International Journal of Emerging Technology and Advanced Engineering
Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 3, Issue 8, August 2013)
Continuously Variable Transmission using the Inner Surface of
a Spherical Rotor and Conceptual Design for the Automotive
Transmission
Hyoungwoo Lee1, Kibong Han2
1,2
Department of Mechatronics Engineering, Jungwon University, South Korea
A half-toroidal type of CVT has been applied since 1999
by Nissan Corporation. But many problems have occurred
such as spin loss, low torque capacity, severe rotor wear
because of a small contact area, large size of CVT, etc.
ISCVT overcomes these problems. The shape of the
ISCVT contact area is always circular compared to the
elliptical area in TCVT. [2-7] Therefore, ISCVT improves
both the power transmissibility and the contact pressure.
A new ISCVT is designed for automobiles with 110 kW
and maximum torque (194 Nm / 4500 RPM). The size of
the ISCVT is 220 mm × 150 mm in the overall diameter
and width. This ISCVT is compared with a toroidal CVT in
terms of the transmission efficiency, maximum shear stress,
and life time; it is shown to be more excellent than TCVT.
Abstract—A new, continuously variable transmission
(ISCVT) that contacts with the inner and outer surfaces of a
spherical rotor is introduced. ISCVT consists of four units, a
pair of driving/driven rotors, a pair of pressure devices, four
traction ball assemblies, and a ratio changer. The four
traction balls are situated between the driving and the driven
rotors. The ratio changer is located inside one pair of rotors
and rotates with a helical groove. The traction ball connector
moves through the groove in the ratio changer and varies the
traction ball angle, which can change the speed of the driven
rotor. The speed ratio of ISCVT is derived by kinematic
analysis and determined by the height of the point of contact.
We applied the moment equation in order to find the
transmission efficiency. ISCVT was applied to a 110 kW
passenger car. We evaluated its applicability to automobile
usage through a numerical investigation. The maximum shear
stresses are formulated by Hertzian contact theory. The lifetime is calculated by the Lundberg-Palmgren method.
Simulation results show that ISCVT with four traction ball
assemblies is very compact with a high power density and
high transmission efficiency. Based on this simulation, we
have designed ISCVT and performed stress analysis to find
shortcomings in ISCVT. The associated transmission
performance is compared with that of toroidal CVT (TCVT)
and shown to be excellent. We also applied ISCVT with
higher capacity for a construction vehicle and truck.
II. BASIC INNER SPHERICAL TRACTION DRIVER CVT
MODEL
A. Layout
The ISCVT is illustrated in Figure 1. The dimension and
geometry of the ISCVT are defined by the following design
parameters. The curvature of the driving rotor is r1. The
traction ball radius is r . n10 , n20 , and n are the direction
vectors of the driving/driven rotors and the traction ball,
respectively. 1 and  2 are the angle between the rotor axis
and the contact point, respectively.  is the tilting angle of
the traction ball. The radius of the driven rotor is r2 . h1 , h2 ,
Keywords—CVT, Inner spherical continuously variable
transmission (ISCVT), transmission performance, OSR,
transmission efficiency, power density, gradability, life-time.
h3 , and h4 are the lengths between the driving rotors,
driven rotors, traction ball axis, and center of the contact
area, respectively.
I. INTRODUCTION
Existing geared vehicle transmissions use a clutch that
isolates the engine power. Geared transmissions result in
shifting shock because the speed is changed during high
speeds of about 3000~6000 RPM. But a continuously
variable transmission (CVT) can continuously transfer
power from low to high speeds. CVT has no shifting
shocks, which makes it quiet and reduces transmission
vibration. Given the continually increasing global concern
of consumers with environmental issues, users demand
higher efficiency. CVT has already achieved a 10 percent
improvement in fuel economy.[1]
B. Operating Principle
The ISCVT consists of four units, a pair of
driving/driven rotors, a pair of pressure devices, four
traction ball assemblies, and a ratio changer.
The power is generated from the engine or motor and
transmitted to the driving rotor. The driving and driven
rotors are connected with a pressure device, as shown in
Figure 2.
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The driving rotor rotates the traction ball by the pressure
device due to rolling contact without slip between the
driving rotor and the traction ball. The speed of rotation of
the driven rotor is changed by tilting the traction ball.
