Admission dependent and independent. Calculation of dependent dimensional tolerances that determine the location of the hole axes. State standard of the Russian Federation
Deviations in the location of surfaces and coordinating dimensions, as well as deviations in dimensions (diameters, widths, etc.) can appear both jointly and independently of each other. Their mutual influence is possible both during the manufacturing process and during the control process. Therefore, it is customary to consider independent and dependent tolerances for the location of surfaces and coordinating dimensions.
Independent clearance- tolerance of relative position or shape, the numerical value of which is constant and does not depend on the actual dimensions of the surfaces or profiles under consideration.
Dependent tolerance of location or shape- this is a variable tolerance, the minimum value of which is indicated in the drawing or technical requirements and which is allowed to be exceeded by an amount corresponding to the deviation of the actual size of the surface of the part from the maximum material limit (the largest maximum shaft size or the smallest maximum hole size). To indicate a dependent tolerance after it numerical value in the frame write the letter M in a circle à.
According to GOST R 50056-92, the concepts of minimum and maximum values of dependent tolerance are established.
Minimum value of dependent tolerance– the numerical value of the dependent tolerance when the considered (normalized) element and (or) base have dimensions equal to the maximum limit of the material.
The minimum dependent tolerance value can be zero. In this case, location deviations are allowed within the tolerance range of the element size. With a zero dependent location tolerance, the size tolerance is the total size and location tolerance.
Maximum dependent tolerance value– the numerical value of the dependent tolerance, when the element in question and (or) the base have dimensions equal to the minimum material limit.
Dependent tolerances are assigned only for elements (their axes or planes of symmetry) that are holes or shafts.
The following dependent shape tolerances exist:
– axis straightness tolerance cylindrical surface;
– tolerance for flatness of surface symmetry of flat elements.
Dependent mutual position tolerances:
– tolerance of perpendicularity of the axis or plane of symmetry relative to the plane or axis;
– tolerance for inclination of the axis or plane of symmetry relative to the plane or axis;
– alignment tolerance;
– symmetry tolerance;
– axis intersection tolerance;
– positional tolerance of an axis or plane of symmetry.
Dependent tolerances of coordinating dimensions:
– tolerance of the distance between the plane and the axis or plane of symmetry;
– tolerance of the distance between the axes (planes of symmetry) of two elements.
Dependent location tolerances are assigned mainly in cases where it is necessary to ensure the assembly of parts mating simultaneously on several surfaces with specified clearances or interferences. The use of dependent tolerances of shape and location reduces the cost of production and simplifies the acceptance of products.
The numerical value of the dependent tolerance can be related:
1) with the actual dimensions of the element in question;
2) with the actual dimensions of the base element;
3) with the actual dimensions of both the base and considered elements.
When indicating a dependent tolerance in drawings in accordance with GOST 2.308-79, the icon à is used.
If the dependent tolerance is related to the actual size of the element in question, the symbol is indicated after the numerical value of the tolerance.
If the dependent tolerance is related to the actual size of the base element, the symbol is indicated after letter designation bases.
If the dependent tolerance is associated with the actual size of the element in question and the dimensions of the base element, then the sign à is indicated twice after the numerical value of the tolerance and after the letter designation of the base.
Dependent tolerances are usually controlled by complex gauges, which are prototypes of mating parts. These gauges are pass-through only and guarantee non-fitting assembly of products. Complex gauges are quite complex and expensive to manufacture, so the use of dependent tolerance is advisable only in serial and mass production.
An independent tolerance for the location of the hole axes is a tolerance whose numerical value is constant for a large number of parts of the same name (for example, a batch of parts) and does not depend on the actual size (diameter) of the hole or (or perhaps “and”) on the size of the base. If there are no indications on the drawing, then the tolerance is considered independent.
The meaning of this concept comes down to the fact that with an independent tolerance during measurement, it is necessary to determine the location error in such a way that the value of the size (diameter) of the hole does not affect the value of the location deviation.
In the previous figures, the location tolerances are independent, i.e. center-to-center distances must be maintained within tolerances specified by positional deviations, or by maximum deviations and do not depend on what the actual diameters of the holes are (but, of course, the holes, in turn, must be made within their permissible dimensions).
