Non-uniform-pitch threaded connection pair

Abstract

A non-uniform-pitch threaded connection pair, comprising an internal thread and an external thread, the pitch of the internal thread and the pitch of the external thread being constant, the pitch of the internal thread being larger than that of the external thread, and the difference between the pitch of the internal thread and the pitch of the external thread being (I), where P is the pitch of the external thread, K is a comprehensive reference coefficient, k is a load coefficient, and kc is a material coefficient. The pitch difference of the connection pair can make the stress borne by thread teeth more uniform.

Description

A non-equidistant thread connection pair Technical Field

[0001] The invention relates to a non-equidistant thread connection pair, belonging to the field of non-standard fasteners. Background Art

[0002] Bolted connection structure (including bolts and nuts) is a relatively common connection method, widely used in the fields of machinery and construction. At present, the pitch of the internal and external threads of the traditional threaded connection pair is equal and is a constant value P. When the threaded connection pair is subjected to load, the external thread is subjected to tension and elongation, and the internal thread is subjected to pressure and compression. The elongation of the external thread near the support surface (taking bolts and nuts as an example, the support surface is the end face of the nut used to tighten the connected parts, that is, the support surface is the end face at the end position of the screwing area of ​​the internal and external threads, and the other end face is the end face at the starting position) becomes larger, and the elongation of the external thread away from the support surface is small. This causes the axial load of the threaded connection pair to be mainly borne by the first three threads close to the support surface, causing significant stress concentration in the first three circles of threads.

[0003] Currently, two design approaches have been proposed to address the stress concentration problem in the first three threads near the support surface: improving the load uniformity of each thread turn and improving the stress uniformity of each thread turn. In both of these technical solutions, at least one of the internal and external threads is a variable-pitch thread, such as the threaded connection with variable clearance disclosed in Chinese invention patent application publication number CN101796312A. The use of a variable-pitch thread significantly increases the difficulty of thread machining, hindering mass production and resulting in low machining efficiency.

[0004] To address the above-mentioned issues, another unequal-pitch solution is available in the prior art, where the pitch of the internal and external threads is constant, but the pitches of the internal and external threads are unequal. For example, Chinese invention patent application publication number CN106438657A discloses a pitch interference fit thread, a threaded connector, and a thread correction tool, wherein the pitches of the internal and external threads are unequal, with the pitch of one thread being 95% to 99% of the pitch of the other. The pitch difference between the internal and external threads, ΔP, equals (0.01 to 0.05)P, resulting in axial interference fit within the thread engagement length. Furthermore, after the threads are fully engaged, the interference fit is greatest at the ends of the engagement length and gradually decreases toward the center. Since the internal and external threads in this technical solution are interference fit, the difficulty of assembling the internal and external threads is increased; in addition, after the internal and external threads are assembled, the two ends will be in contact and the middle will be suspended. When the internal and external threads are not loaded, there is already a large interaction force between the internal and external threads. When the internal and external threads are loaded, the bearing capacity of the thread teeth at the support surface decreases, and the bearing capacity of the thread teeth away from the support surface increases, which will inevitably lead to stress concentration in the few threads away from the support surface.

[0005] In addition, existing non-uniform pitch threads do not consider factors such as load and material. However, when using non-uniform pitch threads, it is necessary to achieve high load uniformity for the internal and external threads based on the load conditions and the performance of the internal and external thread processing materials.

[0006] Summary of the Invention

[0007] The purpose of the present invention is to provide a non-equidistant thread connection pair to solve the problem that the pitch difference between the internal and external threads in the existing thread connection pair exceeds a certain value and factors such as load and material are not considered, resulting in interference fit between the internal and external threads and thus causing stress concentration.

