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04-29-2015, 07:40 PM | #155 |
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DAMPING
With the suspension completely removed this winter, I took the opportunity to have the dampers dyno tested on a Roehrig 2VS shock dyno.
http://roehrigengineering.com/produc...actuators/2vs/ Not mine, unfortunately, so I had to pay for what I am about to reveal. As I had an OE rear shock on hand (for no reason than I hadn't gotten around to throwing it out, although it was still serviceable with about 90,000 km of use). The OEM shock provides an interesting comparison. It is typical to test dampers at shaft speeds from 0 - 10 inches per second and to measure the damping force in lbs (+ve for compression and -ve for rebound). The front struts have 28 adjustment positions and the rear shocks have 32 adjustment positions, so both were tested at a sampling of the range, with more emphasis on the stiffer end. The chassis (and the wheel) do not respond at the shaft speed, but at the shaft speed divided by the motion ratio. Similarly the damping force seen by the chassis (or wheel) is not the force at the damper shaft, but that force times the motion ratio. So rather than providing the raw data from the shock dyno, I have adjusted the forces and velocities by the motion ratio. This makes it possible to meaningfully compare the front damping to the rear. FRONT STRUT Several conclusions can be drawn from the above graph: 1) the Ohlins strut adjustments are effective from 1 (stiffest) to about 14 (mid adjustment), with any setting higher than 14 providing practically no change in damper characteristics. For reference, Ohlins recommends a setting of 10 as a starting point. 2) the Ohlins strut is a single-adjustable that primarily affects rebound damping but with a lesser effect on compression damping. REAR SHOCK 1) like the front strut, the rear adjustments are effective over about half the total adjustment range (about 1 to 15, 32 being maximum). Higher settings were omitted to avoid congesting the graph. Again Ohlins recommends 10 as a starting point. 2) the rear shock has similar damping curves to the front. 3) the OEM rear shock has less compression damping than the Ohlins shock at any setting of the Ohlins shock. 4) the OEM rear shock has less low-speed rebound damping than the Ohlins shock at any setting of the Ohlins shock (which probably explains some of the float experienced on the OE suspension). FRONT AND REAR COMPARED 1) the full-stiff settings on the Ohlins dampers are similar front and rear. 2) the full-soft setting on the Ohlins front strut is considerably softer in rebound than the full-soft setting on the Ohlins rear shock. In my experience a setting of 10 works well both front and rear for street driving, even though my rear springs are twice as stiff as those that came with the Ohlins kit (which would suggest that a higher setting at the rear might be desirable). The other benefit of having the dampers tested is that one can see how well each side is matched to the other, at the same damper setting. My dampers appear well matched, side to side. |
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04-29-2015, 11:13 PM | #156 |
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The Old "Nut and Bolt"
A "nut and bolt" inspection is a pre-track / pre-race / start of season tradition. The idea being to check that every fastener underneath your car is present, accounted for and properly torqued. The method can be anything from a cursory look, to a pull on everything critical with a wrench and whatever feels about right, to a check with a torque wrench on everything that can be reached with a torque wrench.
