CN111881530A - An optimization design method for vibration reduction of aero-engine - Google Patents

Abstract

The invention discloses an optimization design method for vibration reduction of an aero-engine. The steps are as follows: determining parts of casing accessories that need vibration reduction, and establishing a vibration transmission model; The transfer model is transformed into the frequency domain, and according to the expected performance index, the frequency domain analysis method is used to design the optimal control parameters; through simulation, it is confirmed whether the optimal control parameters designed meet the performance requirements, and if not, the optimization design is carried out again. The present invention can attenuate the vibration of multiple parts of the casing accessories, and can achieve complete vibration isolation of specific parts, which has important practical significance.

Description

An optimization design method for vibration reduction of aero-engine

technical field

The invention belongs to the field of aero-engine vibration reduction and noise reduction, and particularly relates to an aero-engine vibration reduction optimization design method.

Background technique

Aero-engines have strict requirements on reliability and safety, and the vibration of the whole machine is one of the important factors to reduce reliability and safety. Therefore, airworthiness provisions at home and abroad stipulate that the vibration of aero-engines should be monitored and controlled. For example, the Civil Aviation Administration of China specifically defines vibration, that is, "Each engine must be designed and constructed so that the engine operates properly over its declared operating range over its entire flight envelope and over its entire RPM and power or thrust range, without causing Excessive stress on any part of the engine due to vibration should not result in the transmission of excessive vibrational forces to the aircraft structure." However, due to the various sources of engine vibration, it is difficult to control the engine in real time, and only Airworthiness requirements are ensured through good structural design. For example, the assembly error can be reduced by dynamic balancing, so as to reduce the periodic vibration of the engine due to the unbalanced mass of the high and low pressure shafts; or through the parameter optimization design of the damper, the vibration can be attenuated along its transmission path.

In practical engineering, another vibration problem is often encountered, that is, the vibration caused by aerodynamic unbalanced force (aerodynamic interference, rotor misalignment and bearing damage, etc.) The vibration of the authority electronic controller, fuel adjustment mechanism, fuel and lubricating oil pipelines is too large. Since aerodynamic imbalance is unavoidable, the current common practice is to use vibration isolation and shock resistance technology, that is, to add vibration isolators to the installation positions of electronic controllers, fuel adjustment mechanisms, etc. to isolate or attenuate the vibration transmission from the engine. The key point of this design method is to select appropriate vibration isolator parameters. For example, Zhao Kui detailed the design method of vibration isolator parameters in his paper "Design Method and Experimental Research of Vibration Isolator for an Aero-Engine Controller" ; Chinese patent application CN103742591A discloses a design method of an adaptive mass damper for a rotating machinery rotor.

However, the above design method only considers the isolation of a single component, so it only involves the parameter optimization design of a single vibration isolator; when multiple components need to be isolated, due to the large number of parameters to be designed, the current method is more complicated.

SUMMARY OF THE INVENTION

In order to solve the technical problems mentioned in the above-mentioned background art, the present invention proposes an optimization design method for vibration reduction of aero-engine.

In order to realize the above-mentioned technical purpose, the technical scheme of the present invention is:

An optimization design method for vibration reduction of aero-engine, comprising the following steps:

(1) Determine the parts of the casing accessories that need to be damped, and establish a vibration transfer model;

(2) Select one of the components and take its parameters as the target to be optimized for design;

(3) Transform the vibration transfer model established in step (1) into the frequency domain, and use the frequency domain analysis method to design the optimal control parameters according to the desired performance index;

(4) Confirm whether the optimal control parameters designed in step (3) meet the performance requirements through simulation, and if not, return to step (2) and carry out the optimization design again.

Further, in step (1), a system identification method or a parameter matching method is used to establish a vibration transfer model.

Further, for the two vibration damping components, the established vibration transfer model is as follows:

为对应的一阶求导，

为对应的二阶求导；

表示气动不平衡力；u(t)为待优化参量。In the above formula, m _u , ku and _cu are the mass, stiffness coefficient and damping coefficient of component 1, respectively; m _s , _{k s} and c _s are the mass, stiffness coefficient and damping coefficient of component 2; zu ( _t ) is the displacement of component 1, z(t) is the relative displacement of component 1 and component 2,

For the corresponding first-order derivative,

for the corresponding second-order derivative;

represents the aerodynamic unbalanced force; u(t) is the parameter to be optimized.

