1 lot = 3 solotnik = 12.7978 g (CIS) 1 funt = 32 lot = 0.409 kg (CIS) 1 pud = 40 funt = 16.38 kg (CIS) 1 berkowetz = 163.8 kg 1 kwan = 100 tael = 1000 momme = 10000 fun = 3.75 kg (J) (J) 1 hyaku kin = 1 picul = 16 kwan = 60 kg (J)

tdw = tons dead weight = lading capacity of a cargo vessel (cargo + ballast + fuel + stores), mostly given in long tons, i.e. 1 tdw = 1016 kg

Energy, work, quantity of heat Work 1 ft lb 1 erg 1 Joule (WS) 1 kpm 1 PSh 1 hph 1 kWh 1 kcal 1 Btu

ft lb = = = = = = = = =

erg

J = Nm = Ws

kpm

PSh

1 13.56 . 106 1.356 0.1383 0.5121 . 10-6 73.76 . 10-9 1 100 . 10-9 10.2 . 10-9 37.77 . 10-15 10 . 106 0.7376 1 0.102 377.7 . 10-9 7.233 98.07 . 106 9.807 1 3.704 . 10-6 6 12 6 3 . . . . 1.953 10 26.48 10 2.648 10 270 10 1 1.014 1.98 . 106 26.85 . 1012 2.685 . 106 273.8 . 103 1.36 2.655 . 106 36 . 1012 3.6 . 106 367.1 . 103 1.581 . 10-3 3.087 . 103 41.87 . 109 4186.8 426.9 778.6 10.55 . 109 1055 107.6 398.4 . 10-6

hph 0.505 . 10-6 37.25 . 10-15 372.5 . 10-9 3.653 . 10-6 0.9863 1 1.341 1.559 . 10-3 392.9 . 10-6

kWh

kcal

0.3768 . 10-6 0.324 . 10-3 27.78 . 10-15 23.9 . 10-12 277.8 . 10-9 238 . 10-6 2.725 . 10-6 2.344 . 10-3 0.7355 632.5 0.7457 641.3 1 860 -3 . 1.163 10 1 0.252 293 . 10-6

Btu 1.286 . 10-3 94.84 . 10-12 948.4 . 10-6 9.301 . 10-3 2510 2545 3413 3.968 1

1 in oz = 0.072 kpcm; 1 in lb = 0.0833ft lb = 0.113 Nm; 1 thermi (French) = 4.1855 . 106 J; 1 therm (English) = 105.51 . 106 J Common in case of piston engines: 1 litre-atmosphere (litre . atmosphere) = 98.067 J

·

41

Physics Power, Energy Flow, Heat Flow, Pressure and Tension, Velocity Power, energy flow, heat flow Power

3

1 erg/s 1W 1 kpm/s 1 PS (ch) 2) 1hp 1 kW 1 kcal/s 1 Btu/s

= = = = = = = =

erg/s

W

1 107 9.807 . 107 7.355 . 109 7.457 . 109 1010 41.87 . 108 10.55 . 109

10-7

kpm/s

PS

. 10-7

0.102 0.102 1 75 76.04 102 426.9 107.6

1 9.807 735.5 745.7 1000 4187 1055

hp

. 10-9

kW . 10-9

0.136 0.1341 1.36 .10-3 1.341 . 10-3 13.33 . 10-3 13.15 . 10-3 1 0.9863 1.014 1 1.36 1.341 5.692 5.614 1.434 1.415

kcal/s

10-10

Btu/s

. 10-12

23.9 94.84 . 10-12 10-3 239 . 10-6 948.4 . 10-6 9.804 . 10-3 2.344 . 10-3 9.296 . 10-3 0.7355 0.1758 0.6972 0.7457 0.1782 0.7068 1 0.239 0.9484 4.187 1 3.968 1.055 0.252 1

1 poncelet (French) = 980.665 W; flyweel effect: 1 kgm2 = 3418 lb in 2

Pressure and tension μbar mbar = = cN/ dN/m2 cm2

Unit 1 μb = daN 1 mbar = cN/cm2 1 bar = daN/cm2

=

1

= =

1 kp/m2 = 1mm = WS at 4 °C 1

p/cm2

=

kp/m2 kp/cm2 mm p/cm2 = at WS

lb sq ft

sq in

sq in

sq in

–

–

–

–

–

–

–

–

0.7501

–

2.089

0.0145

–

–

2089

14.5

lb

long ton sh ton

1

0.001

10.2

1.02

–

106

1000

1

10 197

1020

1.02

98.07

–

–

1

0.1

0.0001

–

–

–

0.2048

–

–

–

–

10

1

0.001

–

0.7356

–

2.048

0.0142

–

–

1000

1

0.01

2048

14.22

–

–

106

105

100

1

73 556

96.78

–

1422

0.635

0.7112

13.6

1.36

0.00136

–

1

–

2.785 0.01934

–

–

10332

1033

1.033

–

760

1

2116

14,7

–

–

–

–

0.3591

–

1

–

–

–

=

980.7

0.9807 10 000

–

98 067

98.07

1 Torr = 1 mm = QS at 0 °C

1333

1 atm (pressure of the = atmosphere)

–

1.333 0.00133

1013

1.013

–

atm

1000

–

–

Torr = mm QS

–

980.7 0.9807

0.0102

kp/ mm2

0.001

1 kp/cm2 = 1 at = (technical atmosphere) 1 kp/mm2

bar = daN/ cm2

4.882 0.4882

0.0102 750.1 0.9869

735.6 0.9678

0.0064 0.0072

1 lb/sq ft

=

478.8 0.4788

–

1 lb/sq in = 1 psi

=

68 948 68.95

0.0689

703.1

70.31

0.0703

–

51.71

0.068

144

1

–

0.0005

1 long ton/sq in (GB)

=

–

–

154.4

–

–

157.5

1.575

–

152.4

–

2240

1

1,12

1 short ton/sq = in (US)

–

–

137.9

–

–

140.6

1.406

–

136.1

–

2000

0.8929

1

1 psi = 0.00689 N / mm2 1 N/m2 (Newton/m2) = 10 μb; 1 barye (French) = 1 μb; 1 pièce (pz) (French) = 1 sn/m2 ≈ 102 kp/m2; 1 hpz = 100 pz = 1.02 kp/m2; 1 micron (USA) = 0.001 mm QS = 0.001 Torr. In the USA, “inches Hg” are calculated from the top, i.e. 0 inches Hg = 760 mm QS and 29.92 inches Hg = 0 mm QS = absolute vacuum. The specific gravity of mercury is assumed to be 13.595 kg/dm3.

Velocity Unit m/s m/min km/h ft/min mile/h

42

= = = = =

m/s

m/min

km/h

ft/min

mile/h

1 0.0167 0.278 0.0051 0.447

60 1 16.67 0.305 26.82

3.6 0.06 1 0.0183 1.609

196.72 3.279 54.645 1 87.92

2.237 0.0373 0.622 0.0114 1 ·

Physics Equations for Linear Motion and Rotary Motion

Definition

Basic formulae

SI Sym Symunit bol

Uniform motion

Linear motion distance moved divided by time

Velocity

m/s

v

Angular velocity

rad/s

ω

v =

s2 – s1 Δs = = const. t2 – t1 Δt

Angle of rotation

rad

Distance moved

m

s

s = v.t acceleration equals change of velocity divided by time

Acceleration

m/s2

a

Angular acceleration

rad/s2

α

a =

v2 – v1 Δv = = const. t2 – t1 Δt

m/s

v

Circumferential speed

m/s

v

Distance moved

m

s

Uniform motion and constant force or constant torque Work

Power

J

W

W

P

Non-uniform (accelerated) motion Force

N

F

In case of any motion

3

+ t angle of rotation ϕ = ω . t angular acceleration equals change of angular velocity divided by time ω2 – ω1 Δω α = = = const. t2 – t1 Δt

motion accelerated from rest: a =

Velocity

angular velocity = angle of rotation in radian measure/time ϕ2 – ϕ1 Δϕ ω = = = const. t2 – t1 Δt

motion accelerated from rest:

v+s t

Uniformly accelerated motion

Rotary motion

v2

2s v = = t 2s t2

α =

v + a @ t + Ǹ2 a @ s

ω t

=

ω2 2ϕ

2ϕ t2

+@t v+r@+r@@t

s =

a . 2 v . v2 t = t = 2 2 2a

angle of rotation α . 2 ω . ω2 t = t = ϕ = 2 2 2α

force . distance moved

torque . angle of rotation in radian measure

W = F.s

W = M.ϕ

work in unit of time = force . velocity

work in unit of time = torque . angular velocity

P+W+F@v t

P+W+M@ t

accelerating force = mass . acceleration

accel. torque = second mass moment . angular acceleration

F = m.a

M = J.α

Momentum (kinetic energy) equals half the mass . second power of velocity

Kinetic energy due to rotation equals half the mass moment of inertia . second power of the angular velocity

Ek + m @ v2 2

Ek + J @ 2 2

Energie

J

Ek

Potential energy (due to force of gravity)

J

Ep

weight . height Ep = G . h = m . g . h

Centrifugal force

N

FF

FF = m . rs . ω2 (rs = centre-of-gravity radius)

·

=

43


Mathematics / Geometry

Page

Calculation of Areas

45

Calculation of Volumes

46

4

44

·

Mathematics / Geometry Calculation of Areas

A = area Square

U = circumference Polygon

A = a2

A + A1 ) A2 ) A3

a + ǸA

=

d + a Ǹ2

Rectangle

a . h1 + b . h2 + b . h3 2

Formed area 2

A+a@b

A + r (2 Ǹ3 * ) 2

d + Ǹa 2 ) b 2

[ 0.16 @ r 2

Circle

Parallelogram

a =

d2 . π = r2 . π 4

A =

A+a@h

4

[ 0.785 @ d 2

A h

U+2@r@ + d@

Circular ring

Trapezium A+m@h

A+

@ (D 2 * d 2) 4

+ (d ) b) b @

m =

a+b 2

A =

a.h 2

Triangle

b =

Circular sector

A =

D–d 2 r2 . π . α_

360_ b.r 2 r . π . α_ b = 180_

= a =

2.A h

Circular segment

Equilateral triangle A+

a2 Ǹ 4

3

A =

3 . a2 . √3 2

d+2@a s + Ǹ3 @ a

Octagon

A + 2a 2 (Ǹ2 ) 1) d + a Ǹ4 ) 2 Ǹ 2 s + a (Ǹ2 ) 1)

·

r2 2

ǐ

α_ . π 180_

– sin α

Ǒ

+ 1 [ r (b * s) ) sh ] 2 s + 2 r sin 2 α s α h = r (1 – cos ) = tan 2 2 4 α_ . π α^ = 180_ b + r @ ^

a d + Ǹ3 2

Hexagon

A =

Ellipse

D.d.π = a.b.π 4 D+d . π U ≈ 2 U + (a ) b) [ 1 ) A =

1 4

ǐ Ǒ

ǐ Ǒ ǐ Ǒ

1 a–b 2 a–b 4 + a+b 64 a + b 1 a – b 6 ..... ] + 256 a + b

45

Mathematics / Geometry Calculation of Volumes

V = volume Cube

O = surface

M = generated surface Frustum of cone

V + a3 O + 6 @ a2 d + a Ǹ3

Square prism

Sphere

V+a@b@c

V + 4 r3 + 1 @ d3 3 6

O + 4 @ r2 + @ d2

Spherical zone

Parallelepiped

V=

V+A@h

(Cavalier principle)

Pyramid

π.h

(3 a2 + 3 b2 + h2)

6

M+2@r@@h

Spherical segment V =

ǒ

π.h 3 2 s ) h2 4 6 + h2 r * h 3 M+2@r@@h + (s 2 ) 4h 2) 4

V =

ǒ

A.h 3

Ǔ

Ǔ

Spherical sector

Frustum of pyramid V=

h 3

Cylinder V =

V + 2 @ h @ r2 @ 3

(A1 + A2 + ǸA 1 @ A 2)

O =

π.r 2

(4 h + s)

Cylindrical ring

d2 . π

h

4

V =

M+2@r@@h

Hollow cylinder

D . π2 . d 2 4

O + D @ d @ 2

O + 2 @ r @ @ (r ) h)

Barrel V =

V =

h.π 4

(D2 – d2)

r2 . π . h

m+

h2 )

ǒ d2 Ǔ

V =

h.π 12

(2 D2 + d2)

Prismatoid

3 M+r@@m O + r @ @ (r ) m)

46

D–d 2 + h2 2

[ 4.189 @ r 3

d + Ǹa 2 ) b 2 ) c 2

Cone

ǐ Ǒ

m =

O + 2 (ab ) ac ) bc )

4

π.h (D2 + D d + d2) 12 . π m (D + d) M = 2 +2@@p@h V =

V + h (A 1 ) A 2 ) 4A) 6

2

·


Mechanics / Strength of Materials

Page

Axial Section Moduli and Axial Second Moments of Area (Moments of Inertia) of Different Profiles

48

Deflections in Beams

49

Values for Circular Sections

50

Stresses on Structural and Fatigue Strength of Structures

51

5

·

47

Mechanics / Strength of Materials Axial Section Moduli and Axial Second Moments of Area (Moments of Inertia) of Different Profiles Cross-sectional area

Section modulus W 1 + bh 2ń 6

1 + bh 3ń 12

W 2 + hb 2ń 6

2 + hb 3ń 12

W 1 + W 2 + a 3ń 6

1 + 2 + a 4ń 12

W 1 + bh 2ń 24 for e + 2 h 3 W 2 + hb 2ń 24

1 + bh 3ń 36

W1 +

5 3 R + 0.625 R 3 8

W 2 + 0.5413 R 3

5

Second moment of area

6b2 + 6bb1 + b21 h2 12 (3b + 2b1)

W1 = for e =

W1 =

1 3b + 2b1 3

2b + b1

1 + 2 +

Ι1 =

h

BH3 – bh3

Ι1 =

6H

W 1 + W 2 + D 3ń 32 [ D 3ń 10

W1 = W2 =

2 + hb 3ń 48

π

D4 – d4

32

D

5 Ǹ 3 R 4 + 0.5413 R 4 16

6b2 + 6bb1 + b21 h3 36 (2b + b1)

BH3 – bh3 12

1 + 2 + D 4ń 64 [ D 4ń 20

1 + 2 +

(D 4 * d 4) 64

or in case of thin wall thickness s: W 1 + W 2 + ń (r ) sń2) [ sr 2

1 + 2 + sr 3 ƪ1 ) (sń2r) 2ƫ [ sr 3

W 1 + a 2bń4

1 + a 3bń4

W 2 + b 2ań4

2 + b 3ań4

W 1 + 1ńa 1

1 +

3 (a b * a32 b2) 4 1 1

or in case of thin wall thickness s: s + a 1 * a 2 + b 1 * b 2 + 2 (a * a 2) + 2 (b * b2) W1 [

a (a ) 3b) s 4

W 1 + 1 ń e + 0.1908 r 3 4 = 0.5756 r with e = r 1 – 3π

ǐ

Ǒ

1 [

2 a (a ) 3b) s 4

ƨ

Ʃ

Ι1 = π / 8 – 8 / (9 π) r4 = 0.1098 r4

axis 1-1 = axis of centre of gravity

48

·

Mechanics / Strength of Materials Deflections in Beams

f, fmax, fm, w, w1, w2 a, b, l, x1, x1max, x2 E q, q0

Deflection (mm) Lengths (mm) Modulus of elasticity (N/mm2) Line load (N/mm) Fl3

w (x) =

3EΙ

ƪ

1–

3

x

.

2

ǐǑƫ x 3

1

+

l

α, α1, α2, αA, αB Angles (°) Forces (N) F, FA, FB Ι Second moment of area (mm4) (moment of inertia)

2

Fl3

f=

l

tan α =

3EΙ

Fl2 2EΙ

FB = F w (x) =

ql4 8EΙ

ƪ

1–

4

x

.

3

l

ǐǑƫ x 4

1

+

3

ql4

f=

l

tan α =

8EΙ

ql3 6EΙ

FB = q . l q0l4

w (x) = FB =

120EΙ q0 . l

ƪ

2

w (x) =

Fl3 . x l

w1 (x1) =

Fl3 . a 6EΙ l

w2 (x2) =

Fl3 . b 6EΙ l

16EΙ F FA = FB = 2

x1max + a Ǹ(l ) b)ń3a for a > b change a and b for a < b

4–5 .

FA = F

b l

ƪ

1–

x l

4 3

ǐǑƫ x 5

+

l

ǐǑƫ x 2

x≤

l

ǐǑ ǒ ǐǑ ǒ

Ǔ

x2 l b 2 x1 1+ – 1 b ab l l

q0l4

f=

30EΙ

l

Fl3

f=

2

x1 ≤ a

tan α =

f=

ƪ ǐ Ǒ ǐǑƫ

5

Fl2 16EΙ

ǐǑǐǑ

ǐ Ǒ ǐ Ǒ

a 2 b 2 f l tan α1 = 1+ 2a l l b

Fl3 3EΙ

Ǔ

a l

24EΙ

tan α =

48EΙ

x22 l l+b a 2 x2 1+ – x2 ≤ b fmax = f a ab 3b l l

FB = F

q0l3

ǐǑǐ

l+b f l tan α2 = 1+ 3a a 2b

Ǒ

ǐ Ǒ

ƪ ǐ Ǒƫ

ǐ Ǒ

Fl3 a 2 Fl2 . a a 1 x 2 Fl3 . x a 4 a a 1– tan α1 = – f= 1– . 1– l 3 l 2EΙ l 3 l 2EΙ l l 2EΙ l l x = ≤ a < l/2 Fl3 . a Fl2 . a x 1 a 2 Fl3 . a x 4 a 2 a 1– tan α2 = – w (x) = fm = 1– 1– 2 l 3 l 8EΙ l 3 l 2EΙ l l 2EΙ l l a ≤ x ≤ l/2 w (x) =

ƪ ǐ Ǒ ǐ Ǒƫ

FA = FB = F

ƪ ǐ Ǒ ǐ Ǒ ǐǑǐ

Ǒƫ

2 a a x1 1 x1 3 a a 2 + 1+ . – 1+ 3 l l l l 3 l l Fl2 . a a Fl3 a 2 2 a tan α1 = x1 ≤ a 1+ f= 1+ . 2EΙ l l 2EΙ l 3 l x2 Fl3 . a . x2 Fl3 . a Fl2 . a 1– w2 (x2) = x2 ≤ l fm = tan α2 = l 2EΙ l l 8EΙ l 2EΙ l FA = FB = F Fl3

w1 (x1) =

w1 (x1) =

2EΙ

ǐǑǐ ǐ Ǒ

Fl3 . a . x1 l 6EΙ l

ƪ ǐ Ǒƫ 1–

x1 2 l

Ǒ

x1 ≤ l

ǐ Ǒƫ

ƪ

x2 2 Fl3 . x2 2a 3a . x2 + – w2 (x2) = l l l l 6EΙ l

FA = F

a l

ql4 . x 24EΙ l q.l

w (x) = FA =

·

2

x2 ≤ a

ǐ Ǒ ƪ ǐǑ ǐǑƫ

FB = F 1 + 1–2

FB =

Fl3 3EΙ

ǐǑǐ Ǒ a a 2 1+ l l

fmax =

Fl3

+

x 3 l

.

9 √3 EΙ

a l

x 2 l

f=

ǐ Ǒ

a l

tan α =

0≤x≤l

fm =

5ql4 384EΙ

tan αA =

Fl2 . a l

6EΙ

tan αB = 2 tan αA

ǐ

Fl2 . a a 2+3 l l

6EΙ

tan α =

Ǒ

ql3 24EΙ

q.l 2

49

Mechanics / Strength of Materials Values for Circular Sections

Polar section modulus: Axial second moment of area (axial moment of inertia): Polar second moment of area (polar moment of area):

5

π . d3 32 π . d3 Wp = 16 π . d4 Ιa = 64

Wa =

Axial section modulus:

d mm

A cm2

Wa cm3

Ιa cm4

6 7 8 9 10 11

0.293 0.385 0.503 0.636 0.785 0.950

0.0212 0.0337 0.0503 0.0716 0.0982 0.1307

12 13 14 15 16 17

1.131 1.327 1.539 1.767 2.011 2.270

18 19 20 21 22 23

Ιp =

π . d4 32

π . d2 4 π . d2 . . l m = 4 kg = 7.85 dm3

Area:

A =

Mass: Density of steel: Second mass moment of inertia (mass moment of inertia):

Mass / I kg/m

J/ I kgm2/m

d mm

A cm2

0.0064 0.0118 0.0201 0.0322 0.0491 0.0719

0.222 0.302 0.395 0.499 0.617 0.746

0.000001 0.000002 0.000003 0.000005 0.000008 0.000011

115 120 125 130 135 140

103.869 113.097 122.718 132.732 143 139 153.938

149.3116 169.6460 191.7476 215.6900 241.5468 269.3916

0.1696 0.2157 0.2694 0.3313 0.4021 0.4823

0.1018 0.1402 0.1986 0.2485 0.3217 0.4100

0.888 1.042 1.208 1.387 1.578 1.782

0.000016 0.000022 0.000030 0.000039 0.000051 0.000064

145 150 155 160 165 170

165.130 176.715 188.692 201.062 213.825 226.980

2.545 2.835 3.142 3.464 3.801 4.155

0.5726 0.6734 0.7854 0.9092 1.0454 1.1945

0.5153 0.6397 0.7854 0.9547 1.1499 1.3737

1.998 2.226 2.466 2.719 2.984 3.261

0.000081 0.000100 0.000123 0.000150 0.000181 0.000216

175 180 185 190 195 200

24 25 26 27 28 29

4.524 4.909 5.309 5.726 6.158 6.605

1.3572 1.5340 1.7255 1.9324 2.1551 2.3944

1.6286 1.9175 2.2432 2.6087 3.0172 3.4719

3.551 3.853 4.168 4.495 4.834 5.185

0.000256 0.000301 0.000352 0.000410 0.000474 0.000545

210 220 230 240 250 260

30 32 34 36 38 40

7.069 8.042 9.079 10.179 11.341 12.566

2.6507 3.2170 3.8587 4.5804 5.3870 6.2832

3.9761 5.1472 6.5597 8.2448 10.2354 12.5664

5.549 6.313 7.127 7.990 8.903 9.865

42 44 46 48 50 52

13.854 15.205 16.619 18.096 19.635 21.237

7.2736 8.3629 9.5559 10.8573 12.2718 13.9042

15.2745 18.3984 21.9787 26.0576 30.6796 35.8908

54 56 58 60 62 64

22.902 24.630 26.421 28.274 30.191 32.170

15.4590 17.2411 19.1551 21.2058 23.3978 25.7359

66 68 70 72 74 76

34.212 36.317 38.485 40.715 43.008 45.365

78 80 82 84 86 88 90 92 95 100 105 110

50

Wa cm3

J=

π . d4 . l . 32

Mass / I kg/m

J/ I kgm2/m

858.5414 1017.8760 1198.4225 1401.9848 1630.4406 1895.7410

81.537 88.781 96.334 104.195 112.364 120.841

0.134791 0.159807 0.188152 0.220112 0.255979 0.296061

299.2981 331.3398 365.5906 402.1239 441.0133 482.3326

2169.9109 2485.0489 2833.3269 3216.9909 3638.3601 4099.8275

129.627 138.721 148.123 157.834 167.852 178.179

0.340676 0.390153 0.444832 0.505068 0.571223 0.643673

240.528 254.469 268.803 283.529 298.648 314.159

526.1554 572.5553 621.6058 673.3807 727.9537 785.3982

4603.8598 5152.9973 5749.8539 6397.1171 7097.5481 7853.9816

188.815 199.758 211.010 222.570 234.438 246.615

0.722806 0.809021 0.902727 1.004347 1.114315 1.233075

346.361 380.133 415.476 452.389 490.874 530.929

909.1965 1045.3650 1194.4924 1357.1680 1533.9808 1725.5198

9546.5638 11499.0145 13736.6629 16286.0163 19174.7598 22431.7569

271.893 298.404 326.148 355.126 385.336 416.779

1.498811 1.805345 2.156656 2.556905 3.010437 3.521786

0.000624 0.000808 0.001030 0.001294 0.001607 0.001973

270 572.555 280 615.752 300 706.858 320 804.248 340 907.920 360 1017.876

1932.3740 2155.1326 2650.7188 3216.9909 3858.6612 4580.4421

26087.0491 30171.8558 39760.7820 51471.8540 65597.2399 82447.9575

449.456 483.365 554.884 631.334 712.717 799.033

4.095667 4.736981 6.242443 8.081081 10.298767 12.944329

10.876 11.936 13.046 14.205 15.413 16.671

0.002398 0.002889 0.003451 0.004091 0.004817 0.005635

380 400 420 440 460 480

1134.115 1256.637 1385.442 1520.531 1661.903 1809.557

5387.0460 6283.1853 7273.5724 8362.9196 9555.9364 10857.3442

102353.8739 125663.7060 152745.0200 183984.2320 219786.6072 260576.2608

890.280 986.460 1087.572 1193.617 1304.593 1420.503

16.069558 19.729202 23.980968 28.885524 34.506497 40.910473

41.7393 48.2750 55.5497 63.6173 72.5332 82.3550

17.978 19.335 20.740 22.195 23.700 25.253

0.006553 0.007579 0.008721 0.009988 0.011388 0.012930

500 520 540 560 580 600

1693.495 2123.717 2290.221 2463.009 2642.079 2827.433

12271.8463 13804.1581 15458.9920 17241.0605 19155.0758 21205.7504

306796.1572 358908.1107 417392.7849 482749.6930 555497.1978 636172.5116

1541.344 1667.118 1797.824 1933.462 2074.032 2219.535

48.166997 56.348573 65.530667 75.791702 87.213060 99.879084

28.2249 30.8693 33.6739 36.6435 39.7828 43.0964

93.1420 104.9556 117.8588 131.9167 147.1963 163.7662

26.856 28.509 30.210 31.961 33.762 35.611

0.014623 0.016478 0.018504 0.020711 0.023110 0.025711

620 640 660 680 700 720

3019.071 3216.991 3421.194 3631.681 3848.451 4071.504

23397.7967 25735.9270 28224.8538 30869.2894 33673.9462 36643.5367

725331.6994 823549.6636 931420.1743 1049555.8389 1178588.1176 1319167.3201

2369.970 2525.338 2685.638 2850.870 3021.034 3196.131

113.877076 129.297297 146.232967 164.780267 185.038334 207.109269

47.784 50.265 52.810 55.418 58.088 60.821

46.5890 50.2655 54.1304 58.1886 62.4447 66.9034

181.6972 201.0619 221.9347 244.3920 268.5120 294.3748

37.510 39.458 41.456 43.503 45.599 47.745

0.028526 0.031567 0.034844 0.038370 0.042156 0.046217

740 760 780 800 820 840

4300.840 4536.460 4778.362 5026.548 5281.017 5541.769

39782.7731 43096.3680 46589.0336 50265.4824 54130.4268 58188.5791

1471962.6056 1637661.9830 1816972.3105 2010619.2960 2219347.4971 2443920.3207

3376.160 3561.121 3751.015 3945.840 4145.599 4350.289

231.098129 257.112931 285.264653 315.667229 348.437557 383.695490

63.617 66.476 70.882 78.540 86.590 95.033

71.5694 76.4475 84.1726 98.1748 113.6496 130.6706

322.0623 351.6586 399.8198 490.8739 596.6602 718.6884

49.940 52.184 55.643 61.654 67.973 74.601

0.050564 0.055210 0.062772 0.077067 0.093676 0.112834

860 880 900 920 940 960 980 1000

5808.805 6082.123 6361.725 6647.610 6939.778 7238.229 7542.964 7853.982

62444.6517 66903.3571 71569.4076 76447.5155 81542.3934 86858.7536 92401.3084 98174.7703

