Giving Slip the Slip: Lockup Torque Converters and Split Torque Automatic Transmissions

Fluid clutches — fluid couplings and torque converters — have many advantages for automotive transmissions, but with those benefits comes a cost: fuel-wasting hydraulic slippage even at cruising speed. Since the 1940s, automakers have come up with a variety of strategies for reducing or eliminating that slip, including series parallel “split torque” transmissions and different types of converter lockup clutches. In this installment of Ate Up With Motor, we take a look at how GM, Ford, Chrysler, Packard, and Studebaker have approached this slippery problem from 1949 through the late eighties.


One of the fundamental differences between a fluid clutch (a fluid coupling or torque converter) and a mechanical plate clutch is slippage. A healthy plate clutch slips only briefly when engaged or disengaged. Once the clutch plate is fully engaged against the pressure plate, both must turn together at the same speed. This is why a disc clutch must be disengaged or the transmission shifted to neutral whenever the vehicle stops. If the engine isn’t free to turn faster than the driveshaft at rest or extremely low speeds, the engine will stall!

By contrast, a fluid clutch always slips at least a little. When the vehicle is at rest, the fluid clutch slips enough to allow the engine to idle with the transmission in gear without stalling. When starting, the fluid clutch’s driving torus (the impeller) may reach a speed of 2,000 rpm or more before the driven torus (the turbine) begins to move at all, a point known as stall. Once the vehicle is moving at a constant speed, the speed difference between the impeller and turbine diminishes, eventually reaching a minimum point known as coupling stage. However, even at coupling stage, the turbine still turns somewhat slower than the impeller. For example, if engine speed is 2,500 rpm, the turbine might only turn 2,375 rpm: hydraulic slippage of 5%. Unlike a reduction gear, this speed reduction doesn’t multiply engine torque. The speed difference between impeller and turbine is simply lost to heat within the operating fluid.

At idle and very low road speeds, slippage is desirable because it keeps the engine from stalling or lugging in gear. Hydraulic slippage also offers some advantages during acceleration: As with a plate clutch, a certain amount of slip makes for a much smoother takeoff. Also, as we explained in the sidebar of our article on GM’s other early automatic transmissions, torque converters utilize the speed difference between the driving and driven torus members to multiply engine torque.

Hydraulic slippage is much less desirable at cruising speed. First, that slip wastes some fuel — you’re “paying” for more engine revolutions (and more power) than reach the driveshaft. Second, the nonlinear relationship between impeller and turbine speed can be troublesome. Turbine speed tends to fall behind impeller speed any time load increases (for example, when going up a steep grade) and an increase in engine speed doesn’t immediately produce a corresponding increase in turbine speed.

Fluid clutches also don’t allow much engine braking. When you coast in gear, the transmission attempts to drive the engine crankshaft, whose inertia causes a braking effect. With a plate clutch, that effect is often quite pronounced, especially in a reduction gear, because of the mechanical connection between the engine flywheel and the transmission input shaft; neither can overrun the other. That isn’t true of a fluid clutch, whose torus members are free to rotate at different speeds. When coasting, the turbine overruns the impeller, which causes slippage, but relatively little braking effect. This is why lifting off the throttle in a car with a fluid coupling or torque converter can feel almost like shifting into neutral.


One way to minimize coupling slip is to supplement the fluid clutch with a plate clutch that engages when the vehicle reaches cruising speed.

There are several ways to arrange such a clutch, but the common object is to create a mechanical connection between the engine and the transmission input shaft (which otherwise is driven by the fluid clutch’s turbine(s)). When that connection is fully engaged, the fluid clutch’s torus members are effectively locked together — they must turn together at engine speed — and transmit no torque. All engine power flows through the mechanical clutch directly to the input shaft. As long as this mechanical lockup remains fully engaged, there is no slip, which improves fuel economy (typically by around 4% at cruising speeds) and provides throttle response and engine braking comparable to a vehicle with a manual gearbox.

Color diagram of three-element torque converter with lockup clutch © 2016–2017 Aaron Severson

The logic of these schematics, which are not to scale and which have been greatly simplified for clarity, is that each color represents a set of components that are integral or otherwise connected so that they always rotate together; clutches, brake bands, planet pinions, and flywheels are shown in black or gray. In this schematic of a basic three-element torque converter with lockup clutch, the torus cover, lockup clutch pressure plate, and impeller (red) all rotate with the engine flywheel. The turbine and lockup clutch driven plate are both splined to the transmission input shaft (medium blue). The one-way stator clutch (fuchsia) allows the stator (light blue) to freewheel in the direction of engine rotation, but reverse rotation locks the stator against the fixed stator shaft (dark gray), providing torque multiplication under some conditions. (Author diagram)

However, such a lockup clutch can’t remain locked all the time. For one, the engine would stall every time the vehicle came to a halt with the lockup clutch engaged unless the driver shifted to neutral each time. Also, engaging a lockup clutch will prevent a torque converter from multiplying torque. Therefore, a lockup clutch needs to be accompanied by some mechanism to selectively engage or disengage the clutch.

On some prewar bus and rail car transmissions with lockup torque converters, the lockup clutch was engaged manually, but automotive lockup clutches are designed to engage and disengage automatically. The usual strategy is to disengage the clutch at idle (or just off idle) and during acceleration and then then engage the clutch for cruising.


Lockup clutches were used in some nonautomotive torque converters before World War II, but the first regular-production automotive application was the 1949 Packard Ultramatic Drive, which established a model for subsequent transmission designs. It was followed in fairly short order by the 1950 Studebaker Automatic Drive, which took a somewhat different approach to the same problem.


Ultramatic’s lockup clutch was a hydraulically operated “wet” clutch located within the torque converter housing, between the engine flywheel and the converter turbine. The clutch plate itself, which was faced on both sides with cork, was splined at its hub to the transmission input shaft. The space ahead of the clutch formed a hydraulic cylinder containing an annular steel piston that acted as a pressure plate. Filling the cylinder with pressurized oil would lock the pressure plate to the clutch plate itself, forcing the clutch and input shaft to rotate with the torus cover at engine speed. If oil pressure to the cylinder was relieved, the pressure plate would no longer be held against the clutch plate, disengaging the clutch and allowing the torque converter to function normally.

1949 Packard Ultramatic transmission torque converter and lockup clutch schematic © 2016–2017 Aaron Severson

The original Packard Ultramatic had a four-element torque converter with a single stator (light blue) interposed between the first- and second-stage turbines, providing a stall ratio of 2.4:1 at 1,600 rpm. (Both turbines, here colored medium blue, were bolted together into a single unit, but had different blade profiles.) The lockup clutch control valve, not pictured, allows pressurized oil to enter the clutch cylinder through a hollow passage within the input shaft, pushing the pressure plate (colored red, like the torus cover and impeller it rotates with) into the engaged position. (Author diagram)

Like the contemporary Buick Dynaflow, Ultramatic was designed to operate most of the time in its direct drive top gear. Since the transmission was largely dependent on the torque converter for torque multiplication, it was crucial that the lockup clutch not engage prematurely and that there be some means to immediately disengage it for acceleration.

Packard achieved that goal by treating the engagement of the lockup clutch like an automatic gear change. A spring-loaded plunger valve controlled oil flow to the lockup clutch’s hydraulic cylinder. The valve was pushed open or closed by opposing pressures generated by a throttle-controlled valve and a centrifugal governor driven by the output shaft. Once governor pressure was sufficient to overcome the combined spring and throttle valve pressure, oil would be applied to the clutch cylinder, engaging the lockup clutch. If governor pressure dropped below that threshold, the clutch would disengage.

On early Ultramatic-equipped Packards, the lockup clutch would not engage until road speed reached at least 15 mph (24 km/h) even on a light throttle. Opening the throttle further could delay engagement to a maximum of 56 mph (90 km/h). To avoid stalling or lugging the engine, the clutch would automatically disengage if road speed fell below 13 mph (21 km/h). There was also a “kickdown” mechanism that allowed the driver to disengage the clutch at road speeds between 13 and 56 mph (21 and 90 km/h) by flooring the accelerator, allowing the torque converter to again multiply torque.