Dependent location tolerance - a tolerance indicated on the drawing or in other technical documents in the form of a minimum value that can be exceeded by a value depending on the deviation of the actual size of the element (hole) and/or base in question from the maximum material limit, i.e. for a hole from the smallest limit hole size.
The dependent location tolerance is highlighted with the symbol M,
standing next to the location tolerance and/or the base.
The full value of the dependent location tolerance is determined by the formula:
,
where is the minimum tolerance value indicated in the drawing (the part of the dependent tolerance that is constant for all parts);
|
|
– additional tolerance value depending on the actual dimensions of the holes.
If the hole is made with maximum size(diameter), then will be maximum and determined as
, ,
where is the hole tolerance.
Interpreting the above, it can be argued that the minimum guaranteed clearance for the passage of a fastener can be increased (which occurs when the actual dimensions of the mating elements deviate from the passage limits), and a correspondingly increased position deviation allowed by the dependent tolerance becomes acceptable.
Let us explain the above using specific examples.
In Fig. 7, and the positional tolerance of the location is independent (there are no indications on the drawing). This means that the center of the ø10H12 hole must be within the circle with a diameter of 0.1mm and not go beyond, regardless of what the actual diameter of the hole is.
In Fig. 7, b the positional tolerance is dependent (this is indicated by the symbol M next to the location tolerance). This means that the minimum position tolerance value is 0.1 mm (for hole diameter ).
As the hole diameter increases, the location tolerance can be increased (due to the resulting gap in the connection). The maximum position tolerance value can be when the hole is made at the upper limit size, i.e. when = 10.15 mm. Eventually
,
and then, i.e. the center of the hole ø 10H12 can be in a circle with a diameter of 0.25 mm.
5.Numerical tolerance values
hole locations
For connection (Fig. 1, a, type A), both plates 1 and 2 to be connected are provided with through holes for the passage of fasteners. For connection type B - through holes only in the 1st plate. The diametrical gap between the fastener and the hole in the plate must ensure free passage of the bolt (rivet) into the hole to ensure assembly. The guarantee can be achieved when the actual hole size is obtained close to the minimum size limit holes, and the shaft (bolt, rivets) - to the maximum limit size (usually, where d is the nominal size of the bolt). The difference between the sizes and is the minimum gap, which is guaranteed, since with a larger gap, the better the assembly will be ensured. The minimum diametrical clearance is taken as the positional tolerance for the location of the holes, and:
– for type A connections: ;
– for type B connections: (gap in only one plate).
Here T is the main positional tolerance in diametrical terms (twice the maximum displacement from the nominal location according to GOST 14140-81).
For standard fasteners, there are developed tables with the diameters of through holes for them and the corresponding smallest (guaranteed) clearances (GOST 11284-75). One of these tables is given in Appendix 1.
2. When setting dimensions, using a “ladder” with reference to the assembly base:
For type A connections – 
;
For type B connections – 
.
In Appendix 2 “Recalculation of positional tolerances for maximum deviations of dimensions coordinating the axes of holes. Rectangular coordinate system” according to GOST 14140-81, numerical values are given maximum deviations depending on the specified positional tolerance for some dimensioning schemes.
Appendix 3 provides examples of converting positional tolerances into maximum deviations for some sizing schemes with tolerance symbols on the drawings.
Dependent tolerance according to GOST R 50056-92 is a variable tolerance of shape, location or coordinating size, the minimum value of which is indicated on the drawing or in the technical requirements and which can be exceeded by an amount corresponding to the deviation of the actual size of the considered and (or) basic element of the part from maximum limit material. According to GOST 25346-89, the maximum material limit is a term referring to that of the maximum dimensions to which the largest volume of material corresponds, i.e. the largest maximum shaft size d max or the smallest maximum hole size D min.
The following permissions can be assigned to dependents:
- shape tolerances:
- - tolerance for straightness of the axis of the cylindrical surface;
- - tolerance for flatness of surface symmetry of flat elements;
- location tolerances (orientation and location):
- - tolerance of perpendicularity of the axis or plane of symmetry relative to the plane or axis;
- - tolerance for inclination of the axis or plane of symmetry relative to the plane or axis;
- - alignment tolerance;
- - symmetry tolerance;
- - axis intersection tolerance;
- - positional tolerance of the axis or plane of symmetry;
- tolerances of coordinating dimensions:
- - tolerance of the distance between the plane and the axis or plane of symmetry of the element;
- - tolerance of the distance between the axes or planes of symmetry of two elements.
