Wiisass

John, those tire temps don't really say much. I mean if your saying that what you ran there was the best setup where the comparison data? Do you have temp data and lap times for running the tires at 40psi hot? Like Cary was saying, I think that measuring tire temps over the course of a lap will just tell you if your setup is close to what it should be for the course. I mean if your inside temps are 50degrees hotter than the outside, then your going to have to adjust something. But if they're in a good range, then you should be fine. IR temps are something that would really help out with suspension tuning, but most people don't have access to stuff like this and the cost on it does add up pretty quickly so it's not for the less serious. But knowing the temp on the surface of the tire during a run at each part of the course will give you a much better idea than measuring temps at the end of the lap. you do have to keep in mind that surface temps don't correlate in any linear relationship with core temps, so the IR temps are essentially measuring something different than core temps. John, another question, are you provided with tire curves for the tires you run? I mean they suggest a pressure based on these curves and what their testing shows.

Doesn't matter to me, I'm just interested to see how far people are going with this stuff.

When I was talking about having them both at the same height. That's just ideal, it makes things a lot easier. Like if you had complete freedom in where the pivots were, but I know this isn't the case with production cars. I might have been mistaken about things being parallel and being close enough to not matter. Without plotting all this stuff, it seems like if the arms are parallel and the inboard and outboard lateral locations are close for the tie rod and lca it will work out pretty well. It's when those locations are different where it will start to hurt you. For example, if the tie rod pivot is an inch further outboard that the lca inboard mount, then the arcs will have different radii and it will create some bumpsteer at the top and bottom of the travel. I think that's right, I'm trying to picture all of this in my head. It should be pretty easy to do this graphically. You would just need the distance from pivot points of the tie rod and lca and then the height differences of the inboard points and the outboard points. And then just use a compass to trace an arc for each. You would then need to draw vertical lines at several different points on each arc, you could do it at points that represent a half inch in vertical wheel travel and then measure the distance. If the distance stays the same between the two lines throughout the range of travel you should have no bumpsteer. But no matter how you do it, if you measure it and it measures zero or very little bumpsteer then it's working fine. Depending on how your adjusting ackerman, you could do it so bumpsteer is minimally affected if at all. By moving the rack forwards and backwards, if you keep the inboard pivot at the same height and lateral position and the outboard stays the same as well, then there shouldn't be much if any increase in bumpsteer from the previous setup. You mentioned something about the LCA, what are you planning on doing with that and was that in reference to the Ackerman changes or to the bumpsteer adjustment? if it's the bumpsteer part, then don't worry about explaining it, but if it's for ackerman, I would be interested in what your trying to do.

C'mon, they're easy questions. It's not a test, maybe a pop quiz. I'm just curious and trying to elevate the discussion to a higher level. I didn't mean to come off like I don't believe you, I just want more information.

So what temps were you getting at pressures of 30/32 and what temps were you supposed to be getting at a pressure of 40? Also how was "grip" quantified? Accelerometer, lap times, driver feedback?

What are you disagreeing with? And you have to consider a lot more when you lower the tire pressure. This will affect how much force the tire can generate, it's stiffness in several different directions, etc. Adjusting the pressure does a lot more than just put more or less area on the ground. And softer compounds, stiffer construction is the way to go.

For zero bumpsteer, ideally, you would have the inner lca pivot at the same height and lateral distance from the centerline of the car as the tie rod pivot. Then the outboard points would also be at the same height and distance from centerline. This will assure that the outboard points move the same amount. Just having the arms parallel won't work. You have to take into consideration the inboard and outboard points for the LCA and the tie rod. You should be able to find the exact location everything needs to be with a simple kinematic analysis. I've seen several different methods for solving this in different suspension books. One way is to determine the instant center of the LCA, tie rod and strut. It's actually for an SLA suspension, but since we're talking MacP, I'm guessing using the line perpendicular to the strut axis as the top line is fine. This will determine the angle that you want the tie rod to be. Once the angle is determined, you need to figure out the length and where the pivots will be. This can be done graphically. Since you're working from a car that already exists, your constrained differently. Are you considering changing the inboard tie rod pivot location? If so, that gives you another place to play with. If not, then you'd be best to just make sure the tie rod follows the line that it should and you might have to accept whatever bumpsteer is now inherent in the suspension.