[0008] To achieve the above-mentioned purpose, the non-equidistant thread connection pair in the present invention adopts the following technical solutions:

[0009] A non-equidistant thread connection pair, comprising an internal thread and an external thread, wherein the pitch of the internal and external threads is constant and the pitch of the internal thread is greater than the pitch of the external thread. Where: P is the pitch of the external thread; K is the comprehensive reference coefficient, and 0.8≤K≤4.5; k is the load coefficient, which is the ratio of the axial load N borne by the external thread to the product of the yield strength σ of the material and the stress cross-sectional area A1 of the external thread, that is, k c is the material coefficient, which is the material's yield strength σ and the material's elastic modulus E w The ratio of

[0010] The beneficial effect of the above technical solution is that: the present invention proposes an improved non-equidistant thread connection pair, the main improvement of which is the pitch difference between the internal and external threads. Compared with the existing technology, the present invention reduces the pitch difference between the internal and external threads, and solves the internal and external thread pitch difference based on the consideration of the material properties of the threaded connection pair and the load it bears. It has been verified that adopting the pitch difference within the above range can make the stress borne by the thread teeth more uniform.

[0011] Furthermore, the pitch difference between the internal and external threads is ΔP<0.01P.

[0012] The beneficial effect of the above technical solution is that it provides an upper limit value of the pitch difference, which facilitates the design and processing of the thread.

[0013] Furthermore, the pitch difference between the internal and external threads ΔP>0.003P.

[0014] The beneficial effect of the above technical solution is that it provides a lower limit value of the pitch difference, which facilitates the design and processing of the thread.

[0015] Furthermore, the pitch difference between the internal and external threads Where l is the length difference between the internal and external threads, n' is an arbitrary number of turns starting from the starting position of the internal and external thread engagement area, 0≤n'≤n, and n is the total number of turns of the internal and external thread engagement area;

[0016] The length difference l between the internal and external threads is the total elongation l of the external thread w Total compression with internal thread l n The sum of Where 0<k1<1, the external thread in the screwing area of ​​the internal and external threads is simplified into an equivalent load-bearing cylinder, f1(n') is the resultant force of the load on the small external cylindrical surface at any n' turns of the equivalent load-bearing cylinder, and the resultant force of the load on all small external cylindrical surfaces is equal to the axial load N;

[0017] Where k2>1, E n is the elastic modulus of the internal thread material, the internal thread in the screwing area of ​​the internal and external threads is simplified into an equivalent hollow cylinder, A2 is the cross-sectional area of ​​the equivalent hollow cylinder, f2(n') is the resultant force of the load on the small internal cylindrical surface at any n' number of turns of the equivalent hollow cylinder, the resultant force of all the loads on the small internal cylindrical surface and the axial load F N Equal, F N Equal to N and opposite in direction;

[0018] therefore Let E w A1 and E n The ratio of A2 is k3, and let k1+k2k3=K, and the formula for the length difference l between the internal and external threads is simplified to:

[0019] Further order but Further conclusion When n'=n, the length difference l between the inner and outer threads is: Furthermore, the pitch difference between the internal and external threads

[0020] The beneficial effect of the above technical solution is that it provides a specific derivation process of the pitch difference ΔP between the internal and external threads, which facilitates the design and processing of the threads.

[0021] Furthermore, the tooth width of the external thread at the mid-diameter is defined as S w The tooth width of the internal thread at the middle diameter of the external thread is S n When the internal and external threads are used together, the total axial clearance δ=PS w -S n , δ is greater than or equal to the length difference l between the internal and external threads, that is, δ ≥ l.

[0022] The beneficial effect of the above technical solution is to avoid excessive difference in pitch between the internal and external threads, which would lead to too small a fitting clearance and then an interference fit, thereby reducing assembly resistance and improving assemblability.

[0023] Furthermore, the tooth width of the external thread at the mid-diameter is defined as S w , notch C at the middle diameter d2 of the external thread n The width is greater than the tooth width S of the internal thread at the middle diameter d2 of the external thread n , then PS w >S n .

