EMD F7 in SCALE

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Steggy
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Re: EMD F7 in SCALE

Post by Steggy »

ccvstmr wrote:...and then, good old-fashioned, ordinary common sense, kick in to fill the gaps where science and engineering couldn't explain why it didn't work! {sorry Bill, couldn't resist.}
:roll: I thought it was the sales department who was responsible for making the excuses when it doesn't work. :roll:
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EMD F7 in SCALE

Post by Steggy »

EMD F7 in SCALE
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The concept of Diesel-electric railroad propulsion dates back about 100 years, and came to maturity in the late 1930s with the introduction of EMD's 5400 horsepower FT demonstrator. Although it was apparent to the early designers that using electricity to transfer power from a Diesel engine to the driving wheels of a locomotive would be very efficient, a satisfactory method of doing so proved to be elusive and came in stages, beginning with the understanding of the substantial differences between a Diesel engine's power vs. RPM characteristic and a direct current generator's power vs. RPM characteristic.

Early attempts to implement Diesel-electric propulsion focused on adapting the Ward Leonard elevator drive control system that had been so successful in skyscrapers. The Ward Leonard drive had to smoothly operate a DC motor against a high inertia load, which suggested to designers that it might be a suitable way to propel a locomotive. It wasn't, mainly because the locomotive's power source, a DC generator being driven at various speeds by a Diesel engine, doesn't behave anything like the utility power source that runs an elevator.

Research, predominantly at General Electric, eventually showed the way, with one of the most important breakthroughs being developed by GE engineer Hermann Lemp. He devised a control system operated by a single lever that correctly excited the generator according to the engine's power output, thus producing the means to smoothly match propulsion loading to engine operating conditions. Lemp's system became the model for all subsequent Diesel-electric propulsion designs, including the propulsion system used on American World War II submarines.

The reason for this history lesson is to highlight that Diesel-electric propulsion is a complicated thing to implement, and gets more difficult when attempts are made to scale it down to the level of a model locomotive. This doesn't mean it can't be done. Early on in doing the research and design for my F-unit, I evaluated the practicality of using electric power transmission for propulsion. As I previously had written, one of my overall goals was:
  • Control response similar to the prototype....I was hoping to achieve something that would behave somewhat as though it was controlled in the same fashion.
So, my thinking went, what I need is a suitable DC generator, suitable DC motors, suitable switchgear, and a means of matching generator loading to prime mover power output. I could design and build the control package, so all I really needed was a generator and some motors. I would, of course, have to work out how the motors would be mounted in the trucks and geared to the axles, but that sort of mechanical stuff isn't too much of a challenge.

The challenge was in procuring a generator and some motors. I quickly discovered that a "suitable generator" was as elusive as a unicorn. All the generators I could find either lacked the required electrical characteristics or seemed to be nearly as big and heavy as the generator in the real F7. It was a similar story with traction motors, as all the ones that could be fitted into the confines of a 1.6" scale truck were permanent magnet types, thus lacking the characteristics of real traction motors. They also were a bit feeble horsepower-wise for my application.

As I pondered this, wondering if I should perhaps reevaluate my goals, I experienced an epiphany of sorts. The propulsion system didn't have to BE electric. It merely had to BEHAVE as though it were electric. More cogitation convinced me that I could achieve the effect of electric propulsion with fluid power, but not with a system that was like anything found in most model locomotives.

A key characteristic of Diesel-electric propulsion is there is no fixed relationship between primer mover RPM and locomotive speed. So I had to figure how to establish that characteristic using, if possible, off-the-shelf components. The solution was to be found in a Reuland Electric hydrodynamic coupling, an item designed for use in electrically-driven machinery to achieve smooth starts against high inertia loads, such as would be encountered in ball mills, conveyors and overhead cranes. Typically, the three-phase motors that power these machines exhibit a very steep starting torque characteristic, which means the motor cannot be directly coupled to the reduction gear assembly if the transmission of damaging torsional shock to the rest of machine is to be avoided.

The Reuland coupling meets this requirement via an ingenious design that keeps oil from the turbines blades until the coupling is spinning fairly rapidly. At that point, oil is forced through ports into the turbine blades by centrifugal force, causing the coupling to gradually "lock up" and the output shaft to gradually catch up with the input shaft. The result is a smooth start, no matter how suddenly initial torque is applied to the input shaft. Being a fluid coupling, of course, it will not achieve 100 percent lockup, which was a characteristic I was seeking.