2685120.0234 2943747.7113 3220623.3401 3516585.7151 3832492.4910 4169220.1722 4527664.1126 4908738.5156

4559.912 4774.467 4993.954 5218.374 5447.726 5682.010 5921.227 6165.376

421.563844 462.168391 505.637864 552.103957 601.701321 654.567567 710.843266 770.671947

Ιa cm4

·

Mechanics / Strength of Materials Stresses on Structural and Fatigue Strength of Structures Diffusion of stress in structural : loading types

static Maximum stress limit: Mean stress: Minimum stress limit:

dynamic o + sch m + schń2 u + 0

alternating o + ) w m + 0 u + * w

oscillating o + m ) a m + v (initial stress) u + m * a

Ruling coefficient of strength of material for the calculation of structural : Resistance to Fatigue strength under Fatigue strength under Resistance to breaking Rm fluctuating stresses σSch alternating stresses σW deflection σA Yield point Re; Rp0.2 Coefficients of fatigue strength σD Stress-number diagram

Damage curve Endurance limit

Fatigue strength under alternating stresses

Fatigue limit

Number of cycles to failure N

In case of stresses below the damage curve initial damage will not occur to the material. Reduced stress on the member v

Permissible stress ≤

perm.

D . b0 . bd S . ßk

Reduced stress σv For the frequently occurring case of combined bending and torsion, according to the distortion energy theory: v + Ǹ 2 ) 3 ( 0 ) 2

Yield point Re

Resistance to deflection σA Fatigue strength under fluctuating stresses σSch Mean stress σm

Alternate area / Area of fluctuation

Design strength of the member =

5

Resistance to breaking Rm Coefficients of strength

Stress-number curve Stress σ

Fatigue strength diagram acc. to SMITH

Example: Tension-Compression

with: σD = ruling fatigue strength value of the material b0 = surface number (≤ 1) bd = size number (≤ 1) ßk = stress concentration factor (≥ 1) S = safety (1.2 ... 2) with: σ = single axis bending stress τ = torsional stress α0 = constraint ratio according to Bach

Alternating bending, dynamic torsion: α0 ≈ 0.7 Alternating bending, alternating torsion: α0 ≈ 1.0 Static bending, alternating torsion: α0 ≈ 1.6

Diameter of component d

·

Surface roughness Rt in μm

for tension/ compression bd = 1.0

Surface number b0

Size number bd

For bending and torsion

Surfaces with rolling skin Resistance to breaking of the material Rm

51


Hydraulics

Page

Hydrostatics (Source: K. Gieck, Technische Formelsammlung, 29th edition, Gieck Verlag, Heilbronn)

53

Hydrodynamics (Source: K. Gieck, Technische Formelsammlung, 29th edition, Gieck Verlag, Heilbronn)

54

6

52

·

Hydraulics Hydrostatics

Pressure distribution in a fluid

p1 + p0 ) g h1 P 2 + p 1 ) g (h 2 * h 1) + p 1 ) g h

Linear pressure

Hydrostatic force of pressure on planes The hydrostatic force of pressure F is that force which is exerted on the wall by the fluid only - i.e. without consideration of pressure p0. F + g y s A cos + g h s A yD =

Ιx Ι = ys + s y sA ys A

;

xD =

Ιxy ysA

m, mm

6 Hydrostatic force of pressure on curved surfaces The hydrostatic force of pressure on the curved surface 1 - 2 is resolved into a horizontal component FH and a vertical component FV. FV is equal to the weight of the fluid having a volume V located (a) or thought to be located (b) over the surface 1 - 2. The line of application runs through the centre of gravity.

Ť FV Ť + g V

(N, kN)

FH is equal to the hydrostatic force of pressure on the projection of surface 1 - 2 perpendicular to FH. Buoyance The buoyant force FA is equal to the weight of the displaced fluids having densities and ’. F A + g V ) g Ȁ VȀ

(N, kN)

If the fluid with density ’ is a gas, the following applies: FA [ g V

(N, kN)

For k density of the body applies: > k the body floats in the liquid = k the body is suspended < k the body sinks

}

S D Ιx, Ιs Ιxy

= = = =

centre of gravity of plane A centre of pressure moments of inertia product of inertia of plane A referred to the x- and y-axes

·

53

Hydraulics Hydrodynamics

Discharge of liquids from vessels Vessel with bottom opening v + Ǹ2 g H . V + A Ǹ2 g H

Vessel with small lateral opening v + Ǹ2 g H s + 2Ǹ H h (without any coefficient of friction) . V + A Ǹ2 g H .

F+Vv Vessel with wide lateral opening

6 . V + 2 b Ǹ2 g (H 2 3

3ń2

* H1

3ń2

)

Vessel with excess pressure on liquid level v+

Ǹ2 ( g H ) p ) ü

.

V+ A

Ǹ2 ( g H ) p ) ü

Vessel with excess pressure on outlet v+ .

Ǹ2

V+ A

pü

Ǹ2

pü

v: g: : pü: ϕ: ε:

discharge velocity gravity density excess pressure compared to external pressure coefficient of friction (for water ϕ = 0.97) coefficient of contraction (ε = 0.62 for sharp-edged openings) (ε = 0.97 for smooth-rounded openings) F: force of reaction . V : volume flow rate b: width of opening

54

·


Electrical Engineering

Page

Basic Formulae

56

Speed, Power Rating and Efficiency of Electric Motors

57

Types of Construction and Mounting Arrangements of Rotating Electrical Machinery

58

Types of Protection for Electrical Equipment (Protection Against and Foreign Bodies)

59

Types of Protection for Electrical Equipment (Protection Against Water)

60

7

·

55

Electrical Engineering Basic Formulae

Ohm’s law: +U R

U+@R

Material

R+U

R + R 1 ) R 2 ) R 3 ) AAA ) R n R + total resistance ƪƫ R n + individual resistance ƪƫ Shunt connection of resistors: 1 + 1 ) 1 ) 1 ) AAA ) 1 R1 R2 R3 Rn R R + total resistance ƪƫ R n + individual resistance ƪƫ Current consumption

Direct current

P+U@

+P U

P = U . Ι . cos ϕ

Three-phase current

Power

Single-phase alternating current

Electric power:

P = 1.73 . U . Ι . cos ϕ Ι =

Ι =

ƪ Ω mm ƫ ƪ Ω mm ƫ m 2

2

Series connection of resistors:

7

m

P U . cos ϕ

P 1.73 . U . cos ϕ

a) Metals Aluminium Bismuth Lead Cium Iron wire Gold Copper Magnesium Nickel Platinum Mercury Silver Tantalum Tungsten Zinc Tin

36 0.83 4.84 13 6.7...10 43.5 58 22 14.5 9.35 1.04 61 7.4 18.2 16.5 8.3

0.0278 1.2 0.2066 0.0769 0.15..0,1 0.023 0.01724 0.045 0.069 0.107 0.962 0.0164 0.135 0.055 0.061 0.12

b) Alloys Aldrey (AlMgSi) Bronze I Bronze II Bronze III Constantan (WM 50) Manganin Brass Nickel silver (WM 30) Nickel chromium Niccolite (WM 43) Platinum rhodium Steel wire (WM 13) Wood’s metal

30.0 48 36 18 2.0 2.32 15.9 3.33 0.92 2.32 5.0 7.7 1.85

0.033 0.02083 0.02778 0.05556 0.50 0.43 0.063 0.30 1.09 0.43 0.20 0.13 0.54

c) Other conductors Graphite Carbon, homog. Retort graphite

0.046 0.015 0.014

22 65 70

Resistance of a conductor: l. l R= = A γ.A R l γ A

56

= = = = =

resistance (Ω) length of conductor (m) electric conductivity (m/Ω mm2) cross section of conductor (mm2) specific electrical resistance (Ω mm2/m) ·

Electrical Engineering Speed, Power Rating and Efficiency of Electric Motors Speed: n =

Power rating:

f . 60 p

Output power

n = speed (min-1) f = frequency (Hz) p = number of pole pairs

Direct current: Pab = U . . η Single-phase alternating current: Pab = U . . cos .

Example: f = 50 Hz, p = 2 n =

50 . 60 2

1)

Three-phase current: Pab = 1.73 . U . . cos .

= 1500 min-1

Efficiency:

+

P ab 1) @ 100 ƪ%ƫ P zu

Example: Efficiency and power factor of a four-pole 1.1-kW motor and a 132-kW motor dependent on the load

7 Power factor cos ϕ

Efficiency η

132-kW motor

1.1-kW motor

Power output P / PN

1) Pab = mechanical output power on the motor shaft Pzu = absorbed electric power ·

57

Electrical Engineering Types of Construction and Mounting Arrangements of Rotating Electrical Machinery Types of construction and mounting arrangements of rotating electrical machinery (Extract from DIN EN 50347) Machines with end shields, horizontal arrangement Design Bearings

Stator (Housing)

Shaft

General design

Design / Explanation Fastening or Installation

B3

2 end shields

with feet

free shaft end

–

installation on substructure

B5

2 end shields

without feet

free shaft end

mounting flange close to bearing, access from housing side

flanged

B6

2 end shields

with feet

free shaft end

design B3, if necessary end shields turned through -90°

wall fastening, feet on LH side when looking at input side

B7

2 end shields

with feet

free shaft end

design B3, if necessary end shields turned through 90°

wall fastening, feet on RH side when looking at input side

B8

2 end shields

with feet

free shaft end

design B3, if necessary end shields turned through 180°

fastening on ceiling

B 35

2 end shields

with feet

free shaft end

mounting flange close to bearing, access from housing side

installation on substructure with additional flange

Symbol

7

Explanation

Figure

Machines with end shields, vertical arrangement Design

Explanation Bearings

Stator (Housing)

Shaft

General design

Design / Explanation Fastening or Installation

V1

2 end shields

without feet

free shaft end at the bottom

mounting flange close to bearing on input side, access from housing side

flanged at the bottom

V3

2 end shields

without feet

free shaft end at the top

mounting flange close to bearing on input side, access from housing side

flanged at the top

V5

2 end shields

with feet

free shaft end at the bottom

–

fastening to wall or on substructure

V6

2 end shields

with feet

free shaft end at the top

–

fastening to wall or on substructure

Symbol

58

Figure

·

Electrical Engineering Types of Protection for Electrical Equipment (Protection Against and Foreign Bodies) Types of protection for electrical equipment (Extract from DIN EN 60529) Example of designation

Type of protection

DIN EN 60529

IP

4

4

Designation DIN number Code letters First type number Second type number An enclosure with this designation is protected against the ingress of solid foreign bodies having a diameter above 1 mm and of splashing water. Degrees of protection for protection against and foreign bodies (first type number) First type number 0 1

Degree of protection (Protection against and foreign bodies) No special protection Protection against the ingress of solid foreign bodies having a diameter above 50 mm (large foreign bodies) 1) No protection against intended access, e.g. by hand, however, protection of persons against with live parts

2

Protection against the ingress of solid foreign bodies having a diameter above 12 mm (medium-sized foreign bodies) 1) Keeping away of fingers or similar objects

3

Protection against the ingress of solid foreign bodies having a diameter above 2.5 mm (small foreign bodies) 1) 2) Keeping away tools, wires or similar objects having a thickness above 2.5 mm

4

Protection against the ingress of solid foreign bodies having a diameter above 1 mm (grain sized foreign bodies) 1) 2) Keeping away tools, wires or similar objects having a thickness above 1 mm

5

6

Protection against harmful dust covers. The ingress of dust is not entirely prevented, however, dust may not enter to such an amount that operation of the equipment is impaired (dustproof). 3) Complete protection against Protection against the ingress of dust (dust-tight) Complete protection against

1) For equipment with degrees of protection from 1 to 4, uniformly or non-uniformly shaped foreign bodies with three dimensions perpendicular to each other and above the corresponding diameter values are prevented from ingress. 2) For degrees of protection 3 and 4, the respective expert commission is responsible for the application of this table for equipment with drain holes or cooling air slots. 3) For degree of protection 5, the respective expert commission is responsible for the application of this table for equipment with drain holes. ·

59

7

Electrical Engineering Types of Protection for Electrical Equipment (Protection Against Water) Types of protection for electrical equipment (Extract from DIN EN 60529) Example of designation

Type of protection

DIN EN 60529

IP

4

4

Designation DIN number Code letters First type number Second type number An enclosure with this designation is protected against the ingress of solid foreign bodies having a diameter above 1 mm and of splashing water. Degrees of protection for protection against water (second type number) Second type number

7

Degree of protection (Protection against water)

0

No special protection

1

Protection against dripping water falling vertically. It may not have any harmful effect (dripping water).

2

Protection against dripping water falling vertically. It may not have any harmful effect on equipment (enclosure) inclined by up to 15° relative to its normal position (diagonally falling dripping water).

3

Protection against water falling at any angle up to 60° relative to the perpendicular. It may not have any harmful effect (spraying water).

4

Protection against water spraying on the equipment (enclosure) from all directions. It may not have any harmful effect (splashing water).

5

Protection against a water jet from a nozzle which is directed on the equipment (enclosure) from all directions. It may not have any harmful effect (hose-directed water).

6

7

8

Protection against heavy sea or strong water jet. No harmful quantities of water may enter the equipment (enclosure) (flooding). Protection against water if the equipment (enclosure) is immersed under determined pressure and time conditions. No harmful quantities of water may enter the equipment (enclosure) (immersion). The equipment (enclosure) is suitable for permanent submersion under conditions to be described by the manufacturer (submersion). 1)

1) This degree of protection is normally for air-tight enclosed equipment. For certain equipment, however, water may enter provided that it has no harmful effect.

60

·


Materials

Page

Conversion of Fatigue Strength Values of Miscellaneous Materials

62

Mechanical Properties of Quenched and Tempered Steels

63

Fatigue Strength Diagrams of Quenched and Tempered Steels

64

General-Purpose Structural Steels

65

Fatigue Strength Diagrams of General-Purpose Structural Steels

66

Case Hardening Steels

67

Fatigue Strength Diagrams of Case Hardening Steels

68

Cold Rolled Steel Strips

69

Cast Steels for General Engineering Purposes

69

Round Steel Wire for Springs

70

Lamellar Graphite Cast Iron

71

Nodular Graphite Cast Iron

71

Copper-Tin- and Copper-Zinc-Tin Casting Alloys

72

Copper-Aluminium Casting Alloys

72

Aluminium Casting Alloys

73

Lead and Tin Casting Alloys for Babbit Sleeve Bearings

74

Conversion of Hardness Values

75

Values of Solids and Liquids

76

Coefficient of Linear Expansion

77

Iron-Carbon Diagram

77

Pitting and Tooth Root Fatigue Strength Values of Steels

77

Heat Treatment During Case Hardening of Case Hardening Steels

78

·

61

8

Materials Conversion of Fatigue Strength Values of Miscellaneous Materials Conversion of fatigue strength values of miscellaneous materials 3)

Tension

Material

Bending

1)

σW

σSch

σbW

Structural steel

0.45 . Rm

1.3 . σW

0.49 . Rm

1.5 . σbW 1.5 . Re

0.35 . Rm

1.1 . τW 0.7 . Re

Quenched and tempered steel

0.41 . Rm

1.7 . σW

0.44 . Rm

1.7 . σbW 1.4 . Re

0.30 . Rm

1.6 . τW 0.7 . Re

0.40 . Rm

1.6 . σW

0.41 . Rm

1.7 . σbW 1.4 . Re

0.30 . Rm

1.4 . τW 0.7 . Re

Grey cast iron

0.25 . Rm

1.6 . σw

0.37 . Rm

1.8 . σbW

–

0.36 . Rm

1.6 . τW

–

Light metal

0.30 . Rm

–

0.40 . Rm

–

–

0.25 . Rm

–

–

Case hardening steel

σbSch

1)

Torsion σbF

τW

τSch

τF

2)

8

1) For polished round section test piece of about 10 mm diameter. 2) Case-hardened; determined on round section test piece of about 30 mm diameter. Rm and Re of core material. 3) For compression, σSch is larger, e.g. for spring steel σdSch ≈ 1.3 . σSch For grey cast iron σdSch ≈ 3 . σSch

Type of load Ultimate stress val values es

62

Tension

Bending

Torsion

Tensile strength

Rm

–

–

Yield point

Re

σbF

τF

Fatigue strength under alternating stresses

σW

σbW

τW

Fatigue strength under fluctuating stresses

σSch

σbSch

τSch

·

Materials Mechanical Properties of Quenched and Tempered Steels Quenched and tempered steels (Extract from DIN EN 10083) Mechanical properties of steels in quenched and tempered condition Diameter

Material

up to 16 mm

above 16 up to 40 mm

Yield point (0.2 Num- Gr) ber N/mm2 min. Re Rp 0.2

Tensile strength N/mm2 Rm

C22

1.0402

350

550 – 700

300

500 – 650

C35

1.0501

430

630 – 780

370

600 – 750

C45

1.0503

500

700 – 850

430

650 – 800

C55

1.0535

550

800 – 950

500

C60

1.0601

580

850 –1000

C22E

1.1151

350

C35E

1.1181

430

C35R

1.1180

C45E




–

–

–

–

–

–

320

550 – 700

–

–

–

–

370

630 – 780

–

–

–

–

750 – 900

430

700 – 850

–

–

–

–

520

800 – 950

450

750 – 900

–

–

–

–

550 – 700

300

500 – 650

–

–

–

–

630 – 780

370

600 – 750

320

550 – 700

–

–

–

–

430

630 – 780

370

600 – 750

320

550 – 700

–

–

–

–

1.1191

500

700 – 850

430

650 – 800

370

630 – 780

–

–

–

–

C45R

1.1201

500

700 – 850

430

650 – 800

370

630 – 780

–

–

–

–

C55E

1.1203

550

800 – 950

500

750 – 900

430

700 – 850

–

–

–

–

C55R

1.1209

550

800 – 950

500

750 – 900

430

700 – 850

–

–

–

–

C60E

1.1221

580

850 –1000

520

800 – 950

450

750 – 900

–

–

–

–

C60R

1.1223

580

850 –1000

520

800 – 950

450

750 – 900

–

–

–

–

28Mn6

1.1170

590

780 – 930

490

690 – 840

440

640 – 790

–

–

–

–

38Cr2

1.7003

550

800 – 950

450

700 – 850

350

600 – 750

–

–

–

–

46Cr2

1.7006

650

900 –1100

550

800 – 950

400

650 – 800

–

–

–

–

34Cr4

1.7033

700

900 –1100

590

800 – 950

460

700 – 850

–

–

–

–

34CrS4

1.7037

700

900 –1100

590

800 – 950

460

700 – 850

–

–

–

–

37Cr4

1.7034

750

950 –1150

630

850 –1000

510

750 – 900

–

–

–

–

37CrS4

1.7038

750

950 –1150

630

850 –1000

510

750 – 900

–

–

–

–

41Cr4

1.7035

800 1000 –1200

660

900 –1100

560

800 – 950

–

–

–

–

41CrS4

1.7039

800 1000 –1200

660

900 –1100

560

800 – 950

–

–

–

–

25CrMo4 1.7218

700

900 –1100

600

800 – 950

450

700 – 850

400

650 – 800

–

–

34CrMo4 1.7220

800 1000 –1200

650

900 –1100

550

800 – 950

500

750 – 900

450

700 – 850

34CrMoS4 1.7226

800 1000 –1200

650

900 –1100

550

800 – 950

500

750 – 900

450

700 – 850

42CrMo4 1.7225

900 1100 –1300

750 1000 –1200

650

900 –1100

550

800 – 950

500

750 – 900

42CrMoS4 1.7227

900 1100 –1300

750 1000 –1200

650

900 –1100

550

800 – 950

500

750 – 900

50CrMo4 1.7228

900 1100 –1300

780 1000 –1200

700

900 –1100

650

850 –1000

550

800 – 950

36CrNiMo4 1.6511

900 1100 –1300

800 1000 –1200

700

900 –1100

600

800 – 950

550

750 – 900

34CrNiMo6 1.6582 1000 1200 –1400

900 1100 –1300

800 1000 –1200

700

900 –1100

600

800 – 950

30CrNiMo8 1.6580 1050 1250 –1450 1050 1250 –1450

900 1100 –1300

800 1000 –1200

700

900 –1100

700

650

850 –1000

600

800 – 950

800 1000 –1200

700

900 –1100

51CrV4

1.8159

900 1100 –1300

Re Rp 0.2


800 1000 –1200

30CrMoV9 1.7707 1050 1250 –1450 1020 1200 –1450

·

Yield point (0.2 Gr) N/mm2 min. Re Rp 0.2

above 100 up to 160 mm Yield point (0.2 Gr) N/mm2 min. Re Rp 0.2

Symbol

Yield point (0.2 Gr) N/mm2 min.