There is a better way. Once you have torqued a fastener to spec, mark it with Torque Seal . This is a lacquer paste that comes in a small tube that you apply to the junction of a fastener and a fitting. It dries to brittle state so that it will crack and fall off if there is any relative motion between the fastener and the fitting. http://www.aircraftspruce.com/catalog/cspages/f900.php If your torque seal is intact, all is good. On a big project, sometimes it is hard to know if you are installing something for good or just for now. It is easy to forget if you have actually torqued a fastener in such cases. Torque Seal applied after the torque wrench is a clear indication that the fastener has been torqued to spec. We have all laughed at the guy who forgot to torque his wheel nuts and lost a wheel. There but for the grace of god go I. This is "human factors" at work, and it can strike the best of us given the right conditions. Driver's meetings are mandatory. What happens if you get interrupted in your wheel change by a driver's meeting. Are you sure you will remember to torque your wheels when you get back? One trick is to always put your torque wrench on the driver's seat when you start a wheel change. That way you can't drive off without seeing it as a reminder the job isn't quite done. Other good habits: - use a checklist and check it off AS YOU DO THE TASKS. - never leave a job half-done without leaving the work in an obvious state (leave the fastener out rather than putting it in finger tight for example) - when you restart a job, start back a couple of steps from where you left off - make a REAL final inspection part of the job - the right solution for a given task depends on the situation. How familiar are you with the task? How critical is it? Can it be completed without interruption? Will it be completed without interruption? For involved tasks, I like a complete checklist showing all critical tasks, torque specs, manual references and wrench sizes. The job is not done until every box is initialled. Here are a couple of examples for the front and rear suspension installation. Sometimes the official torque value can't be found, but you can always make an educated guess and identify it as such. These lists aren't static - they should be revised with experience and new information. |
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04-30-2015, 09:31 AM | #157 | ||
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Strangely, I had actually used the same logic that you've presented when I was trying to determine why the values you had were different and I calculated 342 lbf/in for your front springs (and 399 lbf/in for the 70 N/mm springs). When you wrote 345 "lb/in", I thought that was a keyboard typo, since 5 is above 2 on the keypad. Now I know that was intentional since that's what you measured it at. It also makes more sense why the Swift 70 N/mm springs are referred to as "400 lbs" springs. So, based on your measurements, we should presume that the Swift data is actually supposed to refer to N/mm, which requires that their "kgf/mm" value be increased by a factor of 10 to represent the N/mm value and that their lbs/inch value is worthless and needs to be recalculated (and correctly referred to as lbf/in) from the newly determined N/mm value. Similarly, their Maximum Load values refer to two different units of measure: kgf and lbs. I think more weight (no pun intended) can be given to your summation since the springs are stamped in increments of 10, implying N/mm - unless Swift Springs were implying a decimal point, since they listed their spring rate as __.0 kgf/mm. Will we ever know? Being a Japanese company, Swift Springs would have measured everything using metric units, possibly further evident by the ordering of metric and imperial values, so their "lbs/inch" and inch values were always going to be calculated rather than measured. Furthermore, there seems to be less of an understanding on their part as to what the equivalent imperial unit of measure is for a particular metric unit of measure - notwithstanding their complete misunderstanding of using the correct metric unit of measure, although, that may just be attributable to having been lost in translation. Being a highly reputable spring manufacturer, I would have expected the same attention to detail as is evident in the manufacture of their springs and the quality that they are renown for. I have seen other spring manufacturers that use metric units of measure listing their spring rates using N/mm - let's hope they are correct! Well, that dissection was fun! On to even more fun an interesting things. |
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04-30-2015, 09:34 AM | #158 | ||
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I've taken a few weeks to absorb this, to go back to your previous posts, and to try and determine how your responses can help me to fill the gaps in my understanding, so I've documented my full thought process as well as my eventual realisations - at least it can provide a historical record for me and maybe answer some of the same questions others may have after reading this thread. If you'd like to respond, some of my latter comments may address some of the things I was originally having problems with comprehending so you may like to just reference those as being the correct interpretations. Hopefully it isn't too disjointed.