The Reuland coupling gets its input through a collar that is bolted to the housing and engaged to the output shaft of the prime mover. Hence when the prime mover is started the housing rotates with it. This arrangement has a somewhat obscure secondary benefit over and above the coupling's basic properties, in that the housing and the oil within add rotating mass, which flywheel effect is always a useful thing to have with internal combustion engines, especially one that does not have an even firing order.
Reuland Type 72 Hydrodynamic Coupling Cross-section
Reuland Type 72 Hydrodynamic Coupling Cross-section
Some back-and-forth conversation with Reuland got me the information I needed to get started in figuring out how to adapt one of their couplings to my application—couplings are available in a number of sizes whose torque vs. RPM characteristics vary widely. Apparently, at the time of my inquiry no one had mated one of their couplings with an internal combustion engine, so the only performance data they could give me was based on usage with three-phase motors. It took some research on my part to correlate that data to the Briggs V-twin's behavior, since a typical three-phase, 1800 RPM squirrel cage motor's torque vs. RPM curve is quite a bit different than that of a gas engine of the same power output. As it turns out, my data conversion calculations ended up being reasonably accurate, but that's getting ahead of the story. I ended up with selecting Reuland's type 72 coupling, of which a cross-sectional view is immediately above. Reuland can furnish these units with a range of output shaft sizes, as well as various sized input collars, so mechanical adaptation would not be a problem for me.

Important as the coupling would be in achieving the desired performance, it wasn't the only aspect to be considered. My plan was to direct the output of the coupling to a gear pump, which would pressurize a hydraulic motor in each truck. This part of the propulsion system is relatively conventional, but varies in the manner in which oil flow is controlled. Any hydrostatic propulsion system essentially behaves like shafting and gears (or a belt and pulleys), as the rotational speed of a hydraulic motor is directly proportional to the speed of the pump that is pressurizing it. In other words, a hydrostatic drive is very "stiff," which means a sudden change in torque input to the pump will be immediately reflected at the motor(s) as a sudden change in the torque being applied to the load. If said motor(s) is(are) propelling a locomotive, lurching will occur, as anyone who has ever operated a large scale train and has accidentally move the controls a bit too suddenly knows.

There are several ways to mitigate the lurching issue. In most hydrostatic propulsion systems, torque changes are moderated by diverting some of the pump output back to the reservoir. Unfortunately, that diverted oil flow contains energy which is subsequently wasted as heat. Not only is that inefficient, it shortens the life of components, especial the elastomer seals in the pump and motors. I wanted to avoid that to maximum extent possible.

So, the plan was to use valves that are two-state—open or closed—to engage or disengage propulsion, as well as set direction of travel. By pneumatically controlling them, I could regulate how quickly the valves shifted and thus would achieve smooth operation. My "proof-of-concept" design used two valves: the reverser valve, which has three positions; and the propulsion load valve, which is either opened or closed. Actuation of these valves is via air cylinders with flow control valves to regulate the rate at which the cylinders cycle. Solenoid valves direct air pressure to one cylinder end or the other as required.

The sequence for setting the locomotive into motion is to shift the reverser valve to the desired direction of travel and then gradually close the propulsion load valve. When open, the load valve "short circuits" the traction motor oil circuit so that the motors cannot be pressurized (this arrangement is also a form of automatic drift when the unit is not running). As the valve is closed, pressure builds and the locomotive starts moving. At initial start, the hydrodynamic coupling slips a lot and the train gradually starts to move. As speed increases, the slippage reduces. If starting on an upgrade with a long train, it is often necessary to notch out the throttle before motion begins, just like on a real locomotive.

As one might expect, there is a little more to the machinery than what might be seen under the hood of a Roll Models or Backyard Rails unit. The following pictures, taken as the power assembly was being built up, give you an idea.

Power Assembly #1
Power Assembly #1
Above is the Briggs V-twin ready for power transmission components. Note the four studs protruding from the crankcase cover, which are for attaching the rest of the power assembly. Use of the studs makes assembly a bit easier and also greatly reduces the likelihood of one of the tapped holes being damaged. Also note the remote oil filter adapter on the right hand side of the crankcase, with yellow caps protecting the tubing fittings.