900 –1100

900 1100 –1300

Yield point Tensile (0.2 strength Gr) 2 N/mm N/mm2 min. Rm Re Rp 0.2

8

63

Materials Fatigue Strength Diagrams of Quenched and Tempered Steels Fatigue strength diagrams of quenched and tempered steels, DIN EN 10083 (in quenched and tempered condition, test piece diameter d = 10 mm)

a) Tension/compression fatigue strength c) Torsional fatigue strength

8

Quenched and tempered steels not illustrated may be used as follows: 34CrNiMo6 30CrMoV4

like 30CrNiMo8 like 30CrNiMo8

42CrMo4 36CrNiMo4 51CrV4

like 50CrMo4 like 50CrMo4 like 50CrMo4

34CrMo4

like 41Cr4

28Cr4

like 46Cr2

C45 C22

like C45E like C22E

C60 and C50 lie approximately between C45E and 46Cr2. C40, 32Cr2, C35, C30 and C25 lie approximately between C22E and C45E. Loading type I: static Loading type II: dynamic b) Bending fatigue strength

64

Loading type III: alternating ·

Materials General-Purpose Structural Steels

General-purpose structural steels (Extract from DIN EN 10025)

Number

Symbol acc. to DIN EN 10025

St33

1.0035

S185

U, N

St37-2

1.0037

S235JR

U, N

y Symbol ( iin )

1)

USt37-2 1.0036 S235JRG1

U, N

RSt37-2 1.0038 S235JRG2

U, N,

St37-3U 1.0114 S235JO St37-3N 1.0116 S235J2G3

U N

St44-2

1.0044

S275JR

U, N

Tensile strength Rm in N/mm2 for product thickness

Upper yield point ReH in N/mm2 (minimum) for product thickness

in mm

in mm

<3

≥3 > 100 ≤ 16 ≤ 100

310... 290... 540 510

185

> 16 ≤ 40

175 2)

> 40 ≤ 63

> 63 > 80 > 100 ≤ 80 ≤ 100

–

–

–

235

225

215

205

195

235

225

215

215

215

275

265

255

245

235

360 340 360... 340... 510 470

430... 410... 580 560

St44-3U 1.0143 S275JO St44-3N 1.0144 S275J2G3

U N

St52-3U 1.0553 S355JO St52-3N 1.0570 S355J2G3

U N

510... 490... 680 630

355

345

335

325

315

St50-2

1.0050

E295

U, N

490... 470... 660 610

295

285

275

265

255

St60-2

1.0060

E335

U, N

590... 570... 770 710

335

325

315

305

295

St70-2

1.0070

E360

U, N

690... 670... 900 830

365

355

345

335

325

To be agreed upon

Treatment condi condition

To be agreed upon

Material

8

1) N normalized; U hot-rolled, untreated 2) This value applies to thicknesses up to 25 mm only

·

65

Materials Fatigue Strength Diagrams of General-Purpose Structural Steels Fatigue strength diagrams of general-purpose structural steels, DIN EN 10025 (test piece diameter d = 10 mm)

E360 E335 E295 S275 S235

E360 E335 E295 S275 S235

a) Tension/compression fatigue strength c) Torsional fatigue strength

8

E360 E335 E295 S275 S235

Loading type I: static b) Bending fatigue strength

Loading type II: dynamic Loading type III: alternating

66

·

Materials Case Hardening Steels

Case hardening steels; Quality specifications (Extract from DIN EN 10084)


Symbol Num( in ber )

Treatment condition

1)

Material

For dia. 11 Yield point Re N/mm2 min.

For dia. 30

Tensile strength

Yield point

Rm N/mm2

Re N/mm2 min.

Tensile strength

For dia. 63 Yield point

Rm N/mm2

Re N/mm2 min.

Tensile strength Rm N/mm2

1.0301 1.1121

C10 C10E

390 390

640 – 790 640 – 790

295 295

490 – 640 490 – 640

– –

– –

C15 Ck15 Cm15

1.0401 1.1141 1.1140

C15 C15E C15R

440 440 440

740 – 890 740 – 890 740 – 890

355 355 355

590 – 790 590 – 790 590 – 790

– – –

– – –

15Cr13

1.7015

15Cr13

510

780 –1030

440

690 – 890

–

–

16MnCr5 16MnCrS5 20MnCr5 20MnCrS5

1.7131 1.7139 1.7147 1.7149

16MnCr5 16MnCrS5 20MnCr5 20MnCrS5

635 635 735 735

880 –1180 880 –1180 1080 –1380 1080 –1380

590 590 685 685

780 –1080 780 –1080 980 –1280 980 –1280

440 440 540 540

640 – 940 640 – 940 780 –1080 780 –1080

20MoCr4 1.7321 20MoCrS4 1.7323 25MoCrS4 1.7325

20MoCr4 20MoCrS4 25MoCrS4

635 635 735

880 –1180 880 –1180 1080 –1380

590 590 685

780 –1080 780 –1080 980 –1280

– – –

– – –

685 835

960 –1280 1230 –1480

635 785

880 –1180 1180 –1430

540 685

780 –1080 1080 –1330

835

1180 –1430

785

1080 –1330

685

980 –1280

15CrNi6 18CrNi8

1.5919 1.5920

15CrNi6 18CrNi8

17CrNiMo6 1.6587 18CrNiMo7-6

For details, see DIN EN 10084

C10 Ck10

1) Dependent on treatment, the Brinell hardness is different. Treatment condition

Meaning

C

treated for shearing load

G

soft annealed

BF

treated for strength

BG

treated for ferrite/pearlite structure

·

67

8

Materials Fatigue Strength Diagrams of Case Hardening Steels Fatigue strength diagrams of case hardening steels, DIN EN 10084 (Core strength after case hardening, test piece diameter d = 10 mm)

a) Tension/compression fatigue strength

c) Torsional fatigue strength

8

Case hardening steels not illustrated may be used as follows: 25MoCr4 like 20MnCr5 17CrNiMo6 like 18CrNi8

Loading type I: static Loading type II: dynamic b) Bending fatigue strength

68

Loading type III: alternating ·

Materials Cold Rolled Steel Strips Cast Steels for General Engineering Purposes Cold rolled steel strips (Extract from DIN EN 10132) Material

Tensile strength Rm 1) 2 N/mm maximum

Symbol ( in )

Number


C55 Ck55

1.0535 1.1203

C55 C55E

610

C60 Ck60

1.0601 1.1221

C60 C60E

620

C67 Ck67

1.0603 1.1231

C67 C67S

640

C75 Ck75

1.0605 1.1248

C75 C75S

640

Ck85 Ck101

1.1269 1.1274

C85S C100S

670 690

71Si7

1.5029

71Si7

800

67SiCr5

1.7103

67SiCr5

800

50CrV4

1.8159

50CrV4

740

1) Rm for cold rolled and soft-annealed condition; for strip thicknesses up to 3 mm

8

Cast steels for general engineering purposes (Extract from DIN 1681) Yield point

Material

Re, e Rp 0 0.2 2

Tensile strength Rm

Notched bar impact work (ISO-V-notch specimens) Av ≤ 30 mm

> 30 mm

Mean value J min.

1)

Symbol

Number

N/mm2 min.

N/mm2 min.

GS-38 ( GE200 )

1.0420

200

380

35

35

GS-45 ( GE240 )

1.0446

230

450

27

27

GS-52 ( GE260 )

1.0552

260

520

27

22

GS-60 ( GE300 )

1.0558

300

600

27

20

The mechanical properties apply to specimens which are taken from test pieces with thicknesses up to 100 mm. Furthermore, the yield point values also apply to the casting itself, in so far as the wall thickness is ≤ 100 mm. 1) Determined from three individual values each. ·

69

Materials Round Steel Wire for Springs

Round steel wire for springs (Extract from DIN EN 10218) Diameter of wire

Grade of wire A

70

C

D

Tensile strength Rm in N/mm2

mm

8

B

0.07

–

–

–

2800 – 3100

0.3

–

2370 – 2650

–

2660 – 2940

1

1720 – 1970

1980 – 2220

–

2230 – 2470

2

1520 – 1750

1760 – 1970

1980 – 2200

1980 – 2200

3

1410 – 1620

1630 – 1830

1840 – 2040

1840 – 2040

4

1320 – 1520

1530 – 1730

1740 – 1930

1740 – 1930

5

1260 – 1450

1460 – 1650

1660 – 1840

1660 – 1840

6

1210 – 1390

1400 – 1580

1590 – 1770

1590 – 1770

7

1160 – 1340

1350 – 1530

1540 – 1710

1540 – 1710

8

1120 – 1300

1310 – 1480

1490 – 1660

1490 – 1660

9

1090 – 1260

1270 – 1440

1450 – 1610

1450 – 1610

10

1060 – 1230

1240 – 1400

1410 – 1570

1410 – 1570

11

–

1210 – 1370

1380 – 1530

1380 – 1530

12

–

1180 – 1340

1350 – 1500

1350 – 1500

13

–

1160 – 1310

1320 – 1470

1320 – 1470

14

–

1130 – 1280

1290 – 1440

1290 – 1440

15

–

1110 – 1260

1270 – 1410

1270 – 1410

16

–

1090 – 1230

1240 – 1390

1240 – 1390

17

–

1070 – 1210

1220 – 1360

1220 – 1360

18

–

1050 – 1190

1200 – 1340

1200 – 1340

19

–

1030 – 1170

1180 – 1320

1180 – 1320

20

–

1020 – 1150

1160 – 1300

1160 – 1300

·

Materials Lamellar Graphite Cast Iron Nodular Graphite Cast Iron Lamellar graphite cast iron (Extract from DIN EN 1561) Tensile Brinell Compressive strength 1) hardness strength 2) Rm σdB 1)

Wall thicknesses in mm

Material Symbol

Number

Symbol acc. to DIN 1691

above

up to

N/mm2

EN-GJL-100

EN-JL1010

GG-10

5

40

GG-15

10 20 40 80

EN-GJL-150

EN-GJL-200

EN-GJL-250

EN-JL1020

EN-JL1030

EN-JL1040

HB 30

N/mm2

min. 100 2)

–

–

20 40 80 150

130 110 95 80

225 205 – –

600

GG-20

10 20 40 80

20 40 80 150

180 155 130 115

250 235 – –

720

GG-25

10 20 40 80

20 40 80 150

225 195 170 155

265 250 – –

840

20 40 80 150

270 240 210 195

285 265 – –

960

20 40 80 150

315 280 250 225

285 275 – –

1080

EN-GJL-300

EN-JL1050

GG-30

10 20 40 80

EN-GJL-350

EN-JL1060

GG-35

10 20 40 80

The values apply to castings which are made in sand moulds or moulds with comparable heat diffusibility. 1) These values are reference values. 2) Values in the separately cast test piece with 30 mm diameter of the unfinished casting. Nodular graphite cast iron (Extract from DIN EN 1563) Properties in cast-on test pieces Wall thickness of casting

Material

Symbol

Number


EN-GJS-400-18U-LT EN-JS1049 GGG-40.3

mm from above from above

30 60 30 60

0.2% Thickness Tensile proof of cast-on strength stress test piece Rm Rp0.2 mm

N/mm2 N/mm2

up to 60 up to 200 up to 60 up to 200

40 70 40 70

390 370 390 370

250 240 250 240

EN-GJS-400-15U

EN-JS1072

GGG-40

EN-GJS-500-7U

EN-JS1082

GGG-50

from 30 up to 60 above 60 up to 200

40 70

450 420

300 290

EN-GJS-600-3U

EN-JS1092

GGG-60


40 70

600 550

360 340

EN-GJS-700-2U

EN-JS1102

GGG-70


40 70

700 650

400 380

·

71

8

Materials Copper-Tin and Copper-Zinc-Tin Casting Alloys Copper-Aluminium Casting Alloys Copper-tin and copper-zinc-tin casting alloys (Extract from DIN EN 1982) Material

8

Condition on delivery

0.2% Tensile proof stress 1) strength 1) Rp0.2 min. in N/mm2

Rm min. in N/mm2

Symbol

Number


CuSn12-C-GS CuSn12-C-GZ CuSn12-C-GC

CC483K

G-CuSn12 GZ-CuSn12 GC-CuSn12

Sand-mould cast iron Centrifugally cast iron Continuously cast iron

140 150 140

260 280 280

CuSn12Ni-C-GS CuSn12Ni-C-GZ CuSn12Ni-C-GC

CC484K

G-CuSn12Ni GZ-CuSn12Ni GC-CuSn12Ni


160 180 170

280 300 300

CuSn12Pb2-C-GS CuSn12Pb2-C-GZ CuSn12Pb2-C-GC

CC482K

G-CuSn12Pb GZ-CuSn12Pb GC-CuSn12Pb


140 150 140

260 280 280

CuSn10-Cu-GS

CC480K

G-CuSn10

Sand-mould cast iron

130

270

CuSn7Zn4Pb7-C-GS CuSn7Zn4Pb7-C-GZ CuSn7Zn4Pb7-C-GC

CC493K

G-CuSn7ZnPb GZ-CuSn7ZnPb GC-CuSn7ZnPb


120 130 120

240 270 270

CuSn7Zn2Pb3-C-GS

CC492K

G-CuSn6ZnNi


140

270

CuSn5Zn5Pb5-C-GS

CC491K

G-CuSn5ZnPb


90

220

CuSn3Zn8Pb5-C-GS

CC490K

G-CuSn2ZnPb


90

210

1) Material properties in the test bar Copper-aluminium casting alloys (Extract from DIN EN 1982) Material

Condition on delivery

0.2% Tensile proof stress 1) strength 1) Rp0.2 min. in N/mm2

Rm min. in N/mm2

Symbol

Number


CuAl10Fe2-C-GS CuAl10Fe2-C-GM CuAl10Fe2-C-GZ

CC331G

G-CuAl10Fe GK-CuAl10Fe GZ-CuAl10Fe

Sand-mould cast iron Chilled casting Centrifugally cast iron

180 200 200

500 550 550

CuAl10Ni3Fe2-C-GS CuAl10Ni3Fe2-C-GK CuAl10Ni3Fe2-C-GZ

CC332G

G-CuAl9Ni GK-CuAl9Ni GZ-CuAl9Ni


200 230 250

500 530 600

CuAl10Fe5Ni5-C-GS CuAl10Fe5Ni5-C-GM CuAl10Fe5Ni5-C-GZ CuAl10Fe5Ni5-C-GC

CC333G

G-CuAl10Ni GK-CuAl10Ni GZ-CuAl10Ni GC-CuAl10Ni

Sand-mould cast iron Chilled casting Centrifugally cast iron Continuously cast iron

270 300 300 300

600 600 700 700

CuAl11Fe6Ni6-C-GS CuAl11Fe6Ni6-C-GM CuAl11Fe6Ni6-C-GZ

CC334G

G-CuAl11Ni GK-CuAl11Ni GZ-CuAl11Ni


320 400 400

680 680 750

1) Material properties in the test bar

72

·

Materials Aluminium Casting Alloys

Aluminium casting alloys (Extract from DIN EN 1706) Material

0.2% proof stress Rp0.2

Tensile strength Rm

Symbol

Number

Symbol acc. to DIN 1725-2

in N/mm2

in N/mm2

AC-AlCu4MgTi

AC-21000

G-AlCu4TiMg

200 up to 220

300 up to 320

AC-AlCu4Ti

AC-21100

G-AlCu4Ti

180 up to 220

280 up to 330

AC-AlSi7Mg

AC-42100

G-AlSi7Mg

180 up to 210

230 up to 290

AC-AlSi10Mg(a)

AC-43000

G-AlSi10Mg

80 up to 220

150 up to 240

AC-AlSi10Mg(Cu)

AC-43200

G-AlSi10Mg(Cu)

80 up to 200

160 up to 240

AC-AlSi9Mg

AC-43300

G-AlSi9Mg

180 up to 210

230 up to 290

AC-AlSi10Mg(Fe)

AC-43400

G-AlSi10Mg

140

240

AC-AlSi11

AC-44000

G-AlSi11

70 up to 80

150 up to 170

AC-AlSi12(a)

AC-44200

G-AlSi12

70 up to 80

150 up to 170

AC-AlSi12(Fe)

AC-44300

GD-AlSi12

130

240

AC-AlSi6Cu4

AC-45000

G-AlSi6Cu4

90 up to 100

150 up to 170

AC-AlSi9Cu3(Fe)

AC-46000

GD-AlSi9Cu3

140

240

AC-AlSi8Cu3

AC-46200

G-AlSi9Cu3

90 up to 140

150 up to 240

AC-AlSi12(Cu)

AC-47000

G-AlSi12(Cu)

80 up to 90

150 up to 170

AC-AlSi12Cu1(Fe)

AC-47100

GD-AlSi12(Cu)

140

240

AC-AlMg3(a)

AC-51100

G-AlMg3

70

140 up to 150

AC-AlMg9

AC-51200

GD-AlMg9

130

200

AC-AlMg5

AC-51300

G-AlMg5

90 up to 100

160 up to 180

AC-AlMg5(Si)

AC-51400

G-AlMg5Si

100 up to 110

160 up to 180

·

8

73

Materials Lead and Tin Casting Alloys for Babbit Sleeve Bearings Lead and tin casting alloys for babbit sleeve bearings (Extract from DIN ISO 4381) Brinell hardness 1) HB 10/250/180

Material

0.2% proof stress Rp 0.2 in N/mm2

1)

Symbol

Number

20 °C

50 °C

120 °C

20 °C

50 °C

100 °C

PbSb15SnAs

2.3390

18

15

14

39

37

25

PbSb15Sn10

2.3391

21

16

14

43

32

30

PbSb14Sn9CuAs

2.3392

22

22

16

46

39

27

PbSb10Sn6

2.3393

16

16

14

39

32

27

SnSb12Cu6Pb

2.3790

25

20

12

61

60

36

SnSb8Cu4

2.3791

22

17

11

47

44

27

SnSb8Cu4Cd

2.3792

28

25

19

62

44

30

1) Material properties in the test bar

8

74

·

Materials Conversion of Hardness Values (DIN EN ISO 18265) Tensile strength

Vickers hardness

Brinell hardness

ǐ

. N/mm2 (F≥98N) 0.102

2)

F D2

= 30

N

Ǒ

mm2

Rockwell hardness

Tensile strength

HRB HRC HRA HRD 1)

255 270 285 305 320

80 85 90 95 100

76.0 80.7 85.5 90.2 95.0

335 350 370 385 400

105 110 115 120 125

99.8 105 109 114 119

415 430 450 465 480

130 135 140 145 150

124 128 133 138 143

495 510 530 545 560

155 160 165 170 175

147 152 156 162 166

575 595 610 625 640

180 185 190 195 200

171 176 181 185 190

87.1

660 675 690 705 720

205 210 215 220 225

195 199 204 209 214

740 755 770 785 800

230 235 240 245 250

219 223 228 233 238

96.7

820 835 850 865 880

255 260 265 270 275

900 915 930 950 965 995 1030 1060 1095 1125

Vickers hardness

Brinell hardness

ǐ

. N/mm2 (F≥98N) 0.102

2)

F D2

= 30

N

Ǒ

mm2

Rockwell hardness HRC HRA HRD 1)

1155 1190 1220 1255 1290

360 370 380 390 400

342 352 361 371 380

36.6 37.7 38.8 39.8 40.8

68.7 69.2 69.8 70.3 70.8

52.8 53.6 54.4 55.3 56.0

1320 1350 1385 1420 1455

410 420 430 440 450

390 399 409 418 428

41.8 42.7 43.6 44.5 45.3

71.4 71.8 72.3 72.8 73.3

56.8 57.5 58.2 58.8 59.4

1485 1520 1555 1595 1630

460 470 480 490 500

437 447 (456) (466) (475)

46.1 46.9 47.7 48.4 49.1

73.6 74.1 74.5 74.9 75.3

60.1 60.7 61.3 61.6 62.2

1665 1700 1740 1775 1810

510 520 530 540 550

(485) (494) (504) (513) (523)

49.8 50.5 51.1 51.7 52.3

75.7 76.1 76.4 76.7 77.0

62.9 63.5 63.9 64.5 64.8

91.5

1845 1880 1920 1955 1995

560 570 580 590 600

(532) (542) (551) (561) (570)

53.0 53.6 54.1 54.7 55.2

77.4 77.8 78.0 78.4 78.6

65.4 65.8 66.2 66.7 67.0

92.5 93.5 94.0 95.0 96.0

2030 2070 2105 2145 2180

610 620 630 640 650

(580) (589) (599) (608) (618)

55.7 56.3 56.8 57.3 57.8

78.9 79.2 79.5 79.8 80.0

67.5 67.9 68.3 68.7 69.0

98.1 20.3 60.7 40.3 21.3 61.2 41.1 99.5 22.2 61.6 41.7

660 670 680 690 700

58.3 58.8 59.2 59.7 60.1

80.3 80.6 80.8 81.1 81.3

69.4 69.8 70.1 70.5 70.8

242 247 252 257 261

23.1 (101) 24.0 24.8 (102) 25.6 26.4

62.0 62.4 62.7 63.1 63.5

42.2 43.1 43.7 44.3 44.9

720 740 760 780 800

61.0 61.8 62.5 63.3 64.0

81.8 82.2 82.6 83.0 83.4

71.5 72.1 72.6 73.3 73.8

280 285 290 295 300

266 271 276 280 285

(104) 27.1 27.8 (105) 28.5 29.2 29.8

63.8 64.2 64.5 64.8 65.2

45.3 46.0 46.5 47.1 47.5

820 840 860 880 900

64.7 65.3 65.9 66.4 67.0

83.8 84.1 84.4 84.7 85.0

74.3 74.8 75.3 75.7 76.1

310 320 330 340 350

295 304 314 323 333

31.0 32.3 33.3 34.4 35.5

65.8 66.4 67.0 67.6 68.1

48.4 49.4 50.2 51.1 51.9

920 940

67.5 85.3 76.5 68.0 85.6 76.9

41.0 48.0 52.0 56.2 62.3 66.7 71.2 75.0 78.7 81.7 85.0

89.5

The figures in brackets are hardness values outside the domain of definition of standard hardness test methods which, however, in practice are frequently used as approximate values. Furthermore, the Brinell hardness values in brackets apply only if the test was carried out with a carbide ball. 1) Internationally usual, e.g. ASTM E 18-74 (American Society for Testing and Materials) 2) Calculated from HB = 0.95 HV (Vickers hardness) Determination of Rockwell hardness HRA, HRB, HRC, and HRD acc. to DIN EN 10109 Part 1 Determination of Vickers hardness acc. to DIN 50133 Part 1 Determination of Brinell hardness acc. to DIN EN 10003 Part 1 Determination of tensile strength acc. to DIN EN 10002 Part 1 and Part 5 ·

75

8

Materials Values of Solids and Liquids

Mean density of the earth = 5.517 g/cm3

Values of solids and liquids Substance (solid)

Symbol

Density g/cm3

8

Agate Aluminium Aluminium bronze Antimony Arsenic Asbestos Asphaltum Barium Barium chloride Basalt, natural Beryllium Concrete Lead Boron (amorph.) Borax Limonite Bronze Chlorine calcium Chromium Chromium nickel Delta metal Diamond Iron, pure Grease Gallium Germanium Gypsum Glass, window Mica Gold Granite Graphite Grey cast iron Laminated fabric Hard rubber Hard metal K20

Woods Indium Iridium Cium Potassium Limestone Calcium Calcium oxide (lime) Caoutchouc, crude Cobalt Salt, common Coke Constantan Corundum (AL2O3) Chalk Copper Leather, dry Lithium Magnesium Magnesium, alloyed Manganese Marble Red lead oxide Brass Molybdenum Monel metal Sodium Nickel silver

Nickel Niobium Osmium Palladium Paraffin Pitch Phosphorus (white) Platinum Polyamide A, B

76

Al Sb As

Ba Be Pb B

Cr C Fe Ga Ge

Au C

In Ir Cd K Ca Co

Cu Li Mg Mn

Mo Na Ni Nb Os Pd P Pt

2.5...2.8 2.7 7.7 6.67 5.72 ≈2.5 1.1...1.5 3.59 3.1 2.7...3.2 1.85 ≈2 11.3 1.73 1.72 3.4...3.9 8.83 2.2 7.1 7.4 8.6 3.5 7.86 0.92...0.94 5.9 5.32 2.3 ≈2.5 ≈2.8 19.29 2.6...2.8 2.24 7.25 1.3...1.42 ≈1.4 14.8 0.45...0.85 7.31 22.5 8.64 0.86 2.6 1.55 3.4 0.95 8.8 2.15 1.6...1.9 8.89 3.9...4 1.8...2.6 8.9 0.9....1 0.53 1.74 1.8...1.83 7.43 2.6...2.8 8.6...9.1 8.5 10.2 8.8 0.98 8.7 8.9 8.6 22.5 12 0.9 1.25 1.83 21.5 1.13

Melting point t in °C

Thermal conductivity λ at 20 °C W/(mK)

≈1600 11.20 658 204 1040 128 630 22.5 – – ≈1300 – 80...100 0.698 704 – 960 – – 1.67 1280 1.65 – ≈1 327.4 34.7 2300 – 740 – 1565 – 910 64 774 – 1800 69 1430 52.335 950 104.7 – – 1530 81 30...175 0.209 29.75 – 936 58.615 1200 0.45 ≈700 0.81 ≈1300 0.35 1063 310 – 3.5 ≈3800 168 1200 58 – 0.34...0.35 – 0.17 2000 81 – 0.12...0.17 156 24 2450 59.3 321 92.1 63.6 110 – 2.2 850 – 2572 – 125 0.2 1490 69.4 802 – – 0.184 1600 23.3 2050 12...23 – 0.92 1083 384 – 0.15 179 71 657 157 650 69.8..145.4 1250 30 1290 2.8 – 0.7 900 116 2600 145 ≈1300 19.7 97.5 126 1020 48 1452 59 2415 54.43 2500 – 1552 70.9 52 0.26 – 0.13 44 – 1770 70 ≈250 0.34