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While I don't have the hub mounting stands that you've constructed, I have remeasured my front springs correctly, while under normal vehicle load, with the wheels still mounted and the car sitting on the ground - I manoeuvred a tape measure into the wheel-well behind the wheel/tyre to sit flush with the spring. My measurements should be accurate to within 1mm. Below are those measurements: For starters, even though it's comparing left and right, it's strange that a 5mm difference in static spring length results in only (i) a 1.5mm difference in the exposed thread below the lower spring perch and (ii) a 2mm difference in ride height. I would have thought that the difference would have been considerate of the 0.96 motion ratio and, therefore, equal to ~4.8mm. That aside, the static spring load for each side differs and it also differs to your 3600 N value, which I thought there might be some correlation or relativeness even though it's comparing 70 N/mm and 60 N/mm spring rates. You made a statement: "Every 60 N/mm spring will have a static compression of 60 mm at a load of 3600 N". This makes me think that the static spring load is the known value from which others are calculated, which concurs with your spreadsheet calculations but it doesn't make sense to me considering the process that you used to initially determine that spring rate was dependent on the measurements for the free spring length, static/compressed spring length and spring rate, which implies the static spring load is the calculated value and, therefore, it should be recalculated for different height springs or springs that aren't compressed as much. I can see how the conversion from static spring load to corner weight shows the relationship between the two and so 3600 N makes sense from that perspective, which means that force was required for that corner weight, which was a result of that ride height, which required that 60 N/mm spring to be compressed by 60mm to achieve that... but... I can't make the link between all that and how the static spring compression would still be 60mm when you changed from OE top mounts to GC plates that had different top mount heights or when changing spring lengths. For the top mount change, I thought you would require 26mm less spring compression, which, for your 60 N/mm springs, would result in 1560 N less static spring load, that being 2040 N. For the spring length change from 200mm to 178mm, I thought you would require 22mm less compression, which, for your 60 N/mm springs, would result in 1320 N less static spring load, that being 2280 N. But then anything different to 3600 N static spring load would not result in the same corner weight. This is what I mean... I understand most parts in isolation but not all in unison. The spreadsheet calculations make sense but I'm having trouble understanding the theory. Speaking of the aforementioned spreadsheet, based on a 3600 N static spring load, I calculated what the values would be if the spring rate was changed from 60 N/mm to 70 N/mm (as is the case for my Swift springs) and it resulted in a static spring compression of 51mm (rounded down). For my springs, that have a free length of 178mm, this would result in a static spring length of 127mm, which is close to the 134mm (+4%) static spring length that I measured for my FR spring, which also has a 326mm ride height (without driver though), so I could see that the calculations roughly panned out but I was still stuck on why the 3600 N static spring load is fixed. Lightbulb moment! Now, I couldn't make sense of any of this until I realised that these linear springs are just obeying Hooke's Law, which states that "the restoring force F needed to extend or compress a spring by some distance X is proportional to that distance" and can be represented by F = -kX where F is the static spring load, k is a constant factor characteristic of the spring, i.e. it's spring rate, and X is the static spring compression. This restoring force (i) returns the spring to equilibrium, i.e. its natural state, (ii) is in the opposite direction to the applied force that’s extending or compressing the spring, and (iii) is reflected by the negative sign on the right side of the equation. Therefore, instead of solving for the restoring force, the equation can be changed to solve for the applied force, in which case, the restoring force would be negated resulting in the formula being -F = -kX or F = kX So, at least Hooke's Law helped me to understand the constant relationship between the static spring load, spring rate and static spring compression. I presume it also means that no single value needs to be considered fixed or known as they are all co-dependent and one value, e.g. static spring load, can initially be calculated from the other values and that calculated value can then later be used as the constant in the relationship since it will eventually be the other two values that will, or can, be variable. By using various values in the spreadsheet, I can see that, if the spring characteristics remain the same and the top mount is changed from OE top mounts to GC plates, the ride height will remain the same if the lower spring perch location is changed by the same amount as the change in top mount height. But since that is outside the scope of the spring, Hooke's Law doesn't apply. While writing this, I think I have made one further connection - based on your image below (as long as it is drawn to scale, which I believe it is considering your attention to detail), for a change in top mounts only, i.e. still using 200mm 60 N/mm springs, I now realise that, (i) the lower threaded portion of the strut assembly (where the lower spring perch is located) is in the same position, (ii) the different top mount assemblies slightly affected (perhaps by 10-12mm, something similar to the height of the top nut) the positioning of the shock in that assembly (there is a visible height different between the top of the strut top thread, and various other points of the upper shock, in the two drawings) resulting in less shock compression at static ride height and (iii) the additional height gained from the GC plates allowed the lower spring perch to be raised for increased tyre clearance. As a result, the static spring length is still the same to achieve all