[0024] The beneficial effect of the above technical solution is that it prevents the internal thread tooth width from being too large, which would cause interference fit between the internal and external threads, thereby reducing assembly resistance and improving assemblability. BRIEF DESCRIPTION OF THE DRAWINGS

[0025] FIG1 is a tooth profile diagram of the external thread of the non-equidistant thread connection pair in the present invention;

[0026] FIG2 is a tooth profile diagram of the internal thread of the non-equidistant thread connection pair of the present invention;

[0027] FIG3 is a schematic diagram of the starting position and the ending position of the non-equidistant threaded connection pair during loading in the present invention;

[0028] FIG4 is a diagram showing the engagement state of the internal and external threads of the non-equidistant thread connection pair according to the present invention;

[0029] FIG5 is a schematic diagram showing the principle of preventing tooth biting of the internal and external threads according to the present invention;

[0030] FIG6 is a schematic diagram of the external thread involved in the present invention simplified into an equivalent load-bearing cylinder;

[0031] FIG7 is a schematic diagram of the average elongation of the tiny cylinder and the elongation at the cylindrical surface in FIG6 ;

[0032] FIG8 is a schematic diagram of the internal thread involved in the present invention simplified into an equivalent force-bearing hollow cylinder;

[0033] FIG9 is a schematic diagram showing the average compression amount and the compression amount at the inner cylindrical surface of the tiny hollow cylinder in FIG8 ;

[0034] FIG10 is a schematic diagram of the force applied to the external thread in the screwing area between the internal and external threads of the present invention;

[0035] Figure 11 shows three common variations of the axial force F(n') acting on the external thread.

[0036] Figure 12(a) is the stress cloud corresponding to curve a in Figure 11;

[0037] Figure 12(b) is the stress cloud corresponding to curve b in Figure 11;

[0038] Figure 12(c) is the stress cloud corresponding to curve c in Figure 11;

[0039] FIG13( a ) is a stress cloud diagram of the internal and external threads of the present application when the threads are not loaded;

[0040] Figure 13(b) is a stress nephogram of the internal and external threads when thread 1 is not loaded;

[0041] Figure 13(c) is a stress nephogram of the internal and external threads when thread 2 is unloaded;

[0042] Figure 13(d) is a stress nephogram of the internal and external threads when thread 3 is unloaded;

[0043] Figure 14(a) is a stress cloud diagram of the internal and external threads when the threads are loaded;

[0044] Figure 14(b) is a stress nephogram of the internal and external threads when thread 1 is loaded;

[0045] Figure 14(c) is a comparison of the stress cloud diagrams of the internal and external threads when thread 2 is loaded;

[0046] Figure 14(d) is a comparison of the stress cloud diagrams of the internal and external threads when thread 3 is loaded;

[0047] FIG15 is a schematic diagram of the numbering and average value positions of external thread teeth;

[0048] FIG16 is a diagram showing the average stress of each thread tooth of the external thread of the present application thread and three comparative threads. DETAILED DESCRIPTION

[0049] The features and performance of the present invention are further described in detail below with reference to the embodiments.

[0050] The non-equidistant thread connection pair of the present invention comprises an internal thread and an external thread, the pitch of the internal and external threads are constant and the pitch of the internal thread is greater than the pitch of the external thread, and the pitch difference between the internal and external threads is Where P is the pitch of the external thread. Compared with the existing technology, the present invention reduces the pitch difference between the internal and external threads, and solves the internal and external thread pitch difference based on the consideration of the material properties of the threaded connection pair and the load it bears. It has been verified that the use of the pitch difference within the above range can make the stress borne by the thread teeth more uniform.

[0051] Example 1 of the non-equidistant threaded connection pair of the present invention:

[0052] A non-equidistant thread connection pair includes an internal thread and an external thread. The pitch of both threads is constant, and the pitch of the internal thread is greater than the pitch of the external thread. The thread profiles of the internal and external threads can be metric, MJ, trapezoidal, and circular threads. Circular threads provide a larger axial clearance, facilitating machining and installation. For ease of description, this embodiment uses a metric thread profile as an example.