1949 Packard Ultramatic transmission schematic © 2016–2017 Aaron Severson

In the Packard Ultramatic transmission, the input shaft drives the hub of the direct drive clutch and the input sun gear of the Ravigneaux gearset (all colored medium blue). In Drive, the direct clutch is engaged and both bands are released, causing the input sun gear, the low sun gear (orange), the planet carrier and output shaft (light green) to turn at input shaft speed. The annulus (purple) is held in reverse, idles in low, and rotates with the sun gears in direct drive. (Author diagram)

Unsurprisingly, each engagement or disengagement of the lockup clutch felt like a shift, giving less technically knowledgeable drivers the mistaken impression that Ultramatic was a two-speed automatic. (Until late 1954, the transmission’s planetary gearset always remain in direct drive unless the driver manually selected Low.) With the clutch engaged, Ultramatic also transmitted some drivetrain vibration that the torque converter would otherwise have absorbed. Although this arrangement sacrificed some of the smoothness that was the main rationale for early automotive torque converter transmissions, it made Ultramatic more efficient than the early Dynaflow, at least at cruising speeds.


Introduced for the 1950 model year, Studebaker’s Automatic Drive was developed in partnership with Borg-Warner and was more broadly known as the DG series, after the Borg-Warner Detroit Gear Division that did much of the development work. The DG series transmissions were notionally three-speed automatics (although early iterations were designed to start in second gear) with a torque converter lockup clutch. However, that clutch was used in a distinctly different way than was Packard’s.

1950 Studebaker Automatic Drive (Borg-Warner DG) transmission schematic © 2016–2017 Aaron Severson

In the early Studebaker Automatic Drive (Borg-Warner DG) transmission, the flywheel drives the torque converter impeller, front pump, and lockup clutch pressure plate (red). The turbine drives the front annulus through the input shaft (medium blue). A reverse band surrounds a brake drum connected to the front carrier, which is also integral with the rear annulus (orange). A multi-disc clutch can lock that carrier to the front sun gear, which is also attached to the rear brake drum (dark red). That drum is connected via overrunning clutch (fuchsia) to the rear sun gear (purple), which is also attached to the inner race of a second overrunning clutch (also fuchsia). A third drum (dark green) allows the overrunning clutch’s outer race to be locked in place by the forward band, which is engaged in all forward gears. The rear planet carrier is attached to the main shaft, which is also splined to the lockup clutch and drives the rear oil pump and governor (medium green). (Author diagram)

The DG series transmissions obtained their indirect ratios with a compound planetary gearset. In first and second gears, the converter turbine drove the front annulus (ring gear). In first, the low band and one-way clutches held both sun gears, causing the ring gears to drive the planet carrier and output shaft at 43.3% of turbine speed (1 / 2.31). In second, the low band was released and the multi-disc clutch engaged, causing both ring gears to rotate at turbine speed. Reaction torque locked the rear sun gear against the rear one-way clutch, driving the planet carrier and output shaft forward at 69.7% (1 / 1.44) of turbine speed. Those gear ratios were in addition to the torque multiplication provided by the torque converter.

1950 Studebaker Automatic Drive (Borg-Warner DG) transmission 2nd gear power flow © 2016–2017 Aaron Severson

Some iterations of the Studebaker Automatic Drive/Borg-Warner DG started in first gear, but early transmissions started in second unless the driver manually selected Low. Starting in Drive provided a maximum starting ratio of 3.10:1 (a mechanical ratio of 1.44:1 times a nominal torque converter stall ratio of 2.16:1 at around 1,600 rpm), which was better than the early Ultramatic or Dynaflow, but still made for lazy performance, especially with Studebaker’s six-cylinder engine. (Author diagram)

Shifting to third gear was accomplished by engaging the torque converter lockup clutch. Once engaged, the lockup clutch simply drove the output shaft at engine speed (not turbine speed). This overran the rear unit planet carrier, unlocking the rear sun gear’s one-way clutch and putting the transmission in direct drive. Unlike Ultramatic and many subsequent torque converter automatics, the converter of the DG series was completely inoperative in top gear, so obtaining any additional torque multiplication required a downshift to second gear.

1950 Studebaker Automatic Drive (Borg-Warner DG) transmission 3rd gear power flow © 2016–2017 Aaron Severson

In third gear, the Studebaker Automatic Drive/Borg-Warner DG’s lockup clutch drives the transmission’s output shaft at engine speed, providing a completely mechanical direct drive with no hydraulic slippage. (Author diagram)

Studebaker-Packard (the two companies merged in 1954) dropped both of these transmissions after the 1956 model year. The replacement was a cheaper and somewhat simpler Borg-Warner unit — which Studebaker-Packard called Flightomatic — that deleted the DG’s lockup clutch. Borg-Warner continued to sell the DG series for several more years to some non-U.S. customers, but U.S. automakers would not rediscover the lockup clutch for another 20 years.


General Motors used lockup clutches in some torque converter bus transmissions, but GM’s early passenger car automatics went a different path, obtaining some of the benefits of the lockup clutch through a novel application of a principle called “split torque.”

As some readers already know, in addition to providing direct drive, reduction and/or overdrive gearing, a planetary gearset can also be used as a differential, either splitting a single source of torque along two paths or combining two torque inputs into one. GM engineer Oliver K. (“O.K.”) Kelley applied the latter concept to the original Hydra-Matic as a way of reducing slippage in certain gears.

To understand the Hydra-Matic’s split torque arrangement, it’s important to first review a couple of basic points about the transmission’s unusual mechanical layout. In the early Hydra-Matic, all power flowed through the intermediate shaft, which was driven by the planet carrier of the front planetary gearset and drove both the fluid coupling impeller (the driving torus) and the hub of the rear multi-disc clutch.

Color diagram of 1940–1947 Model 180 Oldsmobile Hydra-Matic © 2016–2017 Aaron Severson

This diagram of the early Hydra-Matic color-codes each group of mechanical components that always rotate together: torus cover, front oil pump, and front annulus (light green); front planet carrier, intermediate shaft, impeller, and rear clutch hub (red); turbine, main shaft, and rear sun gear (medium blue); front sun gear and brake drum (orange); rear annulus and brake drum and reverse sun gear (light blue); reverse annulus and brake drum (dark green); and rear and reverse planet carriers, rear oil pump/governor, and output shaft (purple). Note that it was the intermediate shaft, not the torus cover, that drove the fluid coupling impeller! (Author diagram)

This meant that the impeller always rotated at intermediate shaft speed, which was not necessarily the same as engine speed. The torus cover, which was bolted to the engine flywheel, drove the front oil pump and the annulus of the front planetary gearset at engine speed. However, if the front brake band was engaged, it locked the front unit sun gear so that the rotation of the annulus forced the planet carrier to orbit the now-stationary sun gear at reduced speed. This also multiplied engine torque; with the front brake engaged, intermediate shaft torque was equal to engine torque times the ratio of the front gearset.

In first, second, and reverse, there was no torque split. The intermediate shaft still drove the rear clutch hub, but with the rear clutch disengaged, the hub just spun idly. Therefore, all intermediate shaft torque was applied to the impeller and then hydraulically transmitted to the turbine, the main shaft, and the sun gear(s) of the rear planetary gearset.

Model 180 Hydra-Matic transmission showing power flow in 2nd gear © 2016–2017 Aaron Severson

When the early Hydra-Matic was in second gear (illustrated), the front clutch was engaged, causing the intermediate shaft and impeller (red) to rotate at engine speed. The rear band was engaged and the rear clutch released, so all power flowed through the turbine and main shaft (blue). Power flow in first was similar, but the front band was engaged and the front clutch released, so the impeller turned slower than the engine. (Author diagram)

In third and fourth, the rear clutch engaged, which locked the rear clutch hub to the rear brake drum, forcing them to turn with the intermediate shaft. Since the drum was affixed to the annulus of the rear planetary gearset, the annulus now also rotated at intermediate shaft speed.

However, the intermediate shaft was also still driving the fluid coupling impeller, which continued to transmit torque to the turbine and the main shaft to the rear sun gear(s). Therefore, intermediate shaft torque was now split between the rear sun gear (through the coupling and the main shaft) and the rear annulus (through the rear clutch). The rear gearset’s planet carrier acted as a differential, combining those torque components and applying the result to the output shaft. O.K. Kelley likened this arrangement to a series parallel electrical circuit.

Model 180 Hydra-Matic transmission showing power flow in 4th gear © 2016–2017 Aaron Severson

In fourth gear, both of the early Hydra-Matic’s bands were released and both clutches were engaged, so the intermediate shaft rotated at engine speed, dividing engine torque between the impeller and the rear clutch. (Author diagram)


Since the intermediate shaft was simultaneously driving both the rear sun gear and the rear annulus any time the rear clutch was engaged, intermediate shaft torque was divided between those gears. The proportion of that split depended on the gears’ respective numbers of teeth and thus their gear ratio.

When torque was applied to the sun gear, the inertia of the output shaft (which was affixed to the planet carrier of the rear planetary gearset) exerted reaction torque on the annulus, attempting to turn it backward. However, with the rear clutch engaged, the annulus couldn’t turn backward because the intermediate shaft was driving it forward. The annulus therefore became a reaction member, multiplying the torque the sun gear applied to the planet carrier.