Further, in step (2), the parameter u(t) to be optimized has the following options:

①u(t)=kz _u (t);

②u(t)=kz(t)

Among them, k is the optimal control parameter to be optimized.

Further, the concrete process of step (3) is as follows:

First, the vibration transfer model is transformed from the time domain to the frequency domain by Fourier transform:

where ω is the frequency;

Z(jω), Z _u (jω), U(jω) are the frequency response functions of z(t), zu (t), and _u (t), respectively;

D(jω)≡ω ² Z _r (jω) is the unbalanced vibration force;

det[G(jω)]=(k _u -m _u ω ² +jc _u ω)(k _s -m _s ω ² +jc _s ω)-(k _s +jc _s ω)m _s ω ² ;

Then, draw two circles on the complex plane, called α-circle and β-circle, where α-circle is a circle with (-1,0) as the center and δ ₁ ≤1 as the radius, β- circle is a circle with (-G(jω)) as the center and δ ₂ ≤1 as the radius, where δ ₁ and δ ₂ are the desired performance indicators;

If α-circle and β-circle have intersection, select the optimal point of the intersection part and denote it as α(jω), then design the optimal control parameter k:

The beneficial effects brought by the above technical solutions:

Aiming at the vibration problem of aero-engine casing accessories, the present invention proposes a design method for local parameter design to achieve global vibration reduction. This method can dampen the vibration of multiple parts of the casing accessories, and can obtain complete Vibration isolation is of great value to practical engineering.

Description of drawings

Fig. 1 is the overall method flow chart of the present invention;

Fig. 2 is the schematic diagram of α-circle and β-circle in the present invention;

Fig. 3 is the schematic diagram of α-circle and β-circle in the embodiment;

4 is a schematic diagram of vibration in the embodiment, wherein (a) is a schematic diagram of the vibration Z (jω) of the position of the ignition adjustment mechanism, and (b) is a schematic diagram of the vibration Z _u (jω) of the position of the electronic controller.

Detailed ways

The technical solutions of the present invention will be described in detail below with reference to the accompanying drawings.

The present invention designs an aero-engine vibration reduction optimization design method, as shown in Figure 1, the steps are as follows:

Step 1: Determine the parts in the casing accessories that need to be damped, and establish a vibration transfer model;

Step 2: Select one of the components and take its parameters as the target to be optimized;

Step 3: Transform the vibration transfer model established in Step 1 into the frequency domain, and use the frequency domain analysis method to design the optimal control parameters according to the desired performance index;

Step 4: Confirm through simulation whether the optimal control parameters designed in Step 3 meet the performance requirements.

In this embodiment, the above-mentioned step 1 can be implemented by the following preferred solutions:

The vibration transfer model is usually obtained by a system identification method or by a parameter matching method. Taking the parameter matching method as an example, for the two damping components, the model can be established as follows:

为对应的一阶求导，

为对应的二阶求导；

表示气动不平衡力——该力即为振源，需要通过优化设计以使其衰减甚至隔离。也就是说，需要通过局部参数优化，使得部件1和部件2处的振动同时得到衰减。Among them, m _u , _ku , _cu are the structural parameters of component 1 (mass, stiffness coefficient and damping coefficient); m _s , ks , c _s are the structural parameters of component 2 _; zu ( _t ) is the displacement of component 1, z(t) is the relative displacement of part 1 and part 2,

For the corresponding first-order derivative,

for the corresponding second-order derivative;

Represents aerodynamic unbalanced force - this force is the source of vibration and needs to be attenuated or even isolated through optimal design. That is to say, it is necessary to optimize the local parameters so that the vibrations at Part 1 and Part 2 are damped at the same time.

At this point, the mathematical model of the optimized system can be obtained:

Among them, u(t) is the parameter to be optimized.