Substance (solid) Porcelain Pyranite Quartz-flint Radium Rhenium Rhodium Gunmetal Rubidium Ruthenium Sand, dry Sandstone Brick, fire Slate Emery Sulphur, rhombic Sulphur, monoclinic Barytes Selenium, red Silver Silicon Silicon carbide Sillimanite Soapstone (talcous) Steel, plain + low-alloy stainless non-magnetic Tungsten steel 18W Hard coal Strontium Tantalum Tellurium Thorium Titanium Tombac Clay Uranium 99.99% Vanadium Soft rubber White metal Bismuth Wolfram Cesium Cement, hard Cerium Zinc Tin Zirconium

Substance (liquid) Ether Benzine Benzole, pure Diesel oil Glycerine Resin oil Fuel oil EL Linseed oil Machinery oil Methanol Methyl chloride Mineral oil Petroleum ether Petroleum Mercury Hydrochloric acid 10% Sulphuric acid, strong Silicon fluid

Symbol

Ra Re Rh Rb Ru

S S Se Ag Si

Sr Ta Te Th Ti U V

Bi W Cs Ce Zn Sn Zr

SymSym y bol

Hg

Density

Melting point

g/cm3

t in °C

Thermal conductivity λ at 20 °C W/(mK)

2.2...2.5 ≈1650 ≈1 3.3 1800 8.14 2.5...2.8 1480 9.89 5 700 – 21 3175 71 12.3 1960 88 8.8 950 38 1.52 39 58 12.2 2300 106 1.4...1.6 1480 0.58 2.1...2.5 ≈1500 2.3 1.8...2.3 ≈2000 ≈1.2 2.6...2.7 ≈2000 ≈0.5 4 2200 11.6 2.07 112.8 0.27 1.96 119 0.13 4.5 1580 – 4.4 220 0.2 10.5 960 407 2.33 1420 83 3.12 – 15.2 2.4 1816 1.69 2.7 – 3.26 7.9 1460 47...58 7.9 1450 14 8 1450 16.28 8.7 1450 26 1.35 – 0.24 2.54 797 0.23 16.6 2990 54 6.25 455 4.9 11.7 ≈1800 38 4.5 1670 15.5 8.65 1000 159 1.8...2.6 1500..1700 0.93...1.28 18.7 1133 28 6.1 1890 31.4 1...1.8 – 0.14...0.23 7.5...10.1 300...400 34.9...69.8 9.8 271 8.1 19.2 3410 130 1.87 29 – 2...2.2 – 0.9...1.2 6.79 630 – 6.86 419 110 7.2 232 65 6.5 1850 22

D Density it at g/cm3 0.72 ≈0.73 0.83 0.83 1.26 0.96 ≈0.83 0.93 0.91 0.8 0.95 0.91 0.66 0.81 13.55 1.05 1.84 0.94

°C 20 15 15 15 20 20 20 20 15 15 15 20 20 20 20 15 15 20

Thermal g Boiling point at conductivity λ 1 013MPa 1.013MPa at 20 °C C °C W/(mK) 35 0.14 25...210 0.13 80 0.14 210...380 0.15 290 0.29 150...300 0.15 > 175 0.14 316 0.17 380...400 0.125 65 0.21 24 0.16 > 360 0.13 > 40 0.14 > 150 0.13 357 10 102 0.5 338 0.47 – 0.22

·

Materials Coefficient of Linear Expansion; Iron-Carbon Diagram; Fatigue Strength Values for Gear Materials Coefficients of linear expansion of some substances at 0 ... 100 °C

Coefficient of linear expansion α The coefficient of linear expansion α gives the fractional expansion of the unit of length of a substance per 1 degree K rise in temperature. For the linear expansion of a body applies: l + l 0 @ @ T where Δl: change of length l0: original length α: coefficient of linear expansion ΔT: rise of temperature

Substance

α [10-6/K]

Aluminium alloys Grey cast iron (e.g. GG-20, GG-25) Steel, plain and low-alloy Steel, stainless (18CrNi8) Steel, rapid machining steel Copper Brass CuZn37 Bronze CuSn8

21 ... 24 10.5 11.5 16 11.5 17 18.5 17.5

Iron-carbon diagram Mixed crystals

Melting + δ-mixed crystals

Temperature in °C

Mixed crystals

Melting + γ-mixed crystals

γ-mixed crystals (austenite)

Melting + primary cementite

γ-mixed crystals + sec. cementite + ledeburite

Mixed crystals

Primary cementite + ledeburite

γ-m.c. + sec.cem.

Sec.cem. + pearlite

8 Sec.cem. + pearlite + ledeburite

pearlite

Ledeburite

Mixed crystals (ferrite)

Pearlite

(cubic face centered)

(cementite)

Melting

Primary cementite + ledeburite

Carbon content in weight percentage

(cubic body centered)

Cementite content in weight percentage

Pitting and tooth root fatigue strength of steels Material symbol

Hardness on finished gear HV1

σHlim

σFlim

N/mm2

N/mm2

16MnCr5 20MnCr5 18CrNiMo7-6

720 680 740

1470 1470 1500

430 430 500

Quenched and tempered steels, quenched and tempered

30CrNiMo8 34CrNiMo6 42CrMo4

290 310 280

730 770 740

300 310 305

Quenched and tempered steels, nitrided

34CrNiMo6 42CrMo4

630 600

1000 1000

370 370

Grade of steel Case hardening steels, case-hardened

·

77

Materials Heat Treatment During Case Hardening of Case Hardening Steels Heat treatment during case hardening of case hardening steels acc. to DIN EN 10084 Usual heat treatment during case hardening A. Direct hardening or double hardening

B. Single hardening

C. Hardening after isothermal transformation

Direct hardening from carburizing temperature

Single hardening from core or case hardening temperature

Hardening after isothermal transformation in the pearlite stage (e)

Direct hardening after lowering to hardening temperature

Single hardening after intermediate annealing (soft annealing) (d)

Hardening after isothermal transformation in the pearlite stage (e) and cooling-down to room temperature

a b c d e

Double hardening

carburizing temperature hardening temperature tempering temperature intermediate annealing (soft annealing) temperature tem erature transformation temperature in the pearlite stage

Usual case hardening temperatures Material

8 y Symbol

Number

a

b

Carburizing temperature 1)

Core hardening Case hardening temperature 2) temperature 2)

°C C10 C10E C15

1.0301 1.1121 1.0401

15Cr3 17Cr3 16MnCr5 16MnCrS5 20MnCr5 20MnCrS5 20MoCr4 20MoCrS4 20NiCrMo2-2 20NiCrMoS2-2

1.7015 1.7016 1.7131 1.7139 1.7147 1.7149 1.7321 1.7323 1.6523 1.6526

15CrNi6 18CrNiMo7-6

1.5919 1.6587

°C

c

°C

880 up to 920

880 up to 980

860 up to 900

830 up to 870

Quenchant

780 up to 820

Tempering

°C

With regard to the properties of th component, the t the selection of the quenchant depends on the hardenability or case-hardenability of the steel, the shape and cross section of the work piece to be hardened, as well as on the effect of the quenchant.

150 up to 200

1) Decisive criteria for the determination of the carburizing temperature are mainly the required time of carburizing, the chosen carburizing agent, and the plant available, the provided course of process, as well as the required structural constitution. For direct hardening, carburizing usually is carried out at temperatures below 950 °C. In special cases, carburizing temperatures up to above 1000 °C are applied. 2) In case of direct hardening, quenching is carried out either from the carburizing temperature or any lower temperature. In particular if there is a risk of warping, lower hardening temperatures are preferred.

78

·


Lubricating Oils

Page

Viscosity-Temperature-Diagram for Mineral Oils

80

Viscosity-Temperature-Diagram for Synthetic Oils of Poly-α-Olefine Base

81

Viscosity-Temperature-Diagram for Synthetic Oils of Polyglycole Base

82

Kinematic Viscosity and Dynamic Viscosity

83

Viscosity Table for Mineral Oils

84

9

·

79

Lubricating Oils Viscosity-Temperature-Diagram for Mineral Oils

Kinematic viscosity (mm2/s)

Viscosity-temperature-diagram for mineral oils

9

Temperature (°C)

80

·

Lubricating Oils Viscosity-Temperature-Diagram for Synthetic Oils of Poly-α-Olefine Base


Viscosity-temperature-diagram for synthetic oils of poly-α-olefine base

9

Temperature (°C)

·

81

Lubricating Oils Viscosity-Temperature-Diagram for Synthetic Oils of Polyglycole Base


Viscosity-temperature-diagram for synthetic oils of polyglycole base

9

Temperature (°C)

82

·

Lubricating Oils Kinematic Viscosity and Dynamic Viscosity for Mineral Oils at any Temperature Kinematic viscosity υ Quantities for the determination of the kinematic viscosity VG grade

W40 [–]

m [–]

32 46 68 100 150 220

0.18066 0.22278 0.26424 0.30178 0.33813 0.36990

3.7664 3.7231 3.6214 3.5562 3.4610 3.4020

320 460 680 1000 1500

0.39900 0.42540 0.45225 0.47717 0.50192

3.3201 3.3151 3.2958 3.2143 3.1775

W = m (2.49575 – lgT) + W40

(1)

W

+ 10 10 * 0.8 m [-]: T [K]: W40 [-]: W [-]: υ [cSt]:

(2)

slope thermodynamic temperature auxiliary quantity at 40 °C auxiliary quantity kinematic viscosity

1)

1) T = t + 273.15 [K] Dynamic viscosity η η = υ . . 0.001

(3)

= 15 − (t – 15) . 0.0007

(4)

t [°C]: 15 [kg/dm3]: [kg/dm3]: υ [cSt]: η [Ns/m2]:

9

temperature density at 15 °C density kinematic viscosity dynamic viscosity

Density 15 in kg/dm3 of lubricating oils for gear units

2)

(Example)

VG grade

68

100

150

220

320

460

680

ARAL Degol BG Plus

–

0.888

0.892

0.897

0.895

0.902

0.905

MOBIL Mobilgear 600 XP

0.880

0.880

0.890

0.890

0.900

0.900

0.910

MOBIL Mobilgear XMP

–

0.890

0.896

0.900

0.903

0.909

0.917

CASTROL Optigear BM

0.890

0.893

0.897

0.905

0.915

0.920

0.930

CASTROL Tribol 1100

0.888

0.892

0.897

0.904

0.908

0.916

0.923

2) Mineral base gear oils in accordance with designation CLP as per DIN 51517 Part 3. These oils comply with the minimum requirements as specified in DIN 51517 Part 3. They are suitable for operating temperatures from -10 °C up to +90 °C (briefly +100 °C). ·

83

Lubricating Oils Viscosity Table for Mineral Oils

Approx. Saybolt assignment universal to AGMA seconds lubricant motor(SSU) N° at car motor °C at 40 100 °C 40 °C gear oils (mean 1) oils value)

Mean viscosity (40 °C) and approx. viscosities in mm2/s (cSt) at

ISO-VG ISO VG DIN 51519

A rox. Approx. assignment to previous 20 °C DIN 51502

40 °C

50 °C

cSt

cSt

cSt

Engler

cSt

5

2

8 (1.7 E)

46 4.6

4

13 1.3

15 1.5

7

4

12 (2 E)

68 6.8

5

14 1.4

20 2.0

10

9

21 (3 E)

10

8

17 1.7

25 2.5

15

–

34

15

11

19 1.9

35 3.5

55

22

15

23 2.3

45 4.5

22

16

32

1)

SAE

SAE

5W

10 W 88

32

21

3

55 5.5

137

46

30

4

65 6.5

214

1 EP

219

68

43

6

85 8.5

316

2 2 EP 2.2

15 W 20 W 20

68

345

100

61

8

11

464

3.3 EP

30

92

550

150

90

12

15

696

4 4 EP 4.4

40

865

220

125

16

19

1020

5 5 EP 5.5

50

70 W 75 W

25 46 36 68

80 W

49 100 150

9

220

114 144

85 W

320

169

1340

320

180

24

24

1484

6 6 EP 6.6

460

225

2060

460

250

33

30

2132

7 EP

680

324

3270

680

360

47

40

3152

8 EP

1000

5170

1000

510

67

50

1500

8400

1500

740

98

65

90

140

250

1) Approximate comparative value to ISO VG grades

84

·


Cylindrical Gear Units Symbols and Units General Introduction

Page 86 + 87 88

Geometry of Involute Gears Concepts and Parameters Associated With Involute Teeth Standard Basic Rack Tooth Profile Module Tool Reference Profile Generating Tooth Flanks Concepts and Parameters Associated With Cylindrical Gears Geometric Definitions Pitches Addendum Modification Concepts and Parameters Associated With a Cylindrical Gear Pair Mating Quantities Ratios Summary of the Most Important Formulae Gear Teeth Modifications

88 88 89 89 90 91 91 91 92 93 93 93 94 95 – 97 98 + 99

Load Carrying Capacity of Involute Gears Scope of Application and Purpose Basic Details General Factors Application Factor Dynamic Factor Face Load Factor Transverse Load Factor Tooth Flank Load Carrying Capacity Effective Hertzian Pressure Permissible Hertzian Pressure Tooth Root Load Carrying Capacity Effective Tooth Root Stress Permissible Tooth Root Stress Safety Factors Calculation Example

99 + 100 100 + 101 102 102 102 102 102 103 103 103 + 104 104 104 – 106 106 106 106 + 107

10

Gear Unit Types Standard Designs Load Sharing Gear Units Comparisons Load Value Referred Torques Efficiencies Example

107 107 107 + 108 108 109 + 110 110 110

Noise Emitted by Gear Units Definitions Measurements Determination via Sound Pressure Determination via Sound Intensity Prediction Possibilities of Influencing ·

111 + 112 112 112 + 113 113 113 + 114 114

85

Cylindrical Gear Units Symbols and Units for Cylindrical Gear Units

a

mm

Centre distance

n

min-1

Speed

ad

mm

Reference centre distance

p

N/mm2

b

mm

Facewidth

p

mm

Pitch on the reference circle

pbt

Pitch on the base circle

mm

Bottom clearance between standard basic rack tooth profile and counter profile

mm

pe

mm

Normal base pitch

d

mm

Reference diameter

pen

mm

Normal base pitch at a point

da

mm

Tip diameter

pet

mm

Normal transverse pitch Axial pitch

Pressure; compression

db

mm

Base diameter

pex

mm

df

mm

Root diameter

pt

mm

Transverse base pitch; reference circle pitch

dw

mm

Pitch diameter

e

mm

Spacewidth on the reference cylinder

prP0

mm

Protuberance value on the tool’s standard basic rack tooth profile

ep

mm

Spacewidth on the standard basic rack tooth profile

q

mm

Machining allowance on the cylindrical gear tooth flanks

r

mm

Reference circle radius; radius

ra

mm

Tip radius

rb

mm

Base radius

f

Hz

Frequency

gα

mm

Length of path of

h

mm

Tooth depth

ha

mm

Addendum

haP

mm

Addendum of the standard basic rack tooth profile

rw

mm

Radius of the working pitch circle

haP0

mm

Addendum of the tool’s standard basic rack tooth profile

s

mm

Tooth thickness on the reference circle

hf

mm

Dedendum

san

mm

Tooth thickness on the tip circle

hfP

mm

Dedendum of the standard basic rack tooth profile

sp

mm

Tooth thickness of the standard basic rack tooth profile

hfP0

mm

Dedendum of the tool’s standard basic rack tooth profile

sP0

mm

hp

mm

Tooth depth of the standard basic rack tooth profile

Tooth thickness of the tool’s standard basic rack tooth profile

u

–

hP0

mm

Tooth depth of the tool’s standard basic rack tooth profile

v

m/s

hprP0

mm

Protuberance height of the tool’s standard basic rack tooth profile

w

N/mm

x

–

Addendum modification coefficient

xE

–

Generating addendum modification coefficient

Addendum modification factor

z

–

Number of teeth

10

Working depth of the standard basic rack tooth profile and the counter profile

Gear ratio Circumferential speed on the reference circle Line load

hwP

mm

k

–

m

mm

Module

A

m2

Gear teeth surface

mn

mm

Normal module

As

mm

Tooth thickness deviation

BL

N/mm2

mt

86

mm

Transverse module

Load value ·

Cylindrical Gear Units Symbols and Units for Cylindrical Gear Units

ZX

–

α

Degree

Transverse pressure angle at a point; pressure angle

rad

Angle α in the circular measure ǒ + @ ń180Ǔ

αat

Degree

Transverse pressure angle at the tip circle

αn

Degree

Normal pressure angle

αP

Degree

Pressure angle at a point of the standard basic rack tooth profile

αP0

Degree

Pressure angle at a point of the tool’s standard basic rack tooth profile

αprP0 Degree

Protuberance pressure angle at a point

αt

Degree

Transverse pressure angle at the reference circle

αwt

Degree

Working transverse pressure angle at the pitch circle

β

Degree

Helix angle at the reference circle

βb

Degree

Base helix angle

εα

–

Transverse ratio

εβ

–

Overlap ratio

εγ

–

Total ratio

ζ

Degree

Factor of safety from pitting

η

–

mm

Radius of curvature

aP0

mm

Tip radius of curvature of the tool’s standard basic rack tooth profile

fP0

mm

Root radius of curvature of the tool’s standard basic rack tooth profile

σH

N/mm2

D

mm

Fn

N

Load

Ft

N

Nominal peripheral force at the reference circle

G

N

Weight

HV1

–

Vickers hardness at F = 9.81 N

KA

–

Application factor

KFα

–

Transverse load factor (for tooth root stress)

KFβ

–

Face load factor (for tooth root stress)

KHα

–

Transverse load factor (for stress)

KHβ

–

Kv

–

LpA

dB

Sound pressure level, A-weighted

LWA

dB

Sound power level, A-weighted

P

kW

Nominal power driven machine

RZ

μm

Mean peak-to-valley height

SF

–

SH

–

Construction dimension

Face load factor (for stress) Dynamic factor

rating

of

Factor of safety from tooth breakage

Size factor

Working angle of the involute Efficiency

S

m2

Enveloping surface

T

Nm

Torque

Yβ

–

Yε

–

ratio factor

YFS

–

Tip factor

YR

–

Roughness factor

YX

–

Size factor

σHlim N/mm2

Zβ

–

Helix angle factor

σHP

N/mm2

Allowable Hertzian pressure

σF

N/mm2

Effective tooth root stress

σFlim

N/mm2

Bending stress number

σFP

N/mm2

Allowable tooth root stress

υ40

mm2/s

Zε ZH

– –

Helix angle factor

ratio factor Zone factor

ZL

–

Lubricant factor

Zv

–

Speed factor

Effective Hertzian pressure Allowable stress number for stress

Lubricating oil viscosity at 40 °C

Note: The unit rad ( = radian ) may be replaced by 1. ·

87

10

Cylindrical Gear Units General Introduction Geometry of Involute Gears 1. Cylindrical gear units

10

1.1 Introduction In the industry, mainly gear units with case hardened and fine-machined gears are used for torque and speed adaptation of prime movers and driven machines. After carburising and hardening, the tooth flanks are fine-machined by grinding (or removing material by means of shaping or generating tools coated with mechanically resistant material). In comparison with other gear units, which, for example, have quenched and tempered or nitrided gears, gear units with case hardened gears have higher power capacities, i.e. they require less space for the same speeds and torques. Further, gear units have the best efficiencies. Motion is transmitted without slip at constant speed. As a rule, an infinitely variable change-speed gear unit with primary or secondary gear stages presents the most economical solution even in case of variable speed control. In industrial gear units mainly involute gears are used. Compared with other tooth profiles, the technical and economical advantages are basically: H Simple manufacture with straight-sided flanked tools; H The same tool for all numbers of teeth; H Generating different tooth profiles and centre distances with the same number of teeth by means of the same tool by addendum modification; H Uniform transmission of motion even in case of centre distance errors from the nominal value; H The direction of the normal force of teeth remains constant during meshing; H Advanced stage of development; H Good availability on the market. When load sharing gear units are used, output torques can be doubled or tripled in comparison

Tip line

with gear units without load sharing. Load sharing gear units mostly have one input and one output shaft. Inside the gear unit the load is distributed and then brought together again on the output shaft gear. The uniform sharing of the load between the individual branches is achieved by special design measures. 1.2 Geometry of involute gears The most important concepts and parameters associated with cylindrical gears and cylindrical gear pairs with involute teeth in accordance with DIN 3960 are represented in sections 1.2.1 to 1.2.4. /1/ 1.2.1 Concepts and parameters associated with involute teeth 1.2.1.1 Standard basic rack tooth profile The standard basic rack tooth profile is the normal section through the teeth of the basic rack which is produced from an external gear tooth system with an infinitely large diameter and an infinitely large number of teeth. From figure 1 follows: – The flanks of the standard basic rack tooth profile are straight lines and are located symmetrically below the pressure angle at a point αP to the tooth centre line; – Between module m and pitch p the relation is p = πm; – The nominal dimensions of tooth thickness and spacewidth on the datum line are equal, i.e. sP = eP = p/2; – The bottom clearance between basic rack tooth profile and counter profile is 0.1 m up to 0.4 m; – The addendum is fixed by haP = m, the dedendum by hfP = m + and thus, the tooth depth by hP = 2 m + ; – The working depth of basic rack tooth profile and counter profile is hwP = 2 m.

Counter profile

Datum line Standard basic rack tooth profile Root line Fillet Tooth root surface Tooth centre line Figure 1 Basic rack tooth profiles for involute teeth of cylindrical gears (acc. to DIN 867)

88

Siees MD · 2009

Cylindrical Gear Units Geometry of Involute Gears

1.2.1.2 Module The module m of the standard basic rack tooth profile is the module in the normal section mn of the gear teeth. For a helical gear with helix angle β on the reference circle, the transverse

module in a transverse section is mt = mn / cosβ. For a spur gear β = 0 and the module is m = mn = mt. In order to limit the number of the required gear cutting tools, module m has been standardized in preferred series 1 and 2, see table 1.

Table 1 Selection of some modules m in mm (acc. to DIN 780) Series 1

1

1.25 1.5

Series 2

2

2.5

1.75

3

4 3.5

1.2.1.3 Tool reference profile The tool reference profile according to figure 2a is the counter profile of the standard basic rack tooth profile according to figure 1. For industrial gear units, the pressure angle at a point of the tool reference profile αP0 = αP is 20°, as a rule. The tooth thickness sP0 of the tool on the tool datum line depends on the stage of machining. The pre-machining tool leaves on both flanks of the teeth a machining allowance q for finishmachining. Therefore, the tooth thickness for pre-machining tools is sP0 < p / 2, and for finishmachining tools sP0 = p / 2. The pre-machining tool generates the root diameter and the fillet on a cylindrical gear. The finish-machining tool removes the machining allowance on the flanks, however, normally it does not touch the root circle – like on the tooth profile in figure 3a. Between pre- and finish-machining, cylindrical gears are subjected to a heat treatment which, as a rule, leads to warping of the teeth and growing of the root and tip circles.

5 4.5

6

8 7

10 9

12 14

16

20 18

25 22

32 28

Especially for cylindrical gears with a relatively large number of teeth or a small module there is a risk of generating a notch in the root on finishmachining. To avoid this, pre-machining tools are provided with protuberance flanks as shown in figure 2b. They generate a root undercut on the gear, see figure 3b. On the tool, protuberance value prP0, protuberance pressure angle at a point αprP0, as well as the tip radius of curvature aP0 must be so dimensioned that the active tooth profile on the gear will not be reduced and the tooth root will not be weakened too much. On cylindrical gears with small modules one often accepts on purpose a notch in the root if its distance to the root circle is large enough and thus the tooth root load carrying capacity is not impaired by a notch effect, figure 3c. In order to prevent the tip circle of the mating gear from touching the fillet it is necessary that a check for meshing interferences is carried out on the gear pair. /1/

10

a) Tool datum line

b) Protuberance flank

Figure 2 Reference profiles of gear cutting tools for involute teeth of cylindrical gears a) For pre-machining and finish-machining b) For pre-machining with root undercut (protuberance) Siees MD · 2009

89


Pre-machining

Finish-machining

Machining allowance q

Root undercut a)

Notch b)

c)

Figure 3 Tooth profiles of cylindrical gears during pre- and finish-machining a) Pre- and finish-machining down to the root circle b) Pre-machining with root undercut (protuberance) c) Finish-machining with notch 1.2.1.4 Generating tooth flanks With the development of the envelope, an envelope line of the base cylinder with the base diameter db generates the involute surface of a spur gear. A straight line inclined by a base helix angle βb to the envelope line in the developed envelope is the generator of an involute surface (involute helicoid) of a helical gear, figure 4. The involute which is always lying in a transverse section, figure 5, is described by the transverse

(1)

r = rb / cosα

(2)

rb = db / 2 is the base radius. The angle invα is termed involute function, and the angle

Base cylinder

Involute helicoid

10

invα = tanα −

ζ = + invα = tanα tanα is termed working angle.