At the same time, the torque on the annulus and the inertia of the output shaft exerted reaction torque on the sun gear. Again, the sun gear wasn’t free to turn backward since it was being driven forward by the main shaft. Therefore, the sun gear also acted as a reaction member, multiplying the torque the annulus applied to the carrier.

In both cases, the torque applied to each gear had to be sufficient to overcome the reaction torque on that gear. Otherwise, the gear would resist and potentially stall the engine.

1942 Oldsmobile B-44 Special Sixty club coupe Hydra-Matic badge

The version of Hydra-Matic illustrated in the accompanying diagrams is the earliest production iteration, the Model 180, which was optional on Oldsmobiles beginning with the 1940 model year. Cadillac, which added Hydra-Matic as an option for 1941, used the heavy-duty Model 250. (We assume the model numbers represent the transmission’s nominal net torque capacity in pounds-feet.) If you see photos or diagrams of the early Hydra-Matic, the easiest way to tell the difference is that the Model 180 had a single rear sun gear while the Model 250 had a compound rear gearset with two sun gears. (Author photo)

This may become a little easier to grasp if we apply some actual numbers. Let’s consider, for example, the earliest Model 180 Hydra-Matic, the version offered in 1940–1942 Oldsmobiles. That transmission’s rear planetary gearset had a single sun gear with 45 teeth and an annulus with 69 teeth. With the sun gear driving, the rear clutch disengaged, and the rear brake engaged, the rear gearset had a ratio of 2.53:1 (1 + 69/45).

With those gears, the rotation of the sun gear and the resistance of the planet carrier applied reaction torque to the annulus at a ratio of -1.53:1 (0 – 69/45) — that is, they attempted to turn the annulus backward at about 65% (100% / 1.53) of sun gear speed. To overcome that reaction torque, therefore, the annulus had to receive 1.53 times as much torque as the sun gear did.

Since the annulus and the sun gear were both driven by the intermediate shaft, the sum of the torque on the annulus (let’s call it TA) and the torque on the sun gear (which we’ll call TS) had to equal the torque on the intermediate shaft (TI). So, in mathematical terms:

TI = TS + TA

Since we also know that:

TA = TS * 1.53


TI = TS + (TS * 1.53)

… which simplifies to:

TI = TS * 2.53

We can then solve for TS:

TS = TI / 2.53

… and calculate the percentage of intermediate shaft torque applied to the sun gear:

100% / 2.53 = 39.47%

The percentage applied to the annulus is therefore:

100% – 39.47% = 60.52%

As we mentioned above, each gear acted as a reaction member, multiplying the torque the other gear applied to the planet carrier. However, each gear was receiving only a portion of the input torque, so only that portion was multiplied. In this case, torque applied to the sun gear was multiplied by 2.53:1 (1 + annulus teeth / sun gear teeth, or 1 + 69/45). With the annulus driving, the gear ratio was 1.65:1 (1 + sun gear teeth / annulus teeth, or 1 + 45/69), so torque on the annulus was multiplied by that amount.

Both the sun gear and annulus were acting on the same planet carrier, so the torque the sun gear applied to the rear planet carrier and output shaft (let’s call it TC) had to be the same as the torque the annulus applied to the carrier. Or, in mathematical terms:

TC = TS * 2.533 = TA * 1.652

As we determined above, the torque on the sun gear (TS) was 39.47% of the total, and 39.47% times 2.533 (allowing for rounding) is 100%. The product of the torque on the annulus (TA) was 60.52% of the total, and 60.52% times 1.652 is also 100%. Therefore, output shaft torque (TC) equaled 100% of intermediate shaft torque. That meant that output shaft torque also equaled the sum of torque on the sun gear and torque on the annulus, or:

TC = TS + TA

Model 180 Hydra-Matic transmission showing power flow in 4th gear (with percentages) © 2016–2017 Aaron Severson

The gearing of the original Model 180 Hydra-Matic splits intermediate shaft torque 60.5%/39.5 in third and fourth gears. (Author diagram)

Again, in older Hydra-Matic transmissions, intermediate shaft torque was not necessarily the same as engine torque. In third gear, the front brake band was engaged, so intermediate shaft torque was equal to engine torque times the ratio of the front gearset. In an early Hydra-Matic, the front annulus had 54 teeth and the front sun gear had 24 teeth, so with the annulus driving, the gear ratio was 1.44:1 (1 + 24/54). In third, therefore, intermediate shaft torque (which again we can call TI) was engine torque * 1.444. Torque on the rear annulus (TA) was 60.52% of that, or about 87.4% of engine torque (1.444 * 60.52%). Torque on the rear sun gear (TS) was 39.47% of intermediate shaft torque, or approximately 57% of engine torque (1.444 * 39.47%).

For example, if the engine were generating 150 lb-ft (203.4 N-m) of torque, third gear would divide and multiply that torque as follows:

TI = 150 lb-ft [203.4 N-m] * 1.444 = 216.7 lb-ft [293.8 N-m]
TA = TI * 60.52% = 131.1 lb-ft [177.8 N-m]
TS = TI * 39.47% = 85.5 lb-ft [116 N-m]

Torque on the rear carrier and output shaft (TC) was therefore:

TC = 131.1 lb-ft [177.8 N-m] + 85.5 lb-ft [116 N-m] = 216.7 lb-ft [293.8 N-m]

The overall ratio in third, therefore, was 1.44:1 (216.7 / 150).

In fourth, the front band was off and the front clutch was engaged, so intermediate shaft torque equaled engine torque. If engine torque were 150 lb-ft (203.4 N-m), fourth gear would divide that torque as follows:

TI = 150 lb-ft [203.4 N-m] * 1.00 = 150 lb-ft [203.4 N-m]
TA = TI * 60.52% = 90.8 lb-ft [123.1 N-m]
TS = TI * 39.52% = 59.2 lb-ft [80.3 N-m]

Torque on the output shaft was therefore:

TC = 90.8 lb-ft [123.1 N-m] + 59.2 lb-ft [80.3 N-m] = 150 lb-ft [203.4 N-m]

… and the overall ratio in fourth was 1.00:1 (150/150).


Once you’ve finished recoiling from this unwelcome flashback to algebra class, you may be muttering, “What exactly is the point of all this? And what does it have to do with fluid coupling slippage?”

In the early Hydra-Matic, the main shaft was hydraulically driven: It was splined to the fluid coupling turbine. Therefore, the speed of the main shaft and rear sun gear were always reduced by slip within the coupling, causing them to turn slower than the impeller. The intermediate shaft was mechanically driven, so while there were frictional losses, there was no slippage as long the rear clutch was functioning properly.

This meant that with the rear clutch engaged, the rear annulus was always rotating faster than the rear sun gear. In mathematical terms, the velocity of the annulus (let’s call it VA) had to be greater than the velocity of the sun gear (VS). The rear planet carrier “resolved” this speed difference — that is, rotation of the faster-moving annulus forced the carrier to orbit the slower-moving sun gear at some intermediate speed.

The velocity of the carrier and output shaft (let’s call it VC) was proportional to the ratio of the planetary gears and the speed difference between the annulus and sun gear:

VC = VS + ((VA – VS) / (1 + sun gear teeth / annulus teeth))

For example, let’s suppose that a 1940 Oldsmobile equipped with Hydra-Matic is cruising in fourth gear at an engine speed of 2,500 rpm. Let’s assume for the sake of illustration that the fluid coupling is 96% efficient at coupling stage. Discounting mechanical losses, we can therefore assume that the turbine and main shaft rotate at 96% of impeller speed, or 2,400 rpm. The intermediate shaft rotates at impeller speed, which, since the front gearset is in direct drive in fourth gear, is 2,500 rpm.

With the gearing we described above (i.e., a rear sun gear with 45 teeth and a rear annulus with 69 teeth), we can calculate carrier speed as follows:

VC = 2,400 + ((2,500 – 2,400) / (1 + 45/69))

… or:

VC = 2,400 + (100 / 1.65) = 2,460.5 rpm

In other words, the annulus rotating at 2,500 rpm will force the carrier to orbit the sun gear at a speed of approximately 2,460.5 rpm. This reduces effective hydraulic slip from 100 rpm (4%) at the turbine to about 39.5 rpm (about 1.6%) at the output shaft.

To be clear, this arrangement can’t and doesn’t prevent the coupling from slipping. Think of it rather as a slippage rebate: Hydraulic slip still occurs, but you regain some of the lost rpm in the planetary gears. In this case, the split torque layout reduces the slippage-related speed difference between the engine and the output shaft by about 60.5% — which, not coincidentally, is the percentage of intermediate shaft torque that flows through the mechanical connection to the rear planetary gearset. Kelley’s patent disclosures described this effect as demultiplication of slippage.

