The 1953 model year also saw the introduction of a heavily revised Buick automatic, dubbed Twin-Turbine Dynaflow. Developed by a group of Buick engineers led by Rudolf J. Gorsky, the twin-turbine transmission was again based on concepts originated in O.K. Kelley’s corporate engineering team; most of the underlying patents (in particular U.S. Patents 2,766,641; 2,782,659; and 3,025,720) were in Kelley’s name. The apparent objectives of the new transmission were to provide additional torque multiplication without hurting part-throttle fuel economy, raising the converter stall speed, or compromising the outstanding smoothness that had always been Dynaflow’s principal virtue.
Twin-Turbine Dynaflow retained the original Dynaflow’s Ravigneaux gearbox, but the torque converter was a new four-element design with a single impeller, a single stator, and — as the name implied — two turbines. The first turbine, which faced the impeller, was essentially a metal ring with closely spaced, slot-like radial vanes around its rim. That ring was pressed into a drum-like support shell, within which was the second turbine, a more conventional bladed torus. The stator, which sat within the first turbine ring, was positioned to receive the ‘backwash’ of oil exiting the second turbine.
Within the turbines’ central hub was an additional planetary gearset. (Insofar as the torque converter and gearbox were separate entities, this gearset was part of the converter.) The annulus of the converter gearset was driven by the first turbine support shell. The gearset’s planet carrier was attached at one end to the second turbine and at the other to the gearbox input shaft. The converter gearset sun gear, meanwhile, was connected to the hub of the stator and shared its one-way clutch. Reverse torque on either element would lock both elements, putting the converter gearset in reduction. This multiplied any torque applied to the first turbine by a ratio of 1.60:1 and forced the second turbine to rotate at 62.5% percent of the speed of the first turbine (i.e., first turbine speed divided by 1.6).
The stream of oil from the impeller would first enter the first turbine, where the oil attempted to impart its angular velocity — that is, to apply torque — to the turbine vanes. Thanks to our old pal, Newton’s Third Law of Motion, whatever torque the oil stream exerted on the turbine vanes would apply an equal and opposite reaction torque on the oil. Since the vanes of the first turbine were open at the back, oil passing through them retained its forward momentum, but the reaction torque effectively reduced the oil’s angular velocity. Oil exiting the first turbine would then enter the inlet of the second turbine, pass through its vanes to the turbine outlet, and then curve back through the stator blades to the impeller.
Up to the point of stall (that is, as long as neither turbine was moving), the oil stream would apply all or nearly all of its torque to the vanes of the first turbine. Consequently, at stall and for a brief period thereafter, the vanes of the first turbine exerted so much reaction torque on the oil stream that the oil exited the first turbine spinning in the opposite direction and therefore opposed the rotation of the second turbine, slightly reducing the net torque on the gearbox input shaft.
Once the first turbine began to move, the oil stream exerted progressively less torque on the first turbine’s vanes, which in essence were now trying to run away from the spinning oil stream. (This is a gross simplification of some rather complicated vector math, but we figure you’re probably confused enough already.) The reduced torque on the first turbine’s vanes meant the vanes also exerted progressively less reaction torque on the oil, allowing the oil stream to enter the second turbine with a somewhat reduced but still positive angular velocity (that is, spinning in the same direction as the impeller) and exert a steadily increasing positive torque on the second turbine’s vanes. To put it another way, discounting slippage, any impeller torque not applied to the first turbine would be applied to the second. Again, torque on the first turbine was multiplied by the planetary gears; torque on the second turbine was not.
As long as the sun gear clutch remained locked, the planetary gearset forced the two turbines to maintain a fixed speed ratio. Therefore, the second turbine could not turn faster (or slower) than 62.5% of the first turbine’s speed. Once the rotational speed of the first turbine was close to impeller speed, however, there was enough torque on the second turbine to cause it, and thus the planet carrier, to overdrive the annulus and the first turbine rather than being driven by them in reduction. That unlocked the sun gear/stator clutch (causing all torque multiplication to cease) and allowed the stator, the sun gear, the annulus, and the first turbine to freewheel idly; the first turbine was now turning fast enough that the oil stream could no longer exert a meaningful amount of torque on the turbine vanes. The torque applied to the second turbine, meanwhile, caused the second turbine (and thus the carrier and the gearbox input shaft) to continue accelerating until it was rotating at close to engine speed.
Like any torque converter, the torque multiplication provided by the dual-turbine converter was continuously variable, peaking at stall and gradually diminishing with increasing turbine speeds. However, there was now significantly more area under the curve thanks to the converter gearset’s additional mechanical advantage. Despite the initial interference between the turbines, which hurt the converter’s efficiency at stall, the converter gearset allowed a higher net stall ratio — now 2.45:1 — and somewhat lower stall speeds. The use of two separate turbines also allowed each to be optimized for its respective operating regime, providing more efficient cruising for better fuel economy without sacrificing off-the-line performance.