Power Assembly #2
Power Assembly #2
Above is the power take-off (PTO) adapter assembly attached to the V-twin. The adapter plate was machined from a piece of 3/4 inch thick cold finished steel, and engages four pilots on the crankcase cover to assure concentricity with the crankshaft. The four tie rods were machined from 1-1/8 inch diameter C1144 TGP steel and are seated in counterbores to maintain alignment. The tie rods are secured to the adapter plate with 1/2"-13 flat head socket screws torqued to 110 lb/ft. I used steel to make all of the power assembly components to assure dimensional stability and to make things heavy—weight, of course, is your friend in a locomotive.

Power Assembly #3
Power Assembly #3
Above is with the hydrodynamic coupling installed. The coupling's input collar is a very precise fit on the V-twin's crankshaft, so precise, in fact, that there is a relief hole in the collar to allow for atmospheric pressure compensation when installing or removing the coupling (a rubber mallet is a necessary assembly tool). A 1/4 inch square key made from cold finished steel connects the crankshaft with the coupling. A standard shaft collar on the crankshaft keeps the key from drifting out of the coupling's collar when there is little or no loading on the coupling.

Note the flange bearing on the coupling's output shaft (right side of picture). I used a Browning industrial product here, which is designed for sustained high RPM operation under load and can be maintained with a conventional grease gun. Actual loading on this bearing is small, but a little overkill with a component that is not readily removed and replaced is usually a good idea. Also partially visible at the right hand side is part of the jaw coupling that engages with the pump.

Power Assembly #4
Power Assembly #4
Above is with the coupling support attached. The support is secured to the PTO tie rods with 1/2"-13 grade eight capscrews, torqued to 110 lb/ft, producing a very rigid assembly. Note the counterbores in the coupling support for accepting the next part of the assembly. As with the PTO adapter plate, the coupling support was machined from 3/4 inch thick cold finished steel. The flange bearing on the coupling output shaft is bolted to the other side. Also note the jaw coupling half protruding through the support plate.

The next post will continue this series.
Last edited by Steggy on Mon Oct 26, 2015 11:37 am, edited 3 times in total.
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Music isn’t at all difficult.  All you gotta do is play the right notes at the right time!  :D
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Steggy
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EMD F7 in SCALE

Post by Steggy »

EMD F7 in SCALE
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Continuing from the above post...
Power Assembly #5
Power Assembly #5
Power Assembly #6
Power Assembly #6
In the above views, the pump adapter has been mounted. It is spaced from the coupling support plate by four precisely machined sleeves made from C1144 TGP. The 1/2"-13 grade eight capscrews go completely through the sleeves, are threaded into the coupling support and torqued to 110 lb/ft. Note the Hytrel insert in the jaw coupling and also the bolts inserted through the pump adapter to secure the pump. You can also see how the coupling support bearing has been secured.

Power Assembly #7
Power Assembly #7
Power Assembly #8
Power Assembly #8
In the above views, the pump has been mounted to the pump adapter. This particular pump is a dual-section type originally designed for use on construction equipment. The two sections are of an unequal displacement and are part of the transition function in the propulsion system. The pump has a standard SAE 'B' mounting flange that pilots on the pump adapter. Were I to decide to try out a different pump with, for example, an SAE 'A' mounting flange, I could do so by installing a different pump adapter.

Visible in the first picture on the side of the pump is the inlet fitting, which gets plumbed via high temperature hose to a suction strainer in the oil reservoir. In the second picture, the two discharge fittings, both of which are -8 JIC, are visible. The pump is actually mounted inverted in order to clear the locomotive's frame, which will be more apparent when I post pictures of the running chassis.

Power Assembly #9
Power Assembly #9
In the above view, the control valve assembly has been mounted. The valve assembly base was made from a piece of 1/2 inch thick cold finished steel and is secured to the top of the pump adapter with two 7/16"-14 grade eight capscrews torqued to 75 lb/ft. Visible at the center left of the picture is the reverser valve assembly, which is a reworked manually-controlled spool valve assembly produced by Energy Manufacturing. It is cycled between one of three positions by the two double acting air cylinders mounted above it (in a bit of oddity, shifting the valve toward the rear of the locomotive makes the unit go forward). The black pipe with the red cap in the background is the return discharge from the reverser valve, which is plumbed to a filter on the oil reservoir via high temperature hose.

To the right of the reverser valve are the propulsion load and transition enable valves. These valves are also reworked from stock products so they can be cycled with air cylinders. The flow control valves that regulate the cylinder stroke rate are visible at the ends of the cylinders themselves. Regulation is achieved via the principle of differential pressure, which makes the cycle rate mostly independent of applied air pressure.