[0053] As shown in Figure 1, the tooth width at the external thread pitch P, major diameter d, minor diameter d1, and external thread pitch diameter d2 is S w , the yield strength of the external thread material is σ, the elastic modulus E w .

[0054] As shown in Figures 2 and 4, the internal thread pitch P', major diameter D, minor diameter D1, and the tooth width of the internal thread at the external thread pitch diameter d2 are S n The number of internal thread turns is n (i.e. the total number of turns of the internal and external thread engagement areas), and the elastic modulus of the internal thread material is E n Among them, the notch C at the middle diameter d2 of the external thread n The width is greater than the tooth width S of the internal thread at the middle diameter d2 of the external thread n , namely: PS w >S n , to prevent the internal thread tooth width from being too large, which will cause interference fit between the internal and external threads, thereby reducing assembly resistance and improving assemblability.

[0055] For the convenience of description, the present invention sets a starting position and an ending position, as shown in Figures 3 and 4. Taking the bolt 1 and the nut 2 as an example, in the screwing area of ​​the internal and external threads, the present invention uses one end face of the internal thread perpendicular to the axis as the starting position section Q, and the other end face of the internal thread as the ending position section Z (traditionally considered as the support surface, that is, the end face of the nut 2 used to press the connected parts). The direction from the starting position to the ending position is the same as the direction in which the external thread bears the rated axial load N.

[0056] Since the number of thread turns of the internal thread is n, when the internal and external threads are used together, n turns of the internal thread must be used in conjunction with n turns of the external thread. When no load deformation occurs, starting from the starting position, any n' number of turns (0<n'≤n) of the internal thread will be longer than the external thread.

[0057] Specifically, the total length of the external thread with n' turns is L1, the length of the internal thread with n' turns is L2, and the length difference between the internal and external threads is: l = L2 - L1; the total axial clearance δ when the internal and external threads are used together is PS w -S n. As shown in Figure 5, in the upper figure, the pitch of the internal and external threads is equal, one side of the internal and external threads fit together, and the other side does not contact, and the axial clearance between the internal and external threads is δ. In the lower figure, the internal and external threads fit together on the far left. Since the pitch of the internal thread is greater than the pitch of the external thread, the internal and external threads start to separate from the fitting position. When they are on the far right, the other sides of the internal and external threads gradually approach. At this time, the length difference between the internal and external threads is l. When δ=l, the rightmost side of the internal and external threads just touch; when δ<l, the internal and external threads will have an interference fit, which is not convenient for assembly. Therefore, δ≥l, which can prevent the tooth biting phenomenon caused by insufficient axial clearance, thereby reducing assembly resistance and improving assemblability.

[0058] Furthermore, the length of the external thread L1 = n'P, the length of the internal thread L2 = n'P', and the pitch difference between the internal and external threads ΔP = P'-P, then l = L2-L1 = n'(P'-P) = n'ΔP, that is Therefore, to solve the pitch difference ΔP, it is necessary to find the length difference l between the internal and external threads.

[0059] When the traditional equidistant thread is loaded and engaged, the external thread will be stretched and the internal thread will be compressed and shortened, and the external and internal threads at any position in the engagement area will have different degrees of elongation or compression. w Total compression with internal thread l n The sum of the lengths of the internal and external threads is taken as the length difference l.