this, which means the static spring load is also the same. But there was a slight difference between the top portion of the two drawings, so the question still stands, although, the values are not to the same degree: how can the static spring length still be 60mm when there is a 10-12mm difference after changing to the GC plates? I thought you would require ~15mm less spring compression, which, for your 60 N/mm springs, would result in 900 N less static spring load, that being 2700 N. Possibly, due to the above-mentioned decreased shock compression, the shock has more length/stroke available before bottoming out and the spring could potentially have less static compression, resulting in more length/stroke available before coil binding (although, since the lower spring perch location was also raised to maintain ride height, the spring's potential has been negated at the "expense" of increased tyre clearance). Has there just been a repositioning of the spring in the vertical plane and, therefore, the static spring length doesn't change? Perhaps it's because the ride height is based on the overall static strut length and it doesn't really matter what the component measurements are, as long as the relationship between, and the measurements for, free spring length, spring rate, static spring load, static spring compression and static spring length remain the same or relatively the same. |
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04-30-2015, 09:36 AM | #159 | ||
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Just for reference, I have Vorshlag camber plates that have been set to its maximum camber of -3° - I've been hesitant to slot the towers for an additional -0.5° to -1.0° of camber. |
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04-30-2015, 09:37 AM | #160 | ||||
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Since I mis-measured the static spring compression, I'll need to recalculate the measurements to determine whether a 5" or 6" spring will be more suitable and which, if any, helper spring will be required. Although, even at ~130mm static spring length for the Z60-178-70 7" springs, I have not experienced, and there is no evidence of, coil bind. Considering my previous measurements indicated a static spring length of 138-163mm, whereas it is actually 129mm/134mm, it points to shorter springs. The physical evidence also points to a shorter spring - it just doesn't seem right to have so much static spring compression. My thoughts are that it would be better to have a main/helper spring configuration such that: (i) the main spring is not pre-loaded and the helper spring is fully/mostly compressed at static ride height, (ii) the helper spring extends/decompresses (up to its total free length) at full droop, and (iii) the main spring compresses (without coil binding) during shock/damper compression. Doesn't this make sense? Isn't the spring preload length just the unladen spring length as a result of the lower spring perch location? And, therefore, isn't the issue of lifting the tyre not affected by the spring configuration or amount of preload? i.e. it won't make a difference whether a longer heavily pre-loaded spring is used or a shorter spring plus helper spring combination is used, or how much any of those springs are pre-loaded, the spring will reach its predetermined maximum allowable extended length anyway and it's going to lift the tyre no matter what? |
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04-30-2015, 09:38 AM | #161 | ||
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This would be so it could be compared to a non-helper spring setup and the calculations can be made for the formula, i.e. the static spring compression was 60mm in your car so, if it the spring configuration was changed to a shorter spring plus helper, which had an FCH of 19mm, then the static spring compression of the shorter spring would be 41mm, but we would still need to specify 60mm as the static spring compression. So, wouldn't the useable stroke be additive (with the helper spring's FCH) in nature, understanding that the spring rate is not additive? It seems like the helper spring length/FCH should factor in to some degree but I'm not sure how. |
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04-30-2015, 09:41 AM | #162 | ||
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1.0-2.0 Hz - Sports cars (street) 2.0-2.5 Hz - Dual duty street/track cars 2.5-3.5 Hz - Race cars without downforce The problem is that it is all subjective. What one person deems harsh may be acceptable to another - within reason. While not exact measurements for my car, I have reused the same tyre rate (2200 lbf/in), front bushing rate (7 lbf/in) and rear bushing rate (23 lbf/in) that you quoted. In addition, I have decided to use a uniform 120 lbs for the front/rear unsprung mass, more so because my wheels and brakes are not OE and, while I am not currently in a position to measure the individual components, the lighter wheels and larger (and heavier?) BBK will most likely cancel each other out. Based on this, my current natural frequencies are: Front 1.98 Hz ± 0.03 Hz Rear 1.80 Hz ± 0.04 Hz The quoted figures are left/right averages with the deviation indicating the spread from average for the left and right. My preference is for the rear to be 10% less (or within a range or 0.2-0.3 Hz less) than the front, so the current figures are very close to being within those limits. As per your suspicions, my car is dual duty, although I have made compromises on comfort to target a more track-oriented ride. The Z60-152-100 (6") or Z60-127-100 (5"+Helper) Swift springs options that I was considering would put the front natural frequency at 2.29 Hz ± 0.05 Hz, which is not too extreme for the front but, without also increasing the rear spring rate, the difference between front and rear is ~0.5 Hz, which is more than target. The rear springs would need to be increased to 180-200 N/mm to get back within the desired 0.2-0.3 Hz range. Following are the average front and rear natural frequencies (deviation, front ± 0.04 Hz, rear ± 0.05 Hz) for my car for the specified spring rates: Front N/mm ... Hz 80 ..... 