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.

THE SLIPPAGE PROBLEM

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.

THE LOCKUP CLUTCH CONCEPT

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.

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.

PACKARD ULTRAMATIC AND STUDEBAKER AUTOMATIC DRIVE

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.

PACKARD ULTRAMATIC

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.

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.

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.

STUDEBAKER AUTOMATIC DRIVE

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.

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.

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.

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.

THE SPLIT TORQUE PRINCIPLE

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.

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.

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.

CALCULATING THE TORQUE SPLIT

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.

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 T_A) and the torque on the sun gear (which we’ll call T_S) had to equal the torque on the intermediate shaft (T_I). So, in mathematical terms:

T_I = T_S + T_A

Since we also know that:

T_A = T_S * 1.53

Therefore:

T_I = T_S + (T_S * 1.53)

… which simplifies to:

T_I = T_S * 2.53

We can then solve for T_S:

T_S = T_I / 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 T_C) had to be the same as the torque the annulus applied to the carrier. Or, in mathematical terms:

T_C = T_S * 2.533 = T_A * 1.652

As we determined above, the torque on the sun gear (T_S) 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 (T_A) was 60.52% of the total, and 60.52% times 1.652 is also 100%. Therefore, output shaft torque (T_C) 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:

T_C = T_S + T_A

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 T_I) was engine torque * 1.444. Torque on the rear annulus (T_A) was 60.52% of that, or about 87.4% of engine torque (1.444 * 60.52%). Torque on the rear sun gear (T_S) 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:

T_I = 150 lb-ft [203.4 N-m] * 1.444 = 216.7 lb-ft [293.8 N-m]
T_A = T_I * 60.52% = 131.1 lb-ft [177.8 N-m]
T_S = T_I * 39.47% = 85.5 lb-ft [116 N-m]

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

T_C = 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:

T_I = 150 lb-ft [203.4 N-m] * 1.00 = 150 lb-ft [203.4 N-m]
T_A = T_I * 60.52% = 90.8 lb-ft [123.1 N-m]
T_S = T_I * 39.52% = 59.2 lb-ft [80.3 N-m]

Torque on the output shaft was therefore:

T_C = 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).

DEMULTIPLYING SLIP

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 V_A) had to be greater than the velocity of the sun gear (V_S). 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 V_C) was proportional to the ratio of the planetary gears and the speed difference between the annulus and sun gear:

V_C = V_S + ((V_A – V_S) / (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:

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

… or:

V_C = 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 (although reducing the amount of torque transmitted through the coupling does reduce the amount of slip). Think of it rather as a slippage rebate: Some 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:

V_C = 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!)

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.

CONTROLLED COUPLING HYDRA-MATIC

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.

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.

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.

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, V_S = 95.96% and V_A = 98.93%. Plugging those into the formula on the previous page gives us:

V_C = 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.

LATER GM SPLIT TORQUE AUTOMATICS

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.)

DUAL-PATH TURBINE DRIVE

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.

The annulus had 80 teeth, the sun gears 46 teeth, and the planets each 17 teeth, making the ratio in low gear 1 + 46/80, or 1.575:1. Therefore, the annulus received 63.4% (1 / 1.575) of input torque in second gear, with the remaining 36.6% applied to the sun gears.

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.

TEMPESTORQUE

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 transmission main shaft.

Unlike Powerglide, which obtained direct drive by engaging a clutch to lock 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 was splined to the transaxle input shaft, which rotated at engine speed, not turbine speed.

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.

For 1963, TempesTorque’s final year, the direct clutch was revised so that its hub was splined to the main shaft rather than the input shaft, which eliminated the torque-splitting feature. In the 1963 transaxle, all power flowed through the input shaft to the torque converter and the main shaft in first, second, and reverse, just like Corvair Powerglide units.

ROTO HYDRA-MATIC

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.

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.

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.

RETURN OF THE LOCKUP CLUTCH

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.

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.

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.

CENTRIFUGAL AND VISCOUS LOCKUP CLUTCHES

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.

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.

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.

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.

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.

FORD’S SPLIT TORQUE AUTOMATICS

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.

AUTOMATIC OVERDRIVE TRANSMISSION (AOD)

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.

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.

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.

ATX TRANSAXLE

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.

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 (V_A) was engine speed while sun gear speed (V_S) was turbine speed, the speed of the splitter gearset’s planet carrier (V_C) was therefore:

V_C = V_S + ((V_A – V_S) / (1 + sun gear teeth / annulus teeth))

… or:

V_C = V_S + ((V_A – V_S) / 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.

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.

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).

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.

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.

THE TRIUMPH OF ORTHODOXY

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 warm-up 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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NOTES ON SOURCES

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