This demultiplication effect was not limited to cruising speed. As long as this transmission remained in third or fourth, the partial lockup reduced slip by 60.5% even under acceleration, when the fluid coupling was significantly less efficient.

For instance, let’s suppose the Oldsmobile driver presses the accelerator to pass. Fluid clutches tend to lag a few beats behind the engine in situations like this, so if instantaneous engine speed rises to 3,000 rpm, instantaneous turbine speed might be only 2,600 rpm. In fourth gear, carrier and output shaft speed would therefore be:

VC = 2,600 + ((3,000 – 2,600) / 1.65) = 2,842.1 rpm

This would reduce total slip (excluding mechanical losses) from 400 rpm (13.3%) at the turbine to about 157.9 rpm (5.3%) at the output shaft.

The split torque arrangement also improved engine braking — particularly in third gear, when the braking effect was further multiplied by the front gearset.

One drawback of this arrangement was that the rear planetary gearset was always planetating (that is, the planet gears were turning relative to their carrier) even in top-gear cruising, which incurred more mechanical (frictional) losses — and potentially more noise and vibration — than a conventional direct drive arrangement where the planetary gears all turn at exactly the same speed. The reduced hydraulic losses more than compensated, but a true direct drive top gear with a fully mechanical lockup clutch would have been even more efficient.

Still, you can see why GM’s corporate engineering team decided that wasn’t necessary. The split torque arrangement provided many of the benefits of a lockup clutch without sacrificing desirable fluid coupling advantages such as freedom from lugging and the ability to soak up powertrain vibration.

As Kelley explained in his patent disclosures, the split torque layout essentially allowed Hydra-Matic to have different fluid coupling characteristics in each gear. The coupling could be “loose” in the lower gears, allowing more slippage for smoother takeoffs and less creep at idle, because the split torque layout would effectively make the coupling “tighter” and more responsive in the higher ranges. Since the partial lockup was limited to third and fourth, there was no risk of stalling the engine at idle and therefore no need for the additional hydraulic controls a lockup clutch would have required. (Additional mechanical complexity was the last thing the early Hydra-Matic needed!)

1952–1956 Dual-Range Hydra-Matic power flow in 3rd gear © 2016–2017 Aaron Severson

Although there were many minor changes to postwar Hydra-Matic transmissions, like the 1952 Dual-Range Hydra-Matic illustrated here, the split torque arrangement changed only in the proportion of the split, which was a function of rear unit gearing. With a 2.63:1 rear gearset, the split was 62% mechanical and 38% hydraulic. (Author diagram)

GM’s Detroit Transmission Division, which built Hydra-Matic, used this layout for all single-coupling four-speed Hydra-Matic transmissions. The actual proportion of the torque split varied with the gearing of each application — for instance, Dual-Range Hydra-Matics, whose rear gearset had a single sun gear with 41 teeth and an annulus with 67 teeth, had a torque split of 62%/38% in third and fourth — but the effects and benefits remained substantially the same.


In 1956, the Detroit Transmission Division introduced the second-generation Model 315 Controlled Coupling Hydra-Matic, which by the end of the model year had replaced the Dual-Range Hydra-Matic on most of GM’s passenger car lines (though not on trucks). The new Hydra-Matic applied the split torque principle not once, but twice.

As we’ve discussed in a previous article, the second-generation Hydra-Matic was designed to be smoother in operation than the older single-coupling transmissions, which had been notorious for their firm shifts. One of the many changes to the new transmission was the replacement of the front clutch with a second, smaller fluid coupling, controlled by alternately draining and refilling its oil supply. Just like the multi-disc friction clutch it replaced, the controlled coupling was disengaged (which in this case meant empty) in first and third and engaged (i.e., full) in second and fourth.

Color diagram of 1956–1964 Controlled Coupling Hydra-Matic transmission © 2016–2017 Aaron Severson

The Controlled Coupling Hydra-Matic (which was known by a variety of trade names) was a major update of the earlier Dual-Range Hydra-Matic. We’ve color-coded the components that always rotate together as follows: main coupling torus cover, controlled coupling torus cover, controlled coupling impeller, and front pump (light green); controlled coupling turbine and front sun gear (light blue); front carrier, main coupling impeller, intermediate shaft, and rear clutch hub (red); main coupling turbine, main shaft, and rear sun gear (medium blue); rear annulus and reverse unit sun gear (orange); rear planet carrier, output shaft, and rear pump/governor (purple). The two sprag clutches are indicated in fuchsia with black triangles. (Author diagram)

In dual-coupling Hydra-Matics, the torus cover of the main coupling was bolted to the flywheel, just as in the single-coupling transmissions, and drove the the annulus of the front planetary gearset. The torus cover also drove the torus cover of the small controlled coupling, so the controlled coupling’s impeller rotated at engine speed even when the coupling was empty.

The controlled coupling turbine was affixed to the front unit sun gear (or, to be very technical, the sleeve shaft connecting the sun gear to the front sprag and overrun clutch). When the small coupling was full, engine torque was transmitted through the coupling to the sun gear, driving it forward.

This creates what you’ll hopefully now recognize as a “series parallel” split torque arrangement in second and fourth. In those gears, engine torque was divided between the mechanically driven front annulus and the hydraulically driven front sun gear. The front carrier, which was mounted on the back of the main coupling impeller, resolved the speed difference between those gears and applied their combined torque to the main impeller and the intermediate shaft.

The front annulus of a Controlled Coupling Hydra-Matic had 56 teeth while the front sun gear had 31 teeth. With the annulus driving, that gave the front gearset a ratio of 1.55:1 (1 + 31/56). If both the annulus and sun gear drove, the sun gear received about 35.6% of engine torque while the other 64.4% was applied to the annulus.

1956–1964 Controlled Coupling Hydra-Matic transmission front coupling showing power flow in 2nd and 4th gears © 2016–2017 Aaron Severson

In the Controlled Coupling Hydra-Matic, filling the small coupling split engine torque at the main coupling torus cover. Torque was recombined at the front planet carrier. In fourth, torque was then split again between the impeller and the rear clutch. (Author diagram)

This split served to demultiply any slip in the controlled coupling by 64.4% in the manner we described on the previous page. For example, if the controlled coupling had an efficiency of 97% at cruising speed, the demultiplication effect would reduce effective slip from 3% at the turbine to a little less than 1.1% at the carrier. Discounting mechanical losses, that would mean the front carrier and impeller rotated at about 98.9% of engine speed.

In first and third, when the small coupling was empty, the controlled coupling turbine did not rotate, so all engine torque flowed through the annulus. That torque and the inertia of the front carrier and impeller exerted reaction torque on the front sun gear, locking the front sprags (which kept the sun gear from turning backward) and driving the carrier and the main coupling impeller forward at 64.4% (100% / 1.55) of engine speed.

1956–1964 Controlled Coupling Hydra-Matic transmission front coupling showing power flow in 1st and 3rd gears © 2016–2017 Aaron Severson

In first and third gears in the Controlled Coupling Hydra-Matic, all power flowed through the front annulus, so there was no torque split at the front gearset. In third, however, torque then split at the front carrier. (Author diagram)

Like the single-coupling Hydra-Matic, the Controlled Coupling Hydra-Matic had an intermediate shaft that mechanically connected the main coupling impeller to the rear clutch hub. In third and fourth, when the rear clutch was engaged, torque on the main impeller was split between the mechanically driven rear annulus and the hydraulically driven rear sun gear.

The rear annulus of the dual-coupling Hydra-Matic had 73 teeth and the rear sun gear had 47 teeth, so with the sun gear driving, the rear gearset had a ratio of 2.55:1 (1 + 73/47). That exerted reaction torque on the annulus at a ratio of -1.55:1 (0 – 73/47). Therefore, the rear annulus had to receive 1.55 times as much torque as the rear sun gear did, giving a torque split of 60.8%/39.2%. This served to demultiply slip in the main coupling by 60.8%.

Like the single-coupling Hydra-Matic, fourth gear in the dual-coupling transmission was direct drive, which meant that the controlled coupling was full and the rear clutch was engaged in top gear. That wasn’t ideal from the standpoint of efficiency because it meant that in fourth, power had to flow through both couplings rather than just one.

As illustrated above, if we assume the controlled coupling was 97% efficient at cruising speed, the speed of the impeller and intermediate shaft (discounting mechanical losses) would be 98.9% of engine speed. If we assume the efficiency of the main coupling at cruising speed was also 97%, the speed of the turbine and the main shaft, again discounting mechanical losses, would be just under 96% (98.9% * 97%).

