International Journal of Mechanical and Production Engineering Research and Development (IJMPERD) ISSN 2249-6890 Vol. 3, Issue 4, Oct 2013, 1-10 © TJPRC Pvt. Ltd.

STUDY AND ANALYSIS OF ANGULAR TORQUING OF ENGINE CYLINDER-HEAD BOLTS USING TORQUE-TO-YIELD BOLTS: A CASE STUDY CHAVAN D K1 & UTTARA TIPNIS2 1

Professor, Mechanical Engineering, MMCOE, Pune, Maharashtra, India 2

Graduate Engineering Student, MMCOE, Pune, Maharashtra, India

ABSTRACT Normally for conventional bolt tightening methods, we use an estimate of resisting in the joint for measuring preload. But with angular torquing, we tighten the bolt with an initial specific „snug torque‟, and then tighten it further with an angle to achieve precise clamping. This gives a more accurate tightening of the joint. This is necessary for high performance engines which require the utmost precision of engine head sealing as their engine heads are subject to higher cyclic loads. Hence, we use a new type of bolt called Torque-To-Yield bolt in conjunction with angular torquing to meet this requirement.Torque to yield bolts, also commonly referred to as angle torque or stretch bolts, are used in many of today‟s modern engines predominantly for cylinder head bolts, but also for main bearing caps. Compared to conventional type bolts, TTY bolts offer the engine manufacturer a number of advantages including; greater flexibility of design, reductions in component costs, more accurate assembly and reliability of seal. Engines designed utilizing TTY head bolts require fewer head bolts to achieve the desired clamping loads than those using conventional bolts. With fewer bolts the engine manufacturer has more flexibility in cylinder head and block design as well as reducing the cost of the engine.

KEYWORDS: Bolts, Engine Cylinder-Head, Tightening Beyond Yield, Angular Torquing INTRODUCTION Fasteners function in an engine to hold parts together. Understanding some of the science of fasteners and fastener tightening is necessary for an engine rebuilder who wants to keep fastener failures and engine failures to an absolute minimum. Bolt load applied to the joint by the fasteners seals a head gasket through head lift-off during firing and changes in temperature that occurs as an engine runs. Threaded fasteners can clamp materials together only when they are holding with the proper amount of tension. For this to happen they must be properly tightened. There are a number of tightening methods as follows: Torque Control The method is used for tightening a fastener to variable levels of utilization, although not over the yield point. During tightening roughly 90% of input energy is lost overcoming the mating friction under the head, nut, and mating threads. Only 10% of input energy is converted into bolt stretch. Generally, one of the most serious disadvantages with the torque-control method is the inaccuracy in preload. Usually, variations in the preload of

30–50% can be expected.

Torque and Angle Control This method involves tightening the fastener to a low initial “snug tight” condition and then applying a prescribed amount of turn to develop the required preload. The actual preload will depend on how far the nut is turned as well as how much preload was established prior to the turning.

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Yield Controlled Tightening This method, developed by the SPS organization, is also known under the proprietary name "Joint Control Method". Very accurate preloads can be achieved by this method by minimizing the influence of friction and its scatter. The method has its roots in a craftsman's "sense of feel" on the wrench which allowed him to detect the yield point of the fastener with reasonable precision. With the electronic equivalent of this method, a control system is used which is sensitive to the torque gradient of the bolt being tightened. Rapid detection of the change in slope of this gradient indicates the yield point has been reached and stops the tightening process.

TIGHTENING OVER THE YIELD POINT WITH ANGULAR TIGHTENING New tightening techniques have been developed and are being used more often in several industries with the aim of achieving higher and more consistent preload values. Tightening over the yield point results in preloads being less affected by friction than is the case in elastic tightening. The yield characteristic of a screw determines the preload and its scatter, which is often less than

10%. The torque required to reach the yield point of a screw is very uncertain due to the

usually large variations in screw strength and friction. However, applying a specific angle after an initial torque leads to more consistent preload levels. Indirectly, the angle tightening is a form of length measurement. The elongation of the screw and the compression of the parts are thus measured concurrently. The torque and angle method gives a more uniform preload than the torque-control method, as long as the fastener is brought into its plastic region. However, it will still give a variation in the preload. Compared with the torque-control method, the level of preload is in some cases increased by up to 70%. Following are the main considerations in tightening over the yield point with torque-angle method: Snug Torque The snug torque is the initial torque before application of the tightening angle. The minimum torque needed to close the gaps between the mating parts is called snug torque. Some of the applied torque and force will be used to close the gaps between the cylinder head, gasket and crankcase. Elastic Part of the Tightening Angle The applied tightening angle after snugging can be divided into elastic and a plastic angle. The elastic angular load is difference between yield force and snug force. Thus the elastic angle from snug to yield is estimated. Spring Constant of a Screw in Plastic Region The strain to rupture for a screw of property class 12.9 is approximately 8%, for property class 10.9 approximately 9%, and for 8.8 approximately 12%. If using screws as opposed to necked down bolts, the plastic elongation gained during tightening will be concentrated on the threaded section of the fastener. Therefore, the strain can be converted to elongation by multiplying it with the length of the unengaged threads, Lt. The recommendation here is not to exceed 50% of the elongation (or strain) to rupture when tightening above the yield point, and this also applies to possible retightening or reuse of a screw. An elongation of more than 50% of the elongation to rupture will not result in any increase in preload and is therefore undesirable. It is therefore recommended not to exceed a total elongation of 4%Lt, 4.5%Lt or 6%Lt for screws of property classes 12.9, 10.9 and 8.8, respectively. As a general value, the maximum recommended elongation is here set to 5% of the length of the unengaged threads. The spring constant of a screw tightened into the plastic region is obtained by