[0060] As shown in Figure 6, for the convenience of analysis, the external thread in the screwing area of ​​the internal and external threads is simplified into an equivalent force-bearing cylinder. The two circular sections of the equivalent force-bearing cylinder are the starting position section Q and the ending position section Z. The cross-sectional area A1 of the equivalent force-bearing cylinder is the stress cross-sectional area of ​​the thread. The axial load N borne by the equivalent force-bearing cylinder at the ending position section Z is in the direction of the axial load N from the starting position section Q to the ending position section Z. The acting force between the internal and external threads is simplified to the outer cylindrical surface of the equivalent force-bearing cylinder bearing a surface load. The surface load is in the opposite direction to the axial load N, and the resultant force of the surface load is the same as the magnitude of the axial load N. Taking the starting position as the starting point, the resultant force of the load on the small external cylindrical surface at any n' number of turns is f1(n'), and the resultant force of the load on all small external cylindrical surfaces is equal to the axial load N, that is,

[0061] A small cylinder is cut off at any height h=n'P of the equivalent load-bearing cylinder. The thickness of the small cylinder is dh=Pdn'. The axial load on the lower section of the small cylinder is The average elongation of a small cylinder can be obtained from Hooke's law The total average elongation of the equivalent loaded cylinder at any n' number of turns is

[0062] As shown in Figure 7, the direction of the surface load on the outer cylindrical surface of the micro-cylinder is opposite to the direction of the load N1 on the lower cross-section of the micro-cylinder, which makes the elongation at the outer cylindrical surface of the micro-cylinder smaller than the average elongation dl. w1 , the outer cylindrical surface of the equivalent load-bearing cylinder is the simplified area of ​​the external thread, so the elongation of the outer cylindrical surface should be the elongation of the external thread dl w , then dl w =k1·dl w1 , 0<k1<1, for the convenience of analysis, assuming k1 is a constant value, then taking the starting position as the starting point, the total elongation of the external thread with any n' turns is

[0063] As shown in Figure 8, for the convenience of analysis, the internal thread in the screwing area of ​​the internal and external threads is simplified into an equivalent force-bearing hollow cylinder. The two circular arc annular sections of the equivalent force-bearing hollow cylinder are the starting position section Q and the ending position section Z. The cross-sectional area of ​​the equivalent force-bearing hollow cylinder is A2, and the axial load borne by the equivalent force-bearing hollow cylinder at the ending position section Z is F. N (According to the balance of forces, F N The direction is from the end position to the starting position, and the force between the internal and external threads is simplified to the inner cylindrical surface of the equivalent hollow cylinder under the surface load. The surface load and the axial force F N The direction is opposite, the resultant force of the surface load and the axial force F N The size is the same. Taking the starting section as the starting point, the resultant force of the load on the small inner cylindrical surface at any n' turns is f2(n'). The resultant force of the load on all small inner cylindrical surfaces is equal to the axial load F N Equal, that is

[0064] A small hollow cylinder is cut off at any height h=n'P of the equivalent force-bearing hollow cylinder. The thickness of the small hollow cylinder is dh=Pdn'. The axial load on the lower section of the small hollow cylinder is Average compression of a tiny hollow cylinder with a thickness of dh By integration, we can find the total average compression of the equivalent hollow cylinder at any n' number of turns starting from the starting section Q:

[0065] As shown in Figure 9, due to the surface load direction of the cylindrical surface inside the tiny hollow cylinder and the load F on the lower section of the tiny hollow cylinder, N The direction of action is opposite, making the compression on the cylindrical surface of the tiny hollow cylinder greater than the average compression dl n1The inner cylindrical surface of the equivalent hollow cylinder is the simplified area of ​​the internal thread, so the compression of the inner cylindrical surface is the compression of the internal thread dl n , then dl n =k2·dl n1 , 1<k2, for the convenience of analysis, assuming k2 is a constant value, then starting from the starting position, the total compression of the internal thread with any n' turns is

[0066] As shown in Figure 10, the axial load of the internal thread acting on the external thread teeth at any n' turns starting from the starting position is F(n'). From the simplified mechanical relationship of the equivalent force-bearing cylinder and the equivalent force-bearing hollow cylinder, we can know that F(n') = f1(n') = f2(n'). Therefore, the total elongation of the external thread l w Total compression with internal thread l n The sum is the length difference l between the internal and external threads, that is:

[0067] When the internal and external thread materials, i.e. the nut dimensions, are selected, E n and A2 are both constant, so E w A1 and E n The ratio k3 of A2 is a constant. In order to facilitate the calculation, the elastic modulus and area in the formula are unified into E w and A1, then For further simplification, let K = k1 + k2k3, where K is the comprehensive reference coefficient, and simplify the formula for the length difference l between the internal and external threads to:

[0068] Furthermore, the length adjustment of the gradual pitch thread In the formula, only F(n') and n' are variables, and there are three common situations in which the axial load F(n') acting on the external thread follows the change trend of n'. As shown in Figure 11, F(n') in curve a increases as n' increases, F(n') in curve b is a constant and remains unchanged, and F(n') in curve c decreases as n' increases. According to the mechanical relationship, the cumulative value of the axial load F(n') on the external thread from the starting position to the end position is the cross-sectional axial load N, that is,

[0069] Simulation calculations were performed with three different trends of F(n'), and the calculated stress cloud maps are as follows: Figure 12(a) is the stress cloud map corresponding to curve a, Figure 12(b) is the stress cloud map corresponding to curve b, and Figure 12(c) is the stress cloud map corresponding to curve c.

[0070] High stress areas typically occur at the root of external threads. The magnitude of the stress is influenced by the combined effects of the axial load F(n') on the external threads and the cross-sectional axial load N(n') (i.e., the force applied to the entire cross-section (excluding the threads) for n' turns). When the internal and external threads are deformed under load, the cross-sectional axial load N(n') increases with the number of thread turns n', starting from the starting position, and reaches its maximum value N at n turns. As shown in Figure 11, when the axial force F(n') on the external threads conforms to curve a, F(n') is smaller at the starting position and larger at the ending position. This trend mirrors the variation of N(n'), as shown in Figure 12(a). Y in the figure represents the location with the maximum stress. The figure shows that the stress at the starting position of the external thread is smaller, while the stress at the ending position is larger, indicating a significant stress concentration phenomenon. Specifically, Y is concentrated at the ending position, and the area of ​​Y increases as it approaches the ending position.

[0071] As shown in Figure 11, when the axial force F(n') on the external thread teeth conforms to curve b, F(n') is a constant and unchanged value. However, since the cross-sectional axial load N(n') gradually increases with n', the stress at the root of the external thread teeth still shows a small stress value at the starting position and a large stress value at the end position, but the stress concentration phenomenon is improved. As shown in Figure 12(b), the area of ​​Y at the end position becomes smaller. When the axial force F(n') on the external thread teeth conforms to curve c, F(n') further increases at the starting position and further decreases at the end position, causing the stress at the root of the external thread teeth at the starting position to further increase and the stress at the end position to further decrease. The stress value of the entire external thread tooth root is more uniform, and the stress concentration phenomenon is smaller. As shown in Figure 12(c), Y is evenly distributed from the starting position to the end position.

[0072] Furthermore, F(n') conforms to the stress state of curve c better, and the selected F(n') equation should satisfy curve c. F(n') should meet two conditions: it is a decreasing function in the interval 0≤n'≤n,

[0073] In order to facilitate the calculation, it is assumed that the axial force on the external thread is 0<n'≤n decreases linearly, and Further conclusion When n'=n, the length difference l between the inner and outer threads is: Let k be the load coefficient, which is the ratio of the axial force N borne by the external thread to the product of the yield strength σ of the material and the stress cross-sectional area A1 of the external thread, that is, Let k c is the material coefficient, which is the material's yield strength σ and the material's elastic modulus E w The ratio of When the main design parameters and materials of the bolt are determined, k and k c is a constant value.

[0074] Furthermore, the pitch difference between the internal and external threads 0.8≤K≤4.5.