In fourth, therefore, VS = 95.96% and VA = 98.93%. Plugging those into the formula on the previous page gives us:

VC = 95.96% + ((98.93% – 95.96%) / (1 + 47 / 73)) = 97.77%

Net slip at the output shaft, therefore, would be 2.23%, which would be less than the slippage in either coupling, albeit still more than in a single-coupling Hydra-Matic.

Detroit Transmission Division was concerned enough about the additional slippage that they considered adding a lockup clutch for the controlled coupling, allowing it to be completely locked out when cruising in fourth. However, that feature wasn’t adopted for production, perhaps because the double demultiplication effect made the Controlled Coupling Hydra-Matic at least as efficient with two couplings as many other contemporary automatic transmissions were with only one. In any case, the second-generation Hydra-Matic was intended for big American cars with big V-8 engines and curb weights often exceeding 4,400 lb/2,000 kg, so a small amount of additional hydraulic slippage was not considered a deal-breaker.


The three new automatic transmissions GM introduced in 1961 for the Y-body “senior compacts” each used the split torque principle, albeit in three quite different ways. (Since we’ve previously discussed these transmissions in some detail, we’ll just look at the split torque function of each.)


Buick’s two-speed Dual-Path Turbine Drive, used in the 1961–1963 Buick Special and Skylark, was probably the simplest of the three in this regard. As we’ve previously explained, Dual-Path had a single planetary gearset within the torque converter housing. The converter turbine drove the annulus of that gearset, whose planet carrier drove the output shaft. In first gear, a one-way clutch held the dual sun gears stationary. In second, a multi-disc clutch allowed the engine to drive the sun gears directly while the turbine continued to drive the annulus.

Color diagram of 1961-1963 Buick Dual-Path Turbine Drive transmission © 2016 Aaron Severson

A schematic of Buick’s short-lived Dual-Path Turbine Drive two-speed automatic. The flywheel drove the torus cover, oil pump, and torque converter impeller (red) while the turbine drove the annulus of the single planetary gearset (blue), whose planet carrier drove the output shaft (green). There were two identical sun gears (orange), one connected to a cam-and-roller one-way clutch (fuchsia) that shared its cam with the clutch for the stator (shown here in purple). The direct clutch connected the other sun gear to the torus cover. Note that the torque converter was “backward,” with the turbine facing the engine rather than the rear axle. (author diagram)

Precisely calculating the second-gear torque split is complicated somewhat by the fact that we don’t know how many teeth the planetary gears have. We do know that the gear ratio in first, with the annulus driving and the sun gears held, was 1.58:1. (We would conjecture that the annulus had 66 teeth, the sun gears had 38, and the planets each had 14, which would make the exact gear ratio 1 + 38/66, or 1.5757:1.) That’s enough to estimate that the annulus received about 63.3–63.4% (1 / 1.58) of input torque in second gear, with the remaining 36.6–36.7% applied to the sun gears.

1961–1963 Buick Dual-Path Turbine Drive transmission diagram showing power flow in 2nd gear (high) © 2016–2017 Aaron Severson

Power flow in the Dual-Path Turbine Drive in second gear was through the torus cover to the hydraulically driven annulus and simultaneously through the mechanically driven rear sun gear of the single planetary gearset. This was not a Ravigneaux gearset — the two sun gears were identical and always rotated at the same speed even in a reduction gear. (author diagram)

Since the sun gears were driven by the engine, the demultiplication effect was considerably smaller than in Hydra-Matic — less than 37%. For example, if the engine was rotating at 2,000 rpm and there was 100 rpm of slippage in the converter, output shaft speed (discounting mechanical losses) would be about 1,937 rpm, effectively reducing hydraulic slippage from 5% at the turbine to about 1.8% at the output shaft.


The automatic used in the 1961–1963 Pontiac Tempest and Le Mans, called TempesTorque, was a variation of the two-speed Corvair Powerglide, adapting the Corvair transaxle’s oil pump driveshaft to send power from the curved driveshaft to the torque converter at the back of the transaxle. 1961–1962 editions also had a split torque top gear, a feature Powerglide didn’t share.

TempesTorque and Powerglide, like the older Ultramatic and Dynaflow transmissions, used a Ravigneaux gearset with a single annulus, two sun gears of different sizes (with different numbers of teeth), and six planets (three long, three short) on a planet carrier connected to the output shaft. As with Powerglide, TempesTorque’s driving member was the larger rear sun gear, which was driven by the torque converter turbine through the main shaft.

Unlike Powerglide, which obtained direct drive by also connecting the smaller front sun gear to the main shaft (forcing both sun gears to rotate at the same speed), the direct drive clutch of the 1961–62 TempesTorque served to connect the front sun gear to the central input shaft, which rotated at engine speed, not turbine speed.

Color diagram of 1961–1962 Pontiac TempesTorque transaxle © 2016–2017 Aaron Severson

A Ravigneaux gearset, like that used in Powerglide and TempesTorque, obtains indirect gear ratios by causing the two sun gears to rotate at different speeds. Normally, direct drive is obtained by locking the input shaft to both sun gears so that they turn at the same speed. In 1961–1962 TempesTorque units, direct drive instead locked the low sun gear to the input shaft while the input sun gear rotated with the main shaft. (author diagram)

1961–62 TempesTorque transaxles shared the Powerglide gearset, whose annulus had 79 teeth and whose sun gears had 23 and 28 teeth respectively, giving first and reverse ratios of +/- 1.82:1. We’ll spare you the complex algebra, but in second gear, 54.9% of input torque went to the hydraulically driven rear sun gear while 45.1% went to the mechanically driven front sun gear. For example, at an engine speed of 2,000 rpm with 100 rpm of converter slippage, the carrier would rotate at 1,945.1 rpm, effectively demultiplying hydraulic slippage from 5% at the turbine to about 2.7% at the output shaft.

1961–1962 Pontiac TempesTorque transaxle diagram showing power flow in 2nd gear (high) © 2016–2017 Aaron Severson

In 1961–1962 TempesTorque transaxles, the direct drive second gear engages the direct clutch, allowing the input shaft (red) to drive the low sun gear (purple) at engine speed. Power also continues to flow through the input shaft to the torus cover and impeller (also red), then hydraulically to the turbine and main shaft, which drives the input sun gear (all medium blue). The planet carrier (light green) resolves the speed difference between the two sun gears and drives the differential input gear. (author diagram)

For 1963, TempesTorque’s final year, the direct clutch was revised to connect the front sun gear to the main shaft rather than the input shaft, which eliminated the torque-splitting feature.


Detroit Transmission Division also applied the split torque principle to the simplified third-generation Model 240 and Model 375 Hydra-Matic transmissions (sometimes called “Roto Hydra-Matic”) used in the 1961–1963 Oldsmobile F-85/Cutlass, some Holden and Opel models, and most full-size 1961–1964 Oldsmobiles and Pontiacs.

Unlike its predecessors, the third-generation Hydra-Matic was a three-speed transmission whose single dump-and-fill fluid coupling was transformed into a torque converter by the addition of a torque multiplier member (not a stator) between the impeller and turbine. There were now two interconnected planetary gearsets with interconnected planet carriers, which were in turn connected to both the torque multiplier and the output shaft. The front sun gear and rear annulus were also connected, forcing them to rotate (or not rotate) together at the same speed. A sprag clutch kept them from rotating backward in any forward gear.

Color diagram of 1961–1964 Roto Hydra-Matic transmission © 2016–2017 Aaron Severson

The three-speed Roto Hydra-Matic was essentially a much simplified adaptation of the costlier, more complex four-speed Controlled Coupling Hydra-Matic. Perhaps its most unusual feature was the very small (8-inch/203mm) torque converter, which alternately emptied and filled in different gears. The converter’s torque multiplier — not a true stator — had no one-way clutch and was affixed directly to the carrier shaft. (author diagram)

Roto Hydra-Matic’s operation was broadly similar to that of its dual-coupling predecessor, relying on alternately emptying and refilling its torque converter and engaging or disengaging its multi-disc front clutch. In first, the converter was full and the front clutch released. In second, the converter was empty and the clutch released while in third, the converter was full and the front clutch engaged simultaneously.