Study and Analysis of Angular Torquing of Engine Cylinder-Head Bolts Using Torque-To-Yield Bolts: A Case Study

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assuming a straight line drawn from the ultimate tensile load down to the yield load, both reduced due to torsion. Both loads vary randomly. Plastic Part of the Tightening Angle The maximum allowed tightening angle is determined by the recommended maximum elongation of a screw. The total, maximum allowed tightening angle is summation of Plastic Part of the Tightening Angle and Elastic Part of the Tightening Angle. Recommended Tightening Angle The total tightening angle, w, should be chosen from among the angles 30 deg, 60 deg, 90 deg..., or n*30 deg, where n=1, 2, 3... An angle of 45 deg is also acceptable. This procedure ensures that friction does not cause an uneven bolt loading and that the correct high tension is achieved every time during assembly. Final Elongation The total angular elongation gained by applying the tightening angle is the sum of the elongation resulting from the elastic and plastic angles. The elastic elongation of a screw is due to tightening with the elastic angle. Plastic angle is a function of the plastic spring constant, of a screw. The spring constant of the clamped parts, is assumed to remain constant when tightening from the elastic into the plastic region of a screw.

BOLT JOINT DIAGRAM To help visualize the loading within bolted connections, joint diagrams have been developed. A joint diagram is a means of displaying the load deflection characteristics of the bolt and the material that it clamps. Joint diagrams can be used to assist in visualizing how a bolted joint sustains an external force and why the bolt does not sustain the whole of this force.

Figure 1 The diagram shown above presents the way that the basic joint diagram is constructed. As a nut is rotated on a bolt's screw thread against a joint, the bolt is extended. Because internal forces within the bolt resist this extension, a tension force or bolt preload is generated. The reaction to this force is a clamp force that is the cause of the joint being compressed. The force-extension diagram presented above shows the bolt extension and the joint compression. The slope of the lines represents the stiffness of each part. The clamped joint usually is stiffer than the bolt.

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ENGINE DETAILS Table 1 Bore Diameter (D)

80 mm

Cylinder peak pressure (Ppeak) Young‟s modulus of elasticity of cylinder head (Eh) Stiffness of gasket (kg) Young‟s modulus of elasticity gasket (Eg) Height of Cylinder Head (Hcyl) Cylinder head Gasket Thickness (Tg) Diameter of Head Contact (Dk) Inner Diameter of Boss (Db) Outer Diameter of Boss (Da) Gap between cylinder head and gasket before assembly (δgap) Thermal expansion coefficient of head (δj)

22 N/mm2 1.4 x 105 N/mm2 1.6 x 105 N/mm2 2 x 105 N/mm2 95 mm 1.3 mm 17 mm 12 mm 24 mm 0.20 mm 1.11 x 10-5 m/m/0C

STUD DETAILS Table 2 Parameter M-series (d) Pitch (P) Material class of stud Ratio of residual bolt preload to maximum combustion force Number of studs around the cylinder (n) Young‟s modulus of elasticity of stud (Eb) Yield Strength (S0.2) Friction coefficient of the screw threads (μt) Friction coefficient under the screw head (μh) Pitch diameter of thread (Dp) Mean diameter of stud head (Dm) Embedment (fz) Length of unengaged thread (Lt) Diameter of minimum c/s area (Dmin) Thermal expansion coefficient of stud (δb)

M12 x 1.5 Bolt 12 mm 1.5 mm 12.9 2 6 2.06 x 105 N/mm2 1080 N/mm2 0.14 0.23 11.0258 mm 17.5 mm 0.0127 mm 15 mm 10.1597 mm 1.11 x 10-5 m/m/0C

M14 x 1.5 Bolt 14 mm 1.5 mm 12.9 2 6 2.06 x 105 N/mm2 1080 N/mm2 0.14 0.23 13.0258 mm 20.5 mm 0.0127 mm 15 mm 12.1597 mm 1.11 x 10-5 m/m/0C

CALCULATIONS FOR M14 X 1.5 BOLT, 12.9 CLASS (FOR STANDARD BOLTS USING TORQUEONLY TIGHTENING APPROACH) Stiffness of Cylinder Head (Kh) 

Effective Length (Lk),



Equivalent Area (Aeq),



Stiffness of Cylinder Head (kh),

Study and Analysis of Angular Torquing of Engine Cylinder-Head Bolts Using Torque-To-Yield Bolts: A Case Study

Stiffness of Bolt (kb)

Loss of Preload Due to Gasket Setting Only (Fg)

Loss of Preload Due to Complete Connection (Fs)

Maximum Combustion Force (Fc)

Available Preload after Assembly (F0.2)

Maximum Permissible Tightening Torque (Mper)

Torsional Stress Generated in the Smallest Cross-Sectional Area of the Stud (Sto)

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Tensile Stress Generated in the Smallest Cross-Sectional Area of the Stud (St)

Bolt Force at Yield Strength (F0.2)

Required Preload (Fp,req)

Minimum Torque Required for Given Calculation (Mreq)

The actual torque applied to each bolt on the cylinder head is between 165 to 175 N.