The pressure device shown in Figure 3 consists of an
inclined guide and three balls. The balls are set between the
input guide and the driving rotor guide. The pressure
device generates normal forces that create transmitted
torques. The input torque results in tangential and normal
forces. The tangential force pushes each rotor and generate
contact pressure.
The thrust forces are generated by the inclined guide,
which generates the normal force proportionally. And are
the tangential and thrust components of the resulting force.
The tangential and thrust forces are defined as follows.
Ti
.
ri
Ft 
Fth 
Ft
T /r
 i i .
tan  tan 
N1 
Fth
T1 / r1
.

cos1 tan  cos1
(3)
In the above, Ft and Fth are the tangential force and the
thrust force that are applied at the traction ball, respectively,
and  is the cam lead angle. In the expression above, N1 is
proportioned to the torque, T1
This type of pressure device results in high efficiency
across a range of speed ratios. Since the angle of inclination,
 , is related to the efficiency, it is obtained by an optimal
process. Figure 4 shows the average efficiency vs. the
inclined angle. The efficiency is maximum in the
neighborhood of 30°.
The assembly of the traction ball consists of four traction
balls. Each traction ball has two bearings on the traction
ball shaft and the bearing housing. The traction ball is fixed
by the traction ball jig. The speed of the driven rotor is
changed by varying the tilting angle of the traction ball
shaft. The traction ball rotates according to the pivot point,
O , which is located at the spherical center of the traction
(1)
(2)
ball. Figure 5 shows a free body diagram of the traction
ball.
The work that is applied to the ratio changer when the
ratio is changed from the lowest speed ratio to the highest
value is calculated from the normal forces and the distance
moved.
FIGURE 1 STRUCTURE OF SWASH PLATE PISTON PUMP BASED
AUTOMATIC LUBRICATION SYSTEM FOR COMMERCIAL TRUCKS.
FIGURE 3 PRESSURE DEVICE IN THE DRIVING ROTOR
Wrc  max ( N1  N2 )r (max  min ) .
(4)
In the above, N1 and N 2 are the contact loads at the
driving rotor and the traction ball, while max and min are
the maximum and minimum static traction coefficients.
FIGURE 2 ASSEMBLY OF THE ISCVT (NEUTRAL POSITION).
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max and min are the maximum and minimum phase
angles of the traction ball, respectively. The tilted shape of
the ratio changer is shown in Figure 6.
The ratio changer enables continuously variable
transmission (CVT) for the driven rotor. The traction ball
connector moves through the helical groove shown in
Figure 6. The height of the point of contact changes
according to the movement of the connector.
V1  1  r1  1  (n  r1 ) .
(7)
Assuming that the driving rotor and the traction ball
have no slip, we can apply the velocity constraint on the
driving rotor and the traction ball. The rotational speeds of
the driving rotor and the traction ball can be written as
follows.
V1  V1 , 1  1  (n10  r1 ) .
(8)
(n  r1 )
III. KINEMATIC AND KINETIC ANALYSIS
Similar to Eqs. (5)~(8), the driven rotor speed can be
expressed as:
A. Kinematic Analysis
In order to find the speed ratio of ISCVT, we have
formulated the driving rotor speed, traction ball speed, and
driven rotor speed, respectively. The driving/driven rotor
speeds and the traction ball speed are as follows.
 2 
2  (n20  r2 ) .
(9)
(n 2  r 2 )
The traction balls in the assembly are connected to each
other; hence, the speeds of the traction balls are the same.
Thus, we rearrange Eqs. (8) and (9) as follows.
1   2 , 1  (n10  r1 )  2  (n20  r2 ) .
(n  r1 )
(10)
(n 2  r 2 )
The angular velocities of the driving rotor and the driven
rotor are related by:
FIGURE 4 EFFICIENCY VS. THE INCLINED ANGLE.
FIGURE 6 OPERATION OF THE RATIO CHANGER.
2 n40  (n10  r1 ) n40  (n 2  r 2 ) .

1 n40  (n20  r2 ) n40  (n  r1 )
(11)
Here, r1 is defined in terms of the angle subtended by the
contact point and the radius of the rotor as follows.