In the following embodiments, component 1 is an electronic controller, and component 2 is a combustion adjustment mechanism, and the values of the variables in the above model are as follows:

m_s(kg) m_u(kg) k_s(N/m) k_u(N/m) c_s(Ns/m) c_u(Ns/m) 973 114 42720 101115 1095 14.6

In the present embodiment, above-mentioned step 2 adopts the following preferred scheme to realize:

For the above mathematical model of vibration transfer, the parameter u(t) to be optimized is usually expressed as a function of structural parameters, and the following forms are available:

①u(t)=kz _u (t);

②u(t)=kz(t).

Which form to choose is determined by the actual situation (such as space, weight, etc.); but for both, k is the optimal control parameter that needs to be optimized. That is, a suitable unknown parameter k needs to be designed so that the vibrational responses of z _u (t) and z (t) to z _r (t) must be simultaneously reduced. In the following embodiment, 2 u(t)=kz(t) is selected as the parameter to be optimized, that is, the vibration at both the combustion adjustment mechanism and the electronic controller is attenuated by optimizing the design of the parameters at the combustion adjustment mechanism.

In the present embodiment, above-mentioned step 3 adopts the following preferred scheme to realize:

This step is the core of the present invention, and the specific process is as follows:

First, the vibration transfer model is transformed from the time domain to the frequency domain through Fourier transform:

Among them, Z(jω), Z _u (jω), U(jω) are the frequency response functions of z(t), zu (t), _u (t), respectively; D(jω)≡ω ² Z _r (jω ) is the unbalanced vibration force (Z _r (jω) is the frequency response function of z _r (t), ω is the frequency); the common denominator det(G) is defined as:

det(G)=(k _u -m _u ω ² +jc _u ω)(k _s -m _s ω ² +jc _s ω)-(k _s +jc _s ω)m _s ω ² .

For the above embodiment, the parameters listed in the table in step 1 are put into the above formula, that is, the required frequency domain model is obtained. The corresponding frequency is usually the natural natural frequency at the fuel adjustment mechanism

Second, define the variable G as follows:

In this example, it can be calculated that:

G(jω)=0.3844+0.1311j

Again, draw two circles on the complex plane, called α-circle and β-circle, where α-circle is a circle with (-1,0) as the center and δ ₁ ≤ 1 as the radius; and β -circle is a circle with (-G(jω)) as the center and δ ₂ ≤1 as the radius - as shown in Figure 2. δ ₁ and δ ₂ are expected performance indicators, that is, the expected attenuation of Z(jω) and Z _u (jω) through optimal design. If Z(jω) is required to be reduced by 3dB, then δ ₁ =0.707; _u (jω) is reduced by 6dB, then δ ₂ =0.5.

In this embodiment, δ ₁ =0.65 and δ ₂ =0.5 are specified as the desired performance indicators, that is, it is expected that the vibration Z(jω) at the fuel adjustment mechanism can be reduced by 4dB (0.65) through the optimal design and at the electronic controller. The vibration Z _u (jω) is attenuated by 6dB(0.5). Thus, a schematic diagram of the α-circle and the β-circle in the embodiment is drawn, as shown in FIG. 3 .

Again, for this embodiment, in Figure 3, α-circle and β-circle have intersection (shaded), so the vibration Z(jω) at the fuel adjustment mechanism and the vibration Z _u (jω) at the electronic controller can be made ) at the same time decay; that is, the optimal design is the intersection of the two circles.

Finally, select the optimal point of the intersection part, denoted as α(jω), then the optimal control parameters to be designed can be obtained by the following formula:

In this embodiment, since the center of the β-circle is located in the intersection area of the two circles, α(jω)=-G(jω) can be selected, that is, α(jω)=-0.3844-0.1311j. At this time, k=-21360-7255.6j is calculated according to the above formula, and it can be seen that this design reduces the vibration Z(jω) at the fuel adjustment mechanism by 4dB, while the vibration at the electronic controller Z _u (jω) Attenuates to Zero, that is, the electronic controller achieves complete vibration isolation.