Twin-Turbine Dynaflow’s principal shortcomings were engine braking and passing response. Even with the converter gearset, there was still little engine braking in Drive; the sun gear clutch would automatically unlock if the output shaft overran the engine. The converter gearset was also of marginal usefulness for passing unless turbine speeds fell significantly below engine speed. Shifting to Low range mitigated both these issues, but Low was really too short to be ideal for passing or mountain driving at highway speeds. Dual-Range Hydra-Matic was much more convenient in those situations.
Buick attempted to address that limitation with the 1955 introduction of Variable Pitch Dynaflow. The revised transmission retained the four-element torque converter and converter gearset, but added Kelley’s latest brainstorm (described in U.S. Patent No. 2,999,400): a variable-pitch feature for the stator blades, similar in principle to a variable-pitch propeller. Rather than being affixed to the stator hub in the usual manner, the stator blades were connected via a series of small crank pins to a servo-controlled annular piston (basically a flat metal ring) that could pivot forward or backward, thus rotating each blade on its crank. Hydraulic pressure on the piston normally held the blades at a low angle relative to the oil stream. Flooring the accelerator, or moving the selector to Low or Reverse, opened a control valve to exhaust one side of the stator servo, flipping the piston to its forward position and cranking the stator blades to a more upright angle. Backing off on the throttle would reengage the servo, causing the piston to flip back to its normal position and thus crank the stator blades back to low angle.
With the blades at their low-angle position, the converter traded some torque multiplication — net stall ratio in low was 2.10:1 — for nominal stall speeds as low as 1,400 rpm, reduced throttle lag, and greater efficiency at cruising speeds. Shifting the stator blades to high angle brought the net stall ratio to 2.50:1 and raised the stall speed to a nominal 2,600 rpm for stronger off-the-line performance and better passing response. Changing the stator pitch in this way wasn’t as effective as an additional reduction gear, but it was helpful nonetheless. The principal drawback was that at very high road speeds, forcing the stator blades to high angle would hurt performance more than it helped.
To take fuller advantage of the new stator, the converter gearset sun gear was divorced from the stator hub and given its own sprag clutch, separate from the stator’s cam-and-roller one-way clutch. Having its own clutch allowed the stator to remain locked after the first turbine freewheeled (further fattening the torque multiplication curve) or to re-lock in response to load without necessarily putting the converter gearset back in reduction.
Although Variable Pitch Dynaflow provided slightly better performance and somewhat better fuel economy than the earlier Twin-Turbine Dynaflow, an effective starting ratio of no more than 2.50:1 in Drive was still marginal for the steadily increasing curb weights (and steadily decreasing axle ratios) of mid-fifties Buicks. This was addressed for 1956 with a revised five-element torque converter that incorporated dual stators as well as twin turbines.
The additional stator — confusingly described as the first stator or front stator — was mounted immediately behind the first turbine and looked much like it. However, the stator blades were angled in more or less the opposite direction so as to counteract the reverse torque that had previously compromised the twin-turbine converter’s efficiency near stall speed. Now, oil entering the second turbine at stall increased the torque on the turbine vanes and the planet carrier rather than opposing their rotation. To put it another way, with the turbines stationary or turning slowly, the rotary flow in different points of the converter was now like this:
- From the impeller outlet to the first turbine inlet: with the engine
- From the first turbine outlet to the first stator: opposite the engine
- From the first stator to the second turbine inlet: with the engine
- From the second turbine outlet to the variable-pitch stator: opposite the engine
- From the variable-pitch stator to the impeller inlet: with the engine.
Once the first turbine was turning fast enough that the rotary flow of oil out of the first turbine outlet was no longer opposite the engine’s rotation, the first stator would freewheel on its own sprag clutch.
The variable-pitch stator and converter gearset were retained, but the additional stator increased the converter’s net stall ratio to 3.10:1 at a nominal 1,500 rpm with the variable-pitch stator blades in their low-angle position or 3.50:1 at 2,800 rpm in high position. The variable-pitch stator controls were also modified so that the blades would normally remain at low angle in Low or Reverse rather than automatically switching to high angle in either of those gears.
Even the high-angle stall ratio didn’t quite match the first-gear ratio of the four-speed Hydra-Matic or the step-off ratios of contemporary two-speed torque converter automatics, but the additional torque multiplication made Dynaflow-equipped cars a good deal less sleepy when starting in Drive. More importantly, as far as Buick was concerned, that improved performance was still obtained without any perceptible shift points.