Base cylinder envelope line Involute of base cylinder

pressure angle at a point α and radius r in the equations

Involute

Developed envelope line

Generator Developed base cylinder envelope

Involute of base cylinder

Figure 4 Base cylinder with involute helicoid and generator

90

Figure 5 Involute in a transverse section Siees MD · 2009


1.2.2 Concepts and parameters associated with cylindrical gears 1.2.2.1 Geometric definitions In figure 6 the most important geometric quantities of a cylindrical gear are shown. The reference circle is the intersection of the reference cylinder with a plane of transverse section. When generating tooth flanks, the straight pitch line of the tool rolls off at the reference circle. Therefore, the reference circle periphery corresponds to the product of pitch p and number of teeth z, i.e. π · d = p · z. Since mt = p / π, the equation for the reference diameter thus is d = mt · z. Many geometric quantities of the cylindrical gear are referred to the reference circle. For a helical gear, at the point of intersection of the involute with the reference circle, the trans-

verse pressure angle at a point α in the transverse section is termed transverse pressure angle αt, see figures 5 and 7. If a tangent line is put against the involute surface in the normal section at the point of intersection with the reference circle, the corresponding angle is termed normal pressure angle αn; this is equal to the pressure angle αP0 of the tool. The interrelationship with the helix angle β at the reference circle is tanαn = cosβ · tanαt. On a spur gear αn = αt. Between the base helix angle βb and the helix angle β on the reference circle the relationship is sinβb = cosαn · sinβ. The base diameter db is given by the reference diameter d, by db = d · cosαt. In the case of internal gears, the number of teeth z and thus also the diameters d, db, da, df are negative values. Right flank Tooth trace

Left flank

Reference cylinder Reference circle d da df b h ha hf s Figure 6 Definitions on the cylindrical gear 1.2.2.2 Pitches The pitch pt of a helical gear (p in the case of a spur gear) lying in a transverse section is the length of the reference circle arc between two successive right or left flanks, see figures 6 and 7. With the number of teeth z results pt = π · d / z = π · mt. The normal transverse pitch pet of a helical gear is equal to the pitch on the basic circle pbt, thus pet = pbt = π · db / z. Hence, in the normal section the normal base pitch at a point pen = pet / cosβb is resulting from it, and in the axial section the axial pitch pex = pet / tanβb, see figure 13.

Siees MD · 2009

e p

Reference diameter Tip diameter Root diameter Facewidth Tooth depth Addendum Dedendum Tooth thickness on the reference circle Spacewidth on the reference circle Pitch on the reference circle

10

Figure 7 Pitches in the transverse section of a helical gear

91


10

1.2.2.3 Addendum modification When generating tooth flanks on a cylindrical gear by means of a tooth-rack-like tool (e.g. a hob), a straight pitch line parallel to the datum line of tool rolls off on the reference circle. The distance (x · mn) between the straight pitch line and the datum line of tool is the addendum modification, and x is the addendum modification coefficient, see figure 8. An addendum modification is positive, if the datum line of tool is displaced from the reference circle towards the tip, and it is negative if the datum line is displaced towards the root of the gear. This is true for both external and internal gears. In the case of internal gears the tip points to the inside. An addendum modification for external gears should be carried through approximately within the limits as shown in figure 9. The addendum modification limits xmin and xmax are represented dependent on the virtual number of teeth zn = z / (cosβ · cos2βb). The upper limit xmax takes into the intersection circle of the teeth and applies to a normal crest width in the normal section of san = 0.25 mn. When falling below the lower limit xmin this results in an undercut which shortens the usable involute and weakens the tooth root. A positive addendum modification results in a greater tooth root width and thus in an increase in the tooth root carrying capacity. In the case of small numbers of teeth this has a considerably stronger effect than in the case of larger ones. One mostly strives for a greater addendum modification on pinions than on gears in order to achieve equal tooth root carrying capacities for both gears, see figure 19. Further criteria for the determination of addendum modification are contained in /2/, /3/, and /4/. The addendum modification coefficient x refers to gear teeth free of backlash and deviations. In order to take into tooth thickness deviation As (for backlash and manufacturing tolerances) and machining allowances q (for premachining), one has to give the following generating addendum modification coefficient for the manufacture of a cylindrical gear:

XE = x +

As 2mn · tan αn

+

q

Datum line of tool = straight pitch line

a)

Straight pitch line

b)

Datum line of tool

Straight pitch line

c)

Figure 8 Different positions of the datum line of tool in relation to the straight pitch line through pitch point C. a) Zero addendum modification; x = 0 b) Negative addendum modification; x < 0 c) Positive addendum modification; x > 0

(3)

mn · sin αn

Figure 9 Addendum modification limit xmax (intersection circle) and xmin (undercut limit) for external gears dependent on the virtual number of teeth zn (for internal gears, see /1/ and /3/).

92

Siees MD · 2009


1.2.3 Concepts and parameters associated with a cylindrical gear pair 1.2.3.1 The mating of two external cylindrical gears (external gears) gives an external gear pair. In the case of a helical external gear pair one gear has left-handed and the other one right-handed flank direction. The mating of an external cylindrical gear with an internal cylindrical gear (internal gear) gives an internal gear pair. In the case of a helical internal gear pair, both gears have the same flank direction, that is either right-handed or left-handed. The subscript 1 is used for the size of the smaller gear (pinion), and the subscript 2 for the larger gear (wheel or internal gear). In the case of an unmodified gear pair (a zero gear pair), both gears have as addendum modification coefficient x1 = x2 = 0 (zero gears). In the case of a gear pair at reference centre distance, both gears have addendum modifications (modified gears), that is with x1 + x2 = 0, i.e. x 1 = - x2 . For a modified gear pair, the sum is not equal to zero, i.e. x1 + x2 ≠ 0. One of the cylindrical gears in this case may, however, have an addendum modification x = 0. 1.2.3.2 Mating quantities The gear ratio of a gear pair is the ratio of the number of teeth of the gear z2 to the number of teeth of the pinion z1, thus u = z2 / z1. Working pitch circles with diameter dw = 2 · rw are those transverse intersection circles of a cylindrical gear pair, which have the same circumferential speed at their mutual point (pitch point C), figure 10. The working pitch circles divide the centre distance a = rw1 + rw2 in the ratio of the tooth numbers, thus dw1 = 2 · a / (u + 1) and dw2 = 2 · a · u / (u +1). In the case of both an unmodified gear pair and a gear pair at reference centre distance, the centre distance is equal to the zero centre distance ad = (d1 + d2) / 2, and the pitch circles are simultaneously the reference circles, i.e. dw = d. However, in the case of a modified gear pair, the centre distance is not equal to the zero centre distance, and the pitch circles are not simultaneously the reference circles. If in the case of modified gear pairs the bottom clearance corresponding to the standard basic rack tooth profile is to be retained (which is not absolutely necessary), then an addendum modification is to be carried out. The addendum modification factor is k = (a - ad) / mn - (x1 + x2). For unmodified gear pairs and gear pairs at reference centre distance, k = 0. In the case of external gear pairs k < 0, i.e. the tip diameters of both gears become smaller. In the case of Siees MD · 2009

internal gear pairs k > 0, i.e. the tip diameters of both gears become larger (on an internal gear with negative tip diameter the absolute value becomes smaller).

Figure 10 Transverse section of an external gear pair with ing left-handed flanks In a cylindrical gear pair either the left or the right flanks of the teeth each other on the line of action. Changing the flanks results in a line of action each lying symmetrical in relation to the centre line through O1 O2. The line of action with ing left flanks in figure 10 is the tangent to the two base circles at points T1 and T2. With the common tangent on the pitch circles it includes the working pressure angle αwt. The working pressure angle αwt is the transverse pressure angle at a point belonging to the working pitch circle. According to figure 10 it is determined by cos αwt = db1 / dw1 = db2 / dw2. In the case of unmodified gear pairs and gear pairs at reference centre distance, the working pressure angle is equal to the transverse pressure angle on the reference circle, i.e. αwt = αt. The length of path of gα is that part of the line of action which is limited by the two tip circles of the cylindrical gears, figure 11. The starting point A of the length of path of is the point at which the line of action intersects the tip circle of the driven gear, and the finishing point E is the point at which the line of action intersects the tip circle of the driving gear.

93

10


Driven

Driven

Line of action

Driving Figure 11 Length of path of AE in the transverse section of an external gear pair A Starting point of engagement E Finishing point of engagement C Pitch point

10

1.2.3.3 ratios The transverse ratio εα in the transverse section is the ratio of the length of path of gα to the normal transverse pitch pet, i.e. εα = gα / pet, see figure 12. In the case of spur gear pairs, the transverse ratio gives the average number of pairs of teeth meshing during the time of of a tooth pair. According to figure 12, the left-hand tooth pair is in the individual point of D while the right-hand tooth pair gets into mesh at the starting point of engagement A. The righthand tooth pair is in the individual point of B when the left-hand tooth pair leaves the mesh at the finishing point of engagement E. Along the individual length of path of BD one tooth pair is in mesh, and along the double lengths of paths of AB and DE two pairs of teeth are simultaneously in mesh. In the case of helical gear pairs it is possible to achieve that always two or more pairs of teeth are in mesh simultaneously. The overlap ratio εβ gives the ratio, owing to the helix of the teeth, as the ratio of the facewidth b to the axial pitch pex, i.e. εβ = b / pex, see figure 13. The total ratio εγ is the sum of transverse ratio and overlap ratio, i.e. εγ = εα + εβ. With an increasing total ratio, the load carrying capacity increases, as a rule, while the generation of noise is reduced.

94

Line of action

Driving Figure 12 Single and double region in the transverse section of an external gear pair B, D Individual points of A, E Starting and finishing point of engagement, respectively C Pitch point

Length of path of Figure 13 Pitches in the plane of action A Starting point of engagement E Finishing point of engagement Siees MD · 2009


1.2.4 Summary of the most important formulae Tables 2 and 3 contain the most important formulae for the determination of sizes of a cylindrical gear and a cylindrical gear pair, and this for both external and internal gear pairs. The following rules for signs are to be observed: In the case of internal gear pairs the number of teeth z2 of the internal gear is a negative quantity. Thus, also the centre distance a or ad and the gear ratio u as well as the diameters d2, da2, db2, df2, dw2 and the virtual number of teeth zn2 are negative. When deg a cylindrical gear pair for a gear stage, from the output quantities of tables 2 and 3 only the normal pressure angle αn and the gear ratio u are given, as a rule. The number of teeth of

the pinion is determined with regard to silent running and a balanced foot and flank load carrying capacity, at approx. z1 = 18 ... 23. If a high foot load carrying capacity is required, the number may be reduced to z1 = 10. For the helix angle, β = 10 up to 15 degree is given, in exceptional cases also up to 30 degree. The addendum modification limits as shown in figure 9 are to be observed. On the pinion, the addendum modification coefficient should be within the range of x1 = 0.2 up to 0.6 and from

u > 2 the width within the range b1 = (0.35 to 0.45) a. Centre distance a is determined either by the required power to be transmitted or by the constructional conditions.

10

Siees MD · 2009

95


Table 2 Parameters for a cylindrical gear *) Output quantities: mm mn degree αn β degree z – x – – xE mm haP0

normal module normal pressure angle reference helix angle number of teeth *) addendum modification coefficient generating addendum modification coefficient, see equation (3) addendum of the tool

Item

10

Formula mn cosβ

Transverse module

mt =

Transverse pressure angle

tanαt =

Base helix angle

sinβb = sinβ · cosαn

Reference diameter

d = mt · z

Tip diameter (for k, see table 3)

da = d + 2 mn (1 + x + k)

Root diameter

df = d – 2 (haP0 – mn · xE)

Base diameter

db = d · cosαt

Transverse pitch

pt =

Transverse pitch on path of ; Transverse base pitch

pet = pbt =

Transverse pressure angle at tip circle

cos αat =

Transverse tooth thickness on the pitch circle

st = mt

Normal tooth thickness on the pitch circle

sn = st · cosβ

Transverse tooth thickness on the addendum circle

sat = da

Virtual number of teeth

zn =

tanαn cosβ

π·d = π · mt z

(

π · db = pt · cosαt z

db da π + 2 · x · tanαn) 2

s

( dt

+ invαt – invαat)

z cosβ · cos2βb

*) For an internal gear, z is to be used as a negative quantity.

96

**)

**) For invα, see equation (1). Siees MD · 2009


Table 3 Parameters for a cylindrical gear pair *)

Output quantities: The parameters for pinion and wheel according to table 2 must be given, further the facewidths b1 and b2, as well as either the centre distance a or the sum of the addendum modification coefficients x1 + x2. Item

Formula z2 z1

Gear ratio

u =

Working transverse pressure angle (“a” given)

cosαwt =

Sum of the addendum modification coefficients

x1 + x2 =

Working transverse pressure angle (x1 + x2 given)

invαwt = 2

Centre distance

a =

Reference centre distance

ad =

Addendum modification factor **)

k =

Working pitch circle diameter of the pinion

dw1 =

2 · a = d cosαt 1 cosαwt u+1

Working pitch circle diameter of the gear

dw2 =

cosαt 2·a·u = d2 u+1 cosαwt

Length of path of

gα =

1 2

Transverse ratio

εα =

gα pet

Overlap ratio

εβ =

b · tanβb pet

Total ratio

εγ = εα + εβ

mt

mt 2·a

(z1 + z2) cosαt

z1 + z 2 2 · tanαn

x1 + x2 tanαn + invαt z1 + z2

(z1 + z2)

2 mt 2

(invαwt – invαt)

cosαt cosαwt =

(z1 + z2)

d1 + d2 2

a – ad – (x1 + x2) mn

(

da12 – db12 +

u

u

10 da22 – db22

) – a · sinαwt

b = min (b1, b2)

*) For internal gear pairs, z2 and a are to be used as negative quantities. **) See subsection 1.2.3.2. Siees MD · 2009

97


1.2.5 Gear teeth modifications The parameters given in the above subsections 1.2.1 to 1.2.4 refer to non-deviating cylindrical gears. Because of the high-tensile gear materials, however, a high load utilization of the gear units is possible. Noticeable deformations of the elastic gear unit components result from it. The deflection at the tooth tips is, as a rule, a multiple of the manufacturing form errors. This leads to meshing interferences at the entering and leaving sides, see figure 14. There is a negative effect on the load carrying capacity and generation of noise.

facewidth are achieved. This has to be taken into consideration especially in the case of checks of patterns carried out under low loads. Under partial load, however, the local maximum load rise is always lower than the theoretical uniform load distribution under full load. In the case of modified gear teeth, the ratio is reduced under partial load because of incomplete carrying portions, making the noise generating levels increase in the lower part load range. With increasing load, the carrying portions and thus the ratio increase so that the generating levels drop. Gear pairs which are only slightly loaded do not require any modification. Wheel

Line of action Pinion

Figure 14 Cylindrical gear pair under load 1 Driving gear 2 Driven gear a, b Tooth pair being in engagement c, d Tooth pair getting into engagement

10

Further, the load causes bending and twisting of pinion and wheel shaft, pinion and wheel body, as well as settling of bearings, and housing deformations. This results in skewing of the tooth flanks which often amounts considerably higher than the tooth trace deviations caused by manufacture, see figure 15. Non-uniform load carrying occurs along the face width which also has a negative effect on the load carrying capacity and generation of noise. The running-in wear of case hardened gears amounts to a few micrometers only and cannot compensate the mentioned deviations. In order to restore the high load carrying capacity of case hardened gears and reduce the generation of noise, intentional deviations from the involute (profile correction) and from the theoretical tooth trace (longitudinal correction) are produced in order to attain nearly ideal geometries with uniform load distribution under load again. The load-related form corrections are calculated and made for one load only – as a rule for 70 ... 100% of the permanently acting nominal load – /5, 6, 7/. At low partial load, patterns which do not cover the entire tooth depth and

98

Bending Torsion Manufacturing deviation Bearing deformation Housing deformation Running-in wear Effective tooth trace deviation Fβ = Σf-yβ Load distribution across the facewidth w Figure 15 Deformations and manufacturing deviations on a gear unit shaft In figure 16, usual profile and longitudinal corrections are illustrated. In the case of profile correction, the flanks on pinion and wheel are relieved at the tips by an amount equal to the length they are protruding at the entering and leaving sides due to the bending deflection of the teeth. Root relief may be applied instead of tip relief which, however, is much more expensive. Thus, a gradual load increase is achieved on the tooth Siees MD · 2009

Cylindrical Gear Units Geometry of Involute Gears Load Carrying Capacity of Involute Gears getting into engagement, and a load reduction on the tooth leaving the engagement. In the case of longitudinal correction, the tooth trace relief often is superposed by a symmetric lon-

Profile correction

gitudinal crowning. With it, uniform load carrying along the facewidth and a reduction in load concentration at the tooth ends during axial displacements is attained.

Longitudinal correction

Figure 16 Gear teeth modifications designed for removing local load increases due to deformations under nominal load 1.3 Load carrying capacity of involute gears 1.3.1 Scope of application and purpose The calculation of the load carrying capacity of cylindrical gears is generally carried out in accordance with the calculation method according to DIN 3990 /8/ (identical with ISO 6336) which takes into pitting, tooth root bending stress and scoring as load carrying limits. Because of the relatively large scope of standards, the calculation in accordance with this method may be carried out only by using EDP programs. As a rule, gear unit manufacturers have such a tool at hand. The standard work is the FVA-Stirnradprogramm /9/ which includes further calculation methods, for instance, according to Niemann, AGMA, DNV, LRS, and others. In DIN 3990, different methods A, B, C ... are suggested for the determination of individual factors, where method A is more exact than method B, etc. The application standard /10/ according to DIN 3990 is based on simplified methods. Because of its – even though limited – universal validity it still is relatively time-consuming. The following calculation method for pitting resistance and tooth strength of case-hardened cylindrical gears is a further simplification if compared with the application standard, however, without losing some of its meaning. Certain conditions must be adhered to in order to attain high load carrying capacities which also results in preventing scuffing. Therefore, a calculation of load carrying capacity for scuffing will not be considered in the following. Siees MD · 2009

It has to be expressly emphasized that for the load carrying capacity of gear units the exact calculation method – compared with the simplified one – is always more meaningful and therefore is exclusively decisive in borderline cases. Design, selection of material, manufacture, heat treatment, and operation of industrial gear units are subject to certain rules which lead to a long service life of the cylindrical gears. Those rules are: – Gear teeth geometry acc. to DIN 3960; – Cylindrical gears out of case-hardened steel; Tooth flanks in DIN quality 6 or better, fine machined; – Quality of material and heat treatment proved by quality inspections acc. to DIN 3990 /11/; – Effective case depth after carburizing according to instructions /12/ with surface hardnesses of 58 ... 62 HRC; – Gears with required tooth corrections and without harmful notches in the tooth root; – Gear unit designed for fatigue strength, i.e. life factors ZNT = YNT = 1.0; – Flank fatigue strength σHlim y 1200 N/mm2; – Subcritical operating range, i.e. pitch circle velocity lower than approx. 35 m/s; – Sufficient supply of lubricating oil; – Use of prescribed gear oils of criteria stage 12 acc. to the gear rig test by the FZG-method and sufficient grey staining load capacity; – Maximum operating temperature 95 °C.

99

10

Cylindrical Gear Units Load Carrying Capacity of Involute Gears

If these requirements are met, a number of factors can be definitely given for the calculation of the load carrying capacity according to DIN 3990, so that the calculation procedure is partly considerably simplified. Non-observance of the above requirements, however, does not necessarily mean that the load carrying capacity is reduced. In case of doubt one should, however, carry out the calculation in accordance with the more exact method. 1.3.2 Basic details The calculation of the load carrying capacity is

based on the nominal torque of the driven machine. Alternatively, one can also start from the nominal torque of the prime mover if this corresponds with the torque requirement of the driven machine. In order to be able to carry out the calculation for a cylindrical gear stage, the details listed in table 4 must be given in the units mentioned in the table. The geometric quantities are calculated according to tables 2 and 3. Usually, they are contained in the workshop drawings for cylindrical gears.

Table 4 Basic details Abbreviation

10

100

Meaning

Unit

P

Power rating

kW

n1

Pinion speed

min-1

a

Centre distance

mm

mn

Normal module

mm

da1

Tip diameter of the pinion

mm

da2

Tip diameter of the wheel

mm

b1

Facewidth of the pinion

mm

b2

Facewidth of the wheel

mm

z1

Number of teeth of the pinion

–

z2

Number of teeth of the wheel

–

x1

Addendum modification coefficient of the pinion

–

x2

Addendum modification coefficient of the wheel

–

αn

Normal pressure angle

Degree

β

Reference helix angle

Degree

υ40

Kinematic viscosity of lubricating oil at 40 °C

mm2 / s

Rz1

Peak-to-valley height on pinion flank

μm

Rz2

Peak-to-valley height on wheel flank

μm Siees MD · 2009


In the further course of the calculation, the quantities listed in table 5 are required. They are derived from the basic details according to table 4.

Table 5 Derived quantities Designation

Relation

Unit

Gear ratio

u = z2 / z1

Reference diameter of the pinion

d1 = z1 · mn / cosβ

Transverse tangential force at pinion reference circle

Ft =

Transverse tangential force at pitch circle

F u = Ft ·

Circumferential speed at reference circle

v = π · d1 · n1 / 60000

Base helix angle

βb = arc sin (cosαn · sinβ)

Degree

Virtual number of teeth of the pinion

zn1 = z1 / (cosβ · cos2βb)

–

Virtual number of teeth of the wheel

zn2 = z2 / (cosβ · cos2βb)

–

Transverse module

mt = mn / cosβ

Transverse pressure angle

αt = arc tan (tanαn / cosβ)

Degree

Working transverse pressure angle

αwt = arc cos [(z1 + z2) mt · cosαt / (2 · a)]

Degree

Transverse pitch

pet = π · mt · cosαt

mm

Base diameter of the pinion

db1 = z1 · mt · cosαt

mm

Base diameter of the wheel

db2 = z2 · mt · cosαt

mm

Length of path of

gα =

Transverse ratio

εα = gα / pet

Overlap ratio

εβ = b · tanβb / pet

Siees MD · 2009

– mm

6 · 107 P · π d1 · n1

N

d1 (u + 1) 2·a

N

m/s

mm

10

1 2

(

da12 – db12 +

u

u

da22 – db22

) – a · sinαwt

mm –

b = min (b1, b2)

–

101


1.3.3 General factors 1.3.3.1 Application factor With the application factor KA, all additional forces acting on the gears from external sources are taken into consideration. It is dependent on the characteristics of the driving and driven machines, as well as the couplings, the masses and stiffness of the system, and the operating conditions.

The application factor is determined by the service classification of the individual gear. If possible, the factor KA should be determined by means of a careful measurement or a comprehensive analysis of the system. Since very often it is not possible to carry out the one or other method without great expenditure, reference values are given in table 6 which equally apply to all gears in a gear unit.