FIGURE 5 THE PRINCIPLE OF SPEED CHANGE.
r1  r1 cos1 (n20 )  r1 sin 1 (n30 ) .
(12)
(13)
V1  1  r1  1  (n10  r1 ) .
(5)
n10  r1  r1s in 1  n40  h1  n40 .
V2   2  r2  2  (n20  r2 ) .
(6)
Similar to Eq. (13), we find all the components of the
speed ratio.
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n20  r2  r2 s in 2  n40  h4  n40 .
(14)
n  rc 2  r2 s in 2  n40  h3  n40 .
(15)
n  rc1  r1s in 1  n40  h2  n40 .
(16)
By substituting (13), (14), (15), and (16) in (11), the
overall speed ratio (OSR) between the driving rotor and the
driven rotor is seen to be related to the height of the contact
point. The speed ratio of ISCVT is represented as follows.

h1 h3
.
h2 h4
(17)
In the above,  is the OSR.
B. Kinetic Analysis
A simple contact model based on Hertzian contact
theory [2] is applied in this study. The maximum shear
stress in the traction rotor is proportional to the 2/3rd order
of the geometric factor, which is defined by the sums of the
inverse of the radii of curvature. The traction coefficient for
the fluid used, SANTOTRACK50, is obtained from the
results of the experiment. The referenced property of
traction fluids is shown in Figure 7.
The power that is supplied from the driving rotor at A1
rotates the traction ball shown in Figure 8 and is
transmitted to the driven rotor through A2 . The velocity
differences between the traction ball and the driven rotor
generate the traction slip, which generates  2 . In order to
FIGURE 7 THE CHARACTERISTIC CURVE OF THE TRANSMISSION FLUID,
SANTOTRACK50.
find  2 , the traction ball speed must be calculated to
(A) DRIVING ROTOR.
ascertain the transmission efficiency.
The normal force is distributed over the contact area. In
order to supply normal force to the driving rotor, pressure
is applied by the pressure device. Shear force is generated
at the same time regarding the fluid, SANTOTRACK50.
This shear force results from the difference in velocity
between the driving rotor and the traction ball. This sliding
velocity results in power loss. We found the power loss by
searching for the traction ball velocity,  . To find  , we
FIGURE 8 THE PRINCIPLE OF SPEED CHANGE.
In the above, T1 is the input torque and A1 is the contact
area between the driving rotor and the traction ball. O1P1 is
the direction vector from the center of curvature to the
center of the contact area. PQ
1 1 is the direction vector from
the center of the contact area to the center of an
infinitesimal area.
The shear stress,  1 , is defined in Eq. (19) assuming that
apply the moment equation of the driving axis.
T1  n10  A1 (O1 P1  PQ
1 1 ) 1dA1  0 .
(B) CONTACT AREA.
the  1 direction is the same direction as that of the sliding
(18)
velocity. We define  1 as follows.
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1   p
IV. CONCEPTUAL DESIGN AND PERFORMANCE ANALYSIS
Vsd
.
(19)
A. Design procedure of ISCVT
To show practicability in automobile usage, a 2000 cc
passenger car is considered for numerical investigation.
The main design specifications are the input speed, input
torque, and overall speed ratio. The design specifications
are listed in Table 1.
Vsd
In the above,  1 is the shear stress vector at the
infinitesimal area,  is the traction coefficient, p is the
pressure at the contact area, and Vsd is the sliding velocity.
The friction factor is a function of the normal force and the
creep rate:    ( p, Cr ) . The creep rate, Cr , is defined as
follows.
Cr 
Vsd
TABLE I
DESIGN SPECIFICATIONS OF ISCVT FOR 110 KW
▪ Displacement
▪ Max. power
.
(20)
▪ Max. torque
V1
▪ Input speed range
▪ Input torque range
▪ Overall speed ratio
▪ Driving / driven rotor diameter
range
▪ Range of the radius of traction ball
▪ Range of the height of the traction
ball pivot
▪ Range of the preloaded thrust
forces
▪ Range of the cam lead angle
Through this process,  is derived.
T2 is calculated from the free body diagram of the
driven rotor, which is shown in Figure 9.
The equation is:
T2  n20  A2 (O2 P2  P2Q2 )  2 dA2  0 .