Full dependent tolerance value:
Where T t in - minimum dependent tolerance value specified
in the drawing, mm;
Gdop - permissible excess of the minimum value of the dependent tolerance, mm.
It is recommended to assign dependent tolerances, as a rule, to those elements of parts for which requirements are imposed assembly in connections with guaranteed clearance. Tolerance T t[P are calculated based on the smallest connection gap, and the permissible excess of the minimum value of the dependent tolerance is determined as follows:
For shaft 
For hole
Where d a and /) d - actual dimensions of the shaft and hole, respectively, mm.
The value of G add can vary from zero to the maximum value. d
If the shaft has a valid size dmin, and hole D max , then
For shaft
For hole
Where TdwTD- size tolerance of the shaft and hole, respectively, mm.
In this case, the dependent tolerance has a maximum value:
For shaft
For hole
If the dependent tolerance is related to the actual dimensions of the considered and base elements, then
where Gd 0P.r and Gd 0P.b are the permissible excesses of the minimum value of the dependent tolerance, depending on the actual dimensions of the considered and basic elements of the part, respectively, mm.
Examples of the use of dependent tolerances include:
- - positional tolerance for the location of through holes for fasteners (Fig. 2.17, A);
- - alignment tolerances of stepped bushings and shafts (see Fig. 2.17, b, V), assembled with a gap;
- - tolerance for the symmetry of the location of grooves, for example, keyways (see Fig. 2.17, d);
- - tolerance for perpendicularity of the axes of holes and end surfaces of body parts for glasses, plugs, lids.
Rice. 2.17.A - positional tolerance of holes for fasteners; b, c - coaxiality of the surfaces of the stepped bushing and shaft; G - symmetry of the keyway relative to the shaft axis
Dependent location tolerances are more economical and beneficial for production than independent ones, since they expand the tolerance value and allow the use of less precise and labor-intensive technologies for manufacturing parts, as well as reducing losses from defects. Control of parts with dependent location tolerances is carried out, as a rule, using complex pass-through gauges.
A dependent tolerance of shape or location is indicated in the drawing by a sign, which is placed in accordance with GOST 2.308-2011:
- - after the numerical value of the tolerance (Fig. 2.17, A), if the dependent tolerance is related to the actual dimensions of the element in question;
- - after the letter designation of the base or without the letter designation in the third field of the frame (see Fig. 2.17, b), if the dependent tolerance is related to the actual dimensions of the base element;
- - after the numerical value of the tolerance and the letter designation of the base (see Fig. 2.17, G) or without a letter designation (see.
rice. 2.17, V), if the dependent tolerance is related to the actual dimensions of the considered and base elements.
On January 1, 2011, GOST R 53090-2008 (ISO 2692:2006) came into force. This GOST partially duplicates GOST R 50056-92, in force since January 1, 1994, in terms of standardization and indication on drawings of maximum material requirements (MMR - maximum material reguirement) in cases where it is necessary to ensure the assembly of parts in connections with a guaranteed gap. Minimum material requirements (LMR - least material reguirement), due to the need to limit minimum thickness the walls of the parts had not previously been presented.
The MMR and LMR requirements combine the constraints of dimensional tolerance and geometric tolerance into one comprehensive requirement that more closely matches the intended purpose of the parts. This complex requirement allows, without compromising the performance of the part’s functions, to increase the geometric tolerance of the normalized (considered) part element if the actual size of the element does not reach the limit value determined by the established size tolerance.
The maximum material requirement (as well as the dependent tolerance according to GOST R 50056-92) is indicated on the drawings with a sign, and the minimum material requirement is indicated with a sign (L), placed in a frame to indicate the geometric tolerance of the normalized element after the numerical value of this tolerance and/or symbol bases.
Calculation of geometric tolerance values T m, ensuring the maximum material requirement can be performed similarly to the calculation of dependent tolerances (see formulas 2.10-2.15).
Designating similarly to dependent tolerances T m, geometric tolerances, which are subject to minimum material requirements - T L , can be written:
Where T m in - the minimum value of the geometric tolerance specified
in the drawing, mm;
Tdop - permissible excess of the minimum value of the geometric tolerance, mm.
