Calculating the Loads on the Front Suspension

Suspension FEA

Simplified left front suspension finite element analysis result

It’s time to finalize my parts order, which means I have to pick all the rod ends and spherical bearings for the suspension. Up to now I’ve relied on rules of thumb, which isn’t good enough for a state-of-the-art race car. So I simply opened up one of my higher-level chassis models in Solidworks, one with most of the suspension attached, and created a 3D sketch of the upper and lower left front suspension A-arms, constraining the endpoints of the lines to the centers of the rod ends and spherical bearings. Then another line to simulate the upright and tire, and I’m done with the sketch. Then I created a structural member along each line and opened a static simulation study. In the study I constrained the inboard ends of the A-arms to zero translation and applied a 1400 pound load toward the rear of the car to simulate 4-G braking. Run the study, and voila! axial loads on each member.

Upper Front: 785 pounds

Upper Rear: 1211 pounds

Lower Front: 2639 pounds

Lower Rear: 3096 pounds

This allows me settle on 1/4″ rod ends for the inner upper attachments, 3/8″ for the inner lowers, a 7/16″ spherical bearing for the upper ball joint, and a 5/8″ one for the lower ball joint. All chrome moly alloy. Even these are far stronger than indicated necessary by this analysis, but I don’t want to deviate too far from standard practice on this design. Maybe transitory loads can be far higher?

Calculating the Loads, Round II

At the urging of a poster on ApexSpeed.com who insisted my calculations were inaccurate due to the non-zero joint stiffnesses, I re-ran the analysis with the suspension a-arms modeled as trusses. A truss model allows only axial loads on a member, using hinges at both ends to zero the moments. This gave me results that were the same +/- 2%, which makes no difference given the uncertainty in estimating the operating loads.

However, I also refined my model to include the suspension pushrod and that did indeed give me a different result. With the pushrod in place I could model the vertical load due to the weight of the car and the aerodynamic downforce. I’m now working with a braking force of 1600 pounds and a vertical load of 1230 pounds, which assumes a coefficient of friction of 1.3 for the tire on the surface of the track. This gives me the following axial forces:

Suspension

Left front suspension modeled as a truss with pushrod

Upper Front: 1103 pounds

Upper Rear: 1258

Lower Front: 4432

Lower Rear: 2366

Pushrod: 2357

Note that including the vertical load and increasing the braking load shifted the most highly-loaded member from the lower rear to the lower front, and increased the maximum load from 3096 to 4432 pounds, or 43%. The chrome-moly 3/8″ rod ends I’ve specified for the lower a-arm inboard joints still have a safety factor of 2.14 under these conditions. The posters on ApexSpeed have strongly warned me against using 1/4″ rod ends in the suspension, so existing inventory will be used in a different application and the upper rod ends will be upgraded to 8mm. It appears that only metric rod ends are available locally. Under these assumptions the pushrod rod ends will have to be further upgraded to 10mm, although I hate to have a whole separate part for just the pushrods, so I’ll have to re-examine the assumptions.

Advertisements

Leave a Reply

Fill in your details below or click an icon to log in:

WordPress.com Logo

You are commenting using your WordPress.com account. Log Out / Change )

Twitter picture

You are commenting using your Twitter account. Log Out / Change )

Facebook photo

You are commenting using your Facebook account. Log Out / Change )

Google+ photo

You are commenting using your Google+ account. Log Out / Change )

Connecting to %s