(21)
The total efficiency of ISCVT is defined as follows.
T
 2 2 .
T11
(A) THE DRIVEN ROTOR.
(22)
2000 cc
110kW / 6,000
RPM
194N·m /
4,500RPM
0.1~8000 RPM
0.1~250 Nm
0.09~0.37
100 ~ 200 mm
10~50 mm
50 ~ 100 mm
0.1 ~ 500 N
0.1 ~ 50°
The simulation is performed by the following procedure.
Firstly, we input the design parameters and mechanical
properties. The design variables are the speed ratio of the
gear reducer, the radius of the traction ball, the height of
the pivot for fixing the traction ball assembly, the radius of
the inner driving/driven rotors, the cam lead angles of the
pressure devices, and the preloaded thrust forces.
Secondly, we set the design constraints: the maximum
shear stresses should not exceed 800 MPa; the average
efficiency must exceed 60%; the overall diameter of
ISCVT is less than 510 mm; and the height of the traction
ball assembly must not exceed 65 mm.
After determining the geometry of ISCVT, the
simulation finds the maximum transmission efficiency
through the overall range of the speed ratio. The optimal
process is a direct search method through parametric study.
Figure 10 depicts a flowchart of the simulation.
(B) CONTACT AREA.
FIGURE 9 FREE BODY DIAGRAM OF THE DRIVEN ROTOR.
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FIGURE 11 CAD ON ISCVT.
B. Stress analysis
To analyse the shortcomings of the driving rotor and
traction ball housing, we performed stress analysis. Figures
12 and 13 show the stress contours of ISCVT parts. The
load condition is the maximum torque at 4500 RPM. This
means that the simulation is performed under the worst
conditions.
FIGURE 10 FLOWCHART OF THE SIMULATION.
Table 2 shows the results on the layout and performance
from simulation at the maximum torque condition.
TABLE II
SIMULATION RESULTS FOR ISCVT.
Layout results
▪ Radii of the driving/driven rotors
▪ Radius of the traction ball
▪ Height of the traction ball pivot
125 mm
43.3 mm
52 mm
▪ Cam lead angle
36°
▪ Preloaded thrust force
Performance results
▪ Transmission efficiency
▪ Work for the ratio changer
▪ Life time
▪ Maximum shear stress
▪ Gradability
220 N
93 %
263 Joules
108,000 hour
552 MPa
20°t
FIGURE 12 STRESS CONTOURS OF THE DRIVING ROTOR AND A
TRACTION BALL.
The assembly design is shown in Figure 11 based on the
result of simulation.
FIGURE 13 STRESS CONTOURS OF THE TRACTION BALL JIG AND
HOUSING.
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The results of analysis are listed in Tables 3 and 4. It is
reasonable to use normal carbon steel such as SCM440
(AISI4140) with a yield strength of 1,650 MPa. The results
show that ISCVT can have lower thickness and weight
because of the low stress as compared with the yield
strength.
TABLE III
DRIVING ROTOR AND TRACTION BALLS
Jig
4,104 / 12,005
Tetrahedral
▪ Node / element
▪ Element type
Housing
1,198 / 3,329
9
▪ Material property
E=200 × 10 Pa,
 =0.3
▪ Max disp.
0.02 mm
0.18 mm
▪ Max stress
200 MPa
559 MPa
FIGURE 14 THE AVERAGE POWER EFFICIENCY.
TABLE IV
TRACTION BALL JIG AND HOUSING
▪ Node / element
▪ Element type
Driving rotor
4,104 / 12,005
Tetrahedral
Traction ball
4,179 / 20,097
9
▪ Material property
E=200 × 10 Pa,
 =0.3
▪ Max disp.
0.1mm
0.02mm
▪ Max stress
141MPa
339MPa
C. Performance analysis
The power efficiency is calculated according to an inputspeed range of 0~8,000 RPM and an input-torque range of
0~250 Nm. These are the operating conditions of
automobiles. The efficiency refers to the average value in
the region of the speed ratio. In Figure 14, the efficiency
over the 90% region is widely distributed
The maximum torque of 194 Nm and rated speed of
4,500 RPM are generated by the engine. According to the
speed ratio and input torque, the efficiency distribution is
shown in Figure 15. As the speed ratio changes from the
high-speed mode to the low-speed mode, the efficiency
variation is not significant. This means that the efficiency is
over 90% in the entire range of operation and the highperformance region is also widely distributed.