[0075] Furthermore, the load coefficient k is usually in the range of 0.4<k<0.8, and the material coefficient k of commonly used thread materials is c =0.001~0.02, taking into account k, k c and K limit the pitch difference between internal and external threads to 0.003P<△P<0.01P.

[0076] The following is a combination of specific thread parameters and comparative tests to demonstrate the superiority of the range of values ​​for the pitch difference ΔP between the internal and external threads of the non-equidistant thread connection pair of the present invention.

[0077] Specific implementation method one:

[0078] The external thread profile is a metric thread, with a pitch of P = 4mm, a major diameter d = 42mm, a minor diameter d1 = 37.67mm, and a median diameter d2 = 39.4mm. The tooth width at the median diameter of the external thread is S w =2mm. Elastic modulus E of external thread material w =206GPa, yield strength σ=930Mpa, axial load N=0.7σA1.

[0079] The internal thread major diameter D = 42.42mm, the minor diameter D1 = 38.09mm, and the tooth width of the internal thread at the external thread mid-diameter d2 is S n =1.76mm, the number of internal thread turns is n=7 turns.

[0080] In this embodiment, PS w =2mm>1.76mm,δ=PS w -S n =0.24mm, l=nΔP=0.126mm, δ>l. k=0.7, △P=0.00211×KP.

[0081] The comparative thread profile provided in this embodiment is the same as the thread of the present application, and the pitch differences between the internal and external threads of various threads are shown in Table 1. The pitch difference between the internal and external threads of comparative thread 1 is ΔP = 0.01P, and the pitch difference between the internal and external threads of comparative thread 2 is ΔP > 0.01P (comparative thread 2 is the form disclosed in CN106438657A, in which the internal and external threads are interference fit. Before loading, the internal and external threads interact with each other, generating large stresses at the starting and ending positions, and the middle turns of the thread are in a suspended state). Comparative thread 3 is a conventional equidistant thread.

[0082] Table 1 Pitch difference ΔP between internal and external threads

[0083] Figures 13(a) to 13(d) and 14(a) to 14(d) are stress cloud maps under different axial loads. It should be noted that the simulation model is an elastic model, and the calculated average stress sometimes far exceeds the yield stress. This indicates that there is obvious stress concentration at that location. Figure 13(a) is the stress cloud map of the thread of the present application without axial load, Figure 13(b) is the stress cloud map of the comparative thread 1 without axial load, Figure 13(c) is the stress cloud map of the comparative thread 2 without axial load, and Figure 13(d) is the stress cloud map of the comparative thread 3 without axial load. The stress cloud maps of the four threads without axial load show that the comparative thread 2 is in an interference fit, and local stress exists at the initial and final positions of the thread. Figure 14(a) is a stress cloud diagram of the thread of the present application under the action of axial load N=0.7σA1, Figure 14(b) is a stress cloud diagram of the comparative thread 1 under the action of axial load N=0.7σA1, Figure 14(c) is a stress cloud diagram of the comparative thread 2 under the action of axial load N=0.7σA1, and Figure 14(d) is a stress cloud diagram of the comparative thread 3 under the action of axial load N=0.7σA1. It can be seen from Figures 14(a) to 14(d) that compared with comparative threads 1 to comparative threads 3, the high stress concentration phenomenon of the thread of the present application is significantly improved.

[0084] According to the numbering (1-7) of the external threads and the positions indicated by the bold short vertical lines in Figure 15, the average stress of each thread was extracted, the average stress variation of each thread was plotted, and the stress concentration factor of each thread was calculated (stress concentration factor = maximum stress / total average stress). Figure 16 shows that the stress uniformity of the thread of this application is good, and Table 2 shows that the stress concentration factor of the thread of this application is the smallest.