1961–1964 Roto Hydra-Matic transmission diagram showing power flow in 1st gear © 2016–2017 Aaron Severson

In first, Roto Hydra-Matic sent all power through the torque converter to the rear sun gear (medium blue). With the neutral clutch engaged, the outer race of the sprag clutch was locked to the case, so the inner race — shared by the rear annulus and front sun gear (orange) — could only turn in the direction of engine rotation. The inertia of the output shaft and the rotation of the rear sun gear would lock the sprag, driving the planet carriers and output shaft forward in reduction. There were at least three different rear gear sets for different versions of this transmission, giving geared ratios of 2.93, 2.97, or 3.03:1 in first. (author diagram)

1961–1964 Roto Hydra-Matic transmission diagram showing power flow in 2nd gear © 2016–2017 Aaron Severson

In second, the neutral clutch remained engaged, the front clutch engaged, and the torque converter was drained, sending all power through the converter’s torus cover (red) to the front annulus with no hydraulic slippage. The sprag remained locked, so the front annulus drove the planet carriers and output shaft forward at reduced speed. There were two different front gear sets for different versions of Roto Hydra-Matic, providing a second gear ratio of 1.56 for the bigger Model 375 and 1.58 for the Model 240. We assume the reason for the inconsequential variation was that the Model 375’s gear teeth were slightly wider, reducing the total number of teeth on each gear. (author diagram)

There was no torque split in second, since with the torque converter empty, all power was transmitted mechanically through the torus cover and front clutch. However, in third, power was divided between the mechanically driven front annulus and the hydraulically driven rear sun gear. Since this meant the front annulus was turning faster than the rear sun, its rotation drove the carrier around the slower-moving sun gears, resolving the speed difference.

If we’re doing the math correctly, this demultiplied converter slippage by about 60%, depending on the specific front and rear gearsets used. Under some conditions, a small amount of torque also flowed through the converter’s torque multiplier, which always turned at the same speed as the planet carriers and output shaft.

1961–1964 Roto Hydra-Matic transmission diagram showing power flow in 3rd gear © 2016–2017 Aaron Severson

The torque split of Roto Hydra-Matic’s split torque third gear depended on gearing. With the original Model 375 gears (2.97 first, 1.56 second), the split was 59.9%/40.1% mechanical/hydraulic. With the original Model 240 gears (3.03 first, 1.58 second), the split became 60.6%/39.4%. With the later 2.93:1 rear gears, the split was either 57.6/47.2% or 58.9/41.1%. (author diagram)


GM abandoned its remaining split torque transmissions after the 1964 model year. It appears that at least part of the rationale for the wholesale switch to more conventional automatic transmission layouts was the desire to more fully exploit torque converter multiplication — not only for starting, but also to bridge the gaps between the geared ratios, especially with two-speed automatics. In the sixties and early seventies, having the torque converter available in all gears was a higher priority for most American automakers than a slight improvement in highway fuel economy.

That changed abruptly with the 1973–1974 OPEC oil embargo, which in the U.S. led to the enactment in December 1975 of the Energy Policy and Conservation Act of 1975 (EPCA). Title III of EPCA included the first-ever fuel economy requirements for U.S.-market cars. Those rules, now known as CAFE (for Corporate Average Fuel Economy), mandated a fleet average fuel economy of 18.0 mpg (13.1 L/100 km) for the 1978 model year. Since that was less than two years away, the EPCA rules sent automakers scrambling for quick fixes that wouldn’t interfere with their ability to also meet the latest federal emissions standards.

Chrysler decided to revive the lockup torque converter clutch, which at the time was still used in some truck and bus transmissions and a few European automatics, but to our knowledge hadn’t been used in a U.S.-made passenger car in about 20 years. Although the fuel economy improvements a lockup clutch offered were modest, it was a feature Chrysler could quickly add to most of its existing passenger cars without a massive retooling bill. The addition also benefited an important Chrysler client: American Motors, which purchased Chrysler TorqueFlite automatics for its own vehicles.

Diagram of a Chrysler-style piston-type torque converter lockup clutch © 2016–2017 Aaron Severson

This diagram illustrates the basic principles of Chrysler’s early lockup clutches, although we’ve simplified many details and deliberately exaggerated the scale for the sake of clarity — the piston arrangement was actually very compact, occupying little more space than a non-lockup converter. The turbine hub had splines that allowed the clutch piston to slide forward and backward (left and right, as illustrated here) as hydraulic pressure was added to or released from the space between the piston and the front of the torus housing. (author diagram)

Chrysler’s lockup clutch, added to many (though not all) TorqueFlite transmissions for 1978, performed the same function as the old Packard Ultramatic unit, but with fewer parts. The actual friction surface was arranged in a ring around the inside front cover of the torus housing, between the flywheel and the torque converter turbine. When engaged, a hydraulic clutch piston splined to the turbine hub slid toward the flywheel, locking the piston (and thus the turbine hub) to the torus cover and forcing them to rotate together at engine speed. As with Ultramatic, clutch engagement and disengagement was controlled in the same manner as a gear change, using opposing governor and throttle valve pressures to engage the clutch piston either after or simultaneously with a shift into third gear.

Diagram of a Chrysler-style piston-type torque converter lockup clutch (engaged) © 2016–2017 Aaron Severson

When Chrysler’s piston-type lockup clutch was engaged, hydraulic pressure forced the piston (medium blue) against the friction surface on the back of the torus housing (red). Since the piston was splined to the turbine hub, this allowed the torus cover to drive the turbine and the input shaft. (author diagram)

Other automakers quickly followed suit. GM and Ford had added lockup clutches to all of their passenger car automatics by the 1982 model year. Many Japanese and European automatic transmissions were so equipped by the mid-eighties. Like the Chrysler and Packard units, most lockup clutches of the eighties and nineties were hydraulic, although a growing number used solenoids to open and close the hydraulic valves, allowing clutch engagement to be controlled by computer.


An alternative to the orthodox hydraulic lockup clutch was the centrifugally operated bypass clutch, which was a completely mechanical lockup not requiring any hydraulic controls. The idea had been around for decades and had previously shown up on some nonautomotive torque converters as well as a few passenger car automatics, notably the ZF unit offered in the Peugeot 404. A variety of automakers and automotive suppliers, including GM, Borg-Warner, and JATCO, returned to the concept in the seventies, and centrifugal bypass clutches appeared on a variety of production automatics throughout the eighties.

In the Borg-Warner type, developed in the seventies and used by Ford for its 1982–1986 C5 transmission and some front-wheel-drive transaxles, the actual clutch plate was permanently engaged, connecting the torque converter turbine to a transfer disc ringed with friction shoes. A one-way clutch in the transfer disc’s hub allowed the disc to drive the input shaft whenever the disc turned faster than the turbine.

Diagram of a three-element torque converter with centrifugal lockup clutch © 2016–2017 Aaron Severson

A centrifugal lockup clutch of the type developed by Borg-Warner and used by Ford in the eighties was positioned between the flywheel and turbine, driving the input shaft at engine speed once the turbine was rotating fast enough for inertia to wedge the friction shoes firmly against the inside of the torus cover. (author diagram)

The friction shoes functioned in a manner analogous to the shoes of an expanding drum brake, using the torque converter’s torus cover as the drum. Shoe position was controlled centrifugally by a series of spring-loaded weights. With the turbine stalled, spring loading held the shoes in the disengaged position. Once the turbine started rotating, the weights’ inertia would effectively “throw” the shoes progressively outward, compressing the springs and forcing the shoes against the inner circumference of the torus cover.

Under load, when the speed difference between the impeller and the turbine was large, the shoes would slip against the torus cover’s inner surface, allowing the torque converter to function normally. However, as the torque converter approached coupling stage, reducing the difference between engine and turbine speeds, the shoes would find enough purchase to wedge the transfer disc against the torus cover, forcing disc and cover to rotate together. Since the torus cover was bolted to the flywheel and always turned at engine speed, this caused the transfer disc to overrun the turbine, locking the one-way clutch separating the two and forcing the torus cover, transfer disc, turbine, and input shaft to all rotate together at engine speed.

Diagram of a centrifugal lockup clutch for a torque converter © 2017 Aaron Severson

A head-on view of a centrifugal lockup clutch. The outermost circle (red) represents the torus cover, which is bolted to the engine flywheel/flex plate and rotates at engine speed. The inner disc (light green) represents the transfer disc, which a permanently engaged clutch (not shown) links to torque converter turbine. Friction shoes (black) move radially outward the faster the turbine rotates, pressing against the torus housing. An overrunning clutch (fuchsia) allows the turbine hub (medium blue) to turn faster than the transfer disc, but not slower. (author diagram)

A centrifugal lockup clutch had several advantages over the more conventional hydraulic variety. The most obvious was that it required no separate controls, making it somewhat cheaper than a hydraulically or electro-hydraulically controlled clutch. Also, because the lockup process was completely mechanical, dictated mostly by turbine speed, a centrifugal clutch worked in all forward gears, at least theoretically providing greater fuel economy benefits than lockup clutches that worked only in top gear. Furthermore, engagement was more progressive — and thus less obtrusive — than conventional hydraulic lockup clutches, which tended to engage and disengage with a noticeable thump. One tradeoff was that the friction shoes were subject to more wear, since they would slip any time there was a significant increase in load. Another compromise, at least for Borg-Warner centrifugal clutches, was they didn’t do much for engine braking, since the turbine could overrun the transfer disc.