FIGURE 15 EFFICIENCY AT THE RATED SPEED.
The shear stresses are shown in Figures 16 and 17 at the
rated speed. The maximum shear stress is within 510~555
MPa in the driving rotor and within 510~630 MPa in the
driven rotor at the maximum torque of 193 Nm. Therefore,
a typical carbon steel such as SCM440 (AISI4140) is
sufficient as material for the traction rotor.
Based on the Lundberg-Palmgren method, the life cycle
of ISCVT is calculated and shown in Figure 18. The lifetime at the maximum transmitted torque of 193 Nm being
the worst condition is 80,300 hours. This means that the
ISCVT’s life-time is more than five years at the worst
condition.
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The required work to operate the ratio changer over the
full speed range is shown in Figure 19. The maximum
value is 280 Joules. At this value, a 300 W stepping motor
is sufficient for varying the speed ratio.
For ISCVT, we define the gradability as the maximum
angle of inclination. The gross vehicle weight is 1660 kg
and torque is 193 Nm at 4500 RPM in the lowest speed
mode the allowable angle of inclination of roads over
which the vehicle can drive is approximately obtained as
follows.
 Fthrust
  sin 1 
 Wg
 Ctire 
air CareaV 2 
2Wg
 .

(23)
In the above,  is the inclined angle, Fthrust is the thrust
force due to the rear wheel, W is the gross vehicle weight,
Ctire is the tire resistant coefficient, air is the air density,
Carea is the projection area of the automobile, and V is the
vehicle speed.
Figure 21 shows that ISCVT is capable of ascending
over an inclination of 20° at the maximum torque condition.
FIGURE 16 MAXIMUM SHEAR STRESS IN THE DRIVING ROTOR.
FIGURE 17 MAXIMUM SHEAR STRESS IN THE DRIVEN ROTOR.
FIGURE 19 WORK FOR THE RATIO CHANGER.
FIGURE 20 A MODEL OF GRADABILITY.
FIGURE 18 LIFE-TIME OF THE DRIVING ROTOR.
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The power density based on the volume is defined as the
maximum power divided by the overall volume. An ISCVT
with an efficiency of 93%, minimum life cycle of 80,300
hours, and maximum shear stress of 600 MPa can be
designed to a diameter of 220 mm and width of 150 mm, as
shown in Figure 22(a). The ISCVT power density is
calculated to be about 19.3 kW / .
The comparison between the TCVT and ISCVT are
shown in Figures 23~25. In Figure 23, the average through
the overall speed ratio of the power efficiencies are about
93 %.
Figure 24 shows that the maximum shear stresses in
ISCVT are much less than those in TCVT.
D. Comparison with the toroidal CVT model
To show the superiority of the proposed CVT model, a
toroidal model of similar size as that of the designed
ISCVT is considered, as shown in Figure 22(b). It has three
traction balls that relay motion between the driving and
driven rotors. By applying the design specifications in
Table 1 to the artificial toroidal model, the overall diameter
and width are found to be about 160 mm and 240 mm,
respectively. The transmission performance is calculated
with respect to input torques of 0.0~250 Nm.
FIGURE 23. COMPARISON BETWEEN ISCVT AND TCVT REGARDING
THE EFFICIENCY.
FIGURE 21 INCLINED ANGLE AT THE LOW-SPEED MODE.
FIGURE 24. COMPARISON BETWEEN ISCVT AND TCVT REGARDING
THE MAXIMUM SHEAR STRESS.
(A) ISCVT.
(B) TCVT.
FIGURE 22 THE OVERALL SIZE OF ISCVT WITH FOUR BALLS AND
TCVT WITH THREE BALLS.
FIGURE 25. COMPARISON BETWEEN ISCVT AND TCVT REGARDING
THE LIFE-TIME.
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Figure 25 shows differences in the life-time. For input
torques in the vicinity of 100 Nm, the ISCVT life-time is
twice that of TCVT.