[0085] Table 2 Stress concentration factors of each thread tooth of external thread

[0086] In other embodiments, the total average elongation l of the equivalent force-bearing cylinder at any n' number of turns can also be directly used. w1 To represent the total elongation l of any n' number of turns of external thread w Of course, we can also directly use the total average compression of the equivalent hollow cylinder at any n' number of circles as l n1 To represent the total compression of any n' number of internal threads l n .

[0087] In other embodiments: according to different requirements, when the simplification method is different, the simplified results of the equations of ΔP and l may be different.

[0088] In other embodiments: the gap at the middle diameter of the external thread is equal to the tooth width S of the internal thread n , namely: PS w =S n .

[0089] The above description is only a preferred embodiment of the present invention and is not intended to limit the present invention. The scope of patent protection of the present invention shall be based on the claims. Any equivalent structural changes made using the description and drawings of the present invention shall be included in the scope of protection of the present invention.

Claims

1. A non-equidistant thread connection pair, comprising an internal thread and an external thread, wherein the pitches of the internal and external threads are constant and the pitch of the internal thread is greater than the pitch of the external thread, characterized in that: Pitch difference between internal and external threads Where: P is the pitch of the external thread; K is the comprehensive reference coefficient, and 0.8≤K≤4.5; k is the load coefficient, which is the ratio of the axial load N borne by the external thread to the product of the yield strength σ of the external thread material and the stress cross-sectional area A1 of the external thread, that is, k c is the material coefficient, which is the yield strength σ of the external thread material and the elastic modulus E of the external thread material w The ratio of 2. The non-equidistant threaded connection pair according to claim 1, characterized in that: The pitch difference between internal and external threads △P<0.01P.

3. The non-equidistant thread connection pair according to claim 1, characterized in that: The pitch difference between internal and external threads △P>0.003P.

4. The non-equidistant threaded connection pair according to any one of claims 1 to 3, characterized in that: Pitch difference between internal and external threads Where l is the length difference between the internal and external threads, n' is the number of turns starting from the starting position of the internal and external thread engagement area, 0≤n'≤n, n is the total number of turns of the internal and external thread engagement area; the length difference l between the internal and external threads is the total elongation l of the external thread w Total compression with internal thread l n the sum of Where 0<k1<1, the external thread in the screwing area of ​​the internal and external threads is simplified into an equivalent force-bearing cylinder, f1(n') is the resultant force of the load on the tiny external cylindrical surface at any n' turns of the equivalent force-bearing cylinder, and the resultant force of the load on all tiny external cylindrical surfaces is equal to the axial load N; Where k2>1, E n is the elastic modulus of the internal thread material, the internal thread in the screwing area of ​​the internal and external threads is simplified into an equivalent force-bearing hollow cylinder, A2 is the cross-sectional area of ​​the equivalent force-bearing hollow cylinder, f2(n') is the resultant force of the load on the tiny internal cylindrical surface at any n' number of turns of the equivalent force-bearing hollow cylinder, and the resultant force of all the loads on the tiny internal cylindrical surface and the axial load F N Equal, F N Equal to N and opposite in direction; therefore Let E w A1 and E n The ratio of A2 is k3, and k1+k2k3=K, the formula of the length difference l between the internal and external threads is simplified to: Further order but Further conclusion When n'=n, the length difference l between the inner and outer threads is: Furthermore, the pitch difference between the internal and external threads 5. The non-equidistant thread connection pair according to any one of claims 1 to 3, characterized in that: Define the tooth width of the external thread at the mid-diameter as S w The tooth width of the internal thread at the middle diameter of the external thread is S n When the internal and external threads are used together, the total axial clearance δ=PS w -S n , δ is greater than or equal to the length difference l between the internal and external threads, that is, δ ≥ l.

6. The non-equidistant threaded connection pair according to any one of claims 1 to 3, characterized in that: Define the tooth width of the external thread at the mid-diameter as S w , notch C at the middle diameter d2 of the external thread n Width is greater than the tooth width S of the internal thread at the middle diameter d2 of the external thread n , then PS w >S n .

Images