Table 6 Application factor KA Working mode of the driven machine

Working mode of prime mover

Uniform

Moderate shock loads

Average shock loads

Heavy shock loads

Uniform

1.00

1.25

1.50

1.75

Moderate shock loads

1.10

1.35

1.60

1.85

Average shock loads

1.25

1.50

1.75

2.00 or higher

Heavy shock loads

1.50

1.75

2.00

2.25 or higher

1.3.3.2 Dynamic factor With the dynamic factor Kv, additional internal dynamic forces caused in the meshing are taken into consideration. Taking z1, v and u from tables 4 and 5, it is calculated from Kv = 1 + 0.0003 · z1 · v

10

u2 1 + u2

(4)

1.3.3.3 Face load factor The face load factor KHβ takes into the increase in the load on the tooth flanks caused by non-uniform load distribution over the facewidth. According to /8/, it can be determined by means of different methods. Exact methods based on comprehensive measurements or calculations or on a combination of both are very expensive. Simple methods, however, are not exact, as a consequence of which estimations made to be on the safe side mostly result in higher factors. For normal cylindrical gear teeth without longitudinal correction, the face load factor can be calculated according to method D in accordance with /8/ dependent on facewidth b and reference diameter d1 of the pinion, as follows: KHβ = 1.15 + 0.18 (b / d1)2 + 0.0003 · b

(5)

with b = min (b1, b2). As a rule, the gear unit manufacturer carries out an analysis of the load distribution over the facewidth in accordance with an exact calculation method /13/. If required, he makes longitudinal corrections in order to

102

attain uniform load carrying over the facewidth, see subsection 1.2.5. Under such conditions, the face load factor lies within the range of KHβ = 1.1 ... 1.25. As a rough rule applies: A sensibly selected crowning symmetrical in length reduces the amount of KHβ lying above 1.0 by approx. 40 to 50%, and a directly made longitudinal correction by approx. 60 to 70%. In the case of slim shafts with gears arranged on one side, or in the case of lateral forces or moments acting on the shafts from external sources, for the face load factors for gears without longitudinal correction the values may lie between 1.5 and 2.0 and in extreme cases even at 2.5. Face load factor KFβ for the determination of increased tooth root stress can approximately be deduced from face load factor KHβ according to the relation KFβ = ( KHβ) 0.9

(6)

1.3.3.4 Transverse load factors The transverse load factors KHα and KFα take into the effect of the non-uniform distribution of load between several pairs of simultaneously ing gear teeth. Under the conditions as laid down in subsection 1.3.1, the result for surface stress and for tooth root stress according to method B in accordance with /8/ is KHα = KFα = 1.0

(7)

Siees MD · 2009


1.3.4 Tooth flank load carrying capacity The calculation of surface durability against pitting is based on the Hertzian pressure at the pitch circle. For pinion and wheel the same effective Hertzian pressure σH is assumed. It must not exceed the permissible Hertzian pressure σHP , i.e. σH x σHP . σH = ZE ZH Zβ Zε

σH

1.3.4.1 Effective Hertzian pressure The effective Hertzian pressure is dependent on the load, and for pinion and wheel is equally derived from the equation

KA Kv KHα KHβ

u+1 Ft u d1 · b

(8)

Effective Hertzian pressure in N/mm2

Further: b Common facewidth of pinion and wheel Ft, u, d1 acc. to table 5 KA Application factor acc. to table 6 Kv Dynamic factor acc. to equation (4) KHβ Face load factor acc. to eq (5) KHα Transverse load factor acc. to eq (7) ZE Elasticity factor; ZE = 190 N/mm2 for gears out of steel ZH Zone factor acc. to figure 17 Zβ Helix angle factor acc. to eq (9) Zε ratio factor acc. to eq (10) or (11) With ß according to table 4 applies: Zβ =

cosβ

(9)

With εα and εβ according to table 5 applies: Zε =

ε 4 – εα (1 – εβ) + β for εβ < 1 εα 3

Zε =

1 εα

for εβ y 1

(10)

(11)

1.3.4.2 Permissible Hertzian pressure The permissible Hertzian pressure is determined by σHP = ZL Zv ZX ZR ZW

σHlim SH

(12)

σHP permissible Hertzian pressure in N/mm2. It is of different size for pinion and wheel if the strengths of materials σHlim are different. Factors ZL, Zv, ZR, ZW and ZX are the same for Siees MD · 2009

Figure 17 Zone factor ZH depending on helix angle β as well as on the numbers of teeth z1, z2, and addendum modification coefficients x1, x2; see table 4. pinion and wheel and are determined in the following. The lubricant factor is computed from the lubricating oil viscosity υ40 according to table 4 using the following formula: ZL = 0.91 +

0.25

(1 + 112 ) 2

(13)

υ40

103

10


For the speed factor, the following applies using the circumferential speed v according to table 5: F = KA Kv KFα KFβ · Zv = 0.93 +

0.157

(14)

40 1+ v

σF

The roughness factor can be determined as a function of the mean peak-to-valley height RZ = (RZ1 + RZ2) / 2 of the gear pair as well as the gear ratio u and the reference diameter d1 of the pinion, see tables 4 and 5, from

ZR =

0.513 R z

3

(1 + u ) d1

0.08

(15)

For a gear pair with the same tooth flank hardness on pinion and wheel, the work hardening factor is

ZW = 1.0

(16)

The size factor is computed from module mn according to table 4 using the following formula:

ZX = 1.05 – 0.005 mn

(17)

with the restriction 0.9 x ZX x 1.

10

Ft · YFS Yβ Yε (18) b · mn

Effective tooth root stress in N/mm2

The following factors are of equal size for pinion and wheel: mn, Ft acc. to tables 4 and 5 KA Application factor acc. to table 6 Kv Dynamic factor acc. to equation (4) KFβ Face load factor acc. to equation (6) KFα Transverse load factor acc. to eq (7) Yε ratio factor acc. to eq (19) Helix angle factor acc. to eq (20) Yβ The following factors are of different size for pinion and wheel: b1, b2 Facewidths of pinion and wheel acc. to table 4. If the facewidths of pinion and wheel are different, it may be assumed that the load bearing width of the wider facewidth is equal to the smaller facewidth plus such extension of the wider that does not exceed one times the module at each end of the teeth. YFS1, Tip factors acc. to figure 19. They YFS2 for the complex stress condition inclusive of the notch effect in the root fillet.

σHlim Endurance strength of the gear material. For gears made out of case hardening steel, case hardened, figure 18 shows a range from 1300 ... 1650 N/mm2 depending on the surface hardness of the tooth flanks and the quality of the material. Under the conditions as described in subsection 1.3.1, material quality MQ may be selected for pinion and wheel, see table on page 77. SH

Required safety factor against pitting, see subsection 1.3.6.

1.3.5 Tooth strength The maximum load in the root fillet at the 30degree tangent is the basis for rating the tooth strength. For pinion and wheel it shall be shown separately that the effective tooth root stress σF does not exceed the permissible tooth root stress σFP , i.e. σF < σFP . 1.3.5.1 Effective tooth root stress As a rule, the load-dependent tooth root stresses for pinion and wheel are different. They are calculated from the following equation:

104

Flank hardness HV1

Figure 18 Allowable stress number for stress σHlim of alloyed case hardening steels, case hardened, depending on the surface hardness HV1 of the tooth flanks and the material quality. ML modest demands on the material quality MQ normal demands on the material quality ME high demands on the material quality, see /11/ Siees MD · 2009


αn

= 20 degree

ha0 = 1.35 · mn a0 = 0.2 · mn

αn

YFS

= 20 degree

ha0 = 1.4 · mn a0 = 0.25 · mn αpr0 = 10 degree pr0 = 0.025 · mn

αn

= 20 degree

ha0 = 1.4 · mn

10

a0 = 0.3 · mn αpr0 = 10 degree pr0 = 0.0205 · mn

Figure 19 Tip factor YFS for external gears with standard basic rack tooth profile acc. to DIN 867 depending on the number of teeth z (or zn in case of helical gears) and addendum modification coefficient x, see tables 4 and 5. The following only approximately applies to internal gears: YFS = YFS∞ (≈ value for x = 1.0 and z = 300). Siees MD · 2009

105


With the helix angle β acc. to table 4 and the overlap ratio εβ acc. to table 5 follows: Yε = 0.25 +

0.75 · cos2β εα

β 120 _

Safety factor required against tooth breakage, see subsection 1.3.6.

(19) 18CrNiMo7-6

with the restriction 0.625 x Yε x 1 Yβ = 1 – ε β ·

SF

15CrNi6 16MnCr5

(20)

with the restriction Yβ y max. (1 - 0.25 εβ); (1– β/120). 1.3.5.2 Permissible tooth root stress The permissible tooth root stress for pinion and wheel is determined by σFP = YST YδrelT YRrelT YX

σFlim (SF)

(21)

σFP permissible tooth root stress in N/mm2. It is not equal for pinion and wheel if the material strengths σFlim are not equal. Factors YST, YδrelT, YRrelT and YX may be approximately equal for pinion and wheel. is the stress correction factor of the referYST ence test gears for the determination of the bending stress number σFlim. For standard reference test gears, YST = 2.0 has been fixed in the standard. YδrelT is the relative sensitivity factor (notch sensitivity of the material) referring to the standard reference test gear. By approximation YδrelT = 1.0. For the relative surface factor (surface roughness factor of the tooth root fillet) referring to the standard reference test gear the following applies by approximation, depending on module mn:

10

YRrelT = 1.00 for mn x 8 mm = 0.98 for 8 mm < mn x 16 mm (22) = 0.96 for mn > 16 mm and for the size factor YX = 1.05 – 0.01 mn

(23)

with the restriction 0.8 x YX x 1. σFlim Bending stress number of the gear material. For gears out of case hardening steel, case hardened, a range from 310 ... 520 N/mm2 is shown in figure 20 depending on the surface hardness of the tooth flanks and the material quality. Under the conditions according to subsection 1.3.1, a strength pertaining to quality MQ may be used as a basis for pinion and wheel, see table on page 77.

106

Flank hardness HV1

Figure 20 Bending stress number σFlim of alloyed case hardening steel, case hardened, depending on the surface hardness HV1 of the tooth flanks and the material quality. ML modest demands on the material quality MQ normal demands on the material quality ME high demands on the material quality, see /11/ 1.3.6 Safety factors The minimum required safety factors according to DIN are: against pitting SH = 1.0 against tooth breakage SF = 1.3. In practice, higher safety factors are usual. For multistage gear units, the safety factors are determined about 10 to 20% higher for the expensive final stages, and in most cases even higher for the cheaper preliminary stages. Also for risky applications a higher safety factor is given. 1.3.7 Calculation example An electric motor drives a coal mill via a multistage cylindrical gear unit. The low speed gear stage is to be calculated. Given: Nominal power rating P = 3300 kW; pinion speed n1 = 141 min-1; centre distance a = 815 mm; normal module mn = 22 mm; tip diameter da1 = 615.5 mm and da2 = 1100 mm; pinion and wheel widths b1 = 360 mm and b2 = 350 mm; numbers of teeth z1 = 25 and z2 = 47; addendum modification coefficients x1 = 0.310 and x2 = 0.203; normal pressure angle αn = 20 degree; helix angle β = 10 degree; kinematic viscosity of the lubricating oil υ40 = 320 cSt; mean peak-tovalley roughness Rz1 = Rz2 = 4.8 μm. The cylindrical gears are made out of the material 18CrNiMo7-6. They are case hardened and ground with profile corrections and width-symmetrical crowning. Siees MD · 2009

Cylindrical Gear Units Load Carrying Capacity of Involute Gears Gear Unit Types Calculation (values partly rounded): Gear ratio u = 1.88; reference diameter of the pinion d1 = 558.485 mm; nominal circumferential force on the reference circle Ft = 800425 N; circumferential speed on the reference circle v = 4.123 m/s; base helix angle βb = 9.391 degree; virtual numbers of teeth zn1 = 26.08 and zn2 = 49.03; transverse module mt = 22.339 mm; transverse pressure angle αt = 20.284 degree; working transverse pressure angle αwt = 22.244 degree; normal transverse pitch pet = 65.829 mm; base diameters db1 = 523.852 mm and db2 = 984.842 mm; length of path of gα = 98.041 mm; transverse ratio εα = 1.489; overlap ratio εβ = 0.879. Application factor KA = 1.50 (electric motor with uniform mode of operation, coal mill with medium shock load); dynamic factor Kv = 1.027; face load factor KHβ = 1.20 [acc. to equation (5) follows KHβ = 1.326, however, because of symmetrical crowning the calculation may be made with a smaller value]; KFβ = 1.178; KHα = KFα = 1.0.

The safety factors against tooth breakage referring to the torque are SF = σFP/σF: for the pinion SF1 = 797/537 = 1.48 and for the wheel SF2 = 797/540 = 1.48.

Load carrying capacity of the tooth flanks: Elasticity factor ZE = 190 ǸNńmm 2; zone factor ZH = 2.342; helix angle factor Zβ = 0.992; ratio factor Zε = 0.832. According to equation (8), the Hertzian pressure for pinion and wheel is σH = 1251 N/mm2. Lubricant factor ZL = 1.047; speed factor ZV = 0.978; roughness factor ZR = 1.018; work hardening factor ZW = 1.0; size factor ZX = 0.94. With the allowable stress number for stress (pitting) σHlim = 1500 N/mm2, first the permissible Hertzian pressure σHP = 1470 N/mm2 is determined from equation (12) without taking into the safety factor. The safety factor against pitting is found by SH = σHP/σH = 1470/1251 = 1.18. The safety factor referring to the torque is SH2 = 1.38.

1.4.2 Load sharing gear units In principle, the highest output torques of gear units are limited by the manufacturing facilities, since gear cutting machines can make gears up to a maximum diameter only. Then, the output torque can be increased further only by means of load sharing in the gear unit. Load sharing gear units are, however, also widely used for lower torques as they provide certain advantages in spite of the larger number of internal components, among others they are also used in standard design. Some typical features of the one or other type are described in the following.

Load carrying capacity of the tooth root: ratio factor Yε = 0.738; helix angle factor Yβ = 0.927; tip factors YFS1 = 4.28 and YFS2 = 4.18 (for ha0 = 1.4 mn; a0 = 0.3 mn; αpr0 = 10 degree; pr0 = 0.0205 mn). The effective tooth root stresses σF1 = 537 N/mm2 for the pinion and σF2 = 540 N/mm2 for the wheel can be obtained from equation (18). Stress correction factor YST = 2.0; relative sensitivity factor YδrelT = 1.0; relative surface factor YRrelT = 0.96; size factor YX = 0.83. Without taking into consideration the safety factor, the permissible tooth root stresses for pinion and wheel σFP1 = σFP2 = 797 N/mm2 can be obtained from equation (21) with the bending stress number σFlim = 500 N/mm2. Siees MD · 2009

1.4 Gear unit types 1.4.1 Standard designs In the industrial practice, different types of gear units are used. Preferably, standard helical and bevel-helical gear units with fixed transmission ratio and size gradation are applied. These single-stage to four-stage gear units according to the modular construction system cover a wide range of speeds and torques required by the driven machines. Combined with a standard electric motor such gear units are, as a rule, the most economical drive solution. But there are also cases where no standard drives are used. Among others, this is true for high torques above the range of standard gear units. In such cases, special design gear units are used, load sharing gear units playing an important role there.

1.4.3 Comparisons In the following, single-stage and two-stage gear units up to a ratio of i = 16 are examined. For common gear units the last or the last and the last but one gear stage usually come to approx. 70 to 80% of the total weight and also of the manufacturing expenditure. Adding further gear stages in order to achieve higher transmission ratios thus does not change anything about the following fundamental description. In figure 21, gear units without and with load sharing are shown, shaft 1 each being the HSS and shaft 2 being the LSS. With speeds n1 and n2, the transmission ratio can be obtained from the formula i = n1 / n2

(24)

107

10

Cylindrical Gear Units Gear Unit Types

The diameter ratios of the gears shown in figure 21 correspond to the transmission ratio i = 7. The gear units have the same output torques, so that in figure 21 a size comparison to scale is illustrated. Gear units A, B, and C are with offset shaft arrangement, and gear units D, E, F, and G with coaxial shaft arrangement.

In gear unit D the load of the high-speed gear stage is equally shared between three gears which is achieved by the radial movability of the sun gear on shaft 1. In the low-speed gear stage the load is shared six times altogether by means of the double helical teeth and the axial movability of the intermediate shaft. In order to achieve equal load distribution between the three intermediate gears of gear units E, F, and G the sun gear on shaft 1 mostly is radially movable. The large internal gear is an annulus gear which in the case of gear unit E is connected with shaft 2, and in the case of gear units F and G with the housing. In gear units F and G, web and shaft 2 form an integrated whole. The idler gears rotate as planets around the central axle. In gear unit G, double helical teeth and axial movability of the idler gears guarantee equal load distribution between six branches. 1.4.3.1 Load value By means of load value BL, it is possible to compare cylindrical gear units with different ultimate stress values of the gear materials with each other in the following examinations. According to /14/, the load value is the tooth peripheral force Fu referred to the pinion pitch diameter dw and the carrying facewidth b, i.e. BL =

Fu b · dw

(25)

The permissible load values of the meshings of the cylindrical gear units can be computed from the pitting resistance by approximation, as shown in /15/ (see section 1.3.4), using the following formula:

10

Figure 21 Diagrammatic view of cylindrical gear unit types without and with load sharing. Transmission ratio i = 7. Size comparison to scale of gear units with the same output torque.

Gear unit A has one stage, gear unit B has two stages. Both gear units are without load sharing. Gear units C, D, E, F, and G have two stages and are load sharing. The idler gears in gear units C and D have different diameters. In gear units E, F, and G the idler gears of one shaft have been ed to one gear so that they are also considered to be singlestage gear units. Gear unit C has double load sharing. Uniform load distribution is achieved in the highspeed gear stage by double helical teeth and the axial movability of shaft 1.

108

BL ≈ 7 · 10-6

u u+1

σ2Hlim KA · SH2

(26)

with BL in N/mm2 and allowable stress number for stress (pitting) σHlim in N/mm2 as well as gear ratio u, application factor KA and factor of safety from pitting SH. The value of the gear ratio u is always greater than 1, and is negative for internal gear pairs (see table 3). Load value BL is a specific quantity and independent of the size of the cylindrical gear unit. The following applies for practically executed gear units: cylindrical gears out of case hardening steel BL = 4 ... 6 N/mm2; cylindrical gears out of quenched and tempered steel BL = 1 ... 1.5 N/mm2; planetary gear stages with annulus gears out of quenched and tempered steel, planet gears and sun gears out of case hardening steel BL = 2.0 ... 3.5 N/mm2. Siees MD · 2009


1.4.3.2 Referred torques In figure 22, referred torques for the gear units shown in figure 21 are represented, dependent on the transmission ratio i. Further explanations are given in table 7. The torque T2 is referred to the construction dimension D when comparing the sizes, to the weight of the gear unit G when

comparing the weights, and to the generated surface A of the pitch circle cylinders when comparing the gear teeth surfaces. Gear unit weight G and gear teeth surface A (= generated surface) are measures for the manufacturing cost. The higher a curve, in figure 22, the better the respective gear unit in comparison with the others.

Table 7 Referred Torques Comparison criteria

Definition T2

δ =

Size

Dimension

D3 BL T2 γ = G BL

Weight Gear teeth surface

α =

T2 A3/2

BL

Ratio i a) Torque referred to size

m mm m mm2 kg mm2 m2

Units of the basic details T2 in Nm BL in N/mm2 D in mm G in kg A in m2

Ratio i b) Torque referred to gear unit weight

10

Ratio i c) Torque referred to gear teeth surface

Ratio i d) Full-load efficiency

Figure 22 Comparisons of cylindrical gear unit types in figure 21 dependent on the transmission ratio i. Explanations are given in table 7 as well as in the text. Siees MD · 2009

109


10

For all gear units explained in figures 21 and 22, the same prerequisites are valid. For all gear units, the construction dimension D is larger than the sum of the pitch diameters by the factor 1.15. Similar definitions are valid for gear unit height and width. Also the wall thickness of the housing is in a fixed relation to the construction dimension D /15/. With a given torque T2 and with a load value BL computed according to equation (26), the construction dimension D, the gear unit weight G, and the gear teeth surface A can be determined by approximation by figure 22 for a given transmission ratio i. However, the weights of modular-type gear units are usually higher, since the housing dimensions are determined according to different points of view. Referred to size and weight, planetary gear units F and G have the highest torques at small ratios i. For ratios i < 4, the planet gear becomes the pinion instead of the sun gear. Space requirement and load carrying capacity of the planet gear bearings decrease considerably. Usually, the planet gear bearings are arranged in the planet carrier for ratio i < 4.5. Gear units C and D, which have only external gears, have the highest torque referred to size and weight for ratios above i ≈ 7. For planetary gear units, the torque referred to the gear teeth surface is more favourable only in case of small ratios, if compared with other gear units. It is to be taken into consideration, however, that internal gears require higher manufacturing expenditure than external gears for the same quality of manufacture. The comparisons show that there is no optimal gear unit available which combines all advantages over the entire transmission ratio range. Thus, the output torque referred to size and weight is the most favourable for the planetary gear unit, and this all the more, the smaller the transmission ratio in the planetary gear stage. With increasing ratio, however, the referred torque decreases considerably. For ratios above i = 8, load sharing gear units having external gears only are more favourable because with increasing ratio the referred torque decreases only slightly. With regard to the gear teeth surface, planetary gear units do not have such big advantages if compared to load sharing gear units having external gears only. 1.4.3.3 Efficiencies When comparing the efficiencies, figure 22d, only the power losses in the meshings are taken into consideration. Under full load, they come to approx. 85% of the total power loss for common

110

cylindrical gear units with rolling bearings. The efficiency as a quantity of energy losses results from the following relation with the input power at shaft 1 and the torques T1 and T2

Ť

+ 1 i

T2 T1

Ť

(27)

All gear units shown in figure 21 are based on the same coefficient of friction of tooth profile μz = 0.06. Furthermore, gears without addendum modification and numbers of teeth of the pinion z = 17 are uniformly assumed for all gear units /15/, so that a comparison is possible. The single stage gear unit A has the best efficiency. The efficiencies of the two stage gear units B, C, D, E, F, and G are lower because the power flow es two meshings. The internal gear pairs in gear units E, F, and G show better efficiencies owing to lower sliding velocities in the meshings compared to gear units B, C, and D which only have external gear pairs. The lossfree coupling performance of planetary gear units F and G results in a further improvement of the efficiency. It is therefore higher than that of other comparable load sharing gear units. For higher transmission ratios, however, more planetary gear stages are to be arranged in series so that the advantage of a better efficiency compared to gear units B, C, and D is lost. 1.4.3.4 Example Given: Two planetary gear stages of type F arranged in series, total transmission ratio i = 20, output torque T2 = 3 · 106 Nm, load value BL = 2.3 N/mm2. A minimum of weight is approximately achieved by a transmission ratio division of i = 5 · 4 of the HS and LS stage. At γ1 = 30 m mm2/kg and γ2 = 45 m mm2/kg according to figure 22 b, the weight for the HS stage is approximately 10.9 t and for the LS stage approximately 30 t, which is a total 40.9 t. The total efficiency according to figure 22 d is η = 0.986 · 0.985 = 0.971. In comparison to a gear unit of type D with the same transmission ratio i = 20 and the same output torque T2 = 3 · 106 Nm, however, with a better load value BL = 4 N/mm2 this gear unit has a weight of 68.2 t according to figure 22 with γ = 11 m mm2/kg and is thus heavier by 67%. The advantage is a better efficiency of η = 0.98. The two planetary gear stages of type F together have a power loss which is by 45% higher than that of the gear unit type D. In addition, there is not enough space for the rolling bearings of the planet gears in the stage with i = 4. Siees MD · 2009

Level correction (dB)

Cylindrical Gear Units Noise Emitted by Gear Units

Correction curve A

Frequency (Hz) Figure 23 Correction curve according to DIN 45635 /16/ for the A-weighted sound power level or sound pressure level 1.5 Noise emitted by gear units 1.5.1 Definitions Noise emitted by a gear unit – like all other noises – is composed of tones having different frequencies f. Measure of intensity is the sound pressure p which is the difference between the highest (or lowest) and the mean pressure in a sound wave detected by the human ear. The sound pressure can be determined for a single frequency or – as a combination – for a frequency range (single-number rating). It is dependent on the distance to the source of sound. In general, no absolute values are used but amplification or level quantities in bel (B) or decibel (dB). Conversion of the absolute values is made for the sound pressure using equation Lp = 20 · log(p/p0) [dB]

(28)

and for the sound power using equation LW = 10 · log(P/P0) [dB] Siees MD · 2009

The reference values (e.g. p0 and P0) have been determined in DIN EN ISO 1683. For the sound pressure, the threshold of audibility of the human ear at 2 kHz has been taken as reference value (p0 = 2 · 10-5 Pa). For the conversion of the sound power applies (P0 = 10-12 W). In order to take into consideration the different sensitivities of the human ear at different frequencies, the physical sound pressure value at the different frequencies is corrected according to rating curve A, see figure 23. A-weighted quantities are marked by subscript “A” (e.g. sound pressure Lp; A-weighted sound pressure LpA). Apart from sound pressures at certain places, sound powers and sound intensities of a whole system can be determined. From the gear unit power a very small part is turned into sound power. This mainly occurs in the meshings, but also on bearings, fan blades, or by oil movements. The sound power is transmitted from the sources to the outside gear unit surfaces mainly by structure-borne noise (material vibrations). From the outside surfaces, air borne noise is emitted.

(29)

111

10


The sound intensity is the flux of sound power through a unit area normal to the direction of propagation. For a point source of sound it results from the sound power LW divided by the spherical enveloping surface 4 · π · r2, concentrically enveloping the source of sound. Like the sound pressure, the sound intensity is dependent on the distance to the source of sound, however, unlike the sound pressure it is a directional quantity. The recording instrument stores the sound pressure or sound intensity over a certain period of time and writes the dB values in frequency ranges (bands) into the spectrum (system of coordinates). Very small frequency ranges, e.g. 10 Hz or 1/12 octaves are termed narrow bands, see figure 24.

Sound intensity level [dB(A)]

Bandwidth

Frequency (Hz)

Figure 25 One-third octave spectrum of a gear unit (sound intensity level, A-weighted)

Bandwidth Sound intensity level [dB(A)]

The sound power LWA is the A-weighted sound power emitted from the source of sound and thus a quantity independent of the distance. The sound power can be converted to an average sound pressure for a certain place. The sound pressure decreases with increasing distance from the source of sound.

Frequency (Hz)

Figure 26 Octave spectrum of a gear unit (sound intensity level, A-weighted) (Frequency)

Figure 24 Narrow band frequency spectrum for LpA (A-weighted sound pressure level) at a distance of 1 m from a gear unit.