T add values are determined as follows:
For shaft
For hole
dmin, and the hole Dmax, That
If the shaft has a valid size d max , and the hole Z) min , then
For shaft
For hole
In this case, the geometric tolerance has a maximum value:
For shaft 
For hole
If the geometric tolerance is related to the actual dimensions of the normalized and base elements, then the value of G additional is found from dependence (2.15).
Examples of the application of maximum material requirements are examples of assigning dependent tolerances according to GOST R 50056-92 in Fig. 2.17. An example of the application of the minimum material requirement is shown in Fig. 2.18, A.
Both the maximum material requirements and the minimum material requirements can be supplemented by an interaction requirement (RPR - reciprocity requirement), which allows increasing the tolerance of the size of a part element if the actual geometric deviation (deviation of shape, orientation or location) of the normalized element does not fully use the restrictions imposed by the requirements MMR or LMR. An example of the application of minimum material requirements and the interaction of size 05 tolerance O_ o, oz9 and concentricity tolerance are shown in Fig. 2.18, b, and an example of applying the requirement for maximum material and the interaction of size 16_о,т and perpendicularity tolerance is in Fig. 2.18, V.
Example 2.2. A dependent tolerance for hole alignment 016 +OD8 is set relative to outer surface 04О_о,25 bushing shown in Fig. 2.19.
From the symbol it is clear that the alignment tolerance depends on the actual size of the element, the axis of which is the base axis, i.e. surfaces 04О_ о 25.
Rice. 2.18.A- minimum material; b - minimum material and interaction; V- maximum material and interaction
Rice. 2.19.
The minimum value of the alignment tolerance indicated in the drawing (7pcs = 0.1 mm) corresponds to the maximum limit of the outer surface material, in this case the size d a = d max = 40 mm, i.e. at d a = d max = 40 mm 
If the outer surface has an actual size d a = dmin, The alignment tolerance can be increased:
Intermediate size values d a and their corresponding tolerance values T m are given in table. 2.9, and in Fig. Figure 2.20 shows a graph of the dependence of the alignment tolerance on the actual size of the outer surface of the bushing.
Rice. 2.20.
Values of dependent alignment tolerance, mm(see Fig. 2.20)
So I look at more or less accessible CAD systems such as Kompas, T-Flex, SolidWorks, SolidEdge and, at worst, Inventor, and do not find the basic functionality needed by designers of foundry equipment, mostly for casting metals rather than plastics. Well, this is where these programs have such basic capabilities as: 1. The ability to display transition lines in the drawing conditionally in accordance with clause 9.5 of GOST 2.305-2008 "ESKD. Images - views, sections, sections."
2. The ability to draw up drawings and transfer data to the specification for parts obtained from blanks in accordance with clause 1.3 "Drawings of products with additional processing or rework" according to GOST 2.109-73 ESKD. "Basic requirements for drawings". In SW this is implemented using SWPlus macros, but in other programs how?
3. The ability to automatically obtain views and sections in the casting drawing with thin lines of machined surfaces of the part in accordance with clause 3 of GOST 3.1125-88 - "ESTD. Rules for the graphic execution of elements of foundry molds and castings." In SW2020 this is semi-implemented with an alternate position view (views can show these thin lines, section views cannot). How about this in other programs?
4. The ability to set the radius size to an inclined twist, that is, to an ellipse, which are often present on parts with slopes (castings, forgings). I know that this can be done in SW. How about this in other programs?
5. The ability to specify on a 3D model of a metal part produced by casting with subsequent machining and on 3D casting models the casting accuracy in accordance with GOST R 53464-2009 - "Castings from metals and alloys. Tolerances of dimensions, weights and allowances for machining". And accordingly, automatically obtain tolerances for the dimensions of cast surfaces. This is not in any program. Do the developers dislike foundry workers?
In addition, it would be nice to know the difference between the array in solid and other cads.