Also visible are two of the four solenoid valves that control this mess. Excepting the air brakes, everything on the locomotive is electrically controlled, hence the need for solenoid valves.

More pictures will follow in the next post.
Last edited by Steggy on Sun Oct 25, 2015 7:44 pm, edited 2 times in total.
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Music isn’t at all difficult.  All you gotta do is play the right notes at the right time!  :D
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Steggy
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EMD F7 in SCALE

Post by Steggy »

EMD F7 in SCALE
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Continuing from the above post...
Power Assembly #10
Power Assembly #10
Above is a view of the fireman's side of the power assembly. Visible are two more solenoid valves and the inlet and discharge fittings of the propulsion oil circuit.

Power Assembly #11
Power Assembly #11
Power Assembly #12
Power Assembly #12
Above is are views of the completed power assembly, the first view of the engineer's side and the second one of the fireman's side. All hydraulic and pneumatic piping is DOM stainless steel tubing flared to JIC standards, 1/2 inch diameter .049 inch tubing for hydraulic circuits and 1/4 inch .035 inch wall tubing for pneumatic circuits. I do not use hoses unless flexibility is required, such as at the trucks, as most hydraulic system failures are due to burst hoses. Also, hose introduces unavoidable losses due to the transition between the end fittings and the hose itself. In addition to inserting lower pumping loss into the circuit, tubing has the added advantage of acting like a radiator and dissipating waste heat.

Note how the engine oil filter is mounted on the side of the coupling support on the engineer's side. It is piped back to the remote filter adapter on the V-twin with 5/16 inch diameter .035 inch wall DOM stainless tubing, also using JIC fittings.

Power Assembly #13
Power Assembly #13
This is a closer view of the engineer's side of the power assembly. The long hexagonal part in the plumbing above and to the left of the pump is a check valve.

The complete power assembly as seen above weighs approximately 200 pounds. How it fits into the overall scheme of things will be the subject of later posts.
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Music isn’t at all difficult.  All you gotta do is play the right notes at the right time!  :D
Andrew Pugh
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Re: EMD F7 in SCALE

Post by Andrew Pugh »

Nice piece of work!

Looking forward to your next posts in the series.

I have not seen this approach before, and find it very interesting.

Are you by chance a mechanical engineer?

AP
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Steggy
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EMD F7 in SCALE

Post by Steggy »

EMD F7 in SCALE
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This posting is a brief digression. I don't plan to post a lot of details about how I made this, or reworked that—TMI tends to turn off readers. However, since Chaski's has a lot of "machinery in general" subject matter, someone might find this interesting.

One of the design problems in my propulsion control system that had to be solved was how to pneumatically actuate a three-position spool valve that had been designed for manual operation. The original "proof of concept" system as designed used a piloted three-position solenoid valve (pictured below) that proved to be less than satisfactory. It inserted too much pumping loss into the system, was bulky and was power-hungry. Due to its bulk, I was limited in where I could place it in the power assembly, and ended up with it at a relatively high elevation above the railhead. That raised the power assembly's center-of-gravity more than I liked, as the higher up the locomotive's weight mass, the more likely it is the unit will derail while running on uneven track

Electrically-Operated Reverser Valve
Electrically-Operated Reverser Valve

After test-running the loco for a while with this valve, I decided I needed to find a different type of valve. I made the decision to use a pneumatically-actuated valve to get better flow characteristics—the valve would not need a pilot circuit, as the air cylinder could generate enough force to shift the valve spool while under pressure. Going pneumatic would lower the continuous electrical load on the alternator and thus the prime mover would be relieved of a small amount of parasitic drag. It should be noted that the traction motor contactors responsible for setting direction of travel in the real F7 were pneumatically actuated. So by pneumatically actuating my reverser valve I would be more closely modeling the prototype, although that was coincidental, not intentional.

Before continuing onward, it is useful to know that most solenoid-operated high pressure valves use a pilot circuit within to generate the force required to actuate the valve against system pressure. This design feature helps to keep the solenoid's power consumption down to a reasonable level, but has the unfortunate side effect of producing some restriction and thus stealing energy that should be doing useful work. The valve I had been using was guilty of that behavior and, I thus decided, had to go. Its physical height, as it turned out, was going to be a problem with internal body clearances, further convincing me to find another way.

The particular valve I decided to use is an Energy DCVA type, which is operated by a lever and detented at each position. This valve is a so-called “open center” type, which means that when in the neutral position, all four ports are open to each other, preventing the buildup of pressure at any point. This was exactly the behavior I needed.