10

Histograms occur in the one-third octave spectrum and in the octave spectrum, see figures 25 and 26. In the one-third octave spectrum (spectrum with 1/3 octaves), the bandwidth results from f o / fu =

3

2, i.e. fo / fu = 1.26,

fo = fm . 1.12 and fu = fm / 1.12; fm = mean band frequency, fo = upper band frequency, fu = lower band frequency. In case of octaves, the upper frequency is twice as big as the lower one, or fo = fm . 1.41 and fu = fm / 1.41.

112

The total level (resulting from logarithmic addition of individual levels of the recorded frequency range) is a single-number rating. The total level is the common logical value for gear unit noises. The sound pressure level is valid for a certain distance, in general 1 m from the housing surface as an ideal parallelepiped. 1.5.2 Measurements The main noise emission parameter is the sound power level. 1.5.2.1 Determination via sound pressure DIN 45635 Part 1 and Part 23 describe how to determine the sound power levels of a given gear unit /16/. For this purpose, sound pressure levels LpA are measured at fixed points surrounding the gear unit and converted into sound power levels LWA. The measurement surface ratio LS is an auxiliary quantity which is dependent on the sum of the measurement surfaces. When the gear unit is placed on a reverberant base, the bottom is not taken into consideration, see example in figure 27. Siees MD · 2009


Machine enclosing reference box Measurement surface

Figure 27 Example of arrangement of measuring points according to DIN 45635 /16/ In order to really detect the noise radiated by the gear unit alone, corrections for background noise and environmental influences are to be made. They are estimated by measuring background noises (caused by noise radiating machines in the vicinity) and the characteristics of the room (reverberation time, resonances in the room) and are used as correction values in the sound power calculation. If the background noises are too loud (limit values of correction factors are achieved), this method can no longer be used because of insufficient accuracy. 1.5.2.2 Determination via sound intensity The gear unit surface is scanned manually all around at a distance of, for instance, 10 cm, by means of a special measuring device containing two opposing microphones. The mean of the levels is taken via the specified time, e.g. two minutes. The sound intensity determined in this way is the average sound energy flow penetrating the scanned surface. The sound power can be determined by multiplying the sound intensity by the scanned surface area. Gear units

This method has been standardized in DIN EN ISO 9614-2. Because of the special property of the measuring device – to determine the direction of sound incidence – it is very easy to eliminate background noises. The results correspond to the values as determined in accordance with DIN 45635. As a rule, the sound intensity method is more accurate (less measurement uncertainty) because it is less insensitive to noises and can also be used in case of loud background noises (e.g. in industrial plants). 1.5.3 Prediction It is not possible to exactly calculate in advance the sound power level of a gear unit to be made. However, one can base the calculations on experience. In the VDI guidelines 2159 /17/, for example, reference values are given. Gear unit manufacturers, too, mostly have own records. The VDI guidelines are based on measurements carried out on a large number of industrial gear units. Main influence parameters for gear unit noises are gear unit type, transmitted power, manufacturing quality, and speed. In VDI 2159, a distinction is made between cylindrical gear units with rolling bearings, see figure 28, cylindrical gear units with sliding bearings (highspeed gear units), bevel gear and bevel-helical gear units, planetary gear units, and worm gear units. Furthermore, information on speed variators can be found in the guidelines. Figure 28 exemplary illustrates a characteristic diagram of emissions for cylindrical gear units. Similar characteristic diagrams are also available for the other gear unit types mentioned. Within the characteristic diagrams, 50%- and 80%-lines are drawn. The 80%-line means, for example, that 80% of the recorded industrial gear units radiate lower noises. The lines are determined by mathematical equations. For the 80%-lines, the equations according to VDI 2159 are: Total sound power level LWA

Cylindrical gear units (rolling bearings)

77.1 + 12.3 . log P / kW (dB)

Cylindrical gear units (sliding bearings)

85.6 + 6.4 . log P / kW (dB)

Bevel gear and bevel-helical gear units

71.7 + 15.9 . log P / kW (dB)

Planetary gear units

87.7 + 4.4 . log P / kW (dB)

Worm gear units

65.0 + 15.9 . log P / kW (dB)

For restrictions, see VDI 2159. Siees MD · 2009

113

10


Type: Cylindrical gear units with external teeth mainly (> 80%) having the following characteristic

Power rating: 0.7 up to 2400 kW Input speed ( = max. speed): 1000 up to 5000 min-1 (mostly 1500 min-1) Max. circumferential speed: 1 up to 20 ms-1 Output torque: 100 up to 200 000 Nm No. of gear stages: 1 to 3 Information on gear teeth: HS gear stage with helical teeth ( = 10° up to 30°), hardened, fine-machined, DIN quality 5 to 8

Sound power level LWA

features: Housing: Cast iron housing Bearing arrangement: Rolling bearings Lubrication: Dip lubrication Installation: Rigid on steel or concrete

Mechanical power rating P Figure 28 Characteristic diagram of emissions for cylindrical gear units (industrial gear units) acc. to VDI 2159 /17/

To calculate a sound pressure level from the given sound power values a measuring method is used comparable with that described in DIN 45635. It is assumed that the sound energy is uniquely radiated from the object in all directions and can propagate undisturbed (free sound propagation). This assumption results in the so-called measurement surface sound pressure level, the average sound pressure at a determined distance to the gear unit. The measurement surface sound pressure level LpA at a distance of 1 m is calculated from the total sound power level LpA = LWA – Ls (dB)

10

Ls

= 10 . log S (dB)

(30)

(31)

S = Sum of the hypothetical surfaces (m2) enveloping the gear unit at a distance of 1 m (ideal parallelepiped) Example of information for P = 100 kW in a twostage cylindrical gear unit of size 200 (centre distance in the 2nd gear stage in mm), with rolling bearings, of standard quality: “The sound power level, determined in accordance with DIN 45635 (sound pressure measurement) or according to the sound intensity measurement method, is 102 ± 3 dB (A). Room and connection influences have not been taken into consideration. If it is agreed that measurements are to be made they will be carried out on the manufacturer’s test stand.”

114

Logarithmic regression LWA = 77.1 + 12.3 x log P/kW dB (80%-line) Certainty rate r2 = 0.83 Probability 90%

Note: For this example, a measurement surface sound pressure level of 102 - 13.2 ≈ 89 dB (A), tolerance ± 3 dB, is calculated at a distance of 1 m with a measurement surface S = 21 m2 and a measurement surface ratio LS = 13.2 dB. (Error of measurement according to DIN EN ISO 9614-2 for measurements in the industrial area with accuracy grade 2.) Individual levels in a frequency spectrum cannot safely be predicted for gear units because of the multitude of influence parameters. 1.5.4 Possibilities of influencing With the selection of other than standard geometries and with special tooth modifications (see section 1.2.5), gear unit noises can be positively influenced. In some cases, such a procedure results in a reduction in the performance (e.g. module reduction) for the same size, in any case, however, in special design and manufacturing expenditure. Housing design, distribution of masses, type of rolling bearing, lubrication and cooling are also important. Sometimes, the only way is to enclose the gear units which makes possible that the total level is reduced by 10 to 25 dB, dependent on the conditions. Attention has to be paid to it, that no structureborne noise is radiated via coupled elements (couplings, connections) to other places from where then airborne noise will be emitted. A sound screen does not only hinder the propagation of airborne noise but also the heat dissipation of a gear unit, and it requires more space. Siees MD · 2009


Shaft Couplings

Page


116

Torsionally Rigid Couplings, Flexible Pin Couplings Flexible Claw Couplings

117

Highly Flexible Ring Couplings, Highly Flexible Rubber Tyre Couplings Highly Flexible Rubber Disk Couplings, Flexible Pin and Bush Couplings

118

All-steel Couplings, Torque Limiters High-speed Couplings, Composite Couplings

119

Miniature Couplings, Gear Couplings Universal Gear Couplings, Multiple Disk Clutches

120

Fluid Couplings, Overrunning Clutches, Torque Limiters

121

Couplings for Pump Drives

122

Coupling Systems for Railway Vehicles

123

Coupling Systems for Wind Turbines

124

11

·

115

Shaft Couplings General Fundamental Principles Rigid and Torsionally Flexible Couplings 2. Shaft couplings 2.1 General fundamental principles In mechanical equipment, drives are consisting of components like prime mover, gear unit, shafts, and driven machine. Such components are connected by couplings which have the following tasks: S Transmitting motion of rotation and torques; S Compensating shaft misalignments (radial, axial, angular); S Reducing the torsional vibration load, influencing and displacing the resonant ranges;

S S S S S

Damping torque and speed impulses; Interrupting the motion of rotation (clutches); Limiting the torque (torque limiters); Sound isolation; Electrical insulation.

The diversity of possible coupling variants is shown in the overview in figure 29. A distinction is made between the two main groups: couplings and clutches.

Shaft couplings

Couplings

Rigid

Flexible

Clamp couplings Flange couplings Radial tooth couplings

Externally operated

Torque controlled

Speed controlled

Direction-of-rotation controlled

Clutches

Torque limiters

Centrifugal clutches

Overriding clutches Overrunning clutches

Friction Hydrodynamic couplings Magnetic couplings Friction couplings

11

Clutches

Positive

Tors. rigid

Tors. flexible

Highly flexible

Gear couplings

Steel spring couplings

Rubber tyre couplings

All-steel couplings Universal t couplings

Pin and bush couplings Claw couplings Flexible element couplings

Rubber disk couplings Rubber ring couplings

Parallel offset couplings

Figure 29 Overview of possible shaft coupling designs

116

·

Shaft Couplings Torsionally Rigid Couplings, Flexible Pin Couplings Flexible Claw Couplings

Torsionally rigid couplings Connect two shafts ends torsionally rigid and exactly centered to each other S designed for heavily stressed shafts S not subject to wear and require no maintenance S suitable for both directions of rotation Nominal torque: 1 300 ... 180 000 Nm

on request

Flexible pin couplings Universally applicable coupling for compensating shaft displacements S maximum operational reliability owing to fail-safe device S suitable for plugin assembly and simplified assembly of the design consisting of three parts Nominal torque: 19 ... 62 000 Nm

Brochure MD 10.1

Flexible pin couplings Disconnecting driving and driven machines upon failure of flexible elements (without fail-safe device) S universally applicable since combination with all parts of the N-EUPEX product range is possible Nominal torque: 19 ... 21 200 Nm

Brochure MD 10.1

11

Flexible claw couplings Fail-safe universal coupling S very compact design, high power capacity S very well suitable for plug-in assembly and assembly into bell housing S also with Taper bush for easy assembly and bore adaptation Nominal torque: 13.5 ... 3 700 Nm

Brochure MD 10.1

·

117

Shaft Couplings Highly Flexible Ring Couplings, Highly Flexible Rubber Tyre Couplings Highly Flexible Rubber Disk Couplings, Flexible Pin and Bush Couplings

$ Highly flexible ring couplings Coupling without torsional backlash S can be used for large shaft misalignments S suitable for high dynamic loads, good damping properties Nominal torque: 1 600 ... 90 000 Nm

Brochure MD 10.1

$ Highly flexible rubber tyre couplings Coupling without torsional backlash S compensating very large shaft misalignments S the rubber tyre can be easily replaced without the need to move the coupled machines S easy mounting on the shafts to be connected by means of Taper bushes Nominal torque: 24 ... 14 500 Nm

Brochure MD 10.1

$ Highly flexible rubber disk couplings For connecting machines having a very nonuniform torque characteristic S very easy plug-in assembly S replacement of rubber disk element is possible without the need to move the coupled machines S flange with dimensions acc. to SAE J620d Nominal torque: 330 ... 63 000 Nm

Brochure MD 10.1

11

Flexible pin and bush couplings Fail-safe universal coupling for medium up to high torques, absorbing large shaft displacements S compact design, low weights and mass moments of inertia S suitable for plug-in assembly Nominal torque: 200 ... 1 300 000 Nm

Brochure MD 10.1

118

·

Shaft Couplings All-steel Couplings, Torque Limiters High-speed Couplings, Composite Couplings

All-steel couplings Torsionally rigid coupling without clearance S compensates radial, angular and axial shaft displacements by means of two flexible disc packs S packs made out of stainless spring steel S easy assembly of coupling due to compact disc packs S modular system: many standard types by combination of standard components Nominal torque: 92 ... 1 450 000 Nm Brochure MD 10.1

Torque limiters On reaching the preset disconnecting torque, the torque limiter separates the coupled drive components both during slow and fast rising torques S after the disengagement process the coupling halves are out of , so that a wear-free running down can be realized Nominal torque: 10 ... 75 000 Nm Brochure MD 10.11

! High-speed couplings Were designed for the energy and petrochemical industries and marine propulsion drives S are used for all high-speed purposes where reliable power transmission is required even with unavoidable shaft misalignment S meet the requirements of API 671 Nominal torque: 1 000 ... 535 000 Nm Brochure MD 10.9

"#

11

Composite couplings Corrosion-resistant, extreme light weight coupling for drives with great shaft distances (e.g. cooling tower fan) S up to 6 metres without centre bearing S easy to handle and to install S maintenance-free and wear-free S reduced coupling vibrations Nominal torque: 1 250 ... 7 600 Nm Brochure MD 10.5

·

119

Shaft Couplings Miniature Couplings, Gear Couplings Universal Gear Couplings, Multiple Disk Clutches

& Miniature couplings Designed for applications with very low torques S fields of application: control systems, machine tools, computer technology, tacho drives, measuring and ing systems, printing and packaging machines, stepping and servomotors, test stands Nominal torque: 5 ... 25 Nm

Brochure MD 10.10

%% Gear couplings Double-ted coupling compensating angular, parallel and axial misalignment of shafts S long-term lubrication is ensured by design measures and by using special seals S small dimensions; can be used for high shock loads S available in many types and variants Nominal torque: 1 300 ... 7 200 000 Nm

Brochure MD 10.1

%% Universal gear couplings Double-ted gear coupling with hobbed and crowned external gear teeth and low torsional backlash S largest possible bore range with grease lubricated gear teeth S mounting dimensions in metric and inch measures acc. to international standards Nominal torque: 850 ... 125 000 Nm

Brochure MD 10.1

11

$' Multiple disk clutches Constant torque transmission by means of pressure ensured by springs S many applications possible owing to mechanical, electrical, pneumatic or hydraulic disengaging devices S protects drives against overloading Nominal torque: 10 ... 30 000 Nm

on request

120

·

Shaft Couplings Fluid Couplings, Overrunning Clutches Torque Limiters

$ Fluid couplings Soft starting without shocks and acceleration of large masses during a load-relieved start of motor S torque limitation during starting and overload S excellent vibration separation and shock damping S torque transmission without wear Nominal power ratings: 0.5 ... 2 500 kW

Brochure MD 10.1

% Overrunning clutches Flender overrunning clutches allow to drive shafts and machines first by means of an auxiliary drive at low speed for startup and then by means of the main drive at higher speeds for full-load operation, the auxiliary drive then being shut off by overrunning. Nominal torque: 9 000 ... 100 000 Nm

Dimensioned drawing M 495

Certified according to directive 94/9/EC (ATEX 95) This coupling is particularly suitable for the use in hazardous locations

Torque limiters With SECUREX, Flender provides a unique modular system of mechanical torque limiters. Owing to a variety of possibilities to combine standard components, the functions Protection from overload as well as Compensation of shaft misalignment can be fulfilled with just one compact unit. With the development of SECUREX, Flender has concentrated its experiences gained over decades in the fields of both overload protection and compensation of shaft misalignments in one product line. SECUREX is based on the wide range of Flender’s standard couplings of different basic types and on standardized safety elements. With this combination, economical coupling solutions can be realized. With the modular SECUREX system, Flender has focused on its core competence in the torque range of up to 1,500,000 Nm and benefits from its rich fund of knowledge and experience gained from application- and product-related R&D (e.g. sliding hubs in the wind energy industry, shear pin solutions in rolling mills, torque limiters in extruder applications, etc.). Brochure K 440

·

121

11

Shaft Couplings Couplings for Pump Drives

)* " .*,

Flexible pin couplings D Tried and tested drive element in millions of pump drives D Good value for money, reliable, available worldwide

Type A

D Complete application-oriented assortment! In addition to the fail-safe standard design, a variant without fail-safe device is available especially developed for hazardous locations Types B / BDS - in two parts Types A / ADS - in three parts Types H / HDS - with intermediate sleeve Certified acc. to directive 94/9/EC (ATEX 95) Type BDS

Type H Katalog MD 10.1

All-steel couplings Were specially designed for pump drives

11

D Meet the requirements of API 610 D Design according to API 671, “NON-SPARKING” and certified acc. to directive 94/9/EC (ATEX 95) also available Nominal torque: 100 ... 17 000 Nm

Katalog MD 10.1

122

·

Shaft Couplings Coupling Systems for Railway Vehicles

(#" )* * + ( ,- Input side couplings Membrane coupling, Type MBG D All-steel membrane coupling for the connection of motor and gear unit D Without backlash; compensating relatively small shaft misalignments Max. nominal torque: 3 425 Nm Max. shaft diameter: 86 mm

Gear coupling, Type ZBG D Double-ted grease lubricated gear coupling between motor and gear unit D Compensating extremely large shaft misalignments D Split spacer with crowned gear teeth Max. nominal torque: 15 000 Nm Max. shaft diameter: 100 mm

on request

Output side couplings Articulated t rubber coupling, Type GKG D Double-ted, flexible coupling without backlash, between axle drive and driving wheel shaft

11

D Low wear and low maintenance D Compensating extremely large shaft misalignments, with low restoring forces Max. nominal torque: 13 440 Nm Max. shaft diameter: 260 mm

on request

·

123

Shaft Couplings Coupling Systems for Wind Turbines

(#" )* + . #*/

$ Fluid couplings in combination with other couplings D Fluid coupling with slip between 2 and 3%. Peak torques caused by gusts of wind are compensated D Combination with RUPEX coupling for small shaft misalignments D Combination with articulated t rubber coupling or ARPEX coupling for large shaft misalignments on request

Articulated t rubber couplings Type GKGW with brake disk D Rubber-elastic ball bearings for extremely large shaft misalignments between gear unit and generator D Very low restoring forces D Electrically insulating and structure-borne noise absorbing D Wearing parts and coupling can be removed without the need to move the generator D Optional with torque-limiting slip hub

on request

All-steel couplings

11

D Design with hexagonal or square disc pack for very large shaft misalignments D Optionally with slip hub for limiting the torque load in case of generator short-circuit D Light spacer out of glass-fibre compound material for lightning insulation D Conical bolt connection of disc packs for easy assembly

on request

124

·


Vibrations Symbols and Units

Page 126


127 – 129

Solution Proposal for Simple Torsional Vibrators

129 + 130

Solution of the Differential Equation of Motion

130 + 131

Formulae for the Calculation of Vibrations

131

Mass

131

Mass Moment of Inertia

131

, Symbols, and Units

132

Determination of Stiffness

133

Overlaying of Different Stiffnesses

134

Conversions

134

Natural Frequencies

134

Evaluation of Vibrations

135 + 136

12

·

125

Vibrations Symbols and Units

a

m

Length of overhanging end

T

s

A

m2

Cross-sectional area

T

Nm

Torque

A

m, rad

Amplitude of oscillation

V

m3

Volume

V

–

Magnification factor; Dynamic / static load ratio

x

m

Displacement co-ordinate (translational, bending) Displacement amplitude

AD; Ae c

12

Damping energy; elastic energy Nm/rad Torsional stiffness Translational stiffness; bending stiffnes

Period of a vibration

c’

N/m

d

m

Diameter

x^

m

di

m

Inside diameter

α

rad

Phase angle

da

m

Outside diameter

γ

rad

Phase angle with free vibration

D

–

δ

1/s

Damping constant

Dm

m

ε

rad

Phase displacement angle with forced vibration

e=

2.718

Euler’s number

E

N/m2

Modulus of elasticity

η

–

Excitation frequency / natural frequency ratio

f, fe

Hz

Frequency; natural frequency

m

Deformation

λi

–

f

Inherent value factor for i-th natural frequency

F

N

Force

Λ

–

Logarithmic decrement

F (t)

N

Time-variable force

π=

3.14159

G

N/m2

i

–

kg/m3

Specific density

ϕ, ϕi

rad

Angle of rotation

rad

Angular amplitude of a vibration

Attenuation ratio (Lehr’s damping) Mean coil diameter (coil spring)

Shear modulus Transmission ratio Number of windings (coil spring)

iF

–

la

m4

Second axial moment of area

lp

m4

Second polar moment of area

J, Ji

kgm2

Mass moment of inertia

J*

kgm2

Reduced mass moment of inertia of a two-mass vibration generating system

k

Nms/ rad

Viscous damping in case of torsional vibrations

k’

Ns/m

l

^

.

Peripheral / diameter ratio

rad/s

Angular velocity (first time derivation of )

rad/s2

Angular acceleration (second time derivation of )

h

rad

Vibratory angle of the free vibration (homogeneous solution)

Viscous damping in case of translational and bending vibrations

p

rad

Vibratory angle of the forced vibration (particular solution)

m

Length; distance between bearings

^ p

rad

Angular amplitude of the forced vibration

m, mi

kg

Mass

M (t)

Nm

Time-variable excitation moment

^ stat

rad

M0

Nm

Amplitude of moment

Angular amplitude of the forced vibration under load ( = 0)

M0*

Nm

Reduced amplitude of moment of a two-mass vibration generating system

ψ

–

Damping coefficient acc. to DIN 740 /18/

ne

1/min

Natural frequency (vibrations per minute)

ω

rad/s

Angular velocity, natural radian frequency of the damped vibration

n1; n2

min-1

Input speed; output speed

0

rad/s

Natural radian frequency of the undamped vibration

Ω

rad/s

Radian frequency of the excitation on vibration

q

–

Influence factor for taking into the mass of the shaft when calculating the natural bending frequency

t

s

Time

126

..

Note: The unit “rad” ( = radian ) may be replaced with “1”. ·

Vibrations General Fundamental Principles

3. Vibrations 3.1 General fundamental principles Vibrations are more or less regularly occurring temporary variations of state variables. The state of a vibrating system can be described by suitable variables, such as displacement, angle, velocity, pressure, temperature, electric voltage/ current, and the like. The simplest form of a mechanical vibrating system consists of a mass and a spring with fixed ends, the mass acting as kinetic energy store

Translational vibration generatig system

and the spring as potential energy store, see figure 30. During vibration, a periodic conversion of potential energy to kinetic energy takes place, and vice versa, i.e. the kinetic energy of the mass and the energy stored in the spring are converted at certain intervals of time. Dependent on the mode of motion of the mass, a distinction is made between translational (bending) and torsional vibrating systems as well as coupled vibrating systems in which translational and torsional vibrations occur at the same time, influencing each other.

Bending vibration generating system

Torsional vibration generating system

Figure 30 Different vibrating systems with one degree of freedom Further, a distinction is made between free vibrations and externally forced vibrations, and whether the vibration takes place without energy losses (undamped) or with energy losses (damped). A vibration is free and undamped if energy is neither supplied nor removed by internal friction so that the existing energy content of the vibration is maintained. In this case the system carries out steady-state natural

Amplitude

Vibration

Amplitude

vibrations the frequency of which is determined only by the characteristics of the spring/mass system (natural frequency), figure 32 a. The vibration variation with time x can be described by the constant amplitude of oscillation A and a harmonic function (sine, cosine) the arguments of which contain natural radian frequency ω = 2 · π · f (f = natural frequency in Hertz) and time, see figure 31.

12

Period x A ω t

= = = =

A · sinω · t Amplitude Radian frequency Time

x = A · sin (ω · t + α) α = Phase angle

Figure 31 Mathematical description of an undamped vibration with and without phase angle ·

127

Vibrations General Fundamental Principles

A damped vibration exists, if during each period of oscillation a certain amount of vibrational energy is removed from the vibration generating system by internal or external friction. If a constant viscous damping (Newton’s friction) exists, the amplitudes of oscillation decrease in accordance with a geometric progression, figure 32 b. All technical vibration generating systems are subject to more or less strong damping effects.

