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The book takes the subject from an introductory level through advanced topics needed to properly design, model, analyze, specify, and manufacture cam-follower systems.
Cam Design and Manufacturing Handbook
(Kinetostatic Camshaft Torque)

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   by Robert L. Norton
Published By:
Industrial Press Inc.
Up-to-date cam design technology, correct design and manufacturing procedures, and recent cam research. SALE! Use Promotion Code TNET11 on book link to save 25% and shipping.
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EXAMPLE 9-4

Comparison of Kinetostatic Torques and Forces Among Three Alternate Designs of the Same Cam.

 

Given: A translating roller follower as shown in Figure 9-1a (p. 214) is driven by a force-closed radial plate cam that has the following program:

 

Design 1

Segment 1: Rise 1 inch in 90 ° double harmonic displacement

Segment 2: Fall 1 inch in 90 ° double harmonic displacement

Segment 3: Dwell for 180 °

 

Design 2

Segment 1: Rise 1 inch in 90 ° cycloidal displacement

Segment 2: Fall 1 inch in 90 ° cycloidal displacement

Segment 3: Dwell for 180 °

 

Design 3

Segment 1: Rise 1 inch in 90 ° and fall 1 inch in 90 ° with polynomial displacement

Segment 2: Dwell for 180 °

Camshaft velocity is 15 rad/sec (143.24 rpm); Follower effective mass is

0.0738 lb-sec2/in (blobs); Damping is 10% of critical ( æ = 0.10)

 

Problem: Find the dynamic force and torque functions for the cam. Compare their peak magnitudes for the same prime circle radius.

 

Solution:

 

1 Calculate the kinematic data (follower displacement, velocity, acceleration, and jerk) for each of the specified cam designs. See Chapter 8 to review this procedure.

 

2 Calculate the radius of curvature and pressure angle for trial values of prime circle radius, and size the cam to control these values. A prime circle radius of 3 in gives acceptable pressure angles and radii of curvature. See Chapter 7 to review these calculations.

 

3 With the kinematics of the cam defined, we can address its dynamics. To solve equation 9.1a (p. 215) for the cam force, we will assume a value of 50 lb/in for the spring constant k and adjust the preload Fpl for each design to obtain a minimum dynamic force of about 10 lb. For design 1, this requires a spring preload of 28 lb; for design 2, 13 lb; and for design 3, 10 lb.

 

4 The value of damping c is calculated from equation 9.2i (p. 218). The kinematic parameters x , v, and a are known from the prior analysis.

 

5 Program DYNACAM will do these computations for you. The dynamic forces that result from each design are shown in Figure 9-18 and the torques in Figure 9-19. Note that the force is largest for design 1 at 82 lb peak and least for design 3 at 53 lb peak. The same ranking holds for the torques, which range from 96 lb-in for design 1 to 52 lb-in for design 3. These represent reductions of 35% and 46% in the dynamic loading due to a change in the kinematic design. Not surprisingly, the sixth-degree polynomial design, which had the lowest acceleration, also has the lowest forces and torques and is the clear winner. Open the files E09-04a.cam, E09-04b.cam, and E09-04c.cam in program DYNACAM to see these results.

 

 

 

 

 

 

 

 

 

 

Copyright 2004, Industrial Press, Inc., New York, NY

 

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