CALCULATIONS FOR M12X1.5 BOLT, 12.9 CLASS (FOR ANGULAR TORQUING USING TORQUE-TO-YIELD BOLTS) Stiffness of Cylinder Head (kh) 

Effective Length (Lk),



Equivalent Area (Aeq),



Stiffness of Cylinder Head (kh),

METHOD

Study and Analysis of Angular Torquing of Engine Cylinder-Head Bolts Using Torque-To-Yield Bolts: A Case Study

Stiffness of Bolt (kb)

Force Required to Yield the Bolt (Fyield) 

Nominal Stress Area of Bolt (As),



Force Required to Yield the Bolt,

Moment Required to Yield the Bolt (Myield)

Snug Force (Fsnug) 

Stiffness of Joint (kj)



Snug Force (Fsnug)

Snug Torque (Msnug)

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Elastic Angle (φe)

Reduced Ultimate Tensile Load (Fb,reduced)

Stiffness of Bolt in Plastic Region (Cb) 

Maximum Elongation of Bolt (δmax) As a general value, the maximum recommended elongation is here set to 5% of the length of the unengaged

threads, i.e. δmax = 0.05Lt.



Stiffness of Bolt in Plastic Region (Cb)

Elongation Due to Elastic Angle (δφ, e)

Plastic Angle (φp)

Total Angle (φ)

Study and Analysis of Angular Torquing of Engine Cylinder-Head Bolts Using Torque-To-Yield Bolts: A Case Study

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Elongation Due to Plastic Angle (δφ, p)

Elastic Part of the Angular Load (Fφ, e)

Plastic Part of the Angular Load (Fφ, p)

Loss of Preload Due to Embedment (Fz)

Final Force (Ffinal)

Final Moment (Mfinal)

Maximum Combustion Force (Fc)

Combustion Force per Bolt

CONCLUSIONS 

A standard M14x1.5 bolt of property class 12.9, tightened below the yield point can be replaced by a Torque-toYield bolt of size M12x1.5 and property class 12.9. It can give a torque of 194.7 Nm, as compared to the torque of 175 Nm given by the M14 bolt.

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

Approximately 16% increase in preload is obtained due to the use of the TTY M12 bolt over the standard M14 bolt.



The maximum elongation of the bolt in the elastic and plastic region is 0.321 mm which is less than the limiting value of 5% of length of unengaged thread; i.e. 0.6 mm. Hence, even after increasing the bolt torque and preload of the TTY bolt, it is still safe against excess deformation.



The combustion force to be borne by each bolt, i.e. 18915 N, is also seen to be in safe limits, as the value of preload achieved by each TTY bolt, i.e. 82790.17 N, is greater than the combustion force per bolt. Hence bolt failure due to this force is avoided. So this method can be safely chosen for engine head sealing.

REFERENCES 1.

“An Introduction to the Design and Behavior of Bolted Joints”, Bickford, John H., Marcel Dekker, Inc., New York (1990).

2.

“The Tightening of Bolts to Yield and Their Performance Under Load”, I. Chapman, J. Newnham and P. Wallace, Journal of Vibration Acoustic Stress Reliability Design, Vol. 108, pp. 213-221 (1986).

3.

“Optimal Bolt Preload for Dynamic Loading”, T. A. Duffey, International Journal of Mechanical Sciences, Vol.35, pp. 257-265 (1993).

4.

“Handbook of Bolts and Bolted joints”, John H. Bickford, Marcel Dekker, Inc., pp 591-620.

5.

“Clamp Load Loss Due To Fastener Elongation Beyond Its Elastic Limit”, Nassar, S.A. and Martin, P., 2006, ASME Journal of Pressure Vessels Technology, Vol. 128, Nov. 2006.

6.

“Engineering Fundamentals of Threaded Fastener Design and Analysis”, Shoberg, Ralph S., PCB Load and Torque Company Publication.

7.

“Fastener Design Manual”, Barrett, Richard T., NASA Reference Manual 1228, March 1990.

8.

“Controlled Tightening Over the Yield Point of a Screw: Based on Taylor‟s Series Expansions”, Toth, Göran A., Journal of Pressure Vessel Technology, November 2003.

9.

“Controlled Tightening Over the Yield Point of a Screw: Based on Monte Carlo Simulations”, Toth, Göran A., Journal of Mechanical Design, July 2004, Vol. 126, pp 729-736.

10. “Torque-Angle Formulation of Threaded Fastener Tightening”, Nassar, Sayed A., Yang, Xianjie, Journal of Mechanical Design, February 2008.