Another unusual lockup clutch variation, used in some versions of GM’s TH440-T4 (a.k.a. 4T60) front-wheel-drive transaxle, was the viscous bypass clutch. Developed in collaboration with Eaton Corporation and first used in the 1984 Cadillac de Ville, the viscous coupling was positioned between the torque converter turbine and the flex plate.

Color diagram of a simple viscous coupling © 2017 Aaron Severson

This diagram illustrates the basic layout of a viscous coupling. A series of driving and driven flanges are arranged in concentric rings. The space around and between those flanges is filled with silicone fluid. Any time the input and output shafts rotate at different speeds, they heat the viscous fluid, causing it to bind the flanges more closely together. (author diagram)

As with a conventional plate clutch, hydraulic pressure within the converter pressed a friction surface on the outside of the viscous coupling’s housing against the torus cover, causing the coupling’s driving (input) flange to rotate at engine speed. As engine speed increased, shear within the viscous coupling’s silicone working fluid would bind the driving flanges to the driven flanges, which in turn drove the transmission input shaft. At higher speeds, pressure within the coupling would more or less lock the flanges together, causing them to drive the input shaft at close to engine speed. A control valve also allowed the viscous coupling to be completely disengaged when necessary by forcing the housing away from the torus cover surface.

Color diagram of a torque converter with viscous bypass clutch © 2016–2017 Aaron Severson

The viscous bypass clutch was engaged or disengaged hydraulically, using pressure differential to move the coupling case against or away from the torus cover. Once engaged, the driving flanges (dark green) would begin to rotate, binding them to the driven flanges (medium blue) and turning the transmission input shaft. (author diagram)

Unlike a plate clutch, the viscous coupling still allowed a small amount of internal slippage, which GM and Eaton argued was balanced by the ability to lock up at road speeds as low as 25 mph (40 km/h). Just as importantly, so far as Cadillac was concerned, the viscous bypass clutch’s engagement or disengagement was progressive enough to be imperceptible. However, the viscous clutch was both less efficient and more expensive than the conventional lockup clutch used in other TH440-T4 applications.


During this period, Ford also revived the split torque concept, first for the AOD, Ford’s first four-speed overdrive automatic, and then for the ATX, the company’s first automatic transaxle for front-wheel-drive applications.


The Automatic Overdrive (AOD) transmission, introduced for some full-size FoMoCo cars in the 1980 model year, was derived from Ford’s older three-speed FMX automatic. It used a Ravigneaux compound gearset with three long and three short planet pinions on a common planet carrier, two sun gears (one with 36 teeth, the other with 30), and a single annulus (with 72 teeth) affixed to the output shaft.

Color diagram of a 1980–1991 Ford AOD transmission © 2016–2017 Aaron Severson

Ford’s four-speed AOD transmission had both a turbine-driven primary input shaft (medium blue) and a narrow central direct drive shaft geared to the torus cover (red). In the first three forward gears, the main input shaft drove the rear sun gear (orange) through the forward clutch while the annulus drove the output shaft (dark green). The reaction member was the planet carrier (purple) in first and the front sun gear (light green) in second. In third and fourth, the direct clutch linked the direct drive shaft to the planet carrier. The fuchsia rectangles represent the transmission’s three one-way clutches; the upward-pointing black triangles indicate that the clutches allow rotation only in the direction of engine rotation. (author diagram)

As in the FMX, the torque converter turbine drove the smaller rear sun gear in first, second, and third gears. Two one-way clutches (supplemented in Low by a brake band) could alternately hold the carrier or the larger front sun gear, providing two forward reduction ratios. Where the AOD parted ways with the FMX was in the direct drive third gear. Instead of locking the front sun gear to the input shaft in high, as the FMX did, the AOD used an additional direct clutch to lock the planet carrier to a central shaft that passed through the main input shaft and was driven directly by the engine.

In third, engine torque was split between the rear sun gear, which rotated at turbine speed, and the planet carrier, which rotated at engine speed. Because the carrier always rotated faster than the rear sun gear, the long planets overdrove the larger front sun gear. This forced the short planet pinions to resolve the speed difference and drive the annulus forward at a speed slower than the carrier, but faster than the rear sun gear, demultiplying hydraulic slippage by 58.3%. (We’ll spare you the math, which with Ravigneaux gearsets is very cumbersome.) For example, if the engine was turning 2,000 rpm and converter slippage was 100 rpm, the annulus would rotate at 1,958.3 rpm, reducing net hydraulic slip from 5% at the turbine to 2.1% at the output shaft.

1980–1991 Ford AOD transmission showing power flow in 3rd gear © 2016–2017 Aaron Severson

In the Ford AOD’s split torque third gear, both the forward and direct clutches were engaged, allowing both the rear sun gear and the planet carrier to drive the annulus forward. Input torque was then split between the rear sun gear and the carrier and recombined by the planet pinions. Both bands were released in third while the transmission’s two one-way clutches (both fuchsia) were ineffective, although the stator’s one-way clutch could lock if turbine speed fell significantly below engine speed. (author diagram)

The AOD’s new overdrive fourth gear was created by releasing the forward clutch — thereby disconnecting the rear sun gear from the main input shaft and the turbine — while engaging a brake band to hold the front sun gear stationary. With the direct clutch still engaged, the carrier, rotating at engine speed, overdrove the rear sun gear, forcing the annulus to rotate at 1.5 times engine speed — an ratio of 0.67:1. The torque converter turbine continued to rotate in fourth, but was no longer connected to the planetary gears, so there was no hydraulic slippage. This lockup wasn’t available in any of the lower gears.

1980–1991 Ford AOD transmission showing power flow in 4th gear © 2016–2017 Aaron Severson

In fourth gear, the AOD’s direct clutch remained engaged, but the forward clutch disengaged and the overdrive band held the front sun gear stationary. With the planet carrier driving and one sun gear held, the transmission became a completely mechanical overdrive with a ratio of 0.67:1. (author diagram)


Ford took a different approach with the ATX transaxle, which became optional on the new FWD (Mk3) Ford Escort and Mercury Lynx for 1981 and the Ford Tempo/Mercury Topaz for 1984. Unlike the AOD and the earlier GM split torque transmissions, the early ATX used a separate planetary gearset specifically for torque-splitting purposes.

Color diagram of a 1981–1985 Ford ATX transaxle with split torque converter © 2017 Aaron Severson

Although the Ford ATX was an automatic transaxle for transverse engine/FWD applications, we’ve omitted most of the differential components for clarity. From this angle, the differential would be in front of the unit, driven by the input gear, which in turn was driven by the planet carrier (both shown in purple). The version of the ATX used in the larger Ford Tempo and Mercury Topaz had the same layout, but used additional plates in all three multi-disc clutches for greater torque capacity. (author diagram)

Known in Ford parlance as a “splitter gear,” the additional gearset was located within the torque converter torus housing, between the turbine and the flex plate. It was a simple planetary gearset with a single annulus (with 78 teeth) and three planet pinions surrounding a sun gear with 48 teeth. The flex plate and torus housing drove the splitter unit annulus at engine speed while the transmission input shaft, driven by the turbine, drove the splitter sun gear.

As we’ve explained in the preceding pages, this arrangement served to demultiply hydraulic slippage in the torque converter. Since annulus speed (VA) was engine speed while sun gear speed (VS) was turbine speed, the speed of the splitter gearset’s planet carrier (VC) was therefore:

VC = VS + ((VA – VS) / (1 + sun gear teeth / annulus teeth))

… or:

VC = VS + ((VA – VS) / 1.615)

For example, if the engine were turning 2,000 rpm and there was 100 rpm of converter slippage, carrier speed would be 1,961.9 rpm, demultiplying hydraulic slippage from 5% at the turbine to about 1.9% at the carrier.

1981–1985 Ford ATX transaxle split torque converter diagram © 2017 Aaron Severson

In split torque ATX transaxles, the torus cover, which was bolted to the flywheel, drove the impeller, the oil pump shaft, and the annulus of the “splitter” gearset (all shown in red). The turbine drove the main shaft and the splitter sun gear (both medium blue). The planet carrier of the splitter gearset drove a secondary driveshaft (both light green), which drove the intermediate clutch. (author diagram)

The splitter gearset’s planet carrier drove a hollow sleeve shaft (which for reference we’ll call the “splitter shaft”), passing through the main transmission input shaft. (The sleeve shaft was also hollow because it contained the solid driveshaft for the transmission oil pump, which was located on the opposite side of the transaxle from the torque converter.) Both input shafts carried power to the main planetary transmission, a Ravigneaux gearset with a single annulus (with 86 teeth), two sun gears (one with 52 teeth, the other with 29), and three short and three long planets (each with 17 teeth) on a common carrier. The carrier drove the differential input gears.