2.09 90 ..... 2.19 100 .... 2.29 Rear N/mm ... Hz 150 .... 1.85 160 .... 1.90 170 .... 1.95 180 .... 1.99 190 .... 2.03 200 .... 2.07 Based on my target front/rear differential, the following front/rear spring rate pairings could be used: F 70, R 140 - current configuration F 80, R 140-160 F 90, R 160-180 F 100, R 180-200 I was advised that my Moton dampers are valved to support spring rates up to 1100 lbf/in (~192 N/mm), although, I'm not sure if that was before or after the motion ratio was applied. To provide a safety buffer, this limit could be reduced by a token 100 lbf/in (~9%), resulting in 1000 lbf/in (~175 N/mm). If the damper valving was based on spring rate before the motion ratio was applied, then this would rule out the rear 180-200 N/mm spring rates and, therefore, the front 100 N/mm spring rate, which I was considering. My decision would then revolve around whether I should maintain the existing spring rates or step them up. Based on spring rates and frequencies alone, I would most likely select Front: 90 N/mm, 2.19 Hz Rear: 180N/mm, 1.99 Hz While the rear 180 N/mm (~1028 lbf/in) spring rate does extend into my self-imposed safety buffer, it is only by a small margin. The alternative would be to use a 170 N/mm (~971 lbf/in) 1.95 Hz rear spring rate. Following all this, I would then need to go back and consider stroke to ensure coil-bind was still not going to be an issue. Thank you for your responses, which made me pay more attention to the ride frequencies. I'd be happy to hear your thoughts. |
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04-30-2015, 09:42 AM | #163 | |
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04-30-2015, 12:49 PM | #164 | |
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04-30-2015, 02:02 PM | #165 | |
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Seriously though, you know up front that if Ohlins recommends setting 10 for street use, nothing much higher (softer) than that matters. What does annoy me is that the guy testing my dampers didn't follow my instructions, which said test from 1 to 13 in increments of two. He covered the whole range in larger increments and counted clicks from full soft instead of full hard like I asked. |
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04-30-2015, 02:04 PM | #166 |
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04-30-2015, 02:18 PM | #167 | |
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04-30-2015, 04:25 PM | #168 |
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fe1rx, I had a question regarding the ohlins torque sheet a few posts earlier. Per the ohlins sheet, the subframe side of the rear guide rod and upper arm are listed under the torque at ride height section. per your m3 rear arm analysis, both of these connections are ball jointed, which means they could be torqued at any height. just want to confirm this is the case or is it still better to torque the ball joints at ride height?
when I did my initial install, I torque the rear arms at ride height, but it would be much easier to torque them at full extension if I had to remove them for some reason.
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04-30-2015, 05:51 PM | #169 | |
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The same argument can be made for the M3 front lateral link (wishbone), however I have noted that there is substantial play in the bolted connections (because BMW uses fully threaded bolts in shear joints for some mysterious reason). Particularly up front it is worth having some weight on the suspension to load that connection in the "more camber" direction when tightening the bolt. |
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05-01-2015, 07:07 AM | #170 | |
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If you want to explore the upper end of the envelope, it should be based on what actually works (i.e. based on testing) not just on math. My point in raising the subject of suspension ride frequencies is that it is a tool for seeing if you are in the right ballpark. Let's say you are bottoming out your front suspension. It is natural to think you need to go stiffer up front, but if your ride frequencies are "right", the problem is actually your ride height. The Ohlins kit as delivered is glaringly out of balance, with a rear ride frequency much lower than the front. I suspect this was a necessary concession to the soft rear subframe bushings. Speaking of which, once you go really stiff, other compliances (tires, bushings, suspension arms, chassis) will defeat your attempt to get the ride frequencies you think you are getting. Race cars are caged, ball-jointed and reinforced to reduce these compliances. I suggest finding someone who actually races a race-prepared 135i and find out what springs they are running. Whatever it is, you probably don't want to go that stiff for a dual duty car. For a practical street driven car I would be skeptical of any car that does not permit at least 4" of total front wheel travel. The OE 135i doesn't have much more than that. It can't afford to lose much (in my opinion). |
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05-01-2015, 12:13 PM | #171 | |
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First, regarding spring rate units, rather than rely on some vendor's table, I have always used 2 conversion factors to get to lb/in from the metric system:
1 N/mm = 5.71 lb/in 1 kg/mm = 56.00 lb/in You can find these on "unit converion calculators" on the internet. It's been my practice to check unusual metric units or formulas before I use them. An example is the Puhn formula for sway bars. I derived it from beam formulas before using it. You also get a feel for the limitations. Second, speaking of sway bars, it is somewhat academic to discuss coil spring travel limits, without considering sway bar stiffness, which often is the major contribution to weight transfer, and prevents the struts from seeing full travel when a wheel lifts on a corner. Ex: Quote:
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tracked cars, 68 Triumph GT6, 81 Porsche 924T, 93 Rx7 Last edited by KevinK2; 05-01-2015 at 06:07 PM.. |
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05-02-2015, 08:33 PM | #172 | |
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I just finished one track day and they seem to have gotten perhaps a bit louder, but they still sound very civilized. Noise from road imperfections is not at all objectionable. Bridgestone has done an amazing job with the road manners of this tire. Who knows how they will age, but I think you will see these tires as an improvement over the ZIIs. One odd thing though, my commute to the track includes some grooved highway sections. In some of these sections the tires developed a strange rumble that had me stopping to check out if one was loosing pressure. That section of highway seems to be special though - I have experienced some winter tires (different vehicle) develop a directional uneasiness in the same area. Dry grip is very good. Even at full tread depth they feel like an R-compound tires. This was a working day at the track for me and I was running street brake pads so I had limited opportunity to test them out and no opportunity to experiment with tire pressures. I used 36 psi cold. This is on 235/40R18 all around. Steering response is really direct and confidence inspiring. This was by no means a 10/10ths day though so I can't say how they will stand up to extended lapping in high temperatures. |
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05-03-2015, 04:24 AM | #173 | |
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05-03-2015, 06:29 AM | #174 |
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I am very happy with the change. The car corners flatter, takes a set quicker, and handles transitions better. More important, I have definitely not gone too far aft in terms of roll stiffness distribution because the car is easy to drive with nannies off (my normal mode). With them on, intervention seems much more subtle (although I didn't fully explore this). My hypothesis is that the improved balance and perhaps by reducing the nasty transients due to excessive body roll may result in the DSC seeing less need to intervene. I will try a faster few laps with DSC on today to see if this is in fact the case.
The car seems to handle curbing a bit better than last year, but this is probably in part due to the fact that I am now making proper use of the front bump stops. One wheel bumps are not at all problematic, but this is not too surprising because the front total roll stiffness is still higher than the rear. Before heading onto the track I made a few laps of a skid pad to make sure I wasn't going to be surprised. Limit cornering attitude is very controllable with throttle. The car is very easy to drive, which is what I found with last year's setup too. Clearly the car still understeers at the limit, or it wouldn't be easy to drive, but that characteristic is so mild as to be unobtrusive. I will log some more laps today with the AIM Solo DL. I have some competition data from last year at the same track (on NT01s) so should be able to get a good comparison, although I am hobbled by OEM brake pads today. I will be able to see how much eDiff activity there is when pushing a bit harder by checking the data logs. |
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05-03-2015, 08:20 AM | #175 | |
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05-03-2015, 09:59 PM | #176 |
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Unfortunately this is a bit of apples-to-oranges comparison as the track layouts differ a bit between runs. Here is a GPS trace of my fastest race lap (blue) from last year on this track, and my fastest practice lap (red) from today on the same track, less the carousel on the returning straight. Up to that deviation (turn 10) the data can be compared.
The data is displayed using the segments depicted below with the start line being at the top of the map and the track run clockwise. Speed traces are comparable, with a bit better launch through the start line in the competition. The competition was run with the Cobb tune removed. Today was run at Stage 1 - Drive, which really doesn't make much difference. I run it primarily for the linear throttle. The blue run includes a downshift to 2nd in turn 4, but is run in 3rd in the red run. The competition run (blue) was run with DTC 60/70 pads, which explains the higher in maximum braking g. Maximum lateral g are comparable between the NT01 R-compounds and the RE-71R Extreme Performance street tires. Data logging is with an AIM Solo DL, collecting chassis data through the CAN Bus. I log a math channel I call E Diff Activity, which is the difference between left and right rear brake pressures. Any difference is due to intervention of the E Diff. Not unexpectedly the stiff rear bar does increase E Diff activity in slow corners. Rear brake pressures due to E Diff are in the order of 15 bar. A point of interest, given that the deceleration rates are comparable between the two laps: with OE pads, maximum brake pressures are typically about 20 bar higher than for the DTC pads under normal braking. This being because the track pads have more bite. Last edited by fe1rx; 05-04-2015 at 08:17 PM.. Reason: Today's tune was Stage 1 - Drive, not Stage 1 - Sport |
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