In the same tflex, the array is created quickly and slows down less, but only there the array is a single object. It will not be possible to hide/suppress one of the array components or select a different configuration for it, as in solid. And since the tflexers are hanging out in the solid branch, I’ll cry to them, maybe they’ll tell me something. I need to save drawings in dxf. But tflex, as it turned out, does not convert drawings to a 1:1 scale before exporting and makes polylines or segments with arcs from splines. With splines, I understand that everything is clear, but with scale? Do not suggest scaling in AutoCAD, the age is not the same) You can read about working with arrays (in English) - https://forum.solidworks.com/thread/201949 Which in free and abbreviated translation) means - in most cases it is better to do several arrays instead of one.
It is necessary to produce 73.2 thousand small pins of two
different sizes
: 37 mm and 32 mm at a price of 10 rubles/piece from your material.
Material AISI 431 or 14Х17н2
A productivity of 2-8 thousand hairpins per week is required. PULSAR23_Contact screws_23.07.19.rar P23_Contact screw_37_(2 sheets)_07.23.19.pdf P23_Contact screw_32_(2 sheets).pdf
| I uploaded a cloud to my email address https://cloud.mail.ru/public/heic/ZRvyFHBXn I’ll try to do this, I’m interested in the reason why this assembly does not combine into one of the 3, but 2 thirds easily grew together, but I can’t do the last one insert...or rather I can insert, but it doesn’t work to splice the last one | Rows of dependent tolerances for the location of the axes of holes for fasteners are established by GOST 14140-81. The standard establishes a series of numbers (in accordance with the RalO series), from which the maximum displacement values Δ of the hole axes from the nominal position are selected, and then, according to the formula T = 2D, they are recalculated into the positional tolerance of the axis in diametrical expression T, as indicated in the top row of numbers in table 36. This table shows the values corresponding to the series of dependent tolerances for the location of the axes, the maximum deviations for six typical cases of the location of the axes of holes in the rectangular coordinate system. This table is compiled on the basis of OST 14140-81 data for the commonly used rectangular coordinate system and for the T values of positional tolerances of hole axes that are often found in examples and problems. | Positional tolerance of the axis in diametrical terms T, mm | ||||||||||||||||||||
| 0,2 | 0,25 | 0,3 | 0,4 | 0,5 | 0,6 | 0,8 | 1 | 1,2 | 1,6 | 2 | ||||||||||||
| One hole coordinated relative to the plane (during assembly, the reference planes of the parts to be joined are aligned) | Limit size deviations between the hole axis and the plane | 0,10 | 0,12 | 0,16 | 0,20 | 0,25 | 0,30 | 0,40 | 0,50 | 0,60 | 0,80 | 1,0 | ||||||||||
Continuation of table 36
| Two holes coordinated relative to each other | Maximum size deviations between the axes of two holes | 0,20 | 0,25 | 0,30 | 0,40 | 0,50 | 0,60 | 0,80 | 1,0 | 1,2 | 1,6 | 2,0 | ||||||
| Several holes arranged in one row | Maximum size deviations between the axes of any two holes | 0,14 | 0,16 | 0,22 | 0,28 | 0,35 | 0,40 | 0,55 | 0,70 | 0,80 | 1,1 | 1,4 | ||||||
| Limit deviations of hole axes from the general plane | 0,07 | 0,08 | 0,11 | 0,14 | 0,18 | 0,20 | 0,28 | 0,35 | 0,40 | 0,55 | 0,70 | |||||||
| I uploaded a cloud to my email address https://cloud.mail.ru/public/heic/ZRvyFHBXn I’ll try to do this, I’m interested in the reason why this assembly does not combine into one of the 3, but 2 thirds easily grew together, but I can’t do the last one insert...or rather I can insert, but it doesn’t work to splice the last one | Rows of dependent tolerances for the location of the axes of holes for fasteners are established by GOST 14140-81. The standard establishes a series of numbers (in accordance with the RalO series), from which the maximum displacement values Δ of the hole axes from the nominal position are selected, and then, according to the formula T = 2D, they are recalculated into the positional tolerance of the axis in diametrical expression T, as indicated in the top row of numbers in table 36. This table shows the values corresponding to the series of dependent tolerances for the location of the axes, the maximum deviations for six typical cases of the location of the axes of holes in the rectangular coordinate system. This table is compiled on the basis of OST 14140-81 data for the commonly used rectangular coordinate system and for the T values of positional tolerances of hole axes that are often found in examples and problems. | Normalized deviations of dimensions coordinating the axes of holes | Limit displacement of the axis from the nominal location (i), mm | |||||||||||||||