I found Grape on CC, thanks. So it looks like it's just an FE program, and not a suspension analysis program. So it's something more along the lines of ANSYS. So it's something that's good for rigid parts and not modeling suspension kinematics. About inventor, it sounds like you'll be fine using that. I don't know what you mean about limits to the modifications you can do. What kind of stuff are you planning? Some of the cheaper programs do have limits, but it also depends if they have a template for your or if you have the ability (in the program) to create your own. If you step up to a program like ADAMS, you can do anything. It just has a very steep learning curve. One thing I would like about using inventor would be the complications when trying to graph suspension features. I mean just for an example, graphing camber versus wheel travel to get an idea of what your camber curve will look like. In a decent suspension program, you just have to click a couple things, but in inventor it seems like you would have to do a lot more. But it might be easier than I'm imagining, I haven't had any experience with it in about 5 years and even then it wasn't much. But then again, even with experience in other CAD software and imagining the possibilities, I would choose a simple suspension analysis program, like susprog3d or suspension analyzer, over the CAD program any day, just for the layout.

Does anyone have a link to this grape program, I would like to check it out? I have used Susprog3d, Suspension analyzer, ADAMS and Bill Mitchell's Wingeo before for suspension design and they all have their good and bad sides. Are you using inventor for any reason, it seems a more suspension oriented program would be more beneficial for work like this. It just seems that there's a lot of extra work involved in using something like inventor. But I guess if there's an FE addition to this, you could use it for stresses and loads and such to calculate what size members you would need.

I didn't think you were an idiot, i was just under the impression that you were trying to replace the whole tie rod rather than just the end. That is why I suggested the clevis. Alright, I see what your doing now. Tapping won't be too hard, either pay someone to do it on a lathe or do it at home on the vice and just be careful. If you don't want to be concerned with the bolt breaking, then that's fine. But I'm just bringing it up because when you space it too far the bending in the bolt gets very high. If your putting high forces through the tie rod, the moment your creating when you space down the outboard tie rod point too far can be higher than the bending strength of the bolt. It's a cantilivered beam, it should be something to be concerned with especially when you space it a lot.

Is there going to be a joint in this hex stock? If not I highly recommend that you don't just thread that into the rack. You're going to need some type of joint at the rack so the suspension can move. Also about bending the 5/8" bolt, I've seen it done before on a 240SX with about 1" spaced down. That's why I thought I would mention my concern. It should just be an issue of making the bumpsteer spacers have a larger OD to take a lot of the bending load. When you say 0 ackerman, I'm guessing you mean parallel steer. About moving the outboard tie rod point laterally, I figured there might be some clearance issues, but depending on the length of the steering arm, you should still be able to get a little out of it. Your raising the inboard board LCA point correct?

Cygnusx1, do you still have the sway bar on the car? Or have you completley removed it at this point. From your previous description is seems like a spring and bar problem. You said that it happens at the middle of the turn when the car is closest to its steady state performance. This would mean that the shocks weren't adding much to the effect. So removing the front bar or softening it, if adjustable, would be what I would suggest. The camber thing is also probably an issue, but I don't know S30s well enough to say whether or not. But since it seems like you increased your camber and your still only running 1.25 degrees, I would definitely say you aren't running enough. How much lower than stock is your car? Lowering the car, especially on a strut suspension greatly effects the camber curves. If you lower the car enough, the front view IC will actually be outside of the car giving positive camber gain with jounce. Oh and btw, are you Supra guy as well? Col, how much will the weight of the V8 effect the overall weight and distribution of the car? The weight distribution change will cause the greatest effect. But I don't know how much it will change things. But then again, if you're doing the engine swap before you redo the suspension, all of the effects changing engines can be tuned out through selection of different springs/sway bars/dampers/tires.

For the inside, have you considered having clevises made? If wouldn't be a hard piece to have machined. And then you could use a simple rod with right and left threads on each end with rod ends for your turnbuckle/adjuster. Then you would be able to use the 5/8 rod ends, which would help out a lot with cost because metric rod ends in the us are not anywhere near worth the price. And then you could do what you had planned for outboard bumpsteer adjustment. Also depending on how you do the inboard clevis, you could use that for bumpsteer adjustment as well. Another thing to be concerned about. Depending on how much your spacing the rod from the steering arm, you're going to have to account for the bending in the connecting bolt. If you space it enough the moment can be great and will bend that bolt. Depending on how you design the spacers, most of the load can be taken by them and you shouldn't have any problems, but it is something to keep in mind if you don't want this thing breaking at the track. Now I'm guessing the talk about moving the rack around is just for ackerman. What is the stock ackerman setting? How much are you planning on going or is it just a guess and test method? If you can't move the rack much, you could consider moving the outboard point laterally which will also adjust the ackerman. This is probably something that needs to be done graphically to make sure you're getting what you've planned on.