Variable Pitch Dynaflow — renamed Twin Turbine for 1959 and Turbine Drive for 1960 — received a variety of further refinements, including several revisions to the stator blade pitch (making the stall ratios 3.10:1 and 3.40:1); marginally higher stall speeds; and, from 1961 on, a shorter, slightly lighter case. Turbine Drive was the sole transmission offered on full-size Buicks from 1961 through 1963.
CONTROLLED COUPLING HYDRA-MATIC
Despite the ongoing development of Powerglide and Dynaflow, GM had no intention of abandoning Hydra-Matic, which was still used in substantial numbers by Pontiac, Oldsmobile, Cadillac, and several outside automakers. Aside from GM’s substantial capital investment in tooling and factory space, which the corporation wasn’t about to casually discard, the various users (and many of their customers) had strong feelings about the comparative advantages of Hydra-Matic and its assorted rivals.
In 1952, the Detroit Transmission Division embarked on a $35 million revamp of the four-speed Dual-Range Hydra-Matic. Walter B. Herndon, one of the engineers from Earl Thompson’s original transmission development group, filed a patent covering most of the fundamentals of the redesigned transmission (U.S. Patent No. 2,876,656) in November 1953, with most of the rest covered in a subsequent application by August H. Borman Jr., Forrest R. Cheek, and Milton H. Scheiter in December 1954 (U.S. Patent No. 3,048,055), but the second-generation Hydra-Matic didn’t actually go on sale until the 1956 model year. (We assume the destruction of the Hydra-Matic plant in Livonia in August 1953 was at least partly responsible for the delay.) Development of the production version, formally known as the Model 315 or Controlled Coupling Hydra-Matic, was credited to Detroit Transmission engineers P.J. Rhoads and Kenneth W. Gage; Gage subsequently moved to Buick, where he worked on later iterations of Dynaflow.
To understand the changes to the second-generation Hydra-Matic (called “Jetaway Hydra-Matic” by Oldsmobile, “Strato-Flight Hydra-Matic” and later “Super Hydra-Matic” by Pontiac), it’s helpful to first recap the major elements of the original version. As we’ve previously explained, the early Hydra-Matic had a single fluid coupling and three planetary gearsets controlled using two contracting band-type brakes, two multi-disc clutch packs, and (from 1951 on) a single cone clutch to provide four forward speeds and one reverse. The fluid coupling itself was driven indirectly: The torus cover, which was bolted to the engine flywheel, drove the annulus of the first planetary gearset, whose planet carrier drove a hollow intermediate shaft (surrounding and concentric with the transmission main shaft) that connected the fluid coupling impeller to the clutch assembly of the second gearset (which also partially bypassed the fluid coupling in third and fourth in order to reduce slippage). The fluid coupling’s turbine drove the transmission main shaft and the sun gear(s) of the second planetary gearset.
The redesigned transmission maintained the same basic layout (although some components were repositioned), but replaced the front clutch pack with a second fluid coupling — the eponymous controlled coupling — located immediately behind the torus housing, between the first and second planetary gearsets. The second coupling was smaller than the main coupling and incorporated valves that allowed its oil supply to be completely drained or completely refilled in less than half a second. The rear clutch remained a multi-disc unit, although it was beefed up for greater torque capacity. (The design team considered adding a third fluid coupling to replace the rear clutch, but ultimately decided the benefits weren’t worth the substantial extra cost.)
The second fluid coupling had the same function as the multi-disc clutch it replaced: to put the front planetary gearset in direct drive by causing the annulus, the sun gear, and the planet carrier to rotate together at the same speed, or close to it. The main coupling torus cover drove both the front unit annulus and the impeller of the second coupling (through its torus cover, which also drove the front oil pump). The second coupling’s turbine was connected (via a hollow sleeve shaft) to the front unit sun gear. If the coupling was empty, the impeller simply turned idly and the turbine remained stationary. Refilling the coupling would cause the impeller to drive the turbine — and thus the sun gear — at close to engine speed. (In technical terms, filling the second coupling split the engine’s torque between the annulus, which was driven mechanically, and the sun gear, which was driven hydraulically. The torque was then recombined by the planet carrier.)
The redesigned transmission also deleted the earlier Hydra-Matic’s front brake, replacing it with a sprag-type one-way clutch that performed the same function: holding the front gearset sun gear in place whenever the front clutch was disengaged (or in this case empty). A similar sprag clutch was attached to the annulus of the second planetary gearset. Since the sprag clutches didn’t require any external engagement mechanisms, automatic shifts up or down could now be accomplished by controlling the front coupling and the rear clutch (as shown in the table below) rather than simultaneously coordinating clutch and brake engagements. The sprags also needed no routine adjustment.