| 0,10 | 0,12 | 0,16 | 0,20 | 0,24 | 0,30 | 0,40 | 0,50 | 0,60 | 0,80 | 1,00 | ||||||||
| Maximum deviations of dimensions coordinating the axes of holes (±), mm | ||||||||||||||||||
| Three or four holes arranged in two rows | 0,14 | 0,16 | 0,22 | 0,28 | 0,35 | 0,40 | 0,55 | 0,70 | 0,80 | 1,1 | 1,4 | |||||||
| 0,20 | 0,25 | 0,30 | 0,40 | 0,50 | 0,60 | 0,80 | 1,0 | 1,2 | 1,6 | 2,0 | ||||||||
| One hole coordinated with respect to two mutually perpendicular planes (during assembly, the base planes of the parts being connected are aligned) | Maximum deviations of sizes L 1 and L 2 | 0,07 | 0,08 | 0,11 | 0,14 | 0,18 | 0,20 | 0,28 | 0,35 | 0,40 | 0,55 | 0,70 | ||||||
| Holes coordinated relative to each other and arranged in several rows | Maximum deviations of dimensions L 1; L2; L 3; L 4 | 0,07 | 0,08 | 0,11 | 0,14 | 0,18 | 0,20 | 0,28 | 0,35 | 0,40 | 0,55 | 0,70 | ||||||
| Limit deviations of dimensions diagonally between the axes of any two holes | 0,20 | 0,25 | 0,30 | 0,40 | 0,50 | 0,60 | 0,80 | 1,0 | 1,2 | 1,6 | 2,0 | |||||||
Note: If, instead of the size deviation between the axes of any two holes, size deviations from each hole to one base hole or base plane (i.e. dimensions L 1; L 2 etc.), then the maximum deviation should be halved.
Let's look at examples of using this table.
Example. The two parts are fastened together with five bolts arranged in one row. The nominal dimensions of the center distances are 50 mm. Smallest sizes The diameters of the bolt holes are 20.5 mm. The largest outer diameters of the bolts are 20 mm. Let's consider three options (a, b, c) for setting dimensions in the drawing, shown in Fig. 74.
Solution:
a) a type A connection is given, in which the bolts pass with clearance through the holes in the first and second parts to be connected. The positional deviation for connection type A is Δ=0.5·S min. If the entire smallest gap is used to compensate for the offset, in this example:
S min =20.5-20=0.5 (mm).
The positional tolerance of the hole axes of a given connection can be determined by the formula:
T=k·S min
at k=1 for a connection that does not require adjustment T=1·0.5=0.5 (mm).
From Table 36 we find that E = 0.5 mm is a value included in the standard series and therefore does not require rounding.
The method for setting the positional tolerance of the axes in the drawing is shown in Fig. 74, a. Only the nominal dimensions of the center distances are indicated in the frames. Location tolerance specified conventional sign, its value and symbol (letter M), indicating that it is dependent, are inscribed in a tolerance frame divided into three parts;
b) when normalizing the tolerance of interaxial distances, according to the figure, in which the arrangement of the holes is similar to the example under consideration, we find that the maximum deviation of the size between the axes of any two holes is equal to +0.35 mm, and the maximum deviation of the axes of the holes from common plane±0.18 mm.
Fig.74. Schemes for setting interaxial dimensions
With the indicated placement of interaxial dimensions, as shown in Fig. 74, b, they can be considered as links in a dimensional chain, in which the closing dimension is a size of 200 mm with maximum deviations of ±0.35 mm and a tolerance of T = 0.70 mm. Thus, finding the tolerances (maximum deviations) of the four center distances is reduced to solving the direct problem of a five-link dimensional chain, in which the nominal dimensions of the links and the tolerance of the closing link are known. The problem is solved by the equal tolerance method, since all component links are equal to 50 mm.
The tolerance of each of the interaxial dimensions (links of the dimensional chain) is equal to 0.70/4 = 0.175 mm, and the permissible deviations are approximately ±0.09 mm.