Alright, I read through this whole thread last night, and I know I'm going to miss some stuff, but I did catch a couple mistakes. First off, it's a good guide for struts, I'm currently looking to put something together for another macp car and have been searching for take apart struts. Or at least a monotube strut that I could easily rebuild/revalve. But a couple things that I noticed. Someone mentioned that they thought the HTS plot was in lbs on the y-scale and meters per second on the x-axis. I think that the y-scale is in newtons /10. Which would make the upper limit of the rebound side around 4000 newtons. This makes much more sense if you work out the numbers. Using 400lbs at 0.6mps would only get you a damping coefficient of somewhere around 16, while using 4000N at 0.6mps would get you a coefficient of 38lb/in/s. Which is a much more reasonable value especially if people are trying to run higher spring rates with. Another thing, someone mentioned about the installation ratios and their effect on spring rates. Installation ratio is the ratio of the displacement of the spring/shock/roll bar versus the displacement of the wheel center. When it comes to comparing the spring rate or damping coefficient of the shock/damper to the equivalent rate at the wheel center, the installation ratio needs to be squared. This is important in order to calculate the wheel rate. So for example, you have a 1000lb/in spring with a 0.9 installation ratio. The spring will move 0.9 inches for every inch of wheel travel, but will have a rate of 810lb/in at the wheel center. This rate is known as the wheel rate and allows people to compare different vehicles more objectively. This wheel rate along with the sprung mass value allows you to calculate the natural frequency of the corner. The tire rate also needs to be included in this to get a better idea, but that adds another big variable. Calculating the natural frequencies of the sprung mass is a way of normalizing the suspension model against other cars. Common values for the natural frequency that are important for racers are around 2.0-2.5Hz. The lower end of this is the upper limit of factory sports cars. The range above 3Hz is more reserved for race cars with full aero. The higher the natural frequency, the stiffer the car is. I was wondering if anyone had physically measured the front and rear installation ratios of a 240Z? Also, what are common weights and distributions for the car? What's the normal unsprung weight at each corner? Also, I noticed that 225/250 seems to be a common spring rate upgrade, why is that? Is it just what several people have run with success or are there more reasons behind it. You guys seem to have a really good discussion going, I would just like to keep it going and maybe even get more and more in depth if you guys are ready. Tim

Alright, I read through this whole thread last night, and I know I'm going to miss some stuff, but I did catch a couple mistakes. First off, it's a good guide for struts, I'm currently looking to put something together for another macp car and have been searching for take apart struts. Or at least a monotube strut that I could easily rebuild/revalve. But a couple things that I noticed. Someone mentioned that they thought the HTS plot was in lbs on the y-scale and meters per second on the x-axis. I think that the y-scale is in newtons /10. Which would make the upper limit of the rebound side around 4000 newtons. This makes much more sense if you work out the numbers. Using 400lbs at 0.6mps would only get you a damping coefficient of somewhere around 16, while using 4000N at 0.6mps would get you a coefficient of 38lb/in/s. Which is a much more reasonable value especially if people are trying to run higher spring rates with. Another thing, someone mentioned about the installation ratios and their effect on spring rates. Installation ratio is the ratio of the displacement of the spring/shock/roll bar versus the displacement of the wheel center. When it comes to comparing the spring rate or damping coefficient of the shock/damper to the equivalent rate at the wheel center, the installation ratio needs to be squared. This is important in order to calculate the wheel rate. So for example, you have a 1000lb/in spring with a 0.9 installation ratio. The spring will move 0.9 inches for every inch of wheel travel, but will have a rate of 810lb/in at the wheel center. This rate is known as the wheel rate and allows people to compare different vehicles more objectively. This wheel rate along with the sprung mass value allows you to calculate the natural frequency of the corner. The tire rate also needs to be included in this to get a better idea, but that adds another big variable. Calculating the natural frequencies of the sprung mass is a way of normalizing the suspension model against other cars. Common values for the natural frequency that are important for racers are around 2.0-2.5Hz. The lower end of this is the upper limit of factory sports cars. The range above 3Hz is more reserved for race cars with full aero. The higher the natural frequency, the stiffer the car is. I was wondering if anyone had physically measured the front and rear installation ratios of a 240Z? Also, what are common weights and distributions for the car? What's the normal unsprung weight at each corner? Also, I noticed that 225/250 seems to be a common spring rate upgrade, why is that? Is it just what several people have run with success or are there more reasons behind it. You guys seem to have a really good discussion going, I would just like to keep it going and maybe even get more and more in depth if you guys are ready. Tim