Consequently, in comparison with TCVT, ISCVT is
expected to be excellent in terms of the transmission
performance. In Table 5, the performance of TCVT and
ISCVT is summarized.
V. CONCLUSION
We introduce a new traction drive CVT and apply it to a
110 kW passenger car to evaluate its practicability for
automobiles. We have derived the speed ratio of ISCVT by
kinematics and the moment equation regarding equilibrium.
This paper has presented the conceptual design and
performance analysis through numerical investigation of
ISCVT by focusing on its basic components.
With the proposed mechanism, we can get traction CVT
with very high efficiency (over 90%) in the ranges of the
maximum speed and the maximum torque. ISCVT has high
power density (more than 9.3) even though the material of
the traction rotor is carbon steel, such as SCM440 for
which the yield strength is 1800 MPa. The expected lifetime is derived by the Lundberg-Palmgren method and the
gradability is verified.
TABLE V
COMPARISON BETWEEN TCVT AND ISCVT
▪ Efficiency
▪ Max. shear stress
TCVT
110kW / 6000
RPM
90 %
920 MPa
ISCVT
110kW / 6000
RPM
93.4 %
545 MPa
▪ Life time
7,950 hr
80,300 hr
▪ Weight
▪ Work for the ratio
changer
▪ Power density
15.7 kg
8.2 kg
571 Joules
246 Joules
▪ Maximum power
8.27
kW /
9.30
kW /
E. Analysis of the torque capacity vs. the overall size
To analyze higher power capacities and larger sizes, the
simulation range is expanded in terms of the overall
diameter and input power. The results show that when the
rotor is larger than 250 mm, the performance is high
(efficiency of over 92%), as shown in Figure 26.
The maximum shear stress is considered for largecapacity vehicles. Under 630 MPa, the rotor can be used
for small vehicles, as shown in Figure 27.
The life-times are compared across the various
capacities shown in Figure 28. The simulation result shows
that the fatigue life-time is over 40,000 hours.
FIGURE 27. SIMULATION RESULTS FOR THE MAXIMUM SHEAR STRESS
UNDER HIGH CAPACITIES AND LARGE OVERALL SIZES.
FIGURE 28. SIMULATION RESULTS FOR THE FATIGUE LIFE-TIME UNDER
HIGH CAPACITIES AND LARGE OVERALL SIZES.
FIGURE 26. SIMULATION RESULTS FOR THE EFFICIENCY UNDER HIGH
CAPACITIES AND LARGE OVERALL SIZES.
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[3]
The results show that ISCVT can incline over a 20°
ramp at the maximum torque. The associated transmission
performance is calculated to be excellent in comparison
with TCVT; ISCVT also is shown to be practicable for
mid-sized automobiles. This study also suggests ISCVT for
vehicles with high power capacities such as construction
vehicles or trucks. The results show that ISCVT is
applicable for all kinds of vehicle.
[4]
[5]
[6]
REFERENCES
[1]
[2]
Takashi Imanishi and Shinji Miyata, (2003). Development of the
Next-Generation Half Toroidal CVT. Motion & control No. 14, 2024
Norton, R. L. (2000). Machine Design – An Integrated Approach,
2nd edn., Prentice-hall Inc, N. J..
[7]
695
Ryu, J. H, Seong, S. H. and Park, N. G. (2002). A continuously
variable transmission having a four bar linkage and spherical rotors,
Proc. 2002 KSAE Spring Conference, 2, 889-896, May 30-June 1,
Jeju, Korea.
Y. Zhang, X. Zhang, W. Tobler, (2000). A Systematic Model For the
Analysis of Contact, Side Slip, and Traction of Toroidal Drives,
Journal of mechanical design, 122, 523-528.
Yasuhiro Murakami, Yoshie Arakawa and Makoto Maeda, (1999)
Development and Testing of CVT Fluid for Nissan Toroidal CVT,
Society of Automotive Engineers, Inc, 1-8
Yoshihiro Ono, Adrian P. Lee, (2000). The Durability and
Reliability of Variators for Dual-capacity Full-toroidal CVT. SAE
international. 2000-01-0826
Masaki Nakano, Noboru Maruyama (2000). Development of a Large
Torque Capacity Half-toroidal CVT. SAE international. 2000-010825