The use of the sprag clutches necessitated an alternative means of obtaining neutral and reverse, which both required that the rear annulus be able to turn backward. In the earlier single-coupling Hydra-Matic, that was achieved by simultaneously releasing the rear clutch and the rear brake band, but the new transmission’s rear sprag clutch couldn’t be disengaged that way. Instead, the Controlled Coupling Hydra-Matic interposed a multi-disc neutral clutch between the rear sprag’s outer race and the transmission case. The neutral clutch was engaged in all forward gears, allowing the rear sprag to function normally. In neutral or reverse, with the neutral clutch disengaged, the sprags wouldn’t lock even if turned backward; reverse torque would just cause the neutral clutch hub to rotate backward along with the rear annulus. The front sprag, which had no such mechanism, remained locked in both neutral and reverse as long as the engine was running.
Another complication of the sprag clutches was that they would release on the overrun, so second or third gears provided no more engine braking than fourth and the transmission would freewheel when coasting in first. To compensate, the Controlled Coupling Hydra-Matic retained the rear brake band — now called the overrun band — and added a single-disc overrun clutch that could be engaged to lock the front unit sun gear sleeve shaft. The overrun clutch and overrun band served as auxiliary brakes, supplementing the sprag clutches in Low and D3 (aka S or D-Right) ranges. (The overrun clutch was also locked in reverse.) Neither the overrun clutch nor the overrun band was operative in D4 (aka D or D-Left) range, so there still wasn’t much engine braking in that range. Given the limitations of contemporary drum brakes, selecting D3 or Low for mountain driving or maneuvering on steep grades was prudent.
The redesigned Hydra-Matic now had a Park position on the selector, a first for the Hydra-Matic series. The parking pawl that position controlled wasn’t entirely new: Hydra-Matic had always incorporated a pawl to lock the annulus of the third planetary gearset, originally to provide reduction in reverse and, after the reverse cone clutch was added for 1951, later for use as a parking brake. The parking pawl on the Controlled Coupling Hydra-Matic now acted on the reverse planetary gearset’s planet carrier rather than the annulus and could be used in addition to or instead of a conventional emergency brake acting on the rear drums.
Since both fluid couplings were active in fourth gear, the Controlled Coupling Hydra-Matic also slipped a bit more at cruising speed than did its single-coupling predecessor. The torque split in third and fourth gears mitigated that somewhat, but the redesigned transmission nonetheless sacrificed some fuel efficiency. Interestingly, Herndon’s 1953 patent disclosure included provision for a mechanical lockup clutch to completely eliminate the second coupling’s additional slippage in fourth gear, but that feature was absent from the production transmission.
The Controlled Coupling Hydra-Matic’s principal advantage was significantly smoother shifts than the single-coupling Hydra-Matic could muster. The rear clutch could still produce a mild thump on 2–3 or 3–2 shifts, but it was seldom objectionable and the 1–2 and 3–4 shifts were almost seamless. Shift quality was also more consistent than before — a distinct improvement over the single-coupling Hydra-Matic, which was very sensitive to proper adjustment of its bands and linkages. A bit of straight-line performance was inevitably sacrificed for that smoothness, but after 16 years of complaints about the endemic jerkiness of the single-coupling Hydra-Matic, that was a tradeoff many were prepared to accept.
Unfortunately, owners found that the new Hydra-Matic was somewhat less rugged than the single-coupling transmission it replaced. Particularly on early units, operation of the second coupling could be erratic in extreme temperatures, the aluminum torus cover was prone to cracks, and aggressive driving could damage the sprags of the one-way clutches. A litany of running changes progressively addressed most of those issues, but it’s interesting to note that GMC and Chevrolet trucks stuck with the older Dual-Range Hydra-Matic until the early sixties. (So did Rolls-Royce, which built the Dual-Range Hydra-Matic under license.)
The update also did nothing to reduce Hydra-Matic’s considerable weight, which now ran to some 225 to 240 lb (102 to 109 kg), or make it cheaper to build; it was undoubtedly one of the most costly, if not the costliest, of contemporary automatics. Consequently, there were fewer outside users than before. American Motors purchased some dual-coupling Hydra-Matics (which AMC dubbed “Flashaway”) for 1956–1957 Hudson and Nash models, but subsequently switched to less-expensive Borg-Warner (and later Chrysler) automatics. Even within GM, cost considerations would soon prompt Oldsmobile and Pontiac to adopt cheaper alternatives, although some Cadillac and Pontiac models would retain the Controlled Coupling Hydra-Matic through the 1964 model year.