Carry out feasibility and performance limit analysis according to the method of the present invention, and confirm that the performance index reaches the expected value. For the above embodiment, if the index calls for a 4dB reduction in the vibration Z(jω) at the fuel adjustment mechanism and a 20dB reduction in the vibration Z _u (jω) at the electronic controller, then the design can obviously be confirmed; and if the index calls for the fuel adjustment mechanism The vibration Z(jω) at and the vibration Z _u (jω) at the electronic controller are both reduced by 20dB. After redrawing α-circle and β-circle, because there is no intersection, it shows that the index requirements are too high at this time, and the performance must be reduced indicator requirements. The design is now confirmed by reducing the vibration Z(jω) at the fuel adjustment mechanism by 4dB and the vibration Z _u (jω) at the electronic controller by at least 20dB. The real-time simulation results are shown in Figure 4. It can be seen that, using the method proposed in the present invention, the vibration Z (jω) at the fuel adjustment mechanism is indeed reduced by 4dB, and at the same time, the vibration Z _u (jω) at the electronic controller is completely attenuated, that is, the electronic controller is completely attenuated. Complete vibration isolation. The design not only meets the requirements, but also has complete vibration isolation performance, which is the most desired design in practical engineering.

The embodiment is only to illustrate the technical idea of the present invention, and cannot limit the protection scope of the present invention. Any changes made on the basis of the technical solution according to the technical idea proposed by the present invention all fall within the protection scope of the present invention. .

Claims (5)

1. an aero-engine vibration reduction optimization design method, is characterized in that, comprises the following steps: (1) Determine the parts of the casing accessories that need to be damped, and establish a vibration transfer model; (2) Select one of the components and take its parameters as the target to be optimized for design; (3) Transform the vibration transfer model established in step (1) into the frequency domain, and use the frequency domain analysis method to design the optimal control parameters according to the desired performance index; (4) Confirm whether the optimal control parameters designed in step (3) meet the performance requirements through simulation, and if not, return to step (2) and carry out the optimization design again. 2 . The aero-engine vibration reduction optimization design method according to claim 1 , wherein, in step (1), a system identification method or a parameter matching method is used to establish a vibration transfer model. 3 . 3. according to the described aero-engine vibration damping optimization design method of claim 2, it is characterized in that, adopt parameter matching method, for two damping parts, the vibration transfer model of establishment is as follows:

In the above formula, m _u , ku and _cu are the mass, stiffness coefficient and damping coefficient of component 1, respectively; m _s , _{k s} and c _s are the mass, stiffness coefficient and damping coefficient of component 2; zu ( _t ) is the displacement of component 1, z(t) is the relative displacement of component 1 and component 2,

For the corresponding first-order derivative,

for the corresponding second-order derivative;

represents the aerodynamic unbalanced force; u(t) is the parameter to be optimized. 4. according to the described aero-engine vibration reduction optimization design method of claim 3, it is characterized in that, in step (2), described parameter u (t) to be optimized has following form selection: ①u(t)=kz _u (t); ②u(t)=kz(t) Among them, k is the optimal control parameter to be optimized. 5. according to the described aero-engine vibration reduction optimization design method of claim 3, it is characterized in that, the concrete process of step (3) is as follows: First, the vibration transfer model is transformed from the time domain to the frequency domain by Fourier transform:

where ω is the frequency; Z(jω), Z _u (jω), U(jω) are the frequency response functions of z(t), zu (t), _u (t) respectively; D(jω)≡ω ² Z _r (jω) is Unbalanced vibration force; det[G(jω)]=(k _u -m _u ω ² +jc _u ω)(k _s -m _s ω ² +jc _s ω)-(k _s +jc _s ω)m _s ω ² ;

Then, draw two circles on the complex plane, called α-circle and β-circle, where α-circle is a circle with (-1,0) as the center and δ ₁ ≤1 as the radius, β- circle is a circle with (-G(jω)) as the center and δ ₂ ≤1 as the radius, where δ ₁ and δ ₂ are the desired performance indicators; If α-circle and β-circle have intersection, select the optimal point of the intersection part and denote it as α(jω), then design the optimal control parameter k:

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