The DCVA valve has an ample cast iron body and an internal relief valve. Although it was unlikely that pressure in my propulsion system would ever reach the level required to crack the relief valve, it was there if needed. Having a cast iron body meant that the valve assembly would be dimensionally stable with normal temperature changes, helping to minimize efficiency-robbing internal blow-by. The ends of the spool are sealed with standard sized O-rings, which are held in place by simple retainers—there is no high pressure on the O-rings. The pressure and return ports are 3/4 inch NPT female and the two working ports are 1/2 inch NPT female. I'm no fan of pipe threads in high pressure hydraulics, as they tend to present sealing problems if everything isn't just right, but that is how the valve was made.

Upon getting one of these valves, my first step, of course, was to disassemble it and examine its innards. I discovered that the detent mechanism could be readily detached from the spool, which was good, since I wasn't going to need it in my application. However, before removing the detent mechanism I carefully measured the amount of spool travel between positions, which the dial indicator said was exactly 0.250 inches in both directions. Hmm...quarter inch? How convenient! :D

One of my design requirements was that valve actuation would be solely by air pressure—no springs allowed. Even the best springs weaken with repeated cycling, and vary from one to the next in compression or extension force. So I had to concoct an actuating means that would precisely shift the spool a quarter inch in either direction and automatically seek the center neutral position in the event control power to the pneumatic solenoid valves was lost.

After filling up the wastebasket in my office with sketches, I came up with the idea of using two half inch stroke cylinders offset in such a way that when one is extended and the other is retracted, the valve will be in neutral. Extending the retracted cylinder would shift the valve spool quarter inch out from neutral. Retracting the extended cylinder would shift the valve spool a quarter inch in from neutral. The half inch travel of each cylinder accounts for the quarter inch of spool movement needed plus another quarter inch to give the other cylinder travel room. Conveniently, half inch is the shortest stroke double-acting cylinder available in the bore size I needed—I would not need to devise a means of artificially constraining cylinder stroke. Here's the layout drawing I made of the mechanism:

Reverser Valve Actuation Layout Drawing
Reverser Valve Actuation Layout Drawing

The actual assembly is in the following pictures, although it slightly varies from my layout drawing.

Reverser Valve Actuation #1
Reverser Valve Actuation #1
Reverser Valve Actuation #2
Reverser Valve Actuation #2
Reverser Valve Actuation #3
Reverser Valve Actuation #3

The cylinder mounting bracket, which was cut out of a piece of 3/8 inch cold finished steel, replaces a cast aluminum bracket that both mounted the valve's original actuating lever and held the O-ring retainer into the end of the valve body. The opposite end of the valve body now has a short aluminum cap that replaces the original piece, which is what enclosed the no-longer-present detent mechanism.

A trunnion that is driven by the cylinders is affixed to the exposed end of the valve spool, secured with a common quarter inch hitch pin and spring clip. The trunnion itself consists of a piece of 3/8 inch cold finished steel that has been fitted with a hub machined to be a precise fit on the end of the spool. The hub is a press fit into the trunnion plate and then welded. The press fit assured that the plate would remain exactly perpendicular to the hub bore’s centerline during welding—even a relatively small error would cause the mechanism to bind.

Plungers engage the trunnion to drive it back and forth, the plungers having been machined from type 303 stainless steel. The main part of the plunger threads onto the cylinder rod and a thick retaining washer on the inner end is tightened against the end of the plunger by a jam nut. The two flats machined into the one end of the plunger are for holding and turning it with a wrench to make adjustments. There is 0.010 inch nominal clearance between the plunger's OD and the mating bore in the trunnion plate. A dab of wheel bearing grease suffices for lubrication.

The cylinder bore size was determined by calculating the amount of force that would be needed to shift the valve spool under maximum pressure, which I deemed to be 1800 PSI (that number comes from the fact that pressure peaks at the point where incipient wheelslip is present). The locomotive's nominal main air reservoir pressure is 100 PSI, but could be lower under some circumstances. So I added 25 percent to the 1800 PSI number to account variations in air pressure, as well as any error I might have made in computing the reverser valve spool area that was directly exposed to high pressure. I ended up using 1-1/2 inch bore cylinders, as the net force produced by the rod end of the cylinder, being lower than that produced by the opposite end, was only a couple of percent higher that what I had calculated. Testing showed that the cylinders could successfully shift the valve on as little as 60 PSI of air.