The corresponding dimensioning (in a chain) is shown in Fig. 74, b. The 200 mm size is marked with an asterisk (*), since its error depends on the actual errors of the center distances of 50 mm;
c) in the case when deviations in dimensions coordinating the centers of holes need to be assigned relative to the base (in this example, the base can be the axis of the first hole or the end of the part), the calculation should be carried out based on the fact that the interaxial distances are the closing dimensions in three-link dimensional chains. For example, in a chain consisting of sizes 50, 100 and 50 mm, or in a chain consisting of sizes 100, 150, 50 mm, etc.
The permissible deviations of the distance between the centers of each pair of holes are taken from the table. 36 and equal to ±0.35 mm. Since their tolerances for closing center distances are equal to 0.70 mm, and the tolerances for sizes 50, 100, 150, 200 mm are equal to 0.70/2 = 0.35 mm, that is, the permissible deviations of these dimensions are equal to ±0.18 mm.
The corresponding arrangement of interaxial dimensions in the drawing (alignment with a ladder) is shown in Fig. 74, c.
Analyzing the accuracy of setting interaxial dimensions in Fig. 74, one can be convinced that when setting dimensions from one base, the tolerances on dimensions coordinating the centers of holes can be twice as large as when setting successive interaxial dimensions.
CONCLUSION
The presented material discusses several important issues of interchangeability, which are fundamental in the study of the discipline “Metrology, standardization and certification”:
The ESDP system for smooth cylindrical matings, which is uniform for all branches of mechanical engineering;
Standardization of accuracy of standard connections;
Dimensional analysis;
Calculation of smooth limiting calibers,
These issues are an integral part of the practical activities of designers and technologists.
The published material is a teaching aid and in no case can it be considered as a textbook containing comprehensive information on the above sections of interchangeability. This is evidenced by the peculiarity of the presentation of the material - in the form of questions and answers, concepts and definitions. Small excerpts from the tables of standards explain the specifics of their construction. Many illustrations throughout the chapters and specific numerical examples allow students to test their ability to use reference tables.
An important point associated with the publication of this manual is the lack of a sufficient number of reference books and regulatory documents, necessary for students of design and technology faculties when performing course work provided curriculum given discipline, and
as well as course and diploma projects.
IN textbook The method of calculations associated with dimensional analysis involves performing them “manually”, since performing this work on a computer requires special training. The manual does not include issues related to the interchangeability of corner and conical connections, gears and gears. Due to the characteristics of these connections, their interchangeability, tolerances and fits must be considered with methods and means of their measurement and control, and this is possible when publishing a new manual.
| TABLE OF CONTENTS | |
| PREFACE................................................... ........................................................ .................... | |
| 1. INTERCHANGEABILITY AND ITS TYPES.................................................... ............................ | |
| 2. CONCEPT OF DIMENSIONS, TOLERANCES AND DEVIATIONS................................................... | |
| 3. SIZE TOLERANCE. GRAPHIC REPRESENTATION OF TOLERANCES.................................... | |
| 4. CONCEPT OF 0 LANDINGS. TYPES OF LANDING................................................... ................ | |
| 5. PRINCIPLES OF CONSTRUCTION OF LANDINGS. FIT IN THE HOLE AND SHAFT SYSTEM.................................................... ........................................................ ....................................... | |
| 6. UNIFIED SYSTEM OF ADMISSIONS AND LANDINGS (USDP), ITS STRUCTURE................................................... ........................................................ ...................................... | |
| 7. FITINGS IN THE ESDP SYSTEM FOR SMOOTH CYLINDRICAL JOINTS………………….................................. ........................................................ ......... | |
| SELF-TEST QUESTIONS.................................................................. ...................................... | |
| 8. ACCURACY OF PARTS FORM.................................................... ......................................... | |
| 9. INTERCHANGEABILITY OF PIN CONNECTIONS………………………. | |
| 9.1. PURPOSE AND TYPES OF PIN CONNECTIONS.................................................... | |
| 9.2. PIN FORMS................................................... ........................................................ ...... | |
| 9.3. INSTALLING PINS................................................... ............................................... | |
| 10. INTERCHANGEABILITY OF KEYED CONNECTIONS.................................................... | |