1981–1985 Ford ATX transaxle showing power flow in 1st gear © 2017 Aaron Severson

In first gear, the split torque ATX sent all power through the converter turbine to the main shaft (medium blue) and a one-way clutch (fuchsia) to the smaller rear sun gear (dark red). The larger front sun gear (orange) was held by the low band, forcing the carrier and differential input gear (purple) to rotate forward at reduced speed. In first, the splitter gearset in the converter rotated, but transmitted no power. (author diagram)

The splitter shaft was ineffective in first and reverse, rotating idly while the main shaft drove the smaller input sun gear at turbine speed. In second gear, the multi-disc intermediate clutch engaged to connect the splitter shaft to the annulus of the Ravigneaux gearset. Since the larger rear sun gear was held by a brake band, this caused the smaller sun gear to overrun the input shaft and spin idly on its one-way clutch. Since the splitter shaft was now driving, hydraulic slippage was demultiplied by 61.9% (1 / 1.615).

1981–1985 Ford ATX transaxle showing power flow in 2nd gear © 2017 Aaron Severson

In the ATX transaxle’s second gear, the low band remained engaged, holding the larger rear sun gear (orange) while the intermediate clutch engaged, allowing the splitter shaft (green) to drive the rear annulus (dark blue). The one-way clutch allowed the smaller front sun gear to overrun the main shaft, which continued to turn idly. (author diagram)

In the direct drive third gear, the multi-disc direct clutch engaged, again allowing the main input shaft to drive the smaller sun gear, and the brake band was released, allowing the large sun gear to rotate freely. The intermediate clutch remained engaged, so the splitter shaft continued to drive the annulus at the same time. With both the annulus and small sun gear driving, the planet gears overdrove the large sun gear, forcing it to rotate faster than the engine. This in turn forced the planet carrier to rotate faster than either input shaft, albeit still slightly slower than the engine.

1981–1985 Ford ATX transaxle showing power flow in 3rd gear © 2017 Aaron Severson

In the ATX transaxle’s third gear, the low band released, the intermediate clutch remained engaged, and the direct clutch engaged, locking the main shaft to the smaller front sun gear (dark red). With the annulus (dark blue) driven by the splitter shaft and the small sun gear driven by the main shaft at turbine speed, the large sun gear (orange) was overdriven. The planet carrier (purple) then “resolved” the speed difference, rotating slower than the large sun gear, but faster than either the small sun gear or the annulus. (author diagram)

Although the math is again very cumbersome, this arrangement demultiplied converter slippage by a whopping 93.4% in third. For example, if engine speed were 2,000 rpm with 100 rpm of converter slippage, carrier speed (discounting mechanical losses) would be 1,993.4 rpm, reducing hydraulic slippage from 5% at the turbine to a mere 0.003% at the differential. That was efficient enough that the ATX could forgo a lockup clutch without a noticeable sacrifice in fuel economy, an important consideration for Ford’s cheapest U.S.-market models.


Ford’s split torque revival was relatively brief. Later versions of the ATX transaxle abandoned the splitter gear and dual input shafts for either a conventional hydraulic lockup (which Ford abbreviated FLC, for “full lockup clutch”) or, in some applications, a centrifugal lockup clutch (CLC) similar to that of the C5 transmission. The change coincided with the availability of more powerful engines in the Escort/Lynx and Tempo/Topaz lines, which suggests that the rationale may have been to facilitate increases in the transaxle’s torque capacity. (Tempo/Topaz versions of the ATX already had extra clutch plates in each multi-disc clutch.)

Similar concerns led to the elimination of the split torque feature of the four-speed AOD in the early nineties. The AOD was adequately strong for early eighties engines, but as power and torque increased throughout the decade, the secondary input shaft became a notable weak link. When the electronically controlled AOD-E debuted in 1991, it had only a single input shaft with no third-gear torque split. The same was true of the closely related 4R70 and 4R70W that replaced the AOD-E in 1993.

By then, all or nearly all domestic and most non-U.S. passenger car and light truck automatics had lockup clutches, most of them hydraulically operated and electronically controlled, differing only in minor details. The centrifugal lockup clutch eventually went the way of the split torque units, since its purely mechanical operation didn’t offer the fine-tuned control of electronically controlled electro-hydraulic clutches, whose operation can be better tailored to different operating conditions. For example, an electronically controlled clutch can be programmed to remain unlocked during warmup or if either the engine or transmission is in danger of overheating, which a centrifugal clutch cannot.

This period of comparative orthodoxy — mostly four-speed overdrive automatics with electro-hydraulic lockup clutches — turned out to also be relatively brief. Subsequent automatic transmission design has diverged along several paths, including relatively conventional planetary transmissions with an ever-growing number of gears, at least three types of continuously variable transmissions, and dual-clutch semiautomatic transmissions. All these are beyond the scope of an article that is already considerably longer and more complicated than we originally intended, so we’ll just say that for automatics that have torque converters, computer-controlled lockup clutches are now the established norm and are likely to remain so.

As for split torque layouts, those have become common for hybrid electric vehicles, but that subject, like many others, will have to wait for another day.


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  1. Well Aaron . . another masterpiece. This and the GM automatic history are probably the most definitive descriptions of these technologies on the Internet, excepting pure design and engineering treatises. Well done and thank you; this must have been an enormous amount of work.

  2. I have yet to read this monumental work in depth. Whether it will add to my working knowledge is debatable, but my brain will benefit from the workout.
    Aaron is probably now the best informed person in the world regarding the history of automatic transmission development

    1. I appreciate the compliment, but I’m really not! This is a remarkably broad, convoluted, and idiosyncratic field and there’s a LOT I don’t know. For people who want a broader overview, I would recommend a book by Philip G. Gott entitled Changing Gears: The Development of the Automotive Transmission, published by the SAE as part of their Historical Series in 1991. (At this point, an updated, expanded edition wouldn’t go amiss, given all the subsequent development in CVTs and automatics with five or more speeds.)

  3. Smashing read, great job!

  4. I can’t imagine the hours of work which you must have put into understanding these various transmissions, to say nothing of writing up a description that a simpleton like me could (mostly) understand. Another fascinating article, thank you for all your effort!

    One thing I’ve always wondered about was if any manufacturers looked into Wilson pre-selector gearboxes as a basis of an automatic. Wilson pre-selectors were pretty well established technology, although not common, by the late ‘30s. Obviously some sort of mechanism would have been required to determine what gear to select and when to actually shift, but starting with a Wilson ‘box at least some of the problems would have been solved. But I’ve never heard of anyone going that route.

    1. The GM team that designed Hydra-Matic was certainly familiar with the Wilson preselector. In fact, Cadillac’s chief engineer ordered an early Daimler Double Six with the Wilson and Laurence Pomeroy’s Fluid Flywheel for evaluation purposes. However, Wilson gearboxes were quite bulky and complex because the nature of their operation required a separate set of epicyclic gears for each ratio, including reverse. With automated hydraulic operation and combinations of brakes and clutches, it was possible to get the same results more efficiently.

      1. Interesting—thanks for the information!

        1. I haven’t studied the Wilson preselectors in any great detail, but if you’re curious, the applicable U.S. patents are US1404675 and US1796904. As you’ll see if you look at the first one, the original iteration had three speeds forward and one reverse, for which it requires four epicyclic gearsets. A Simpson gearset (which I’ll be discussing in great detail in the next few days) provides the same number of ratios from only two gearsets, and a single Ravigneaux gearset can give you four forward speeds and reverse if you have enough clutches. So, you can see how those would be preferred from a standpoint of cost and packaging!

        2. For comparison, a four-speed Wilson pre-selector has four planetary gearsets, four sets of brake bands, and a cone clutch, which is a lot of pieces.

  5. sir im having an issue with my 93 f150 aod. its the mechanically controlled aod. works great no real problems. but the question is when i put my buddys obd code finder on it. the only readings i got was for all the electronically controlled aod. there were around 6 defaults that popped up. i called a trans shop and he had no answer,but it sounded strage to him. if someone would have put in a used ecu ,fron a electric controlled aod. would it work,yet throw out aod default codes. i take it your a writer and not a trans guy ,but maybe someone could answer the question.

    1. I’m not able to provide repair or maintenance advice, sorry!

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