I had achieved the goal of building a reverser valve assembly that used only air pressure for actuation.
Last edited by Steggy on Tue Jan 10, 2023 12:04 pm, edited 1 time in total.
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Music isn’t at all difficult.  All you gotta do is play the right notes at the right time!  :D
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Steggy
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Re: EMD F7 in SCALE

Post by Steggy »

Andrew Pugh wrote:Nice piece of work!
Thanks!
Are you by chance a mechanical engineer?
No. :)
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Music isn’t at all difficult.  All you gotta do is play the right notes at the right time!  :D
Pontiacguy1
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Re: EMD F7 in SCALE

Post by Pontiacguy1 »

Sounds like you've got what is basically a torque converter/fluid clutch in between the motor and the hydraulic pump. I guess I didn't realize that any were made that were that size and would work for that application, other than something small like for a go-kart or something similar.
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Re: EMD F7 in SCALE

Post by Steggy »

Pontiacguy1 wrote:Sounds like you've got what is basically a torque converter/fluid clutch in between the motor and the hydraulic pump.
It's a hydrodynamic (fluid) coupling, not a clutch or a torque converter. A clutch is a purely mechanical device in which fluid plays no role in power transmission. A torque converter is a variation of a fluid coupling with at least three elements (pump, turbine and stator) and multiplies torque during periods of high slippage, making it very different than a fluid coupling. During the design phase of this project I was trying to find a torque converter whose size and characteristics were suitable for this application. Unfortunately, the smallest non-automotive converter I could find had a stall speed of 1800 RPM with 35 horsepower input, which was too high to be workable.
I guess I didn't realize that any were made that were that size and would work for that application, other than something small like for a go-kart or something similar.
Reuland Electric makes two couplings that are smaller than the one I am using.
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Music isn’t at all difficult.  All you gotta do is play the right notes at the right time!  :D
GeorgeT
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Re: EMD F7 in SCALE

Post by GeorgeT »

Really nice work, would love to see it run.

Why did you directly mount everything to the gas engine? I would be concerned about all the vibration coming for the engine being applied to the hydraulics and the locomotive frame?

Also looking forward to seeing your progress.
GeorgeT
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Steggy
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Re: EMD F7 in SCALE

Post by Steggy »

GeorgeT wrote:Why did you directly mount everything to the gas engine?
When a mass like the hydrodynamic coupling is spinning at several thousand RPM, concentricity with the crankshaft centerline is critical if undue stress on rotating components is to be avoided. The only way to guarantee concentricity is to pilot everything from the registers that are cast into the crankcase PTO cover. Any other method would not be as accurate.

Aside from mechanical considerations, mounting everything to the engine makes it possible to conveniently remove and install the power assembly as a unit.
I would be concerned about all the vibration coming for the engine being applied to the hydraulics and the locomotive frame?
One of the nice things about the Briggs V-twin (and the V-twins made by Kohler, Honda, et al) is its low vibration levels. All reciprocating engines develop two significant types of vibration: torsional and inertial. Torsional vibration is seen at the crankshaft as a periodic variation of the crankshaft's average angular velocity and thus directly affects the load. Torsional vibration is more apparent at low RPM and high loading. Flywheel mass can assist in counteracting torsional vibration.

In a conventional hydrostatic propulsion system, the pump would be directly coupled to the crankshaft and hence would be subjected to the full level of torsional vibration. In the case of my F7's propulsion system, the hydrodynamic coupling absorbs most of the torsional vibration, so it isn't seen by the pump. Also, the coupling adds rotating mass, which further dampens torsional vibration.

Inertial vibration caused by the reciprocating motion in the engine is more difficult to avoid because it is transferred from the crankcase into whatever structure is supporting the engine, which in this case would be the locomotive's frame. The V-twin's inertial vibration is very low for a two cylinder engine and thus is not much of a concern. Incidentally, if the engine is solidly attached to the frame and the frame itself is very rigid and massive the perceived vibration level will be low, because the frame's mass will act as a shock absorber.
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Music isn’t at all difficult.  All you gotta do is play the right notes at the right time!  :D
Andrew Pugh
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Re: EMD F7 in SCALE

Post by Andrew Pugh »

BigDumbDinosaur wrote: When a mass like the hydrodynamic coupling is spinning at several thousand RPM...
Wow, how high are you spinning that thing? :mrgreen: Looking forward to your next part of the series! Have you shared your name, or shall we continue to refer to you as BigDumbDinosaur?

Andrew
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