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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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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.




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,, and in program DYNACAM to see these results.











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


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