Displacement x

a) Undamped vibration (δ = 0)

b) Damped vibration (δ > 0)

c) Stimulated vibration (δ < 0) Time t Figure 32 Vibration variations with time (A = initial amplitude at time t = 0; δ = damping constant)

12

If the vibrating system is excited by a periodic external force F(t) or moment M(t), this is a forced or stimulated vibration. With the periodic external excitation force, energy can be supplied to or removed from the vibrating system. After a building-up period, a damped vibrating system does no longer vibrate with its natural frequency but with the frequency of the external excitation force. Resonance exists, when the applied frequency is at the natural frequency of the system. Then, in undamped systems the amplitudes of oscillation grow at an unlimited degree, figure 32 c. In damped systems, the amplitude of oscillation

128

grows until the energy supplied by the excitation force and the energy converted into heat by the damping energy are in equilibrium. Resonance points may lead to high loads in the components and therefore are to be avoided or to be quickly traversed (example: natural bending frequency in high-speed gear units). The range of the occurring amplitudes of oscillation is divided by the resonance point (natural frequency = excitation frequency, critical vibrations) into the subcritical and supercritical oscillation range. As a rule, for technical vibrating systems (e.g. drives), a minimum frequency distance of 15% or larger from a resonance point is required. Technical vibrating systems often consist of several masses which are connected with each other by spring or damping elements. Such systems have as many natural frequencies with the corresponding natural vibration modes as degrees of freedom of motion. A free, i.e. unfixed torsional vibration system with n masses, for instance, has n-1 natural frequencies. All these natural frequencies can be excited to vibrate by periodic external or internal forces, where mostly only the lower natural frequencies and especially the basic frequency (first harmonic) are of importance. In technical drive systems, vibrations are excited by the following mechanisms: a) From the input side: Starting processes of electric motors, system short circuits, Diesel and Otto engines, turbines, unsteady processes, starting shock impulses, control actions. b) From transmitting elements: Meshing, unbalance, universal-t shaft, alignment error, influences from bearings. c) From the output side: Principle of the driven machine, uniform, nonuniform, e.g. piston compressor, propeller. As a rule, periodic excitation functions can be described by means of sine or cosine functions and the superpositions thereof. When analysing vibration processes, a Fourier analysis may often be helpful where periodic excitation processes are resolved into fundamental and harmonic oscillations and thus in comparison with the natural frequencies of a system show possible resonance points. In case of simple vibrating systems with one or few (maximum 4) masses, analytic solutions for the natural frequencies and the vibration variation with time can be given for steady excitation. For unsteady loaded vibrating ·

Vibrations General Fundamental Principles Solution Proposal for Simple Torsional Vibrators systems with one or more masses, however, solutions can be calculated only with the aid of numerical simulation programmes. This applies even more to vibrating systems with non-linear or periodic variable parameters (non-linear torsional stiffness of couplings; periodic meshing stiffnesses). With EDP programmes, loads with steady as well as unsteady excitation can be simulated for complex vibrating systems (linear, non-linear, parameter-excited) and the results be represented in the form of frequency analyses, load as a function of time, and overvoltages of resonance. Drive systems with torsionally flexiFixed one-mass vibration generating system

ble couplings can be designed dynamically in accordance with DIN 740 /18/. In this standard, simplified solution proposals for shockloaded and periodically loaded drives are made, the drive train having been reduced to a two-mass vibration generating system. 3.2 Solution proposal for simple torsional vibrators Analytic solution for a periodically excited one(fixed) or two-mass vibration generating system, figure 33.

Free two-mass vibration generating system

Figure 33 Torsional vibrators J, J1, J2 c k M (t)

ö .

ö.. ö

mass moment of inertia [kgm2] torsional stiffness [Nm/rad] viscous damping [Nms/rad] external excitation moment [Nm], time-variable angle of rotation [rad], ( ϕ = ϕ1 – ϕ2 for two-mass vibration generating systems as relative angle) = angular velocity [rad/s] (first time derivation of ϕ) = angular acceleration [rad/s2] (second time derivation of ϕ)

= = = = =

Differential equation of motion:

Natural radian frequency (undamped): ω0

One-mass vibration generating system:

ω0 = (32)

0 +

(

(

M (t) .. . )k )c@+ J J J 20 2

Two-mass vibration generating system with relative coordinate:

M(t) ) k @) c @+ J1 J* J* .

(

2

with + 1 * 2

J* =

J 1 . J2 J1 + J2

·

fe = ne =

20 (34)

(35)

Ǹ

c@

[rad/s]

J1 ) J2 J1 @ J2

(36)

ƪradńsƫ

(37)

Natural frequency:

(33)

(

..

c J

ω0 2π ω0 . 30 π

[Hz]

(38)

[1/min]

+ k + damping constant J

(39)

[1/s]

(40)

ω0 = natural radian frequency of the undamped vibration [rad/s] fe = natural frequency [Hertz] ne = natural frequency [1/min]

129

12

Vibrations Solution Proposal for Simple Torsional Vibrators Solution of the Differential Equation of Motion 3.3 Solution of the differential equation of motion

Damped natural radian frequency:

+ Ǹ 20 * 2 + 0 @ Ǹ1 * D 2

(41)

Periodic excitation moment M(t) + M 0 @ cos @ t

Attenuation ratio (Lehr’s damping): D

D =

ψ k . ω0 δ = = ω0 4π 2.c

M0 = amplitude of moment [Nm] Ω = exciting circuit frequency [rad/s] (42)

ψ = damping coefficient on torsionally flexible coupling, determined by a damping hysteresis of a period of oscillation acc. to DIN 740 /18/ and/or acc. to Flender brochure. ψ =

damping energy elastic deformation energy

(43)

=

AD Ae

Reference values for some components:

D = 0.001...0.01

shafts (material damping of steel)

D = 0.04...0.08

gear teeth in gear units

Total solution: + h ) p

(44)

a) Free vibration (homogeneous solution öh )

ϕh = A . e – δ

.t .

cos ( ω . t – γ )

Constants A and γ are determined by the start. ing conditions, e.g. by h = 0 and h = 0 (initialvalue problem). In damped vibrating systems (δ > 0) the free component of vibration disappears after a transient period.

D = 0.04...0.15 (0.2) torsionally flexible couplings

b) Forced vibration (particular solution öp )

D = 0.01...0.04

M* 1 p + c0 @ Ǹ(1 * 2) 2 ) 4D2 @ 2

gear couplings, all-steel couplings, universal t shafts

@ cos ( @ t * ε) Static spring characteristic for one load cycle

(45)

(46)

Phase angle: tan ε = Frequency ratio:

2.D.η 1 – η2

+

(47)

(48)

0

One-mass vibration generating system: M0 * + M0 Two-mass vibration generating system: J2 . M M0 * = 0 J1 + J2

(49)

(50)

12 c) Magnification factor

Figure 34 Damping hysteresis of a torsionally flexible component

ϕp = V+

130

M0 * c

.

V

.

cos (Ω

.

t – ε)

(51)

^ p 1 + ^ + M (52) Ǹ(1 * 2) 2 ) 4D2 @ 2 stat M *0 ·

Vibrations Solution of the Differential Equation of Motion Formulae for the Calculation of Vibrations ^ p

The magnification factor shows the ratio of the dynamic and static load and is a measure for the additional load caused by vibrations (figure 35).

= vibration amplitude of forced vibration

Magnification factor V

Phase displacement angle ε

^ stat = vibration amplitude of forced vibration at a frequency ratio η = 0.

Figure 35 Magnification factors for forced, damped and undamped vibrations at periodic moment excitation (power excitation). Magnification factors V and phase displacement angle ε.

Frequency ratio + +

0

3.4 Formulae for the calculation of vibrations For the calculation of natural frequencies and vibrational loads, a general vibration generating system has to be converted to a calculable substitute system with point masses, spring and damping elements without mass.

3.4.1 Mass m = · V [kg] V = volume [m3] = specific density [kg/m3] 3.4.2 Mass moment of inertia J =

ŕ r dm: 2

Mass moment of inertia

Cylinder D

general integral formula Torsional stiffness

J=

.π.L 32

.

D4

c=

π.G 32 L

.

D4

J=

.π.L 32

.

(D4 – d4)

c=

π.G 32 L

.

(D4 – d4)

J=

.π.L 160

.

D15 – D25 D1 – D2

c=

3.π.G 32 L

L Hollow cylinder d

D L

Cone D1

D2 L

Hollow cone

J=

D1 d1

.π.L

.

160

d2 D2 L

·

–

ǐ

(D13 . D23)

.

2

(D1 + D1D2 + D22)

12 5

5

D1 – D2 D1 – D2

d15 – d25 d1 – d2

Ǒ

c=

3.π.G

(D1

.

3.

D2

3)

(D12 + D1D2 + D22)

32 L –

(d13 . d23) (d12 + d1d2 + d22)

131

Vibrations , Symbols and Units

Table 8 Symbols and units of translational and torsional vibrations Term

Quantity

Unit

Explanation

Mass, Mass moment of inertia

m J

kg kg · m2

Translatory vibrating mass m; Torsionally vibrating mass with mass moment of inertia J

Instantaneous value of vibration (displacement, angle)

x ϕ

m rad *)

Instantaneous, time-dependent value of vibration amplitude

x max, x^ , A max, ^ , A

m rad

Amplitude is the maximum instantaneous value (peak value) of a vibration.

m/s rad/s

Oscillating velocity; Velocity is the instantaneous value of the velocity of change in the direction of vibration. The d’Alembert’s inertia force or the moment of inertia force acts in the opposite direction of the positive acceleration.

Amplitude

.

Oscillating velocity

..

Inertia force, Moment of inertia forces

m@x .. J@

Spring rate, Torsional spring rate Spring force, Spring moment Attenuation constant (Damping coefficient), Attenuation constant for rotary motion

c’ c c’ . x c.ϕ

Nm Linear springs N · m/rad N In case of linear springs, the spring recoil N·m is proportional to deflection.

k’

N · s/m In case of Newton’s friction, the damping force is proportional to velocity and Nms/rad attenuation constant (linear damping).

k

N N·m

δ = k’ / (2 . m) δ = k / (2 . J)

1/s 1/s

Attenuation ratio (Lehr’s damping)

D = δ/ω0

–

For D < 1, a damped vibration exists; for D ≥ 1, an aperiodic case exists.

Damping ratio

x^ n ń x^ n)1 ^ n ń ^ n)1

– –

The damping ratio is the relation between two amplitudes, one cycle apart.

–

+ In (x^ n ń x^ n)1) + In (^ n ń ^ n)1)

Damping factor (Decay coefficient)

Logarithmic damping decrement Time Phase angle

12

x .

+2@@D Ǹ1 * D 2 t

s

α

rad

Phase displacement angle

ε = α1 − α2

rad

Period of a vibration

T = 2 . π / ω0

s

Frequency of natural vibration Radian frequency of natural vibration

f = 1/T = ω0 /(2 . π)

Hz

ω0 = 2 . π . f

rad/s

0 + Ǹcńm 0 + ǸcńJ

rad/s rad/s

d + Ǹ 20 * 2

rad/s

Ω η = Ω/ω0

rad/s –

Natural radian frequency (Natural frequency) Natural radian frequency when damped Excitation frequency Radian frequency ratio

The damping factor is the damping coefficient referred to twice the mass.

Coordinate of running time In case of a positive value, it is a lead angle. Difference between phase angles of two vibration processes with same radian frequency. Time during which a single vibration occurs. Frequency is the reciprocal value to a period of vibrations; vibrations per sec. Radian frequency is the number of vibrations in 2 . π seconds. Vibration frequency of the natural vibration (undamped) of the system. For a very small attenuation ratio D < 1 becomes ωd ≈ ω0. Radian frequency of excitation Resonance exists at η = 1.

*) The unit “rad” may be replaced with “1”.

132

·

Vibrations Formulae for the Calculation of Vibrations

3.4.3 Determination of stiffness Table 9 Calculation of stiffness (examples) Example

Stiffness

Symbols

Coil spring c’ =

Torsion bar

ƨ Ʃ

G . d4

N m

3

8 . Dm . iF

ƨ Ʃ

G . Ιp

c =

iF = number of windings G = shear modulus 1) d = diameter of wire Dm = mean coil diameter

Nm rad

l

π . d4

Shaft: Ιp =

32 π

Hollow shaft: Ιp =

32

4

4

( da – di )

Ιp = second polar moment of area l = length d, di, da = diameters of shafts

Tension bar c’ =

Cantilever beam

c’ =

F f

=

ƨ Ʃ

E.A l

N m

l3

Hollow shaft: Ιa =

A

ƨ Ʃ

3 . E . Ιa

Shaft: Ιa =

E

N m

F f

π . d4 64 π 64

Ιa 4

4

( da – di )

= modulus of elasticity 1) = cross-sectional area

= force = deformation at centre of mass under force F = second axial moment of area

Transverse beam (single load in the middle)

c’ =

F f

=

48 . E . Ιa l3

ƨ Ʃ N m

Transverse beam with overhanging end c’ =

F f

=

3 . E . Ιa a2 . (l + a)

ƨ Ʃ N m

l a

= distance between bearings = length of overhanging end

1) For steel: E = 21 S 1010 N/m2; G = 8.1 S 1010 N/m2 ·

133

12


Measuring the stiffness: In a test, stiffness can be determined by measuring the deformation. This is particularly helpful if the geometric structure is very complex and very difficult to acquire.

frequency of 10 Hz (vibrational amplitude = 25% of the nominal coupling torque). The dynamic torsional stiffness is greater than the static torsional stiffness, see figure 36.

Translation: cȀ + F ƪNńmƫ f

3.4.4 Overlaying of different stiffnesses To determine resulting stiffnesses, single stiffnesses are to be added where arrangements in series connection or parallel connection are possible.

(53)

F = applied force [N] f = measured deformation [m] Torsion: T ƪNmńradƫ c+

(54)

Series connection: Rule: The individual springs in a series connection carry the same load, however, they are subjected to different deformations.

T = applied torsion torque [Nm] ϕ = measured torsion angle [rad]

1 1 1 1 1 c ges + c 1 ) c 2 ) c 3 ) AAA ) c n

Measurements of stiffness are furthermore required if the material properties of the spring material are very complex and it is difficult to rate them exactly. This applies, for instance, to rubber materials of which the resilient properties are dependent on temperature, load frequency, load, and mode of stress (tension, compression, shearing). Examples of application are torsionally flexible couplings and resilient buffers for vibration isolation of machines and internal combustion engines. These components often have non-linear progressive stiffness characteristics, dependent on the direction of load of the rubber material.

Parallel connection: Rule: The individual springs in a parallel connection are always subject to the same deformation. c ges + c 1 ) c 2 ) c 3 ) AAA ) c n

(55)

(56)

3.4.5 Conversions If drives or shafts with different speeds are combined in one vibration generating system, the stiffnesses and masses are to be converted to a reference speed (input or output). Conversion is carried out as a square of the transmission ratio: Transmission ratio: i =

Slope = static stiffness

Slope = dynamic stiffness

12 Figure 36 Static and dynamic torsional stiffness

For couplings the dynamic stiffness is given, as a rule, which is measured at a vibrational

134

reference speed n1 = n2 speed

(57)

Conversion of stiffnesses cn2 and masses Jn2 with speed n2 to the respective values cn1 and Jn1 with reference speed n1: c n1 + c n2ńi 2

(58)

J n1 + J n2ńi 2

(59)

Before combining stiffnesses and masses with different inherent speeds, conversion to the common reference speed has to be carried out first. 3.4.6 Natural frequencies a) Formulae for the calculation of the natural frequencies of a fixed one-mass vibration generating system and a free two-mass vibration generating system. Natural frequency fe in Hertz (1/s): ·

Vibrations Formulae for the Calculation of Vibrations Evaluation of Vibrations One-mass vibration generating system: Torsion: fe =

1 2π

c J

Two-mass vibration generating system: (60)

fe + 1 2

Ǹ

c

J1 ) J2 J1 @ J2

(61)

c = torsional stiffness [Nm/rad] J, Ji = mass moments of inertia [kgm2] Translation, Bending : f e + 1 2

Ǹ mcȀ

(62)

fe + 1 2

ǸcȀ

m1 ) m2 m1 @ m2

(63)

c’ = translational stiffness (bending stiffness) [N/m] m, mi = masses [kg] b) Natural bending frequencies of shafts ed at both ends with applied masses with known deformation f due to the dead weight. fe +

q 2

Ǹgf

[Hzƫ

Table 10 λ-values for the first three natural frequencies, dependent on mode of fixing Bearing application

(64)

g = 9.81 m/s2 gravity

λ1

λ2

λ3

1.875

4.694

7.855

4.730

7.853

10.966

π

2π

3π

3.927

7.069

10.210

f = deformation due to dead weight [m] q = factor reflecting the effect of the shaft masses on the applied mass q = 1 shaft mass is neglected compared with the applied mass q = 1.03 ... 1.09 common values when considering the shaft masses q = 1.13 solid shaft without pulley c) Natural bending frequencies for shafts, taking into dead weights (continuum); general formula for the natural frequency in the order fe, i.

ǒ Ǔ @ Ǹ@EA

f e,i + 1 @ i 2 l

2

ƪHzƫ

(65)

λi = inherent value factor for the i-th natural frequency l = length of shaft [m] E = modulus of elasticity [N/m2] Ι = moment of area [m4] = density [kg/m3] A = cross-sectional area [m2] d = diameter of solid shaft [m] ·

For the solid shaft with free bearing on both sides, equation (65) is simplified to: f e,i + @ d 8

ǒ il Ǔ @ ǸE 2

ƪHzƫ

(66)

i = 1st, 2nd, 3rd ... order of natural bending frequencies. 3.5 Evaluation of vibrations The dynamic load of machines can be determined by means of different measurement methods. Torsional vibration loads in drives, for example, can be measured directly on the shafts by means of wire strain gauges. This requires, however, much time for fixing the strain gauges, for calibration, signal transmission and evaluation. Since torques in shafts are generated via bearing pressure in gear units, belt drives, etc., in case of dynamic loads, structure-borne noise is generated which can be acquired by sensing elements at the bearing points in different directions (axial, horizontal, vertical). Dependent on the requirements, the amplitudes of vibration displacement, velocity and acceleration can be recorded and evaluated in a sum (effective vibration

135

12


velocity) or frequency-selective. The structure-borne noise signal reflects besides the torque load in the shafts also unbalances, alignment errors, meshing impulses, bearing noises, and possibly developing machine damages. To evaluate the actual state of a machine, VDI guideline 2056 1) or DIN ISO 10816-1 /19, 20/ is consulted for the effective vibration velocity, as a rule, taking into structure-borne noise in the frequency range between 10 and 1,000 Hertz. Dependent on the machine structure (resilient or rigid foundation) and power transmitted, a distinction is made between four machine groups (table 11). Dependent on the vibration velocity, the vibrational state of a machine is judged to be “good”, “acceptable”, “still permissible”, and “non-permis-

sible”. If vibration velocities are in the “nonpermissible” range, measures to improve the vibrational state of the machine (balancing, improving the alignment, replacing defective machine parts, displacing the resonance) are required, as a rule, or it has to be verified in detail that the vibrational state does not impair the service life of the machine (experience, verification by calculation). Structure-borne noise is emitted from the machine surface in the form of airborne noise and has an impact on the environment by the generated noises. For the evaluation of noise, sound pressure level and sound intensity are measured. Gear unit noises are evaluated according to VDI guideline 2159 or DIN 45635 /17, 16/, see subsection 1.5.

Table 11 Boundary limits acc. to VDI guideline 2056 1) for four machine groups

Machine groups

K

Range classification acc. to VDI 2056 (“Effective value of the vibration velocity” in mm/s)

Including gear units and machines with input in ut power ower ratings of ...

... up to approx. 15 kW without special foundation.

Still Nonpermissible permissible

Good

Acceptable

up to 0.7

0.7 ... 1,8

1.8 ... 4.5

from 4.5 up

up to 1.1

1.1 ... 2.8

2.8 ... 7.1

from 7.1 up

... from approx. 15 up to 75 kW without special foundation. M

12

... from approx. 75 up to 300 kW and installation on highly tuned, rigid or heavy foundations.

G

... over 300 kW and installation on highly tuned, rigid or heavy foundations.

up to 1.8

1.8 ... 4.5

4.5 ... 11

from 11 up

T

... over 75 kW and installation on broadly tuned resilient foundations (especially also steel foundations designed according to light-construction guidelines).

up to 2.8

2.8 ... 7

7 ... 18

from 18 up

1) 08/97 withdrawn without replacement; see /20/

136

·


Page Bibliography of Sections 10, 11, and 12

138 + 139

13

·

137

Bibliography

/1/

DIN 3960: Definitions, parameters and equations for involute cylindrical gears and gear pairs. March 1987 edition. Beuth Verlag GmbH, Berlin

/2/

DIN 3992: Addendum modification of external spur and helical gears. March 1964 edition. Beuth Verlag GmbH, Berlin

/3/

DIN 3993: Geometrical design of cylindrical internal involute gear pairs; Part 3. August 1981 edition. Beuth Verlag GmbH, Berlin

/4/

DIN 3994: Addendum modification of spur gears in the 05-system. August 1963 edition. Beuth Verlag GmbH, Berlin

/5/

Niemann, G. und Winter, H.: Maschinenelemente, Band II, Getriebe allgemein, Zahnradgetriebe-Grundlagen, Stirnradgetriebe. 3rd edition. Springer Verlag, Heidelberg, New York, Tokyo (1985)

/6/

Sigg, H.: Profile and longitudinal corrections on involute gears. Semi-Annual Meeting of the AGMA 1965, Paper 109.16

/7/

Hösel, Th.: Ermittlung von Tragbild und Flankenrichtungskorrekturen für Evolventen-Stirnräder, Berechnungen mit dem FVA-Programm “Ritzelkorrektur”, Zeitschrift Antriebstechnik 22, (1983) Nr. 12

/8/

DIN 3990: Calculation of load capacity of cylindrical gears. Part 1: Introduction and general influence factors Part 2: Calculation of pitting resistance Part 3: Calculation of tooth strength Part 4: Calculation of scuffing load capacity Beuth Verlag GmbH, Berlin, December 1987

/9/

FVA-Stirnradprogramm: Vergleich und Zusammenfassung von Zahnradberechnungen mit Hilfe von EDV-Anlagen (jeweils neuester Programmstand), FVA-Forschungsvorhaben Nr. 1, Forschungsvereinigung Antriebstechnik, Frankfurt am Main

/10/

DIN 3990: Calculation of load capacity of cylindrical gears. Application standard for industrial gears. Part 11: Detailed method; February 1989 edition Part 12: Simplified method; Draft May 1987 Beuth Verlag GmbH, Berlin

/11/

DIN 3990: Calculation of load capacity of cylindrical gears. Part 5: Endurance limits and material qualities; December 1987 Beuth Verlag GmbH, Berlin

13

138

·

Bibliography

/12/

FVA-Arbeitsblatt zum Forschungsvorhaben Nr. 8: Grundlagenversuche zur Ermittlung der richtigen Härtetiefe bei Wälz- und Biegebeanspruchung. Stand Dezember 1976, Forschungsvereinigung Antriebstechnik, Frankfurt am Main

/13/

FVA-Ritzelkorrekturprogramm: EDV-Programm zur Ermittlung der Zahnflankenkorrekturen zum Ausgleich der lastbedingten Zahnverformungen (jeweils neuester Programmstand), FVA-Forschungsvorhaben Nr. 30, Forschungsvereinigung Antriebstechnik, Frankfurt am Main

/14/

Niemann, G.: Maschinenelemente 2. Bd., Springer-Verlag Berlin, Heidelberg, New York, 1965

/15/

Theissen, J.: Vergleichskriterien für Grossgetriebe mit Leistungsverzweigung, VDI-Bericht 488 “Zahnradgetriebe 1983 - mehr Know how für morgen”, VDI-Verlag, 1983

/16/

DIN 45635: Measurement of noise emitted by machines. Part 1: Airborne noise emission; Enveloping surface method; Basic method, divided into 3 grades of accuracy; April 1984 edition Part 23: Measurement of airborne noise; Enveloping surface method; Gear transmission; July 1978 edition Beuth Verlag GmbH, Berlin

/17/

VDI-Richtlinien 2159: Emissionskennwerte technischer Schallquellen; Getriebegeräusche; Verein Deutscher Ingenieure, July 1985

/18/

DIN 740: Flexible shaft couplings. Part 2. Parameters and design principles. August 1986 edition; Beuth Verlag GmbH, Berlin

/19/

VDI-Richtlinien 2056: Beurteilungsmaßstäbe für mechanische Schwingungen von Maschinen. VDI-Handbuch Schwingungstechnik; Verein Deutscher Ingenieure; October 1964; (08.97 withdrawn without replacement)

/20/

DIN ISO 10816-1: Mechanical vibration - Evaluation of machine vibration by measurements on non-rotating parts. August 1997 edition; Beuth Verlag GmbH, Berlin

13

·

139

TRANSLATION of Technical Handbook 5th Edition 03/2009 Copyright by FLENDER AG Bocholt

140

·

The information provided in this brochure contains merely general descriptions or characteristics of performance which in actual case of use do not always apply as described or which may change as a result of further development of the products. An obligation to provide the respective characteristics shall only exist if expressly agreed in the of contract. All product designations may be trademarks or product names of Siemens AG or supplier companies whose use by third parties for their own purposes could violate the rights of the owners.

A. Friedrich Flender AG P.O. Box 1364 46393 BOCHOLT

Subject to change without prior notice Order No.: E86060-T5701-A101-A1-7600 Dispo 18500 BU 0309 3.0 Ro 148 En Printed in © Siemens AG 2009

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