| 10.1. KEYED CONNECTIONS................................................................... ................................... | |
| 10.2. TOLERANCES AND FITTS OF KEYED CONNECTIONS.................................................... | |
| 10.3. TOLERANCES AND FITMENTS OF SHAFT WITH HOLE.................................................... ....... | |
| 11. INTERCHANGEABILITY OF SPLINED CONNECTIONS................................................... | |
| 11.1. GENERAL INFORMATION................................................... ........................................................ .... | |
| 11.2. SYSTEM OF TOLERANCES AND FITTS OF SPLINE CONNECTION………… | |
| 11.3. DESIGNATION ON THE DRAWINGS OF SPLINED CONNECTIONS AND SPLINED PARTS......................................................... ........................................................ ............ | |
| 12. TOLERANCES AND FITTS OF ROLLING BEARINGS.................................................... . | |
| 12.1. GENERAL INFORMATION................................................... ........................................................ ... | |
| 12.2. TOLERANCES AND FITS OF ROLLING BEARINGS ACCORDING TO CONNECTING DIMENSIONS................................................................... .................................... | |
| 12.3. SELECTION OF ROLLING BEARING LITCHINGS.................................................................... ...... | |
| 12.4. DESIGNATION OF BEARING LANDINGS IN THE DRAWINGS.................... | |
| 13. INTERCHANGEABILITY OF THREADED CONNECTION PARTS................................... | |
| 13.1. GENERAL PROVISIONS................................................... ................................................... | |
| 13.2. METRIC THREAD AND ITS PARAMETERS.................................................... ............. | |
| 13.3. GENERAL PRINCIPLES FOR ENSURING THE INTERCHANGEABILITY OF CYLINDRICAL THREADS................................................................... ........................................................ ... | |
| 13.4. FEATURES OF TOLERANCES AND FITTS OF METRIC THREADS………….. | |
| 14 ROUGHNESS AND WAVY SURFACES.................................................... | |
| 14.1. GENERAL PROVISIONS................................................... ................................................... | |
| 14.2. STANDARDING THE ROUGHNESS OF SURFACES.................................................. | |
| 14.3. SELECTION OF ROUGHNESS PARAMETERS.................................................... ............. | |
| 14.4. SURFACE ROUGHNESS DESIGNATION.................................................... | |
| 14.5. SURFACE WAVENESS AND PARAMETERS FOR ITS NORMALIZATION................................................................. ........................................................ .................... | |
| 15. SMOOTH CALIBERS AND THEIR TOLERANCES.................................................... ............................ | |
| 15.1. CLASSIFICATION OF SMOOTH CALIBERS.................................................... ............. | |
| 15.2. TOLERANCES OF SMOOTH CALIBERS.................................................... ............................... | |
| 16. SELECTION OF UNIVERSAL MEASUREMENT TOOLS FOR ASSESSING LINEAR DIMENSIONS................................................... ........................................................ ............. | |
| 16.1. GENERAL INFORMATION................................................... ........................................................ .... | |
| 16.2. MAXIMUM MEASUREMENT ERROR AND ITS COMPONENTS........... | |
| 17. INTERCHANGEABILITY ACCORDING TO DIMENSIONS INCLUDED IN THE DIMENSION CHAINS................................................... ........................................................ ............................................... | |
| 17.1. BASIC CONCEPTS, TERMS, DEFINITIONS AND NOTATIONS…… | |
| 17.2. CALCULATIONS OF DIMENSION TOLERANCES INCLUDED IN DIMENSION CHAINS................................................................... ........................................................ ........................................... | |
| 18. CALCULATION OF DIMENSIONAL CHAINS DETERMINING TOLERANCES FOR DISTANCES BETWEEN HOLES..................................................... .................................... | |
| 18.1. GENERAL PROVISIONS................................................... ........................................................ | |
| 18.2. TOLERANCES FOR LOCATION OF HOLE AXES FOR FASTENING PARTS................................................................... ........................................................ .................................... | |
| 18.3. CALCULATION OF DEPENDENT DIMENSIONAL TOLERANCES DETERMINING THE LOCATION OF THE HOLE AXES......................................................... .................................... | |
| CONCLUSION................................................. ........................................................ ...................... |
Sergey Petrovich Shatilo
Nikolai Nikolaevich Prokhorov
Vladislav Valikovich Chorny
Sergey Vitalievich Kucherov
Galina Fedorovna Babyuk
